lecAyr.Jr* 7\>r&7 
 
A MANUAL 
 
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 MARINE ENGINEERING 
 
STANDARD WORKS FOR ENGINEERS. 
 
 Iim.TEENTH EDITION, Thoroughly Revised. In Pocket Size. Leather. 
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A M ANTIAL 
 
 OF 
 
 MARINE ENGINEERING: 
 
 COMPRISING 
 
 THE DESIGN, CONSTRUCTION, AND WORKING OF 
 MARINE MACHINERY. 
 
 BY 
 
 A. E. SEATOX, 
 
 70RMERLY LECTURER ON MARINE ENGINEERING TO THE RoYAL NAVAL COLLEGE, GREENWICH; 
 
 MEMBER OF THE INSTITUTION OF CIVIL ENGINEERS J MEMBER OF THE INSTITUTION 
 
 OF NAVAL ARCHITECTS; MEMBER OF THE INSTITUTION OF MECHANICAL 
 
 ENGINEERS: MEMBER OF THE NORTH-EAST COAST INSTITUTION 
 
 OF SHIPBUILDERS AND ENGINEERS ; MEMBER OF 
 
 THE INSTITUTE OF METALS, ETC. 
 
 TUlitb numerous Cables ano illustrations rccuceb from "CdorKfiig 
 
 IDrawings ano pbotoiuapbs. 
 
 EIGHTEENTH EDITION, THOROUGHLY REVISED, GREATLY ENLARGED, 
 AND MOSTLY RE- WRIT J EN TO DATE. 
 
 ■ 
 
 LONDON: 
 CHARLES GRIFFIN & COMPANY, LIMITED. 
 
 NEW ?ORK: D. VAN NOSTRAND CO. 
 
 19 18. 
 
 [All Sights Reserved.] 
 

 Engineering 
 Library 
 
 /l/7<zc/i* . Z^e/o/. 
 
 . 
 
 • . ■ ■ ■ . • .-• 
 
PREFACE TO THE EIGHTEENTH EDITION. 
 
 The demand for this new Edition came when the War, with all 
 its pomps and circumstances, has made it difficult to produce- — 
 both to Author and Publishers. Since the issue of the last Edition 
 the changes in engineering practice have been many and great 
 In some measure this is due to the special demands arising out 
 of the "War conditions, but it is largely due to the advance that 
 goes on now, day by day, from the better knowledge of science 
 gained by diligent research, and by the better application of it, 
 whereby that experience is gained which engenders confidence as 
 well as stimulates invention, and this produces improvements of 
 many kinds. 
 
 The economic side of engineering, however, is asserting itself 
 to a degree that never obtained in pre- War times, and it will 
 surely remain as a predominating factor in all our every-day 
 calculations for many years to come, so that we are compelled 
 now to approach and to determine problems on lines not con- 
 templated formerly. To be successful it will be necessary to 
 cast aside all prejudices, to treat lightly the precedents, and to 
 concentrate the solving of them — each on its own merits — by 
 giving full heed to the physical and economic conditions only. 
 
 Necessity has ever been the mother of invention. To-day it 
 will be likewise the remover of prejudice as well as the alma 
 
 41 15 
 
VI PREFACE. 
 
 mater of research to all her children, so that they may thrive 
 in a way they never have done hitherto in this country. 
 
 D. O. R A. exercises a powerful influence over authors and 
 publishers, whereby they are restrained from making public any 
 of the wonderful advances achieved during the past four years, 
 or to allude to the inventions whereby so much has been accom- 
 plished on the sea by the genius that otherwise might have 
 remained dormant. Nevertheless this Edition does contain much 
 that is new, and what was old has been renovated and brought 
 up to date. 
 
 Special Appendices have been added which deal with the 
 Heavy Oil Engine, Geared Turbines and Superheaters, as now in 
 general use, and altogether the attempt is made to maintain the 
 character of the Manual as far as circumstances will permit. 
 
 A. E. SEATON. 
 
 Westminster, September, 1918. 
 
ORIGINAL PKEFACE. 
 
 The following Work has been prepared to supply the existing want of a 
 Manual showing the application of Theoretical Principles to the Design 
 and Construction of Marine Machinery, as determined by the experience 
 of leading engineers, and carried out in the most recent successful practice. 
 The data on which it is based, now first thrown into form for publication, 
 have been collected during many years of study and practical work. It 
 is hoped that the volume will be found useful by the engineer and draughts- 
 man engaged in practice as a Handbook of Reference, and by the student, 
 launched for the first time on the intricacies of Marine Construction, as 
 a guide, supplying to some extent bis lack of experience. 
 
 The rules and formulas introduced (which have been divested as far as 
 possible of complexity, and given in the simplest form attainable) may be 
 used by any one who designs with some regard to theory, and, by varying 
 the constants, be made to suit his own ideas of strength and stiffness. It 
 may, perhaps, be thought by some that in certain instances details have been 
 entered into with unnecessary minuteness ; but it should be remembered, 
 on the other hand, that not every engineer has the contents of a well-filled 
 drawing-office to fall back upon in cases of doubt and difficulty. 
 
 It is hardly necessary to premise that it is wholly impossible to reconcile 
 the practice of the naval designer, who thinks more of efficiency and weight 
 than of cost, with that of the mercantile engineer, who studies efficiency 
 and cost with but small regard to weight, and, therefore, few rules can be 
 driven which shall absolutely suit both. However, the manufacturer of 
 machinery for the Merchant Service might follow with advantage much that 
 has been proved to be good in naval practice, and the Naval Authorities 
 might again, on their part, borrow from the Mercantile Marine a few 
 suggestions which would render a warship, while no less efficient than at 
 present, perhaps somewhat less intricate for those who have to work her. 
 
 In conclusion, the author can but express a hope that the publication of 
 these notes, imperfect as they necessarily are, may tend to make a little 
 clearer some of the technicalities of Marine Design and Construction, and 
 so help forward, in however slight a degree, the application of scientittc 
 investigation to those problems which the marine engineer is called upon, 
 day by day, to solve. 
 
 Hull, January With, 1883. 
 
GENERAL CONTENT S 
 
 CHAPTER I. 
 
 General Introduction, . . . . pp. 1-18 
 
 Fundamental Principles — Paddle-Wheels — The Screw — Hydromotors — Motive Power 
 —Steam used Expansively — Early Marino Engines — Propellers — Multiple Screws. 
 
 CHAPTER II. 
 
 Resistance of Ships and Indicated Horse-Power Necessary for Speed, pp. 19-55 
 
 Value of Trial Trips — Resistance of a Ship — Chief Cause of Resistance — Residual 
 Resistance — Coefficient of Fineness — Skin Resistance — Wetted Skin of a Ship — Kirk's 
 Analysis — Mumford's Calculation of Wetted Skin — Seaton's Modification — Form for 
 Required Speed — Seaton's Rule for Limitation of Speed — Power necessary — Tank 
 Experiments — Rankine's Rule — Kirk's Analysis — Progressive Trials — Curve of Revolu- 
 tions — Curve of Slip — Sea Performances of Steamers — Progressive Trials for Effect of 
 Depth of Water — Taylor's Formula for Depths of Water on Trials. 
 
 CHAPTER III. 
 
 Marine Engines : Their Types and Variations of Design, . pp. 56-108 
 
 Various Types — Paddle-wheel Engines — Beam Engines — Side Lever Engines — 
 Oscillating Engines — Vertical Direct-acting Engines — Twin-cylinder Engines — Diagonal 
 Paddle Engines — Screw Engines — Horizontal — Trunk Engines — Return Connecting-rod 
 Engines — Vertical Direct-acting Engines — Comparison of Present and Past Practice — 
 Arrangement of Cylinders — Compound Engine — Three-, Four-, and Six-cylinder Com- 
 pounds — Single and Two-crank Engines — Three-, Four-, Five-, Six-, and Eight-crank 
 Engines — Oil Engines — Diesel — Semi-Diesel — Reversal of Oil Engines — Of Propeller — ■ 
 Turbine Machinery — Experiments with Turbines — Combination of Turbines with Reci- 
 procators — Geared Turbines — Design of Oil Engines — Double-acting Diesel — Two-cycle 
 Diesel — Fuel Consumption of the latter — Gas Engines. 
 
 CHAPTER IV. 
 
 Steam used Expansively, . . . pp. 109-135 
 
 Reciprocating Engines — Turbines — Expansion in Stages — Moist Steam — Adiabatic 
 Expansion — Clearance — Effect of Clearance and Cushioning — Mean Pressure in a Com- 
 pound Engine — Actual Mean Pressure in Practice — Frictional Resistance in Stop Valves 
 and Pipes — Wire-drawing — Liquefaction during Expansion — Timing of h'xhaust — Lead 
 — Calculation of Expected Mean Pressure — Graphic Method. 
 
 CHAPTER V. 
 
 Steam used after Expansion— Turbines, . . pp. 136-151 
 
 Modern Reaction Turbine — Modern Impulse Turbine — Combination Turbines- 
 Shape of Passages — Compound Turbines — Design of Screw for Turbine Ships — Efficiency 
 
X CONTENTS. 
 
 of Turbine — Power developed by Turbine — Rateau's Formula for Steam Consumption — 
 Sea Experiences with Turbines— Torsion Meters — Amsler's Meter — Fottinger's Meter — 
 Hopkinson-Thring Meter — Bevis-Gibson Meter — Collie's Meter — Denny-Johnson Meter — 
 Shaft Horse-power — Steam Engine Indicator. 
 
 CHAPTER VI. - 
 
 Efficiency of Marine Engines, . . pp. 152-173 
 
 Water consumed per Shaft Horse-power — Mechanical Efficiency — Vertical Engines — 
 Efficiency of Turbines and Reciprocators — Results of Trials with Triple-compound 
 Engine — Yarrow's Experiments with Torpedo Boat — Steam Efficiency — High Pressure : 
 Its Advantages and Disadvantages — Efficiency of the Engine as a Machine — Friction of 
 Piston — Of Stuffing-boxes, Guides, and Slides — Friction at Shaft Journals — Of Valve 
 Motions — Loss from Pumps — Inertia of Moving Parts — Losses due to Mechanical Defects 
 and Physical Causes — Experiments with Superheated Steam — Resistance of Pipes and 
 Passages. 
 
 CHAPTER VII, 
 
 Engines — Simple and Compound, . . -pp. 174-194 
 
 Elementary Steam Engine — Genesis of Compound Engine — Expansive Engine — 
 Effects of Increase of Pressure — Progress made by Early Engineers — Receiver Compound 
 Engine — Expansive and Compound Engines Compared — Division of Work — Direct- 
 expansion Compound Engine — Requisites in the Marine Engine — Comparative Theoretical 
 Efficiency of Various Marine Engines — Further Comparison of Efficiency of Engines — 
 Experiments in the Mercantile Marine — Triple-expansion Compound Engine — Increased 
 Pressure of Steam — Compound System of Cylinders. 
 
 CHAPTER VIIL 
 
 Horse-power — Nominal, Indicated, and Shaft or Brake. . pp. 195-209 
 
 Nominal Horse-power — Lloyd's N.H.P. — Estimated Horse-power — Indicated Horse- 
 power — Indicator Diagram — Mean Pressure — -Shaft Horse-power— -Thrust Horse- power 
 — Brake Horse-power — Indicated Thrust — Tow-rope Horse-power- -Net Horse-power — 
 Piston Speed and Revolutions — Rate of Revolution of Marino Engines — Revolutions — 
 Length of Stroke. 
 
 CHAPTER IX. 
 
 General Design and the Influences which effect It, . . pp. 210-227 
 
 General Design and Arrangements — The Condenser — Inclination — Effect of External 
 Causes— Supply of Materials — Influence of Tonnage Laws — Of the Board of Trade — 
 Balancing and Avoidance of Vibration — Pitch of Screw and Vibration — Suction, Cavi- 
 tation, etc. — Steel Castings — Aluminium — Duralumin — Future of the Reciprocator. 
 
 CHAPTER X. 
 
 The Cylinder and its Fittings, . . -pp. 228-261 
 
 Size — Rates of Expansion — Diameter of Cylinders — Back Pressure in L.P. Cylinder — 
 Intermediate M.P. Cylinder — Arrangement of Cylinder — Size of Steam Ports — Triple and 
 Quadruple Engines — Ratio of Cylinders — Drop Valves — Main Steam Pipe — Area through 
 Valves— Steam Ports and Passages — Opening of Port to Steam — Exhaust Passages — 
 Cylinder Liner — Width of Steam Ports — Piston Valves — "Drop" Valves — Double- 
 
CONTENTS. xi 
 
 ported Valves — Steam Jackets — Boring Holes — Auxiliary Valves — Escape or Relit i 
 Valves — Drain Cocks — Receiver Space — Column Pacings and Feet -Holding-down Bolts 
 — Horizontal Cylinders — Oscillating Cylinders — Cylinder Covers — Cylinder Cover Studs 
 and Bolts — Cylinder Flanges — Clearance of Piston — Valve- box Covers — small Doors 
 and Covers — Lagging and Clothing of Cylinders— L. P. Cylinder Body — Stufling-boxcs and 
 Glands. 
 
 CHAPTER XI. 
 
 The Piston — Piston-rod — Connecting-rod, . . pp. 262-285 
 
 Piston — Ramsbottom's Rings — Common Piston Rings — Piston Springs — Cameron's 
 I 'a tent — Mather & Piatt's Patent — Buckley's Patent — Qualter & Hall's Patent — 
 Rowan's — Restrained Packings — Body of Piston — Pistons of Ordinary Marine Engines — 
 Details of Construction — Junk-ring Bolts— Safety Rings and Lock Bolts — Solid Packings 
 — Diameter of Piston-rod — Piston-rod Ends — Cast-steel Piston-rod Crosshead — Piston-rod 
 Guides — Surface 01 Guide Block — Crossheads and Gudgeons — Connecting-rods — Con- 
 necting-rod Bolts and Brasses — Caps of Connecting-rod Brasses — Gudgeon End of Rod. 
 
 CHAPTER XII. 
 
 Shafting, Cranks and Crank-shafts, etc., . pp. 286-333 
 
 Shafting of Modern Engine — Crank-shaft of Oil Engine — Tunnel or Intermediate 
 Shafting — Alternating Stresses — Effect of Stresses on Materials — Safe Working Stress — 
 Twisting Moment — Resistance to Twisting — Torsional Stiffness of a Shaft — Bending 
 Moment — Equivalent Twisting Moment — Crank-shafts — Curve of Twisting Moments — 
 Momentum of Moving Parts — Overhung Crank — Paddle Shafts — Crank-shaft of Screw 
 Engine — Built Crank-shaft — Couplings — Surface of Crank-pins and Shaft Journals — 
 Drivers — Taper Bolts — Cross Keys — Propeller Shafts — Outer Bearing — Screw-shaft End 
 — Stern Bush — Stern Tube — Thrust Shaft — Thrust — Friction Loss on Thrust Block — 
 Revolutions per Square Inch on Thrust Block — Diameter of Thrust Collars — Length of 
 the Bearings of Tunnel Shafting — Diameter of Shafts (Rules for Lloyd's, B.O.T., Bureau 
 Veritas) — Shafts for Paddle Steamers — Turbine Shafting — Shafts of Oil Engine — Steel 
 Shafts— Diameter of Propeller Shaft — Details of Crank-shafts — Hollow Shafts — Shafts 
 for Oil Engines. 
 
 CHAPTER XIII. 
 
 Foundations, Bed-plates, Columns, Guides, and Framing, pp. 334-343 
 
 Bed-plates and Foundations — Main Bearings — Caps or Keeps for Main Bearings — 
 Main-bearing Bolts — Brasses — Columns — Guide-plates — Framing — Entablature of Oscil- 
 lating and Steeple Engines. 
 
 CHAPTER XIV. 
 
 The Condenser, . . pp. 344-372 
 
 The Common or Jet Condenser — The Amount of Injection Water — The Area of 
 Injection Orifice — Snifting Valve — Surface Condenser — Condenser Tubes — Condenser 
 Efficiency — The Effect on Economy of Consumption — Surface Condenser Efficiency — 
 Effect of Air mixed with Steam — Air Leaks to Condenser — Flow of Vapour and Water — 
 Cooling Surface — Allowances of Cooling Surface — Cooling Surface per Horse-power — 
 Greatest Length of Tube — Condenser Tubes — Tube-plates — Tube Packings — Steam Side 
 and Water Side of Tubes — Spacing of Tubes — Body of the Condenser — Construction of 
 the Surface Condenser — Shape of the Modern Condenser — Stiffness of Flat Surfaces — 
 Quantity of Cooling Water — Passage of Circulating Water — Size of Circulating Pump 
 Pipes — Size of Inlet and Discharge Pipes — Extra Supply Cock — Manholes — Drain Cocks 
 — Testing — Cementing — Evaporators. 
 
x ii CONTENTS. 
 
 CHAPTER XV. 
 
 Pumps pp. 373-411 
 
 Air-pump — Single-acting Vertical — Edwards' Air-pump — Double-acting Air-pump — 
 Weir's Dual Pump — Air-pump Indicator Diagrams — Efficiency of Air-pumps — Air-pumps 
 with and without Foot Valves — Size of Air-pump — Capacity — Air-pump for Jet-condenser 
 — For Surface Condenser — Rotary Air-pumps — Parsons' Vacuum-augmentor — Pump 
 Rods — Pump Buckets — Valves — Coe & Kinghorn's Patent — Thompson's — Beldam's — 
 Area through Valve Seats — Suction Pipes from the Condenser to the Air-pump — Circu- 
 lating Pumps — Single-acting Pump — Double-acting Pump — Size of Circulating Pump — 
 Circulating Pump Rods, Buckets, and Valves — Valve Area — Diameter of Suction Pipes — 
 Rotary Pumps — Centrifugal Pumps — Pumps of R.M.S. " Mauretania " — Feed Pumps — 
 Sea Water — Net Feed Water — Relief Valves — Valves and Valve-boxes — Air Vessels — 
 Feed Pipes — " Pet Valves " — Feed Tank — Feed-pump Rod — Bilge Pumps — Capacity of 
 Bilge Pumps — Directing Boxes — Mud Boxes — Sanitary Pump — Ejectors — Power for 
 Circulating Pumps. 
 
 CHAPTER XVL 
 
 Valves and Valve Gear, . • . pp. 412-443 
 
 Seaward's Valves — Locomotive, Trick, Double-ported and Treble-ported Valves — 
 Travel of Flat Valves — Piston Valves — Relief Frames — Double Valves — Dawe & Holt's 
 Patent — Common Relief Frame — Martin and Andrews' Valve — Through Exhaust Valve 
 Back Guides and Springs — Balance Pistons— Joy's Assistant Cylinder — Valve Rods or 
 Spindles — Valve-rod Bolts — Valve-rod Guides — Proportions of Slide Valves — Double- 
 ported Valve — Link Motion — Slot Link — Position of Suspension Pin — Size of Slot Link — 
 Single-bar Link — Double-bar Links — Size of Bar Links — Single Eccentric Gear — Hack- 
 worth's Dynamic Valve Gear — Marshall's Valve Gear — Joy's Valve Gear — Sell's Valve 
 Gear — Expansion Valves — Grid-iron Expansion Valves — Outside Cut-off Valves — Inside 
 Cut-off Valves — Piston Expansion Valves— Expansion Valves for Compound Engines — 
 Eccentrics — Eccentric Straps — Eccentric Rods — Reversing Gear — Diameter of Weigh- 
 shaft. 
 
 CHAPTER XVII. 
 Valve Diagrams, * . • . pp. 444-456 
 
 Motion of the Piston — Lead — Inside Lap— Effect of " Notching-up " — Expansion 
 Valve on an Independent Face, Central Position Ports closed — Ditto. Open — Expansion 
 Valve working on the Back of Main Valve, Cutting-off at Inside Edge, Variation in Cut-off 
 by Varying Travel — Expansion Valve on Back of Main Valve, Cutting-off at Outside Edge 
 — Construction of Valve Diagrams — Effect on Indicator Diagrams of Crank Sequence. 
 
 CHAPTER XVIII. 
 
 Propellers, . . . . .pp. 457-501 
 
 Fundamental Principle — -Stream Water — Effects of Wake Currents on Screw — The 
 Paddle Wheel — Nominal or Apparent Slip — Real Slip — Path of Blade Tips — Dimensions 
 and Form of Propeller — Paddle with Fixeil Floats — Feathering Floats — Control of Floats 
 — Diameter of Feathering Wheel — Area of Float — Number and Proportions of Floats — 
 Immersion of Float — Bosses — Paddle Arms — Position of Gudgeons — Outer Bearing — 
 The Screw Propeller — Surface — Smith's Original Screw — Developed and Projected 
 Surfaces — Diameter of Screw — Indicated Thrust — Rankinc's Rule for Thrust — Pressure 
 per Square Inch on Propeller Blade — Diameter of Screw suited to Ship — Limit to Diameter 
 — Pitch of Screw — Pitch Ratio — Surface Ratio — Acting Surface — Thrust — Thickness of 
 
CONTENTS. Xlll 
 
 Blade — Propeller Boss — Blades— Number of Blades — Shape of Blades — Section of Blades 
 — Studs or Bolts — Diameter of Bolts — Material for Blades — Weight — Feathering Screws 
 — Bevis' Patent — Lifting Screws. 
 
 CHAPTER XIX. 
 Sea Cocks and Valves, . . .pp. 502-510 
 
 Kingston Valve — Discharge Valves — Bilge Valves — Communication Boxes — Bilge 
 Suction Piping — Water Service — Expansion Joints — Safety Collars — Flanges. 
 
 CHAPTER XX. 
 
 Auxiliary Machinery, .... pp. 511-537 
 
 Classes of Auxiliary Machinery — Exhaust Steam — Intermittent and Constant Working 
 — Electric Light Engines — Use of Electric Current for all Auxiliary Work — Loss of Steam 
 — Feed Pumps — Engine - room Pumps — Reversing Gears — Ventilating and Forced 
 Draught Fans — Hvdraulic Pumps and Accumulators — Air Compressor — Brown's Steam 
 Tiller — Refrigerating Machines — Auxiliary Condenser. 
 
 CHAPTER XXI. 
 
 Boilers, Fuel, etc., Evaporation, . . .pp. 538-556 
 
 Efficiency of Furnace — Chimney Draught — Fuel — Patent Fuels — Oil Fuel — Oil 
 Burners — Value of a Fuel — Rate of Combustion — Artificial Draught — Howden's System 
 of Forced Draught — Ellis and Eaves' System — Quantity of Fuel burnt on Grate — Size 
 of Funnel — Combustion Heating Surface — Size and Height of Funnel — Evaporation — 
 Tube Surface — Evaporative Power — Efficiency of Boiler. 
 
 CHAPTER XXII. 
 
 Boilers— Tank Boiler Design and Details, . , pp. 557-585 
 
 Forms of Tank Boiler — Cylindrical Boilers with Two Furnaces — With Three Furnaces- 
 ■ — With Four Furnaces — Naval Boilers of To-day — Small Boilers — Double-ended Boiler — 
 Oval Boilers — Holt's Boiler — Dry Combustion Chamber Boiler — Gunboat Boilers - 
 Vertical Cylindrical Boiler — Locomotive Boiler — Wet Bottom Locomotive Boilers — 
 Double-ended Locomotive Boilers — Seaton's Locomotive Boiler — Dimensions of a Boiler 
 — Area of Fire Grate — Consumption of Fuel — Heating Surface — Tube Surface — Total 
 Heating Surface — Efficiency of Heating Surface — Efficiency of Boilers — Area through. 
 Tubes — Capacity of Boiler Shell — Steam Space — Area of Uptake or Funnel Sections — 
 Capacity of Funnels — Examples of Boiler's Details. 
 
 CHAPTER XXIII. 
 
 Water-tube Boilers, . . . . pp. 586-633- 
 
 Development of Grate Area — Tubes — Circulating Tubes — Stays — Steam Drums— The 
 Two Types — Herreshoff Boiler — Babcock & Wilcox — Normand — Ferguson & Fleming 
 — Seaton's Boiler — Yarrow — Blechynden-Wnite-Forster — Thornycroft — Mumford — Du 
 Temple — Reed — Belleville — Diirr — Niclausse — Stirling — Hohenstein— Miyabara— Thorny- 
 croft-Marshall — Amount of Water in Water-tube Boiler — Feed Arrangements— Fire 
 Places — Consumption Trials. 
 
X1V CONTENTS. 
 
 CHAPTER XXIV. 
 
 Boilers— Construction and Detail. . . pp. 634-668 
 
 Testing by Hydraulic Pressure — Admiralty Rules, Bureau Veritas, German Govern- 
 ment Rules for Testing — Boiler Shell, Cylindrical — Riveting — General Rules for Riveted 
 Joints — Treble-riveting — Butt Joints with Double Butt Straps — Treble-riveted Butt 
 Joint — Thickness and Breadth of Butt Straps — Treble-riveted Lap Joints — Double Butt 
 Straps and Double-riveting — Butt Joints with Double Straps and Single-riveted — Do. 
 Double-riveted — Butt Joints with Double Straps, Treble-riveted — Treble-riveted (zig- 
 zag) Butt Joint — Do. Quadruple-riveted — Welded Joints — Circumferential Seams — 
 Methods of Work— Material — Allowance for Wear — Boiler Ends — Riveting — Quality of 
 Plate — Thickness of End Plates — Furnaces — Rule for Thickness of a Plain Furnace — 
 Methods of Stiffening Furnaces — Methods of Connecting Furnaces to End Plates — Com- 
 bustion Chambers — Tubes — Stay Tubes — Stays — Continental and British Practice — 
 Water Spaces — Manholes — Weight of Boilers. 
 
 CHAPTER XXV. 
 
 Boiler Mountings and Fittings, . , pp. 669-690 
 
 Smoke-box — Funnel — Furnace Fronts and Doors — Fire Bars — Martin's Patent Bars 
 — Henderson's Patent Door and Bars — Crude Petroleum Residuals — Stop Valve — 
 Safety Valve — Size of Safety Valve — Internal Pipes — Feed Valves — Automatic Feed 
 Valves — Blow-off Cock — Scum Cock — Water Gauge — Steam Gauge — Sentinel Valve — 
 Air Valve — Weir's Hydrokineter — Steam Whistles — Separator — Boiler Clothing — 
 Cameron's Patent Lagging. 
 
 CHAPTER XXVI, 
 
 Fitting in ol Machinery, Starting and Reversing of Engines, etc., pp. 691-717 
 
 Fitting Machinery into a Ship — Boring the Stern Post — Engine Seatings — Thrust 
 Block Seating — Pedestals for Tunnel Shafting — Boiler Seatings or Bearers — Fitting 
 Machinery on Board the Ship — Holding-down Bolts — Staying Engines — Ramming Chocks 
 and Stays — Boiler Seats — Copper Pipes — Starting and Reversing Engines — Brown's 
 Patent Reversing Gear — Steam Gear for Reversing of Propellers driven by Turbines — 
 Blenkinsop's Arrangement for Manoeuvring — Ships with Four Screws driven by Turbines 
 — Regulation and Stop Valves — Steam Turning Gear — Ash Hoists — Governors — Dunlop's 
 Governor — Steam Governor — Durham & Churchill's Velometer — Coutt's and Adamson's 
 Governor — Westinghouse Governor — Gauges — Lubricators and Impermeators — Cadman's 
 Patent Lubricators — Centrifugal Lubricators — Sight Feed Lubricator — Mechanical 
 Impermeator — Drain Pipes — Jacket Drains — Feed Heaters — Weir's Feed Heater — 
 Evaporators — Ladders — Gratings and Platforms — Feed Filters — Stokehole Ventilators. 
 
 CHAPTER XXVII. 
 
 Weight and other Particulars of Machinery relating thereto, . pp. 718-744 
 
 Weight of Engines and Boilers — Warships— Torpedo Boat and Destroyer — Battle- 
 ships and Cruisers — First, Second, and Third Class Cruisers — Piston Speeds — Boiler and 
 Mean Pressure — Materials — Design — Effect of Loads and Stresses — Parts subject to 
 Intermittent Stresses — To Alternating Stresses— Shearing and Torsion — Elastic Limit- 
 Safe Working Stresses — Stretch under Tension — Paddle Engines — Naval, Express, and 
 Cargo Ship Engines— High-speed Craft — Boilers — Cost — Auxiliaries and Appurtenances 
 — Spare Gear — Fuel — Weight of Modern Machinery — Relation of Weight to Tonnage — 
 X. Atlantic Service — Steamers on Routes to India, Cape, and Australia — Battleships and 
 Cruisers — Scouts and Destroyers — Cross-Channel Steamers — Shallow Water Steamers — 
 Paddle-wheel Steamers — Turbine Steamers. 
 
COXTK.NTR. 
 
 XV 
 
 CHAPTER XXVIII. 
 
 Effect of Weight— Inertia and Momentum— Balancing the Same, . pp. 715-801 
 
 Fixed Parts — Moving Parts — Momentum — Balancing : — Preliminary Definitions — 
 Harmonic Motion — Inertia of Parts of an Engine — Of Connecting-rod — Of Valve Gear — 
 Of Air Pomp — Single-crank and Valve Gear balanced by Rotating Weights only — Do. 
 by Bob Weights and Rotating Weights combined — To Balance an Engine of any Number 
 of Cranks — Curves of Free Force and Couples for a Multi-crank Engine — Four-crank 
 Engine — Yarrow-Schlick-Tweedy System — Modified Principle of Balancing — Effect of 
 Inertia ami Weight — Stresses due to Inertia — Design of Balance Weights. 
 
 CHAPTER XXIX. 
 
 Materials used by the Marine Engineer, . . pp. 802-835 
 
 Cast Iron — Different Kinds of Pig — Cold- blast Iron — Mixtures — Specific Gravity of 
 Cast lion — Strength — Wrought Iron — Rolled Bar — Merchant, Best, etc., Bars — Weight 
 of Bars — Forgings — Steel — Bessemer, Siemens-Martin — Steel for Boiler Construction — 
 Admiralty, Board of Trade, Lloyd's, British Corporation Rules — Sizes of Bars and Plates 
 — Price — Steel Forgings — Castings — Nickel Steel — Chrome-Vanadium Steel — Soft Steel 
 for Solid- drawn Tubes — Manganese Steel — Copper — Tin — Zinc or Spelter — Lead — 
 Aluminium — Duralumin — Antimony, Alloys, Brass, Muntz Metal — Naval Brass — Tube 
 Metal — Gun-metal — Admiralty and Phosphor Bronzes — Crotarite — Immadium — Tur- 
 b.ulium — Stone's Bronze — Bull's Metal — Melloid — Delta Metal — Parsons' White Bronze 
 — Babbit's White Metal — Admiralty White Metal — Plumtine — Magnolia — Fenton's 
 White Metal — Stone's White Bronze — German Silver — Richard's Plastic Metal — Tables 
 on Strength, Composition, Properties, and Prices of Metals, etc. — Effect of Temperature 
 on Strength of Metals. 
 
 CHAPTER XXX. 
 
 Oil and Lubricants — Engine Friction, . . pp. 836-844 
 
 Consumption of Oil — Kinds of Oil — Animal Oils — Vegetable Oils — Mineral Oils — 
 Density — Viseosity — Boiling, Setting, and Flash Points — Compound Oils — Soaps — 
 Adulteration — Solid Lubricants — Tallow — Grease — Coefficients of Friction. 
 
 CHAPTER XXXI. 
 
 Tests and Trials : Their Objects and Methods, . . pp. 845-858 
 
 General Considerations — Factors or Margin of Safety — Admiralty Rules for Testing 
 Machinery — Italian Government — British Corporation — Lloyd's — Board of Trade — 
 Germanischer Lloyds — Trials under Steam — Admiralty Ship Trials — Merchant Ship 
 Trials — Modern Steamship Trials — Progressive Trials — Do. at Sea — Endurance Trials — 
 Mercantile Marine. 
 
 APPENDIX A. 
 
 The Diesel Oil Engine, PP- 859-875 
 
 Efficiency — Rate of Revolution — Two-cycle Engines — Double-acting Oil Engines — 
 Junker Engine — Reversing Oil Engines — Manoeuvring — Twin-screw Oil Ship " Fiona " 
 — Doxford's Double-piston Engine — Examples of Oil-engined Ships — Semi-Diesel 
 Engines. 
 
iVi CONTENTS. 
 
 APPENDIX B. 
 Lloyd's Register Rules for Internal Combustion Engines, . pp. 876-884 
 
 APPENDIX C. 
 Bureau Veritas Rules for Internal Combustion Engines, . „ . pp. 885-886 
 
 APPENDIX D. 
 Turbines on Shipboard, . . . . . . . pp. 887-896 
 
 APPENDIX E. 
 Superheated Steam and Superheaters, . . . . . pp. 897-903 
 
 APPENDIX F. 
 
 Board of Trade Rules for Boilers, ..... pp. 904-911 
 
 Surveyor's Duty and Responsibility — Working Pressure — Girders for Flat Surfaces 
 — Compression Stress on Steel Tube Plates — Cylindrical Boilers — Rules. 
 
 APPENDIX G. 
 
 Lloyd's Rules for determining the Working Pressure to be allowed in New 
 
 Boilers, ........ pp. 912-916 
 
 APPENDIX H. 
 
 Rules of the British Corporation for the Working Pressure, Thickness of 
 
 Plates, and Sizes of Stays for Steel Boilers. . . . pp. 917-921 
 
 APPENDIX I. 
 
 Bureau Veritas Rules for Boiler Shells, Working Pressure or Thickness of 
 
 Plates, and for Sizes of Stays, . . . . -pp. 922-930 
 
 APPENDIX J. 
 Admiralty Rules for Steel Boilers and Materials and their Tests, . . pp. 931-935 
 
 APPENDIX K. 
 German Government Rules tor Boilers. . . . . .pp. 936-944 
 
 APPENDIX L 
 Lloyd's Rules lor Electric Light Installations on Board Vessels, . . pp. 945-948 
 
CONTENTS. XVli 
 
 APPENDIX M. 
 Refrigerating Machinery and Appliances, Lloyd's Registers' Rules, . . pp. 940-954 
 
 APPENDIX N. 
 Rules for Vessels Fitted for Burning and Carrying Liquid Fuel, . • .p. 955 
 
 APPENDIX 0. 
 Board of Trade Rules tor Safety Valves, . . . . . pp. 956-960 
 
 APPENDIX P. 
 Bureau Veritas Rules for Boiler Safety Valves and Metal Tests, etc., . . p 961 
 
 APPENDIX Q. 
 Admiralty Rules for Testing Materials for Machinery, . . . pp. 962-964 
 
 APPENDIX R. 
 
 Recommendations for Main Boilers of the British Marine Engineering Design 
 
 and Construction Committee, . . ... • • PP- 965-967 
 
 Index, ......... P- 969 
 
LIST OF TABLES. 
 
 TABLE 
 
 I. Showing Progress of Naval Engineering — Screw Propulsion, . 
 II. „ „ Engineering in the Mercantile Marine, . 
 
 III. John's Coefficients for Computing Effective Horse-power, . 
 
 IV. Coefficients for Computing I.H.P. by the old Admiralty Formula, 
 V. Coefficients (Prismatic) of Displacement suitable for Various Speeds, 
 
 VI. Relation of Power to Speed (Sir W. H. White), 
 
 VII. Results of Trials of Ships at Speeds from 9 to 12 Knots, . 
 VIII. „ „ „ „ „ 12 to 15 „ 
 
 IX. „ ,, ,, „ 15 to 18 -„ . 
 
 X. „ „ „ „ 18 to 21 ,, 
 
 XI. „ „ „ „ „ 21 and upwards, . 
 
 XII. „ „ „ fitted with Turbines, 
 
 XIII. „ „ „ „ Paddle-wheels, 
 
 XIV. Relation of Power and Displacements, .... 
 XlVa. Comparative Trials of Ships fitted with Turbines and Reciprocators, 
 
 XV. Water Consumption of Engines of H.M.S. " Amethyst " and " Topaz' 
 Compared, ........ 
 
 XVI. Results of Measured Mile Trials of s.s. " Otaki," 
 XVII. Steam used Expansively, ...... 
 
 XVIII. Factors for Estimatimg Mean Pressure of Steam in Practice, 
 XIX. Results of Trials of Naval Twin-screw Ships, . 
 XX. „ „ Merchant Steamships, . . . 
 
 XXI. „ „ Two-stage Compound Engines, . . 
 
 XXII. „ „ Three-stage Three-crank Engines, . 
 
 XXIII. „ „ „ Four-crank Engines, . 
 
 XXIV. „ „ Four-stage Expansion Engine, , . 
 XXV. Efficiency of some German Engines, . . . . 
 
 XXVI. Yarrow's Experiments on Efficiency, .... 
 XXVII. Tyacke's Experiments on Efficiency, ..... 159, 
 XXVIII. Work done Theoretically by One Pound of Steam, . . 
 XXIX. Examples of Typical Engines, ..... 
 
 XXIXa. Rates of Revolution of Screw Propellers, . . . 
 
 XXIX6. Strokes of Engines as in Common Practice, 
 XXX. Trials of a Typical Cargo Steamer, .... 
 
 XXXI. Factors for obtaining Sizes of Steam, Exhaust Pipes, etc., 
 XXXII. Weight of Steam passed through Pipes with a Drop of 1 Lb., 
 
 XXXIII. Stuffing-boxes and Glands, Sizes of, ... . 
 
 XXXIV. Ratio of Maximum to Mean Twisting Moments at Various Cut-offs, 
 XXXV. Board of Trade Factors for Shafts, .... 
 
 XXXVI. Lloyd's Factors for Shafts of Steam Reciprocators, . 
 XXXVII. and XXXVIIa. Lloyd's Factors for Shafts of Oil Engines, . 
 XXXVIII. British Corporation Factors for Shafts, .... 
 
 XXXIX. Factors for Shafts per Author's Formula, 
 XXXIXa. „ Paddle-shafts, per Author's Formula, 
 
 XXXIX6. „ Crank-shafts of Screw Engines, per Author's Formula, 
 
 XL. Effect of Vacuum on Steam Consumption, 
 
 XLI. Temperature, Latent Heat, and Volume of Steam at very Low Pressure, 
 XLII. Number of Tubes in a Condenser per Square Foot of Tube Plate, 
 XLIII. Trials of Battleship " Ibuki "—"Curtis Turbines, 
 XLIV. Ratio of Cooling Water to Steam Condensed, . 
 
 PACK 
 
 17 
 
 18 
 26 
 27 
 28 
 41 
 42 
 43 
 44 
 45 
 46 
 47 
 48 
 49 
 92 
 
 93 
 97 
 111 
 125 
 130 
 131 
 132 
 133 
 134 
 135 
 156 
 159 
 , 160 
 161 
 176 
 208 
 209 
 232 
 247 
 248 
 260 
 298 
 321 
 322 
 323 
 324 
 329 
 331 
 332 
 362 
 363 
 367 
 368 
 370 
 
XX 
 
 LIST OF TABLES. 
 
 TABLE 
 
 XLV. Ratio of L.P. Cylinder to Circulating Pump Capacities, 
 XLVI. Centrifugal Pumps for Circulating Cooling Water, 
 XLVII. Particulars of Screw Propellers, Solid and Loose-bladed, 
 XLVIII. Thickness of Copper Pipes for Various Purposes, 
 XLIX. Steam and Fuel Consumption of R.M.S. " Lusitania," 
 
 L. Particulars of Admiralty Type of Fans for Air Circulation, . 
 LI. Fuels of all Kinds (Solid), ...... 
 
 Llcr. Fuels (Liquid), ....... 
 
 LII. Comparison of Various Designs of Boilers, .... 
 
 LIII. Allowance of Total Heating Surface in Practice, 
 LIIIo. Capacity of Funnels, ...... 
 
 LIV. Particulars of Marine Boilers made under Old Rules, 
 LV. „ „ „ „ „ New Rules, 
 
 LVI. Basin and Sea-going Trials of H. M.S. " Sheldrake," 
 LVII. Particulars of Surface, Weight, etc., of Express Boilers, 
 LVIII. Special Trials of Belleville Boilers in H.M.S. " Sharpshooter," 
 LIX. Results of Steam Trials with Niclausse Boilers, 
 LX. Trials of the Miyabara Boiler, ..... 
 
 LXI. Water Consumption of Main and Auxiliary Machinery — H.M.S. 
 "Hermes," . . . . 
 
 LXII. Water Consumption of Main and Auxiliary Machinery — H.M.S. 
 
 "Diana," . . . 
 
 LXIII. Results of Trials of Various Modern Marine Boilers in Lbs. per Hour, . 
 LXV. Trials of Various Naval Ships — Water Tube compared with Tank 
 Boilers, ........ 
 
 LXV1. Trials of Various Marine Boilers, ..... 
 
 LXVII. Tests of Boiler Material prescribed by Various Authorities, . 
 LXVIII. Strengths of Various Kinds of Riveted Joints, 
 LXIX. Single-riveted Lap Joints, ...... 
 
 LXX. Particulars of Double-riveted Lap Joints of Plates, . . 
 
 LXXI. ,, Treble-riveted Lap Joints of Plates, . . 
 
 LXXII. „ Treble Special Riveted Lap Joints of Plates, . 
 
 LXXIII. „ Double -riveted Butt Joints of Plates, . 
 
 LXXIV. „ Double Special Riveted Butt Joints of Plates, 
 
 LXXV. „ Treble Ordinary Riveted Butt Joints of Plates, 
 
 LXXVI. „ Treble Special Riveted Butt Joints of Plates, . 
 
 LXXVI1. Relative Thickness of Shell Plates for Different Tensile Strengths, . 
 LXXVIII. Values of Factor F in Formula for Shell Plates, 
 LXXIX. Boiler Tubes, Standard Sizes and Thicknesses, 
 LXXX. Particulars and Scantlings of Modern Cylindrical Boilers, 
 LXXXI. Details of Riveting for Funnels, Casings, etc., . . 
 
 LXXXII. All-round Reversing Gears, Particulars of, . 
 LXXXIII. Limits of Safe Working Stresses on Various Metals, . 
 LXXXIV. Conditions obtaining with Machinery of Modern Ships, 
 LXXXV. Particulars of Machinery of Various Naval Ships, 
 LXXXVI. ,, ,, ,, Ocean Express Steamers, 
 
 LXXXVII. ,, ,, „ Passenger Cargo Steamers, 
 
 LXXXVIII. ,, ,, „ Small Passenger Steamers, 
 
 LXXXIX. „ „ „ Paddle-wheel Steamers, 
 
 XC. ,, „ „ Turbine Screw Steamers 
 
 XCI. Weights of Moving Parts of Balancing Engines, 
 XCII. Admiralty Tensile Tests for Boiler Steel, 
 X ( ' 1 1 1 . Board of Trade Tensile Tests for Boiler Steel, 
 XCIV. Composition of Metal and their Alloys, 
 XCV. ,, Various White (Bearing) Metals. 
 
 XCVI. Phvsical Properties of Various Metals, 
 XCVII. Current and Past Prices of Various Materials (1912). 
 XCVI 1 1. Standard Tests for Steel adopted in U.S. America, . 
 XCIX. Constituents of Solid Matter in River and Sea Water, 
 C. „ ,, „ deposited in Boilers, . 
 
 CI. Effect of Temperature on Metals, 
 CII. Specific Gravity of Various Oils at "80° F., 
 
 PAGE 
 
 397 
 410 
 478 
 509 
 513 
 522 
 544 
 545 
 575 
 579 
 583 
 584 
 585 
 596 
 612 
 616 
 619 
 623 
 
 627 
 
 628 
 628 
 
 630 
 
 . 631 
 
 . 635 
 
 . 638 
 
 . 639 
 
 . 644 
 
 . 644 
 
 . 645 
 
 . 647 
 
 . 647 
 
 . 647 
 
 . 648 
 
 . 650 
 
 . 652 
 
 . 661 
 
 . 667 
 
 . 671 
 
 . 699 
 
 . 724 
 
 . 735 
 
 . 740 
 
 . 740 
 
 . 742 
 
 . 744 
 
 . 7446 
 
 . 744c 
 
 . 779 
 
 . 809 
 
 . 810 
 
 823, 824 
 
 . 825 
 
 826, 827 
 
 . 828 
 
 829, 830 
 
 . 832 
 
 . 832 
 
 . 834 
 
 . 838. 
 
LIST OF TABLES. 
 
 XXI 
 
 TATttK PAQB 
 
 (III Density of Various Oils at Different Temperatures, . . . s:{'.» 
 
 CIV. Viscosity of Various Oils at Different Temperatures, . . . 839 
 
 CV. Boiling, Setting, and Flash Points of Various Oils, . . . siu 
 
 CVI. Prices of Various Oils, . . . . . .sin 
 
 ('VII. Coefficients of Friction of Various Materials, . . . 84.'! sll 
 
 CVIII. Examples of Oil Engine-driven Ships, .... 879 
 
 CIX. Methods of Transmission from Turbine to Screw, . . . 895 
 
 CX. Steam Consumed per S.H. P. -hour (various conditions). . . 898 
 
 CXI. Superheated Steam through Pipes, Lbs. per Minute, . . . 903 
 
 CXII. Board of Trade Factors of Safety for Boilers, . . . 907 
 
 (XIII. Constants for Joints of Boiler Plates with Different Kinds of Riveting, 911 
 
 CXIV. Values of 2 R in the Formula of the Bureau Veritas for Boiler Shells, 923 
 
 CXV. Value of Bureau Veritas Multiplier Factor for Flat Plates, . . 927 
 
 CXVI. Board of Trade Allowances for Safety Valve Area*;, . , . 957 
 
LIST OF ILLUSTRATIONS. 
 
 FIO. 
 
 DESCRIPTION OF ILLUSTRATIONS. 
 
 1. Thornycroft's Stern for Shallow-draught Screw Ships, . 
 
 2. Yarrow's Drop-flap Shallow-draught Screw Ships, . 
 
 3. Paddle Steamer " Charlotte Dundas," 1802, . 
 Machinery and Wheels of Paddle Steamer " Comet," 1812, 
 Modern American Lake Paddle Steamer, 
 Turbine Steamer " Duchess of Argyll," 
 Engines and Paddle Wheels of a Stern Wheeler, 
 p.s. " Empress Queen," 11,000 I.H.P., 
 Early Screw Propellers of Smith and Ericsson, . 
 Kirk's Analysis — Block Model, 
 Curve of I.H.P. for Varying Speeds, 
 Curves of Power, etc., of s.s. " Lusitania " on Trials, . 
 Trials of H.M.S. " Cossack " in Water of Different Depths, 
 Harold Yarrow's Experiments in Water of Different Depths, 
 
 4. 
 
 5. 
 
 6. 
 
 7. 
 
 8. 
 
 9. 
 10. 
 11. 
 12. 
 13. 
 14. 
 15. 
 16. 
 17. 
 18. 
 19. 
 20. 
 21. 
 22. 
 23. 
 24. 
 
 N. D. Lloyds Co.'s „ „ 
 
 Thornycroft Four-Cylinder Engine using Light Oils, 
 Belliss and Morcom Heavy Oil Engine, 
 Compound Diagonal Paddle Engine — Two Cranks, 
 
 M „ „ — Cranks Coupled, 
 
 Three-crank Compound Paddle Engine, 
 
 „ Triple-compound Paddle Engine, 
 American Steamer Beam Engine, 
 Engine of the " Comet," 1811-12, 
 
 25. Side-lever Engine, 
 
 26. Engines of p.s. " Regent," 1816, 
 27 and 27a. Oscillating Engine, . 
 
 28. Steeple Engine, 
 
 29. Twin-cylinder Engine, 
 
 30. Diagonal Compound Paddle Engines, U.S.A., 
 
 31. Trunk Engines, 
 31a. Return Connecting-rod Engine, 
 
 32. Three-crank Triple-expansion Engine (Naval), 
 
 33. Vertical Quadruple-expansion Engine (Express), 
 33a. „ „ „ (Mercantile Ordinary), 
 
 34. Engines of H.M.S. " Salmon " (Destroyer), 
 
 35. High-speed Cruiser Engines, .... 
 
 36. Triple-compound Engines for Fishing Craft, . 
 
 37. Five-cylinder Quadruple-expansion Engines of s.s. " Inchdune," 
 
 38. Eight-cylinder „ „ „ <( " K.W.," ;> 
 
 39. General Arrangement of Turbine, etc., of R.M.S. " Lusitania," 
 
 40. Diagram showing Results of Trials of H.M.S. " Amethyst " and " 
 
 41. Combination of a Central Turbine with Twin Reciprocators. . 
 
 42. „ Twin Turbine's „ „ 
 
 43. Longitudinal Section through Engine-room of R.M.S. " Olympic,' 
 
 44. Plan of Engine-room of R.M.S. " Olympic," . 
 
 45. The Wheel Gearing of s.s. " Vespasian," 
 
 46. Combination of Twin Turbines with a Single Reciprocator, 
 
 47. Oil Engine, Single-acting, Marine Type, Six Cylinders, 
 
 48. „ Double-acting „ „ 
 
 49. Mirrlees-Diesel Single-acting Marine Engine, . 
 
 50. Marine Oil Engine, Two-cycle Single-acting, Diesel System, . 
 50a. Section of „ 
 
 Topaz," etc. 
 
 PA OK 
 
 2 
 
 3 
 
 6 
 
 7 
 
 8- 
 
 9 
 
 11 
 
 14 
 
 IS 
 
 32 
 
 38 
 
 60 
 
 51 
 
 51 
 
 53 
 
 54 
 
 58 
 
 59 
 
 61 
 
 62 
 
 63 
 
 64 
 
 65 
 
 66 
 
 67 
 
 67 
 
 68, 69 
 
 71 
 
 72 
 
 73 
 
 74 
 
 74 
 
 75 
 
 76 
 
 78 
 
 79 
 
 80 
 
 81 
 
 84 
 
 87 
 
 90 
 
 91 
 
 95 
 
 96 
 
 98 
 
 99 
 
 100 
 
 101 
 
 102 
 
 104 
 
 105 
 
 106 
 
 107 
 
XXIV 
 
 LIST OF ILLUSTRATIONS. 
 
 JIG. DESCRIPTION OF ILLUSTRATIONS. PAGE 
 
 51. Two-cycle Oil Engine, showing Three Fundamental Stages, . . .108 
 
 52. Turbo-motor, Single-stage, on De Laval System, . . . .110 
 52a. ,, Two-stage or Compound, De Laval System, . . .110 
 
 53. Rankine's Graphic Method for determining Mean Pressure, . . .112 
 
 54. Theoretical Indicator Diagram showing compression, . . .114 
 
 55. ,, ,, for Triple-expansion Engine, . . . 129 
 55a. Diagrams as in Practice from Theoretical Diagrams, . . . 129 
 
 56. Reaction Turbine, Equivalent Nozzles, . . . . .136 
 
 57. Impulse „ „ „ ..... 137 
 57a. Diagram showing Stages of Pressure Reduction, .... 139 
 576. ,, ,, Variation in Velocity due to Pressure Reduction, . . 139 
 
 58. Ljungstrom Turbine for Electric Propulsion, ..... 140 
 
 59. Curtis Turbine as fitted in a Japanese Battleship, . . . . 141 
 59a and 596. Zoelly Turbine for Marine Work, . . . . .142 
 
 60. Fottinger Torsion Meter, . . . . . • .144 
 60a. Diagrams from a Fottinger Torsion Meter, . . . . .144 
 
 61. Hopkinson and Turing's Torsion Meter mounted on Shaft, . . . 146 
 61a. „ „ „ Diagram of, . . . . 147 
 
 62. Collie's Mechanical Torsion Meter, ...... 149 
 
 63. Crank Effort Diagram from Torsion Meter of G. Hamilton Gibson, . . 151 
 
 64. Friction Trials with Belliss and Morison's Engines, .... 158 
 64a. Diagram showing Effect of Superheat on Consumption, . . .171 
 
 ' 65. Indicator Diagrams and their Combination of some Quadruple Engines, . 193 
 
 66. Diagram showing the Various Stages in the Use of Steam Expansively, . 194 
 
 67. Indicator Diagram, . . . . . . . .199 
 
 68. Weir's Condenser as fitted to a Quadruple-expansion Engine, . . 213 
 
 69. Single-crank Compound Engine — Two Cylinders, . . . .217 
 
 70. Single-crank Triple-expansion Engine, ..... 218 
 
 71. Triple-expansion Engine with Special Valve Gears, . . . . 219 
 
 72. „ „ with Four Cranks, Balanced, . . . 220 
 
 73. Various Arrangements of Cylinders of Triple-expansion Engines, . 236, 237 
 
 74. „ „ ,, Quadruple-expansion Engines . . 239 
 
 75. Cylinders of a Two-crank Quadruple-expansion Engine, . . . 240 
 
 76. „ Three-crank Triple-expansion Engine, .... 242 
 
 77. H.P. and L.P. Cylinders of a Naval Engine, . . . 244, 245 
 
 78. Section through a Cylinder, ....... 250 
 
 78a. Admiralty Method of fitting Liners, ...... 250 
 
 79. Shortened Cylinder with Port in Cover, ..... 257 
 
 80. Combination Metallic Packing, ...... 259 
 
 80a, 806. Stuffing-boxes and Glands, ...... 259 
 
 81. 81a. Piston with Ramsbottom Rings, ...... 264 
 
 82. Piston with Common Packing and Junk Rings, .... 264 
 
 83. „ Mather & Piatt's Packing Ring and Springs, . . . 264 
 
 84. ,, Buckley's Packing Ring and Springs, .... 264 
 
 85. „ Quarter & Hall's Packing Ring and Springs, . . . 264 
 ' 86. „ Cameron's Wave Spring and Common Rings, . . . 264 
 
 87. Methods of Jointing Ends of Spring Rings of Pistons, . . . 264 
 
 88. Packing Rings of Piston restrained on Admiraltv Method, . . . 266 
 
 89. Forged Steel Piston, . . . . . . .267 
 
 90. Cast Steel Piston, . . . . . . . .207 
 
 91. Piston-rod with T End and Double Slipper Guide Shoes, . . - 270 
 
 92. „ „ „ Single Slipper Guide Shoes, . . . 270 
 
 93. „ formed with Head and Slipper Foot and Loose Gudgeon, . . 270 
 
 94. „ Crosshead with Gudgeons in One Piece, .... 274 
 
 95. „ „ „ „ Steel Casting, . . . 275 
 
 96. Diagram showing Thrust of Connecting-rod, ..... 275 
 
 97. Crosshead and Guide Block for Double Piston-rods, .... 278 
 
 98. Connecting-rod with Solid Ends — Gudgeon shrunk in Place, . . . 279 
 
 99. „ „ Fitted Brasses at each End, .... 2T'.t 
 
 100. „ „ „ „ „ ConicalJaws, . . 284 
 
 101. Indicator Diagram of a Diesel Oil Engine (Two-stroke Cycle), . . 287 
 
LIST OF ILLUSTRATIONS. 
 
 XXV 
 
 FIO. DESCRIPTION' OF ILLUSTRATIONS. 
 
 L02. Diagram showing Turning Moment at various Angles of the Crank, 
 
 103. Diagram of the Curve of Twisting Moments of a Single Crank, 
 103a. Diagram of Combined Twisting Moments of Two Cranks, 
 
 104. Crank Effort Diagram of a Three-crank Engine — Triple-compound, 
 
 105. The Cranks of a Paddle-wheel Engine, 
 
 106. Built-up Crank-shaft, 
 
 107. Naval Crank-shaft. . 
 
 108. Ccdervall's Screw Shaft Fittings, 
 
 109. Michel's Thrust Block, 
 
 110. Thrust Block with Adjustable Collars, 
 110a. Overhung Crank (Bureau Veritas). 
 1106. Double Ciank (Bureau Veritas), 
 
 111. Crank-shaft Bearing, . 
 
 112. Improved Form of Crank -shaft Bearing, 
 
 113. Solid Steel Main Bearing Frame — Naval Engines, 
 
 114. Cast Steel Main Bearing Frame — Naval Engines, 
 
 115. Weir Uniflux Condenser for a Turbine— 11,000 S.H.P. 
 
 116. Cylindrical Condenser on Morison's System, . 
 
 117. Ordinary Marine Engine with Contraflo Condenser 
 117a. Diagram showing Flow of Steam in the Contraflo System, 
 
 118. Contraflo Condenser with Feed Temperature Regulator, 
 118a. Morison's Diagram for Air saturated with Water Vapour, 
 1186. Indicator Diagram showing Effect of High Vacuum, 
 
 119. Condenser Tube fitted with Wooden Ferrules, 
 
 120. „ ,, „ Screwed Glands and Tape Packings, 
 
 121. „ ,, „ Safety-screwed Glands, 
 
 122. Air Pump of Ordinary Type, . 
 
 123. Edwards' Air Pump, .... 
 
 124. Weir's Dual Air Pumps, 
 124a. ,, ,, Diagrammatic Section of, 
 
 125. Indicator Diagram of Air Pumps, 
 
 126. Air Pumps, Motor-driven, 
 
 127. Condenser with Kinetic Air Pumps, . 
 127a. Parsons' Vacuum Augmentor, 
 128-132. Air-pump Valves of Kinds. 
 
 133. Impeller or Wheel of a Centrifugal Pump, 
 
 134. 17-inch Circulating Pump, 
 134a. Circulating Pumps by Allen for a Battleship and s.s. " Mauretania," 
 
 135. Common Locomotive Slide Valve, 
 
 136. The Trick Valve, .... 
 
 137. Common Double-ported Flat Valve, . 
 
 138. Piston Slide Valve, .... 
 
 139. Self-balanced Dual Valve, 
 
 140. Thorn's Patent Piston Valve, . 
 
 141. Dawe & Holt's Patent Relief Frame for Flat Valves, 
 
 142. Common Naval Relief Frame for Flat Valves, 
 
 143. Spring and Packing Rings, Relief Frame for Flat Valves, 
 
 144. Martin & Andrews' Valve Relief Frame, 
 
 145. Church's Patent Valve Relief Frame, . 
 
 146. Valve with Exhaust through its Back, 
 
 147. Joy's Assistant Cylinder for Slide Valves, 
 1476. ,, „ Indicator Diagrams of, 
 
 148. Valve-rod Guide, 
 
 149. Proportions of Locomotive Slide Valves, 
 
 150. Slot Link and Rod Ends, 
 
 151. Double-bar Link with Rods inside, 
 
 152. „ „ outside, . 
 
 153. Marshall's Patent Valve Gear, 
 
 154. Joy's Patent Valve Gear, 
 
 155. Effect of Notching-up on Indicator Diagrams, 
 
 156. Diagram of the Piston Path, . 
 
 PAGE 
 
 295 
 
 296 
 
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 307 
 
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 441 
 
 444 
 
XXVI 
 
 LIST OP ILLUSTRATIONS. 
 
 fio. 
 
 157. 
 158. 
 159. 
 160. 
 161. 
 J62. 
 
 DESCRIPTION OF ILLUSTRATIONS. 
 
 Zeuner's Diagram for the Motion of a Slide Valve, 
 
 Valve Diagram showing Effect of " Notching-up," 
 
 Link Motion " Notched-up," .... 
 
 Diagram of the Motion of an Independent Expansion Valve 
 
 „ „ of an Expansion Valve on the Back of the Main Valve, 
 
 „ ,, of another Expansion Valve on the Back of the Main 
 
 Valve, .... 
 
 163. Adjustable Expansion Valve on Back of the Main Valve, 
 
 164. Effect of Sequence of Cranks on the Indicator Diagrams, 
 
 165. Paddle-wheel with Feathering Floats, 
 
 166. Locus of the Flat of a Feathering Wheel, 
 
 167. Feathering Paddle-wheel with Wooden Floats, 
 
 168. „ „ „ Steel-plate Floats, 
 
 169. „ ,, ,, • Gear on Inner Side of Wheel, . 
 
 170. Smith's Screw Propeller as first fitted in the Navy, 
 
 171. Diagram showing Curve of Indicated Thrust, . 
 
 172. Diagram showing Blade Pressures of Screw Propellers, 
 
 173. Solid Cast-iron Screw Propeller, 
 
 174. Bronze Screw Propeller with Loose Blades, Ordinary Surface Ratio, 
 
 175. „ „ „ „ Large Surface Ratio, 
 
 176. General Form of Modern Mercantile Marine Screw, 
 176a. Typical Sections of Screw Propeller Blades, . 
 
 177. Hirsch's Patent Screw of 60°, 
 
 178. Solid Parsons' Bronze Screw of R.M.S. " Mauretania,' 
 
 179. Damaged Bronze Screws, 
 
 180. Bevis Patent Feathering Screw, 
 
 181. Kingston Valve — Typical Design, 
 
 182. Details of Inlet Valves on Steel Ships, 
 
 183. Double Ram Flywheel Auxiliary Pump, 
 
 184. Section of Worthington Duplex Pump with Externally-packed Rams, 
 
 185. Weir's Automatic Pump, ..... 
 
 186. ,, ,, with Float, Tank, and Automatic Gear, 
 
 187. Compound Engine for Electric Generating — Enclosed Self -lubricating, 
 
 188. „ „ „ „ — Single-crank, 
 
 189. „ „ „ „ — Two-crank, Naval, 
 189a. „ „ „ „ „ Section of, 
 
 190. Turbo Electric Installation for Ship Lighting, etc., 
 
 191. Air Compressors for High-pressure Naval Work, 
 191a. „ „ „ „ Section of, 
 
 192. Steam Steering Gear with Hand Gear (vertical), 
 
 193. „ . „ „ „ (horizontal), 
 
 194. „ „ as fitted in Warships, 
 '195. „ „ Brown's Patent, 
 
 196. Refrigerating Apparatus for Warships, 
 
 197. ,, ,, Merchant Ships, . 
 
 198. Section of Hall Refrigerating Apparatus, 
 
 199. Auxiliary Condenser for Merchant Ships, 
 
 200. „ „ Warships, 
 
 201. Holden's Apparatus for Burning Oil Fuel, . " 
 201a. Lassac Lovekin Apparatus for Burning Oil Fuel, 
 
 202. Furnace fitted with Apparatus for using Oil Fuel. 
 
 215. Goldsworthy Gurney's Water-tube Boiler of 1827, 
 
 216. Ward's High-pressure Coil Boiler, 
 
 217. Babcock & Wilcox Boiler, Large Tube, Naval Type, 
 "218. ,. „ Headers for Small Tubes, 
 218a. „ „ „ „ Large 
 
 219. „ „ Small Tube, Naval Type, 
 
 220, 221. „ „ Mixed Sizes of Tubes, Naval Type, . 698, 
 
 222. Normand Water-tube Boiler, . 
 
 223. Fleming & Ferguson's Water-tube Boiler, 
 ■224, 224a. Seaton's Water-tube Boiler, 
 
 PAGE 
 
 445 
 450 
 450 
 451 
 453 
 
 454 
 
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 597 
 
 599 
 
 600 
 
 602 
 
 603 
 
LIST OF ILLUSTRATIONS. 
 
 xxvn 
 
 pia. DESCRIPTION OF ILLUSTRATIONS. 
 
 225. Yarrow's Large Tube Water-tube Boiler, 
 
 226. „ Small Tube Water-tube Boiler, 
 
 227. „ Boiler as made in Japan, 
 
 228. Blechynden's Water-tube Boiler, 
 
 229. Thornycroft's Water-tube Boiler, 
 
 230. Mumford's Large Type of Water-tube Boiler. 
 
 231. Reed's Water-tube Boiler for all Classes of Ships, 
 
 232. Modern Belleville Boiler, with Economiser, . 
 232a. Tube and Headers of a Belleville Boiler, 
 
 233. Modem Niclausse Marine Boiler, 
 
 234. Stirling Marine Boiler, .... 
 
 235. Hohenstein Boiler, . " . . 
 
 236. Miyabara Boiler, ..... 
 
 237. Thornycroft-Marshall Boiler, 
 
 237a. Special 9-rivet Quadruple Joint, . . 
 
 2376. „ 11 -rivet „ „ . 
 
 238-240. Methods of Flanging Boiler Ends, 
 
 241. Furnace Stiffened with a Bowling Hoop, 
 
 with an Adamson Joint, . 
 by Differential Diameter, 
 by Corrugation (Fox's Plan), 
 by Ridges (Purves' Plan), 
 by Curved Grooves (Morison's Plan), 
 by Corrugation (Holmes' Plan), . 
 „ (Deighton's Plan), 
 
 Removable Furnace (Ashlin's Design), 
 
 Methods of Connecting the Furnace to End Plate, 
 
 242. 
 243. 
 244. 
 245. 
 246. 
 247. 
 248. 
 249. 
 250. 
 250a. „ „ „ Tube Plate 
 
 251. „ „ „ End Plate, 
 251a. „ „ „ Tube Plate, . 
 
 252. Admiralty Ferrule for Fire Ends of Tubes, . 
 
 253. Funnel and Casing for Naval Ships, Yachts, etc., 
 
 254. „ „ Merchant Ships (Ordinary), 
 
 255. Naval Bulkhead Self-closing Stop Valve, . , 
 
 256. Spring Safety Valves as fitted in Merchant Ships, 
 
 257. Adams' Form of Valve and Seat of Safety Valves, 
 
 258. Cockburn's Form of Valve and Seat of Safety Valves, 
 
 259. Safety Valves as fitted in Naval Ships (Short Type), . 
 
 260. „ „ ,, (Naval), Combined with 
 
 261. Feed Check Valve (Improved Form) as fitted to Boilers, 
 
 262. „ and Stop Valve Combined, 
 262a. ,, Valve, Automatic, on Mumford's Plan, . 
 
 263. Weir's Hydrokineter, or Water Circulator, 
 
 264. Steam Syren and Organ Whistle Combined, . 
 264a. Steam Syren — Admiralty Type, 
 
 265. Steam Reversing Gear, Brown's Original Pattern, 
 265a. „ „ ,, Improved Self-contained, 
 2656. „ ,, „ and Hand Gear Combined, 
 
 266. Reversing Gear for Twin-screw Turbines, 
 
 267. Steam Manoeuvring Gear for very Large Turbines, 
 
 268. Engine Stop and Regulating Valve for Small Engines, 
 
 269. „ „ ,, Balance for Large Engines 
 
 270. „ ,, ,, Equilibrium (Cockburn's) 
 
 271. Ash Ejector, Steam-worked (See's), . 
 
 272. Combination Governor (Murdoch's), . 
 
 273. Centrifugal Lubricator for Crank Pins, 
 
 274. Automatic Drainer, .... 
 
 275. Feed-water Heater and Automatic Regulator (Weir's 
 275a. „ ,, (Live Steam), 
 
 276. 277. Fresh Water Evaporator and Distiller, . 
 
 278. Fresh Water Supply Installation as in Battleships, 
 
 279. Grease Extractor and Feed-water Filter, 
 
 Stop V 
 
 alve, 
 
 713, 
 
 PACK 
 
 (ill I 
 
 605 
 
 606 
 
 607 
 
 608 
 
 610 
 
 till 
 
 6H 
 
 615 
 
 617 
 
 620 
 
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 622 
 
 624 
 
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 713 
 
 714 
 
 716 
 
 716 
 
XXVIU 
 
 LIST OF ILLUSTRATIONS. 
 
 no. 
 
 281. 
 
 282. 
 283. 
 284. 
 285. 
 
 286. 
 
 287. 
 
 DESCRIPTION OF ILLUSTRATIONS. 
 
 Indicator Diagram of a High-pressure Cylinder, 
 
 » Weight, 
 ,, ,> »> >> >> »> inertia, 
 
 Quadruple Engine, Two Cranks, Balanced (M' Alpine System), 
 Cranks and Curves of Forces for Quadruple Engine (Yarrow-Schlick-Tweedy), 
 Triple-compound Engine with Two Cranks (Wigzell's System), 
 
 in Balancing 
 
 Diagrammatic Single-crank Engine showing Forces, . 
 288-319. Diagrams illustrating the Problems involved 
 Engines, ..... 
 
 320. Heavy Oil Engine, Single-acting Two-cycle (by M. A. N. Co.) 
 
 321. Cylinder Section of the above Engine, 
 
 322. Heavy Oil Engine — Diesel System, Italian. . 
 
 323. „ „ — Junker's Design, 
 
 324. „ „ — Westgarth-Carel System, s.s. " Eveston,' 
 
 325. Emrines (Diesel) of Twin s.s. " Fiona," 
 
 326. Section of a Cylinder of a Westgarth-Carel Oil Engine, 
 
 327. Cylinders of Oil Engines, 
 
 328. Doxford's Double-piston Oil Engine, 
 328a. Semi-Diesel Engine (Kromhout), 
 
 329. Geared Turbine, 
 
 330. ,, s.s. " Normannia," 
 
 331. Double Reduction Gear for Turbines, 
 
 332. „ ,, Floating Frames 
 
 333. Fottinger's Turbine Transformer, 
 
 334. Superheaters in a Three-furnace Boiler, 
 
 335. Robinson's Superheater Header, 
 
 336. Treble-riveted Butt Joints with Half Number of Rivets in Outer Rows that 
 
 are in the Inner Ones, .... 
 
 337. Treble -riveted Butt Joint with Inner Strap taken by Inner 
 
 only, ...... 
 
 338. Hlustration of Irregular Staying of a Flat Surface, 
 
 339. Gauge Cock required by German Government for Testing, 
 
 PAGE 
 
 747 
 
 747 
 
 747 
 
 7c 3 
 
 754 
 
 755 
 
 . 756 
 
 . 757 
 
 Marine 
 
 777-8C0 
 . 861 
 . 862 
 . 863 
 . 864 
 . 869 
 . 870 
 . 871 
 . 872 
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 . 888 
 . 889 
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 . 894 
 . 9C0 
 . 902 
 
 Rows of Rivets 
 
 924 
 
 924 
 926 
 939 
 
I 
 
 A MANUAL 
 
 OF 
 
 MAEINE ENGINEERING 
 
 CHAPTER I. 
 
 GENERAL INTRODUCTION. 
 
 Fundamental Principles. — The first object aimed at by the marine engineer, 
 is to propel a floating body through the water at a certain speed ; the second, 
 so to construct the propelling apparatus that the motion may readily be 
 reversed ; and the third, to adopt such an arrangement of propeller and 
 engine as shall be convenient for the floating body and the service on which 
 it is employed. 
 
 The principle on which nearly all marine propellers work is the pro- 
 jection of a mass of water in the direction opposite to that of the required 
 motion. The only exception to this rule is the case of ferry steamers and 
 some river craft, where a chain or rope lying in the bed of the river passes 
 over a wheel or barrel in the ship itself. 
 
 The water, in modern practice, is projected by — (1) One or more screws 
 at the end of the ship (as will be described under the heading of propellers) ; 
 (2) by one or more paddle-wheels outside of the ship ; or (3), by a form 
 of wheel in the inside of the ship, which is generally spoken of as a /^-pro- 
 peller, as the water issues in jets from orifices in the ship's side. 
 
 The latter, however, is seldom used, although it has some features that 
 make it attractive in particular cases. The paddle-wheel, too, is slowly 
 dying out ; and although there is reason to believe that for river service, 
 especially in the tropics, and certain special duties, it will survive, it is, 
 nevertheless, gradually being displaced by the screw in some of these, while 
 in all other services, even in the shallow water ones which at one time seemed 
 reserved for its use, it is practically gone. 
 
 The paddle-wheel with feathering floats has certain qualities of its own, 
 which render it more serviceable than the screw in particular cases ; for 
 example : — 
 
 In tug boats quick manoeuvring is of extreme importance. The sudden 
 stopping, and equally sudden and certain starting, of the boat is most desir- 
 able ; this cannot be done with the screw, nor can the turning round in 
 
 1 
 
2 -';.'■ '. .' MANUAL OF MARINE ENGINEERING. 
 
 confined spaces be- accomplished even with twin screws so dexterously as 
 wit^' .disconnected paddle- wheels. 
 
 In river steamers running in very shallow water the paddle-wheel, especially 
 when fitted at the stern, has distinct advantages over the screw, even when 
 the latter is placed in a well or enclosure, as done by Mr. Yarrow and Sir 
 John Thoinycroft, inasmuch as it can free itself or be easily freed of weeds, 
 .s less liable to injury, and when injured is easily repaired. 
 
 In steamers making frequent and short calls at piers and wharfs, the 
 paddle-wheel permits of a more rapid service on rough and windy days than 
 can screws, especially the small screws of the turbine-driven ship. 
 
 On the other hand, the paddle-wheel is a heavy and somewhat clumsy 
 instrument exposed to wind and sea, and so liable to damage, even when 
 protected with boxes and guarded with sponsons, fenders, etc. ; it can, 
 however, be repaired by simple means, and even when badly damaged can 
 
 Fig. 1. — Thornycroft's Stern for Shallow Draught Screw Ships. 
 
 be generally sufficiently repaired by the ship's staff to permit of proceeding 
 on the voyage. 
 
 Its position in the ship and the space occupied by the paddle engine 
 interferes with the general economy ; the machinery is heavier and more 
 cumbersome than that driving a screw developing equal power, inasmuch 
 as the speed of revolution is necessarily restricted ; the wheel exerts a thrust 
 on a part of the ship less calculated to take pressure than that to which 
 the screw applies ; but with the modern steel ship, however, that can be 
 remedied, and provided for in a way which was not so easy with the wooden 
 ship of the past. 
 
 The paddle-wheel was the propeller of the first steamers put to practical 
 use, and to-day its efficiency is little short of that of the best screws. 
 
 The modern paddle engines, with their long stroke and choice design, 
 have brought this about to a very great extent. 
 
 The screw is much smaller than even the small high-revolution paddle- 
 
[NTRODUCTION — THE SCREW. 3 
 
 •wheel of to-day ; it is wholly immersed and largely protected by the quarters 
 of the ship. It is generally of so small a diameter as to be wholly below 
 the water line, but in shallow draft ships this is not always possible ; it is, 
 however, totally immersed when working by surrounding it with portions 
 of the body of the ship, so that it revolves in a channel, as done by Sir John 
 Thornycroft (tig. 1), or by dropping a flap behind it, as done by Mr. Yarrow 
 (tig. 2). The thrust from the screw is applied lower down, and nearer the 
 centre line of the ship's resistance, than is that of a paddle-wheel, so that 
 the tipping moment is quite small, and the thrust block is attached to a part 
 of the ship's structure eminently calculated to take the force. 
 
 V 
 
 60,000 
 
 % 40.000 
 
 Z SO, 000 
 
 ? 20.000 
 I 
 <fc 10,000 
 
 i£- 
 
 *&' 
 
 -^°& 
 
 &[.<m 
 
 He 
 
 fio? 
 
 _^1 
 
 
 \\t 
 
 itf9 
 
 ** 
 
 V 
 
 -A 
 
 7 a a w 
 
 Speed ( Statute Miles per Hour) 
 
 II 
 
 50,000 
 
 1 10,000 
 
 £783 
 Speed [Statute Miles perHourJ 
 
 Fig. 2. — Yarrow's Drop Flap for Shallow Draught Screw Ships. 
 
 There being no restriction to the number of revolutions of the screw, 
 the engines can thereby have a high piston speed, and be comparatively 
 light, cheap, and made to occupy less space, and moreover can be placed 
 in positions more convenient to the general arrangements of the ship. It 
 used to be urged that the screw caused much vibration and consequent 
 discomfort to those in the ship ; that it was inherent in the screw, however, 
 was erroneous, and it is known now that vibration was as much due to the 
 momentum of the moving parts of the engines as to the screw ; and. while 
 it is admitted that even a good screw may cause horizontal vibration at 
 the stern, owing to the difference in pressure of water on the lower and upper 
 
4 MANUAL OF MARINE ENGINEERING. 
 
 blades, it is only with the two-bladed variety that it is pronounced, and even 
 then only with those of large diameter and small submersion is it excessive. 
 With well designed three- and four-bladed screws driven at fairly high revolu- 
 tion by engines carefully balanced and well preserved, the vibration is virtually 
 nil. That the screw is liable to foul itself with ropes, nets, etc., which only 
 a diver can remove, and that it may damage itself by striking dock walls, 
 semi-submerged wreckage, etc., so as to necessitate dry docking, or tipping 
 must be admitted ; but this seldom happens. 
 
 Hydromotors. — There is, however, another method by which ships have 
 been, and may be, propelled distinguished from all others by the absence 
 of any propelling instrument. In the past many proposals have been made 
 and patented for ejecting a stream of water from the stern of a ship which 
 has been taken in at the bow or through the bottom. 
 
 As far back at 1729, John Allen proposed this method, and in 1793 James 
 Ramsey constructed and tried on the Thames a boat fitted with a pump 
 and pipes to do this : the speed attained, however, was only 4 knots per 
 hour. It was not, however, till 1849 that John Ruthven patented what 
 has since been known as the jet propeller, and associated with his name, 
 although he was not the first inventor of the system. In this case, however, 
 the impeller of the centrifugal pump is really the propeller. In the case 
 of the s.s. " Hydromotor," built and equipped in Holland in 1876, the 
 stream was ejected from a pair of cylindrical receivers by steam pressing 
 on the surface of the water which had been caused to flow into them by 
 reason of the vacuum formed by condensing the steam of the previous dis- 
 charge ; in fact, they were a pair of pulsometers. The enterprise was a 
 failure, and no one is likely ever again to attempt to propel a sea-going ship 
 by such extravagant means — extravagant in steam consumption and in 
 space occupied both by receivers and water passages. 
 
 Motive Power. — Both the screw and the paddle-wheel revolve around 
 their axes, consequently the engine employed to move them must have 
 circular motion, and inasmuch as both instruments are reversible, and can 
 thereby reverse the motion of the ship, the engines should be capable of 
 reversing. The engine must also be able to work efficiently at all speeds 
 varying from the maximum to the slowest ; for although most mercantile 
 ships usually run at " full speed " — that is, nearly approaching the maximum 
 — naval ships, cruising yachts, and some other vessels travel at varying 
 speeds, and seldom for a long period at full speed. Such engines are by 
 preference connected direct to the propeller shaft, but they may be again, as 
 they have been in the past, geared to them by means of spur wheels and 
 pinions, in order that there may be no compulsion to drive the propeller 
 at the same revolutions as the engines, or vice versa. The first screw steamers 
 had engines similar to those in paddle-wheel ships, and moving at the same 
 revolutions, and it is a fact that in the well-known trials of the s.s. " Rattler " 
 versus the p.s. " Alecto," the engines in these ships were of the same size, 
 design, and construction, but the screw of the '"Rattler" revolved four 
 times to one of the engines. ::: 
 
 Such engines as are suitable for the purposes of marine propulsion may 
 be worked by means of steam, or by gas generated from coal or oil. Hitherto 
 steam has been exclusively employed on sea-going craft, and oil vapour on. 
 
 * Vide "The Screw Propeller," p. 202 
 
INTRODUCTION — MARINE STEAM ENGINES, 6 
 
 small craft in home waters only. The success of the oil engine in such craft. 
 and the fact that engines using only crude and non-dangerous oils can now 
 he obtained, will no doubt lead to their use on large ships, and a more extended 
 service now they are made reversible, so that the propellers can be reversed 
 without wheel or other objectionable gearing, which when of small size do 
 not develop defects so rapidly as in the case when larger power is put through 
 them. The Diesel engine, which can be reversed by means of compressed 
 air. and uses heavy oil, has been fitted to several ships, and is now being 
 supplied to both the mercantile and naval marine in quite large sizes, by 
 multiplying the number of cylinders. The suction gas plant, which on 
 shore is used with fairly satisfactory results, may not be found so attractive 
 on a service where the working day is twenty-four hours, and the working 
 week seven days. It is, moreover, heavier than an oil engine installation. 
 
 The steam-worked machinery is more flexible than the oil and gas-worked, 
 is less liable to derangement from shocks, and more easily governed in a 
 sea way. But perhaps the most important feature to remember in making 
 a comparison is the fact that steam is water vapour produced by heat gene- 
 rated by the combustion of anything that will burn, so that for the boiler 
 of a steam engine fuel of sorts can be found in all parts of the world, whereas 
 for the internal combustion engine only certain oils and certain coals can be 
 used, and if they are not obtainable the engine is useless. There is, moreover, 
 another feature that cannot be altogether overlooked, and that is, while 
 the steam engine works with a back pressure of only 1 or 2 lbs. per square 
 inch, the gas engine has over 15 lbs. 
 
 Marine steam engines are of two kinds, the one works by means of the 
 elastic pressure of steam during expansion, the other by the kinetic energy 
 of the steam developed on expansion ; in the one the action is static, in the 
 other it is dynamic ; and just as in hydraulics, there is the common wheel 
 with its buckets filled with water, whose weight causes it to turn ; likewise 
 the ordinary ram which is forced outwards against a load, or as the recipro- 
 cating hydraulic engine when work is done by the pressure due to head, while 
 the Pelton wheel is moved by a jet of water impinging on its vanes with 
 velocity generated by the fall of the water, so there is the ordinary recipro- 
 cating expansive engine, and the turbine in which the steam acquires a high 
 velocity by expanding into and through a nozzle, which directs its flow on 
 to the blades of a rotor, where it gives up its energy by producing motion 
 in it against resistance. The efficiency of a reciprocating engine is little 
 affected by the velocity of flow of steam, be its piston speed high or low ; 
 the revolution for maximum efficiency of the turbine must be such that the 
 peripheral speed is half that of the flowing steam. This means that the revolu- 
 tions of a simple turbine must be exceedingly high or the diameter of the 
 rotor very large ; and for large power both. By methods adopted by Mr. 
 Parsons and others, the rate of revolution has been brought down to reason- 
 able limits for high-speed ships, while preserving a good efficiency ; but it 
 is still too high for such ships as are engaged in cargo carrying if they are 
 to have screws of decent efficiency and effectiveness. Mr. (now Sir Charles) 
 Parsons has met this difficulty, therefore, by fitting pinions to the shafts of a 
 pair of turbines, and geared them to a spur-wheel on the screw shafting. 
 
 The employment of a quick-revolution engine to drive the paddle shaft 
 by means of a wheel gearing was practised by the Butterley Company so far 
 
G MANUAL OF MARINE ENGINEERING. 
 
 back as 1823, but it was not followed till almost the end of the nineteenth 
 century, when it was revived by one or two firms for river steamers, in order 
 to use the quick-running triple-compound three-crank engine. 
 
 Wheel-gearing has found little or no favour in the eyes of the marine 
 engineer since the geared screw engines of the middle of the nineteenth 
 century were given up, and no doubt there will be much prejudice against 
 it now, although it may be fairly urged that to-day the teeth are double- 
 helical machine-cut, made of more suitable material, having no backlash, 
 and the pinion driving instead of being driven (v. fig. 45). Experience- 
 condemned the system in the past, and it is the successful experience of Mr. 
 Parsons which will revive it to-day. 
 
 With oil engines there seems to be, for practical reasons, a limit to the 
 diameter of the cylinder (20 inches at present), and all increases of power 
 is obtained by an increase in the number of cylinders. Moreover, since 
 they are generally single-acting, and act at most once in two, and generally 
 in four, strokes, their size for the power developed is large compared with 
 that of the steam engine. The first oil engine was exhibited by its inventor 
 at Cambridge 90 years ago. In 1817 Neepce proposed to use volatile oils 
 
 Fig. 3.— Paddle Steamer "Charlotte Dundas," 1802. 
 
 in his explosive engine for propelling ships, and an internal combustion 
 engine was used so far back as 1825 to drive a screw propeller in a boat on 
 the Thames. The system is even now only emerging from its infancy, and 
 the room for developments is extensive. In the submarine ship there have 
 been the greatest developments, as these engines are almost the only ones 
 possible. 
 
 Steam used expansively. — The earliest marine engines were naturally 
 near akin to those on shore, but it is very interesting and noteworthy that 
 the one in the " Charlotte Dundas " of 1802 was a horizontal double-acting 
 engine having a connecting-rod from the piston-rod end to the crank-pin, 
 and, therefore, much in advance of and differing from entirely the engine 
 of James Watt. The honour of designing the engine (see fig. 3) is due to- 
 William Symington, who patented the fitting of the connecting-rod, 1787, 
 Pickard having taken out his patent for the crank and connecting-rod of 
 the beam engine in 1780. 
 
 The " Comet," which was the first steamship in this country to earn 
 money by conveying passengers and goods between Glasgow and Greenock 
 
INTKODI'CTION — .MARINE STEAM ENGINES. 
 
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8 
 
 MANUAL OF MA1UNE ENGINEERING. 
 
 in 1812, had a modification of the beam engine of the type known later on 
 as a Side Lever (v. fig. 4). These ships, as the ones that followed them, had 
 engines using steam at or a little above the atmospheric pressure, and ex- 
 hausting into a jet condenser with an air pump, etc. Later on, after the screw 
 
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 propeller had been adopted, some small ships, and even a few large ones, in 
 the Royal Navy^had non-condensing engines using steam at 50 to 65 lbs. 
 pressure generated in cylindrical tubular boilers. Fig. 6 is a modern steamer 
 as now built for service on the Clyde estuary, and doing the work that the 
 
INTRODUCTION — MARINE STEAM ENGINES. 
 
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10 MANUAL OF MARINE ENGINEERING. 
 
 "Comet" established; she is, however, driven by three screws and turbines 
 instead of four paddles and a one-cylinder engine, and carries more passengers 
 in a trip than the " Comet" in a couple of months. 
 
 In the early marine engines, as in those on land, the cycle was a simple 
 one. Steam was admitted to one or more cylinders direct from the boiler 
 during about 70 to 85 per cent, of the stroke, when working at full power ; 
 at or near the end of the stroke it was allowed to escape to the condenser ; 
 the cylinder on that side of the piston remained in open connection with 
 the condenser during about 85 per cent, of the return stroke, and conse- 
 quently its surface was exposed to the cooling action of the comparatively 
 cold vapour remaining in it. The earliest'screw ships in the Navy had their 
 boiler safety valves loaded to 5 lbs. per square inch ; by 1851 the load had 
 been increased to 14 lbs., and two years later 20 lbs. was taken as the standard. 
 In 1855 several special ships were fitted with non-condensing engines of con- 
 siderable size supplied with steam from 50 to 65 lbs. pressure generated in 
 cylindrical boilers of the so-called Scotch type, but the ordinary ships with 
 condensing engines continued to work with steam generated in box boilers 
 at 20 lbs. pressure. In 1861, the new ironclads had boilers with safety 
 valves loaded to 25 lbs. per square inch, and four years after a further increase 
 was made to 30 lbs., which was the usual or standard pressure till the box 
 boiler ceased to be made. 
 
 The cut-off in the cylinders of these ships when running at full speed 
 was never less than 60 per cent, of the stroke, so that the maximum rate of 
 expansion was no more than 1*67 under these circumstances ; but with 
 the increase in pressure from 25 to 30 lbs. came the supply of expansion 
 valves, whereby a much earlier cut-off could be obtained with a corre- 
 sponding economy in fuel consumption. With such valves a rate of expan- 
 sion of 4*0 could be obtained, so that the terminal pressure at which the steam 
 exhausted to the condenser was under 10 lbs. absolute. In the mercantile 
 marine about 1870 the boiler pressure was increased to 50 lbs. for expansive 
 simple engines, and the cut-off at full speed was about 30 per cent, of the 
 stroke, giving a rate of expansion of 3*33 ; a few ships with these engines 
 had boilers loaded to 60 lbs., but in a general way 50 lbs. was the highest 
 pressure with simple engines. The pressure at exhaust was then about 
 15 lbs. absolute at full speed. 
 
 The compound engine in which the steam, after doing duty in the first 
 cylinder, exhausts into another, instead of delivering to the condenser, was 
 the invention of Hornblower, a Cornish mining engineer, in 1771, improved 
 by Wolff in 1804, and first used on shipboard by Randolph and Elder in the 
 screw steamer " Brandon " in 1854 with a boiler pressure of only 22 lbs. 
 This same enterprising firm supplied the first compound engines for H.M. 
 Navy in 1863, and fitted them in H.M.S. " Constance." It is worthy of note 
 that their designer was Professor Rankine, and they worked with a boiler 
 pressure of 32*5 lbs. per square inch, and that the rate of expansion at full 
 speed was about 5, and the referred mean pressure was 11*3 lbs. per square 
 inch, which is not bad, considering that the theoretical would be not more 
 than 22 lbs., or 5T5 per cent, efficiency. 
 
 It was not, however, till about 1870 that the compound engine was 
 accepted as suitable and desirable for marine purposes, but after experience 
 had proved its economy in consumption of steam, and that the expansive 
 
INTRODUCTION — MAU1NE STKAM ENGINES. 
 
 11 
 
 economy 
 engineers 
 
 engine using high-pressure steam developed mechanical troubles from which 
 it was free, it soon superseded all other engines, and was the accepted type 
 until it was found that, with the increase of boiler pressure from (><> lbs. to 
 100 lbs., their 
 
 gam in 
 was not what 
 had reason to 
 expect from such increases. 
 In 1881 the late Dr. 
 Kirk, whose name is as- 
 sociated with that of John 
 Elder, of the Randolph and 
 Elder firm, and himself a 
 most able and enteif>rising 
 engineer, developed the idea 
 of multiple cylinders, or, as 
 we would now say, of ex-> 
 panding by stages ; he built 
 a three-stage engine, where- 
 by the steam exhausted 
 from the first engine into 
 the second, and from the 
 second to the third engine. 
 Dr. Kirk had, however, in 
 1874 fitted an engine of this 
 kind in s.s. " Propontis," 
 with cylinders 23 to 41 
 inches, and 62 inches dia- 
 meter by 42-inch stroke. 
 Unfortunately, the water- 
 tube boilers of this ship 
 proved dangerous, and 
 caused the whole installa- 
 tion to be doubted. Looking 
 at the subject another way, 
 it may be said that he in- 
 troduced a third cylinder 
 between the ordinary high- 
 pressure and low - pressure 
 cylinders, so as to avoid the 
 big drop in pressure between 
 the high and low. But the 
 principle involved was one 
 of stage, whereby there was 
 a decrease in difference 
 between the temperature 
 during admission to, and 
 
 that at emission of steam from each cylinder. There was, however, a 
 mechanical gain as well, due to the decrease in initial and general loads 
 on the pistons of the compound systems as compared with those obtaining 
 in the expansive engines. The pressures of steam adopted by Dr. Kirk 
 
12 MANUAL OF MARINE ENGINEERING. 
 
 and by the author himself in the early triple-expansion engines was low 
 compared with those now used ; in fact, the pressure was little beyond that 
 with which the compound engine was then working on shipboard. Since 
 that time working pressures have gone on increasing till 180 lbs. was a common 
 practice for triples in the mercantile marine, and on the introduction of the 
 water-tube boiler into the Navy, steam of 250 lbs. pressure was and is used 
 by the triple- compound engine. In the mercantile marine now steam up 
 to 225 lbs. is generated in the cylindrical or tank boiler for use by quadruple 
 expansion engines with advantage ; these engines were introduced by two 
 or three of our leading manufacturers in 1885. The rate of expansion had 
 gone on increasing, till now, with quadruple-expansion engines it is common 
 practice to work at 16, while with the triple engine working with a boiler 
 pressure of 180 lbs. the rate of expansion will be 12 to 13 in the mercantile 
 marine, and about 11 to 12 in H.M. Navy at full power. 
 
 Until quite recently all attempts at economy in consumption of fuel 
 were in the direction of increase of pressure and increasing the number of 
 cylinders. The upper end of the steam expansion diagram was the field in 
 which further gains were looked for, and the dictum of the old chief engineers 
 that there was no economic gain in working with a vacuum over 24 inches 
 was accepted without a close enough examination of the foundations for 
 such a statement, hence few, if any, engineers turned their attention to the 
 possibilities of the gain to be got at the lower end (v. fig. 66). The adoption 
 of the turbine for marine propulsion by Mr. Parsons opened his eyes to this 
 by the great increase in efficiency of his machines when working with the high 
 vacuum so cheaply obtained with an unlimited supply of cooling water 
 and a really good air pump on shipboard ; and further, the general use 
 of feed heaters on shipboard has quite removed the bugbear of cold-feed 
 water for the boilers. Before 1865, with the old common jet condenser, 
 the vacuum was seldom more than 24 inches, whereas with the surface con- 
 denser, which soon became the common practice after that date, 28 inches 
 was not uncommon, and 27 inches easily maintained ; the Admiralty require- 
 ment was that the vacuum on trial should be within 3 inches of the barometer. 
 In passing, however, it may be mentioned that the pressure in the con- 
 denser is in no way subservient to that of the atmosphere, although those 
 registered by the common vacuum gauges, of course, are, for they really 
 only exhibit the difference between the pressure of the air and that in the 
 condenser. To-day gauges can be obtained which indicate the exact pressure 
 in a condenser, without regard to the atmospheric pressure, and should be 
 always used. ^ 
 
 It was not an unknown thing forty years ago to maintain, in a surface 
 condenser, a vacuum of 29 inches with high barometer (30'5 inches), but 
 it was only with one of exceptionally good design, and it, the pumps, and 
 everything in the best of order that so good a vacuum was got. To-day, 
 with the better design of both condenser and pumps, and the relay system 
 or other means for obtaining high efficiency of air pumps, 29 inches can 
 be maintained with comparative ease, consequently a rate of steam expansion 
 that was useless formerly may now be worked with advantage and with 
 considerable gains both in power and economy, especially with turbines. 
 
 The reciprocating engine has its limitations, one of which is the size and 
 nature of the valves for admitting and releasing the steam. It is practically 
 
INTRODUCTION MARINE STEAM ENGINES. 13 
 
 impossible to provide means in this direction to ensure high efficiency to the 
 low-pressure cylinder of a compound system ; wire drawing and clearance 
 losses are great, and the difference in pressure between the cylinder and 
 the condenser necessarily great compared with those of a turbine. Hence 
 the last addition to a marine engine, whereby steam efficiency is considerably 
 enhanced, is the low-pressure turbine taking its steam from the exhaust 
 pipe of the low-pressure cylinder, and expanding it from about 10 lbs. to 
 1 lb. absolute, or even less. It is claimed by Messrs. Denny that the increase 
 of power and an economy of fuel has been found as high as 15 per cent. 
 
 The compound turbine so much in use to-day in H.M. Navy and in express 
 steamships (v. fig. 6) of high speed is capable of a rate of expansion far beyond 
 that of a reciprocating engine ; in fact, there is no practical limit to it. But 
 in the early stages of the compound turbine the thermal efficiency is not so 
 high as is the triple compound engine ; it follows then that the maximum 
 efficiency will be obtained by the combined arrangement of a reciprocator 
 with a low-pressure turbine, taking the steam from the low-pressure cylinder 
 at something like 15 lbs. pressure absolute, which is about the terminal pres- 
 sure of such an engine working at 200 lbs. boiler pressure. 
 
 Propellers. — The " Charlotte Dundas " (fig. 3), the first ship to be pro- 
 pelled by steam in a practical way, had one paddle-wheel at the stern. The 
 " stern wheeler," as she is now called, remains as the surviving representative 
 of the paddle ship in the construction of steam craft for river purposes, 
 especially those working in the tropics, where she is the favourite, and appears 
 likely to continue as such. The first steamer in America, the " Claremont," 
 of 1807, and the " Comet " in this country in 1811, as the first British steamer 
 to be put to commercial purposes, had a pair of side wheels — that is, a wheel 
 on each side with an axle or shaft common to the two. It should, however, 
 be noted that the " Comet " had two pair of wheels (v. fig. 4) when first 
 tried, a system followed by Sir Edward Reed in 1874 for special reasons 
 when he designed the swing saloon steamship " Bessemer." The pair of side 
 wheels continued to be the practice for general purpose down to recent times, 
 and is still followed when such ships are built for the special services already 
 alluded to. The last of the ocean-going paddle ships was the " Scotia," 
 6,871 tons displacement, 362*5 feet long, having engines of 4,950 I.H.P., 
 driving a pair of wheels 40 feet diameter, weighing 156 tons. The largest 
 modern American paddle steamship is the " Priscilla," 424 feet long and 
 5,200 tons displacement; she has engines developing 9,345 I.H.P., which 
 drive her at a speed of over 19 knots per hour. Fig. 5 shows the paddle 
 steamer as now constructed in America for lake and river service to convey 
 passengers and parcels expeditiously. The largest modern British paddle 
 steamship (v. fig. 8) is the " Empress Queen," 360 feet long, 2,900 tons 
 displacement, attaining a speed on service of 2T7 knots with 11,440 I.H.P. 
 The fastest British paddle steamship is " La Marguerite," 330 feet long, 
 1,868 tons displacement, and having a speed of 22 - 3 knots. 
 
 Screw propellers, brought into practical use by Bennet, Woodcroft, 
 Francis P. Smith, and John Ericsson, in 1836 (v. fig. 9), are now the most 
 important instruments of propulsion, and, although much time, genius, 
 and money have been expended in exploiting screws of the most varied 
 forms, engineers have to-day settled down to the almost universal use of the 
 screw, such as used by Smith and his friends in 1840 on the s.s. " Archimedes " 
 
14 
 
 MANl'AL OF MAHINK ENGINEERING. 
 
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INTRODUCTION — MARINE STEAM ENGINES. 
 
 15 
 
 — that is, in the essentials — viz., making it a portion of a true helix and 
 of moderate diameter and acting surface. So far as shape is concerned, 
 the practice now is very like that of Robert Griffiths of 1805 — that is, the 
 blade is narrow at the tips compared with the breadth at middle, and the boss 
 spherical of considerable diameter. The prevailing form for express and 
 naval ships is, of course, a modification of the Griffiths, inasmuch as their 
 tips are somewhat fuller and rounded, so that the blade is the shape of the 
 longitudinal section of a domestic hen's egg ; and in some cases where the 
 diameter is small for the power, the blade is even circular. In fact, all of 
 them are such as would be formed by taking the Griffiths, and reduce the 
 diameter by rounding the tips until the maximum breadth is nearer the top. 
 The number of screws has increased from the one situated in a gap in the 
 deadwood of the ship to the four of the modern high-speed warship driven 
 
 Francis Pettit Smith, 1836. 
 
 Francis Pettit Smith, 1838. 
 
 John Ericsson, 1836. 
 
 Fig. 9. 
 
 by turbines. The twin screw, an obvious development, especially if required 
 to compete with the paddle-wheel in shallow waters, was first introduced 
 by the Rennies in 1854 for service on the Nile ; since then they have replaced 
 the single screw in all express and ocean-going high-speed steamers. With 
 such twin screws, not only is total immersion attained and the " feed " to them 
 unobstructed, but the safety of the ship considerably insured, inasmuch as 
 the liability to total breakdown is reduced by a-half, and a means of steering 
 is provided, should the rudder or its gear be disabled. Two screws, one 
 at the bow and another at the stern with a line of shafting common to them 
 have been used frequently for tugs and ferry boats with advantage, and 
 four screws similarly arranged have been used for the same services. 
 
 Three screws, one in the deadwood and one on each side, somewhat 
 ahead of the middle one, have been used in France, Germany, and Italy, and 
 the United States, but the system did not find favour in this country till 
 
16 MANUAL OF MARINE ENGINEERING. » 
 
 the turbine wa3 used as the prime mover. The development is a natura 
 one, where large power is required in a comparatively shallow vessel, and 
 also with the view of keeping the engine within moderate dimensions. Italy 
 was the first to give the system a practical trial in 1886 on the cruiser 
 " Tripoli " ; France followed on a large scale with the " Dupuy de Lome " 
 in 1890. In 1892 the United States of America adopted the system for 
 the " Columbia," of 7,375 tons displacement, 22'8 knots speed. In the same 
 year the German Government fitted the cruiser " Kaiserin Augusta," of 
 6,330 tons, and 22"5 knots, with three screws, while in 1896 Russia adopted 
 the system for the " Rossia," a large cruiser of 12,130 tons and 14,500 I.H.P. 
 In our own Navy, the " Amethyst," cruiser of 3,000 tons, having Parsons 
 turbines operating on three screws, was the first attempt, if the experiment 
 with H.M.S. " Meteor " in 1855 is excepted. To-day the rule in the British 
 Navy is to fit four screws* to all but the very small ships ; the newest cruisers 
 of 75,000 H.P. are so fitted, as are the new class of " Destroyers." In the 
 mercantile marine our largest and fastest mail steamers have four screws ; 
 the smaller ones, three when driven by turbines or combined reciprocators 
 and turbines. The newest and largest ship, s.s. " Olympic," has thus three 
 screws (v. figs. 43 and 44). With reciprocators only the twin screw is now 
 almost the universal rule for all express steamers, and for cargo steamers of 
 the largest size. 
 
 Multiple Screws. — In some ships of special design, or for special service, 
 a larger number of screws have been fitted, and their arrangement varied ; 
 for example, the saucer-shaped Russian warships " Popofiskas " had six 
 screws, and the Russian ice-breaker, " Ermack," has three screws at the 
 stern and one at the bow. 
 
 Mr. Parsons has also in some few cases fitted more than one screw on 
 each shaft, and more recently the German Naval Authorities have tried, 
 on the cruiser " Lubeck," eight screws, two on each of her four shafts ; the 
 results of this experiment are by no means satisfactory, being much inferior 
 to those of sister ships having twin screws driven by reciprocating engines. 
 
 Table I. contains examples of naval ships' engines typical of the 
 periods since the introduction of the screw propeller, and shows the progress 
 made in the use of steam expansively. 
 
 Progress in the use of steam expansively in the mercantile marine during 
 the past 50 years, as shown by the typical examples in Table II. 
 
 * Quite recently, however, there is a tendency in both the British and Foreign Navies 
 to revert to the twin-screws for the smaller class of ship, and even in second-class cruisers, 
 with a complete turbine for each propeller, and since fcae introduction of gearing the 
 tendency is to twin-screws for all kinds of ships. 
 
INTRODUCTION — MARINK STKAM KNGINKS. 
 
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VALUE OF TRIAL TRIPS. 19 
 
 CHAPTER II. 
 
 RESISTANCE OF SHIPS AND INDICATED HORSE-POWER 
 NECESSARY FOR SPEED. 
 
 Although, strictly speaking, it is not the province of the engineer to deter- 
 mine the power necessary to drive a ship at a certain speed, but rather that 
 of the naval architect, still it is a point of great importance to the engineer, 
 and one with the investigations of which he should be fully acquainted. 
 Circumstances sometimes require, indeed, that the engineer shall name the 
 power, as the naval architect may submit that, inasmuch as he is unaware of 
 the efficiency of the particular engine to be supplied, he cannot say what 
 indicated horse-power will be necessary, but only what effective horse-power. 
 Moreover, the subject is one possessing great interest at all times, and some- 
 times of the utmost importance to the engineer, as the deficiency of speed 
 obtained at the measured mile from that anticipated may be attributed to 
 the inefficiency of the engine and propeller. This charge may be, and often 
 has been, proved to be true ; but, on the other hand, it may be without 
 foundation, the blame really belonging to the designer, who has given the 
 ship lines unsuited to the speed. 
 
 Value of Trial Trips. — Trial trips are now conducted, both in the mer- 
 cantile marine and the Royal Navy, with more care and interest than obtained 
 formerly ; and it is not sufficient to prove at the measured mile merely 
 that the ship has done the speed expected, or that the engines have developed 
 the power for which they were designed. Both engineers and naval archi- 
 tects desire to determine whether the speed has been obtained with the 
 minimum of power, and the engineer can satisfy himself on a most important 
 point — viz., the efficiency of the propeller, and, to some extent, the efficiency 
 of the machinery, while the owner, if it be a private ship, is enabled to judge 
 whether he is paying for what he calls " big horses " or " little horses." 
 Another point (and one most important to the owner) which, to some extent, 
 is determined on a trial trip is — at what expenditure of fuel a ton of dis- 
 placement is carried over a mile. It is not an unknown thing to find that 
 the engine which burns least fuel per I.H.P., does not compare so favourably 
 with others when measured by this latter standard. The apparent con- 
 tradiction here is not very difficult to understand when fully looked into ; 
 it may be, perhaps, best comprehended by taking extreme cases. Suppose 
 the blades of the screw are set so as to have no pitch ; the engine will work, 
 develop a certain power necessary to overcome its own resistance and that 
 of the screw, but it will not drive the ship an inch ; the coal consumption 
 per I.H.P. will probably be somewhat heavier than that of the same engine 
 when working with half its load, but still may be light. Now place the 
 blades fore and aft, so that the pitch is infinity, and although there may be 
 
20 MANUAL OF MARINE ENGINEERING. 
 
 now a large development of power, there will be no appreciable speed — 
 theoretically, none at all. In both these extreme cases the consumption 
 per I.H.P. may be very satisfactory, but the satisfaction would not be 
 experienced by the owner. It is manifest, then, that between these two- 
 extreme limits of pitch there is some value and one position of blade which 
 will give the best result, so far as economy of fuel for load propelled is con- 
 cerned. Not only is the pitch of propeller an important function in all 
 calculations relating to the speed of ships, but the diameter has a very im- 
 portant bearing also on the subject, and more than was generally thought 
 previous to the remarkable trials of H.M.S. " Iris." 
 
 The Resistance of a Ship passing through water is not easily determined- 
 beforehand, as it may vary from more than one cause, and in a way often 
 unanticipated, as has been seen during the trials of the very fast torpedo- 
 boats and destroyers. The investigations of the late Dr. Froude on this 
 subject have shown that the older theories were sometimes erroneous, and 
 the old-established formulae unreliable ; and perhaps the best source of. 
 information on the intricacies of this somewhat complex subject is to be 
 found in the many able papers read by him, and others since his day, before 
 the Institution of Naval Architects and other learned societies. 
 
 When the screw or paddle first commences to revolve, the ship makes no 
 headway, and it is only after some seconds have elapsed that motion is 
 observable. The engine power has, during that period, been employed in 
 overcoming the resistance to motion which all heavy bodies possess, and 
 which is called the vis inertia. When the engine is stopped at the end of the 
 voyage, the ship will continue to move, and come gradually to rest, unless 
 otherwise, retarded by the reversal of the engine or by check ropes. The 
 ship is then said to have " way on her," a phrase which, in scientific language, 
 means that she possesses stored-up energy, called momentum, which is given 
 out, when the engine stops, in overcoming the resistance of the water to 
 the passage of the ship through it. This energy was stored up at starting 
 of overcoming the inertia, and remains stored until there is any retardation 
 in velocity. In this way the weight of the ship helps to preserve a uniformity 
 of motion, as that of a flywheel does to an engine, and, therefore, it is im- 
 portant that tug-boats should have weight as well as power, to prevent 
 towing in the jerky fashion so often observable. When the vis inertia has. 
 been overcome, the power of the engine is directed to overcoming the resist- 
 ance of the water, and wind if there be any, and in accelerating the velocity 
 of the ship ; as the speed increases, the resistance much more increases, 
 until the surplus power available for acceleration becomes nil, and the whole 
 engine power is absorbed in overcoming the internal resistances, or those- 
 belonging to the engine itself and the propeller, and the external, or that 
 of the ship. 
 
 The Resistance of the Water is Twofold. — First, the ship in moving 
 forward has to displace a certain mass of tvater of the same weight as itself, 
 and the water has to fill in the void which would otherwise be left by the 
 ship. The work done here is measurable by the amount of water, and since 
 it is equal to the displacement of the ship, displacement becomes a factor in 
 the calculations of resistance. But to effect this displacing and leplacing of 
 water with the least amount of energy, it is necessary to do it gently — to 
 set the particles of water gradually in motion at the bow, and let them come 
 
KESIDUAL RESISTANCE. 21 
 
 gradually to rest at the stern. If it is not clone gently, and the water is 
 rudely separated, a wave is formed on either side, showing that energy has 
 been spent in raising the water of this wave above its normal level. Although 
 every ship, however well designed to suit the intended speed, causes these 
 waves of displacement, it is the object of the naval architect to reduce their 
 magnitude as much as possible. 
 
 The chief cause of resistance to the passage of a ship through the water 
 is, however, the friction between the surface of the immersed portion and 
 the water. Resistance from this cause is generally spoken of as skin resistance, 
 and is in well-formed ships much greater than the resistance due to other 
 causes. However fine a ship may be, there is, of necessity, a certain area 
 of skin exposed to the water, and though the displacement be very small 
 indeed, and the section transverse to the direction of motion reduced to a 
 minimum, it is found that a considerable amount of power is required to 
 propel the ship through the water, and that, roughly, the power is propor- 
 tional to the wetted surface at the same speeds. It is from this cause that 
 the older rules for speed, involving only displacement, or area of midship 
 section, together with speed as variables, are found to be so misleading. 
 
 Residual Resistance is the term generally used to express the sum of 
 all other resistances to be overcome in propelling a ship through the water, 
 and includes that due to wave-making, eddy-making, etc. In tank experi- 
 ments this is differentiated from skin resistance, and ascertained with accuracy, 
 but it may be calculated with a very fair approximation to it by a formula 
 arrived at by Mr. D. W. Taylor, U.S.A., and published by him there. It 
 is as follows, viz. : — 
 
 Residuary resistance™ lb, -™->™ 
 
 where b is the block coefficient ; 
 
 D is the displacement in tons ; 
 
 V is the speed in knots ; 
 
 L is the length on the water-line in feet. 
 
 V 2 
 The formula is applicable only to speeds for which y- is less than 1'2. As 
 
 will be seen, it conforms to the law of comparison, and, compared with other 
 formulae, gives extremely good results for all classes and sizes of ships. It 
 has been applied to a large number of different ships, and the general results 
 of the application are satisfactory. 
 
 Most of the examples taken for comparison have been warships, where 
 it was found in applying the formula that a slight modification gave better 
 results. This modification consists in taking the length as the extreme 
 length of immersed vessel in place of that on the water-line, and in calculating 
 the block coefficient on the extreme length instead of that between per- 
 pendiculars, as usually taken. For merchant ships the two lengths are 
 generally the same, but for warships the difference in the two is appreciable. 
 Mr. A. W. Johns found that for vessels in which the block coefficient varies 
 from *6 to -65, the calculated results are generally correct for a speed such 
 
 V 2 
 
 that y - i s about '85. Below this speed, the results are generally about 
 
22 MANUAL OF MAUINE ENGINEERING. 
 
 7h per cent, larger, whilst above this speed the results are somewhat smaller 
 than the experimental results. 
 
 For vessels in which the block coefficient varies from - 5 to "55, the calcu- 
 
 V 2 
 
 lated results are generally correct for a speed given by-y- equal to 1. Above 
 
 this speed, the results are slightly smaller, whilst below they are generally 
 about 10 per cent, greater than the experimental results. 
 
 The resistance of the ordinary ship roughly varies as the square of the 
 speed, so long as the highest speed does not exceed that for which the ship 
 is suited, and so long as the wave formation is not so great as to cause 
 considerable variation in the trim of the ship. When such conditions prevail, 
 as they do largely in modern very high-speed ships of small size as are 
 Destroyers and Scouts, the variation in resistance does not follow so simple 
 a rule, and, moreover, the fluctuation is apparently capricious, as may be 
 seen by examining the figures given by Sir William White for destroyers 
 (q.v.). Assuming, however, as we may in the case of ordinary ships, that 
 E varies as S 2 , then to complete the expression, which will give a definite 
 value to the resistance of a given ship, it was necessary to multiply the 
 product of the above two variables by a quantity found from practice ; and 
 if the law were absolutely correct, this quantity should have a fixed value, 
 whatever the size and form of the ship, and would be a " constant " multiplier 
 for all cases in choosing values for C and K. Actual values for them can 
 be found in the tables of performances of ships on trial trips. 
 
 Coefficient of Fineness. — To determine the form of a ship, as to whether 
 it is " fine," " fairly fine," or " bluff," it is usual to compare the displace- 
 ment in cubic feet with the capacity of a box of the same length and breadth, 
 and of depth equal to the draught of water ; the coefficient by which the 
 capacity of such a box must be multiplied to give the displacement being 
 called the coefficient of fineness. Thus 
 
 Coefficient of fineness = T =; rr- T , 
 
 L x B x W 
 
 D being the displacement in tons of 35 cubic feet of sea-water to the ton ; 
 L the length between perpendiculars in feet ; B the extreme breadth of 
 beam in feet ; and W the mean draught of water in feet, less the depth of 
 the keel. Strictly speaking, the length should be measured from the stem 
 to aft part of body-post on the water-line, instead of to aft part of rudder- 
 post ; but as this dimension is not easy to ascertain without referring to 
 the plans, and the calculation is made for the sake of comparison, rather 
 than as an accurate computation, no inconvenience will arise from this, so 
 long as all the ships under comparison are measured in the same way. 
 
 It will be easily seen that the above coefficient only expresses a relation 
 between the cubic contents of the immersed portion of the ship and a box 
 of the same dimension, and gives no certain clue to the fineness of the water- 
 lines, which is really what is wanted for consideration iu dealing with the 
 question of power for speed. 
 
 Two ships may have the same dimensions and the same displacement, 
 and, consequently, the same coefficient of fineness, and yet one may have 
 bluff lines and the other fine — the difference arising from the latter having 
 a flat floor, and the former a hi^h rise of floor. To take an extreme case,. 
 
THR WKTTED SKIN OF A SHIP. 23 
 
 the fine ship might have a rectangular midship section, and the bluff one a 
 triangular one ; and if the " coefficient of fineness " was 0*5, the bluff ship 
 would have rectangular water-planes, while those of the fine ship would be 
 two triangles base to base. 
 
 Now, if a coefficient be obtained by comparing the displacement with 
 the volume of a prism, whose base is the midship section, and height the 
 length of the ship, it will indicate the general fineness of water-lines, and 
 form a guide in the choice of the constants for speed calculations. 
 
 D x 35 
 
 The Prismatic or Coefficient of water-lines = 
 
 area of immersed mid-section x L 
 
 The Skin Resistance is, in all classes of ships, the most serious, although 
 in destroyers at full speed (say 30 knots) it only amounts to 45 per cent, 
 of the total, as compared with 80 per cent, at 12 knots. With cruisers it 
 is as much as 80 per cent, at 20 knots, and more than 70 per cent, at the full 
 speed of 23 knots, while at 12 knots it is 90 per cent. 
 
 From experiments made by Dr. Froude with varnished surfaces, such as 
 given by the modern spirit mixed anti-fouling compositions applied to ship's 
 bottoms to discover «, the index of V, the following was deduced : — 
 
 Resistance = j X A X V n . 
 
 j is a factor which varies with the length of the ship. 
 A the area exposed to water rubbing. 
 V the velocity in feet per second. 
 
 n had a value 2 close to the bow, T85 at 20 feet from it, and 1*83 at 
 50 feet. The average value throughout may be taken at T83. 
 
 At a speed of 10 feet per second the mean resistance was found to amount 
 to 0*25 pound per square foot. 
 
 Taking as an example an express steamer, 400 feet long, at a speed of 
 20 knots, what will be the resistance per 100 square feet of wetted skin ? 
 Here j = '00886, and V = 34. 
 
 R = 0-00886 x 100 X 34 1 8S = 563 lbs. 
 
 The power required to draw this surface through the water at the velocity 
 will be 
 
 Power = 563 x 34 X 60 = 1,148,520 foot-lbs., or 34*8 H.P. 
 
 This, of course, is the net horse-power, and not that developed by the engine 
 driving the propeller. If, however, the efficiency of machinery, propeller, 
 etc., were, say, 0'7, then 
 
 The gross I.H.P. = 34'8 -^ 0'7 = 49-7. 
 
 The Wetted Skin of a Ship should be accurately measured, but this is 
 a somewhat long and troublesome business, and, moreover, necessitates 
 having the full lines of the ship, which, as a rule, are not available in the 
 initial stages of design, and certainly not accessible to the engineer as a rule. 
 There are, however, methods of obtaining the area with sufficient accuracy 
 for the purpose of estimating the indicated horse-power necessary to drive 
 a ship at a required speed, and certainly is accuracy sufficiently close if the 
 
24 MANUAL OF MARINE ENGINEERING. 
 
 allowances have been obtained from practice with wetted skins calculated 
 by the same methods. 
 
 Kirk's Analysis is a system introduced by the late Dr. Alexander Kirk, 
 whereby the qualities of ships can be compared and incidentally the fitness 
 of a proposed ship for the speed required. By the means he provided the 
 wetted skin is found, and for ships with a high rise of floor, as were the rule 
 at the time the area so calculated was within 3 per cent, of the actual, and 
 often almost identical with it. With modern ships the error is often as much 
 as 5 per cent., and even with ships having deep bilge keels and large fins 
 for the side screws the error is 3 per cent. 
 
 Mumford's method of calculating wetted skin gives fairly accurate results 
 with ships of normal form and proportions, but with shallow draft ships 
 and flat bottoms the actual surface is somewhat in excess of that given by 
 his rule, which is as follows : — 
 
 Wetted skin = L (1-7 d + b B). 
 
 L is the length between perpendiculars, d is the depth of immersed midship 
 section, B is the greatest beam, and b is the block coefficient of displacement. 
 Seaton's Modification of Mumford's method is as follows : — 
 
 Wetted skin = (c X d X L) -\ . 
 
 a 
 
 Where c is a coefficient = 2 X area immersed mid-section -fBxrf. For 
 flat bottom shallow-draft ships c is 2*0, while for ships with a high rise of 
 floor, as usual in yachts and fast-sailing ships, c is 1*6. For ordinary ships 
 with a draught of water not less than one-quarter the beam c is 1*8.* 
 
 D is the displacement in tons, and d the mean moulded draft of water. 
 
 Seaton's Method for sea-going ships, whose draft of water is more than 
 one-quarter the beam permits of a ready computation of wetted skin with 
 the same in formation ; here L is the length, B is the beam, and d the 
 moulded draft. 
 
 (1) K = L -=- (055 B+d). 
 
 (2) F = 42 4/K. 
 
 (3) Wetted skin = F X D s — that is = 42 ^K X Di 
 
 The following are the values of F for variations in K : — 
 
 When K is 4, the value of F is 59-4 
 ,, o, ,, „ o2*9 
 
 „ 6, „ „ 65-7 
 
 ,, 7, „ „ 68-3 
 
 When K is 8, the value of F is 70-6 
 Q 72-8 
 
 10, „ „ 74-7 
 
 11, „ „ 76-4 
 
 The form suitable for a required speed is gauged by the coefficient of 
 fineness of water lines, called usually the prismatic coefficient, as already 
 stated; it also may be obtained from the block coefficient by dividing it by 
 the coefficient of mid-ship section, when knowing the area of midship section. 
 
 _ . . „ . Displacement X 35 
 
 Prismatic coefficient = -j ^r- — — — -. -r • 
 
 Area mid-section X lengtn 
 
 This may be called the criterion of form. 
 
 * To ships with multiple screws c is increased by 3i per cent., so that the multiplier 
 is 207 instead of 20. 
 
DETERMINING THE POWER. 25 
 
 The author has devised a rule for guidance in designing a ship, as also 
 to provide a means whereby engineers may avoid trying to do impossible, 
 or, at least, non-economic things in the way of forcing a ship beyond her 
 capacity for speed. 
 
 Seaton's Rule for Limitation of Speed is as follows : — 
 
 Suitable prism coefficient F = 04 £JL -»- £/S. 
 
 When L is the length of ship in feet, and S the speed in knots, being the 
 highest at which the ship can be driven without an excessive expenditure 
 of power. That is, with a length of ship L, 
 
 The maximum economic speed S = (04 £/L -=- F)i 
 
 In a general way the resistance will vary very closely with the square of 
 the speed with ordinary ships, which are not driven at a speed higher than 
 that given by this rule. 
 
 The least length of ship for a given speed and coefficient of fineness can 
 also be ascertained, thus : — 
 
 Minimum length L = ( — j • 
 
 Table V. gives the prism coefficient suitable to the length of a particular 
 ship for a given speed ; that is to say, the actual displacement of the ship 
 of a given length and for a certain speed, should not be greater than that 
 given by multiplying the displacement of a prism of the same section as 
 the midship section of the ship and the same length by the factor given. 
 
 For example, — A ship 300 feet long, for a speed of 15 knots, should have 
 a displacement not greater than *674 of the prism displacement — that is to 
 say, she must have a coefficient of fineness of waterlines of - 674. 
 
 To determine the power necessary to drive a ship at the required speed, 
 the facts already stated must be ascertained — \jz., general dimensions, dis- 
 placement, area of midship section, and the wetted skin. The first investi- 
 gation must be to discover the highest speed possible under these conditions, 
 and that is not less than that required. 
 
 Having done this, a calculation should be made of the maximum total 
 resistance R in pounds. The efficiency of the machinery and propeller 
 must also be known, and if the general efficiency is E, and S the speed in 
 feet per minute — 
 
 Gross indicated horse-power = =;XSv 33,000. 
 
 The total resistance is made up of three components : — 
 
 (1) That due to the skin friction, called frictional resistance. 
 
 (2) That arising from the making of waves and eddies, called residual 
 resistance. 
 
 (3) That due to the action of the propeller on the hull, called the aug- 
 mented resistance. 
 
 In the case of a paddle steamer the velocity given to the water by the 
 wheels is higher than that of the water flowing past it, and increases the skin 
 friction both before the wheels when the water is flowing into, and said to 
 be feeding the race, and abaft the wheels in the race. But with a screw 
 
26 
 
 MANUAL OF MARINE ENGINEERING. 
 
 steamer there is also the increased velocity caused by the feed, but a greater 
 loss is due to the decrease in pressure at the stern, owing to the action of the 
 screw pushing the water away. So great is this in bluff ships that the water 
 flows into the space behind the stern on each side of the race, and so causes 
 an eddy stream to follow the ship. 
 
 The following table has been calculated by Mr. Johns, of R.C.N. Con- 
 structors, as applicable to all modern ships with clean, fresh-painted bottoms, 
 and may be used for estimating the net horse-power necessary for over- 
 coming the skin resistance. The horse-power to overcome the residual 
 resistance can be calculated by means of Taylor's formula (p. 21). The 
 two results added together will give the total net horse-power called E.H.P. 
 If the propulsive efficiency is 06, then — 
 
 I.H.P. = E.H.P. - 0-6. 
 
 TABLE III. — Coefficients for Computing Effective Horse-power 
 
 REQUIRED TO OVERCOME SKIN FRICTION BASED ON Mr. FrOUDE's CON- 
 STANTS. AS GIVEN BY M.R. A. W. JOHNS. 
 
 If S is the wetted surface in square feet, then 
 
 E.H.P. = / . S, where / has the values given below. 
 
 
 
 
 
 Length of Ship in Feet. 
 
 
 
 
 Speed in 
 Knots. 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 100 
 
 150 
 
 200 
 
 250 
 
 300 
 
 350 
 
 400 
 
 450 
 
 500 
 
 25, . 
 
 •2516 
 
 •2477 
 
 •2458 
 
 •2444 
 
 •2434 
 
 •2415 
 
 •2415 
 
 •2407 
 
 •2399 
 
 24, . 
 
 •2242 
 
 •2207 
 
 •2190 
 
 •2178 
 
 •2169 
 
 •2160 
 
 •2152 
 
 •2145 
 
 •2138 
 
 23, . 
 
 •1988 
 
 •1957 
 
 •1942 
 
 •1931 
 
 •1923 
 
 •1916 
 
 •1908 
 
 •1902 
 
 •1895 
 
 22, . 
 
 •1753 
 
 •1726 
 
 •1713 
 
 •1703 
 
 •1696 
 
 •1690 
 
 •1683 
 
 •1677 
 
 •1672 
 
 21, . 
 
 •1537 
 
 •1514 
 
 •1502 
 
 •1494 
 
 •1487 
 
 •1481 
 
 •1476 
 
 •1471 
 
 •1466 
 
 20, . 
 
 •1340 
 
 •1319 
 
 •1308 
 
 •1301 
 
 •1296 
 
 •1291 
 
 •1286 
 
 •1281 
 
 •1277 
 
 19, . 
 
 •1159 
 
 •1141 
 
 •1132 
 
 •1126 
 
 •1121 
 
 •1171 
 
 •1112 
 
 •1108 
 
 •1105 
 
 18, . 
 
 •0995 
 
 •0979 
 
 •0972 
 
 .0966 
 
 •0962 
 
 •0958 
 
 •0955 
 
 •0951 
 
 •0948 
 
 17, . 
 
 •0846 
 
 •0833 
 
 •0827 
 
 •0822 
 
 •0819 
 
 •0815 
 
 •0812 
 
 •0810 
 
 •0807 
 
 16, . 
 
 •0713 
 
 •0702 
 
 •0697 
 
 •0693 
 
 •0690 
 
 •0687 
 
 •0685 
 
 •0682 
 
 •0680 
 
 15, . 
 
 •0594 
 
 •0585 
 
 •0580 
 
 •0577 
 
 •0575 
 
 •0573 
 
 •0570 
 
 •0568 
 
 •0567 
 
 14, . 
 
 •0489 
 
 •0481 
 
 •0478 
 
 •0475 
 
 •0473 
 
 •0471 
 
 •0469 
 
 •0468 
 
 •0466 
 
 13, . 
 
 •0397 
 
 •0390 
 
 •0387 
 
 •0385 
 
 •0384 
 
 •0382 
 
 •0381 
 
 •0379 
 
 •0378 
 
 12, . 
 
 ; 
 
 •0315 
 
 •0312 
 
 •0309 
 
 •0308 
 
 •0307 
 
 •0305 
 
 •0304 
 
 •0303 
 
 •0302 
 
 In the above table skin friction is taken as varying as V 1 ' 825 . 
 In Table IV. are the values of C in the old formula I.H.P. = 
 
 D« 
 
 C 
 
 for ships whose length varies from 100 feet to 900 feet, and the proposed 
 speed from 10 knots to 28 knots, and designed with a form suitable for the 
 speed — that is, the fineness of the water-lines of the ship is such that the 
 prismatic coefficient of displacement will be not greater than given by the 
 formula 0-4 4/L- VS. 
 
 If the ship is finer than determined by this criterion, the value of C may 
 be increased somewhat ; also in the case of a ship having engines of high 
 olficiency, such as possessed by most turbines, and high-class reciprocators 
 
VALUKS OF COEFFICIENT. 
 
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DETERMINING THE POWER. 29 
 
 having all the pumps disconnected from the main engines, and with efficient 
 propellers C, may be somewhat higher. In the latter case the increase will 
 be generally about 5 to 7 k per cent. If, on the other hand, the efficiency 
 of the machinery is from any cause low, as it used to be with the horizontal 
 engines, and even with some of the vertical ones with air, circulating and 
 feed pumps driven by the main engine ; and if the efficiency of the screw 
 is for some reason low, then the values given in the table are rather too 
 high. 
 
 Further, it may not be assumed that such coefficients will be developed 
 from the trials of such ships designed for and fitted with engines to drive 
 them at high speeds when running at low ; for example, it is a common 
 experience to find the highest value at a speed 10 to 20 per cent, below the 
 highest trial speed, and a decrease in value with the decrease of speed below ; 
 this is due to the fall in efficiency of the engines and propeller, both being 
 too large for the power developed ; as a matter of fact, the value of C is 
 a measure of the general efficiency of a ship. 
 
 However, the calculations of resistance and net horse-power is scarcely 
 the province of the engineer, and even the naval architect has found that 
 experiments with models of ships in a tank are the most reliable way of 
 ascertaining resistance and power. The Admiralty have employed a tank 
 and trained staff of experimenters for more than 30 years, first under the 
 guidance of the late Dr. W. Froude, since under that of his gifted son, Dr. 
 R. E. Froude. Until recently tanks were a luxury enjoyed only by a few 
 large wealthy shipbuilding companies ; thanks, however, to the munificence 
 of the eminent engineer, Sir A. F. Yarrow, there is now, at Bushey, a tank 
 equipped with the very best apparatus open to all who desire to have 
 experiments made with the models of proposed ships. 
 
 Tank Experiments with models of ships are very interesting, and of great 
 importance to the builders of vessels out of the common order of things 
 as to form and speed, as it is only by such means that their exact resistance 
 under varying conditions can be computed with such accuracy and reliance 
 as to permit of an exact provision being made of the power for the propulsion 
 of the ships such models represent. In this way the designers of cruisers, 
 scouts, destroyers, etc., whose form is uncommon and speed high, can deter- 
 mine the horse-power necessary for them, as also that for the very high-speed 
 express steamers now required for service on channels and oceans. 
 
 Dr. William Froude's experiments with H.M.S. " Greyhound " and her 
 model led to his establishing the laws which govern the true relation between 
 ships generally and their models, as also those between one ship and another 
 ship whose forms are similar but their dimensions different. Dr. R. E. 
 Froude has for many years followed on with the work begun by his father, 
 and from time to time has published his investigations and their results by 
 reading papers at the Annual Meetings of the Institution of Naval Architects, 
 in whose transactions they may be found recorded, and read with advantage. 
 The fuller consideration of the subject, however, is one outside the scope of 
 this work, except to say that Dr. R. E. Froude in this country, and Mr. 
 Taylor in America, have developed methods whereby screw propellers may be 
 examined and tested by their models, and their efficiency measured, not only 
 per se, but when working at the stern of a model of the ship for which the screw 
 itself is intended. By such experiments the effect on the ship of the screw 
 
30 MANUAL OF MARINE ENGINEERING. 
 
 working astern of it is also ascertained — that is, the augmented resistance due 
 to the screw. 
 
 The tank is a canal wide enough, generally about 20 feet, and deep enough 
 for such models as are used to pass through it without abnormal resistance ; 
 the model itself is made of paraffin wax to a suitable scale, and towed by 
 mechanical means at a speed given by the following formula, when L and S are 
 the length and speed of the proposed ship, and I and s that of the model : — 
 
 Then speed of model s = a/^ X S. 
 
 If the model is made to the scale of a quarter of an inch to the foot, it 
 will be one forty-eighth of the length of the ship, and consequently the speed 
 practically is one-seventh that of the ship itself. The towing apparatus is on 
 a travelling platform athwart the tank, which is caused to move at the speed 
 required by electrical driving gear. The tension on the tow rope is carefully 
 gauged and registered automatically, and gives the resistance of the model 
 when free from the screw. The screw is then fixed to an apparatus on the 
 same platform in rear of the model, and submerged so as to come into the 
 exact position relative to the model that the real screw would be to the 
 real ship. It is caused to revolve at the rate of revolution due to the speed 
 of the ship and designed slip, but without propelling or even touching the 
 model ; its thrust is carefully measured, and the torque or power necessary 
 to turn it also noted. The tension on the tow rope under these new conditions 
 is also recorded, and compared with the preliminary records. The model is 
 also tried in the same way at lower speeds, so that it has progressive trials 
 similar to those the ship will or may have. By comparing the thrust with 
 the torque at each speed, the screw's efficiency is ascertained, and can be 
 plotted on a curve ; by doing the same by the tension of the tow line and 
 torque, the general efficiency of the ship can be compared in the same way, 
 and by referring to the tests without the screw the augment of resistance 
 due to the screw can be determined. In the case of the tension and thrust 
 the speed in feet per second or minute is used as a multiple, while in that 
 of the torque 2 ir X revolutions in the same time is the multiplier to give 
 the power usefully employed and that developed. The thrust multiplied 
 by the speed is the measure of the screw as a " pusher " ; the tow rope tension 
 multiplied by the same speed is the useful work done, and, therefore, the 
 measure of the screw as a propeller. Let T r be the tow rope tension, T the 
 screw thrust, and t the torque, R the revolutions, and s the speed in feet 
 for time unit. Then 
 
 Efficiency of screw 
 
 Efficiency of propulsion 
 
 For engineers' use there are several rules, which may be employed with 
 advantage, and which will give the indicated horse-power under normal 
 conditions with a fair amount of accuracy. 
 
 (1) Professor Rankine's Rule may be mentioned, although its employment 
 is restricted ; it is as follows : — 
 
 T X* 
 
 6-28 R x t 
 
 X ■)* /\ S _L y 
 
 T^7 or T 
 
PROFESSOR RANKINE'S RULE. 31 
 
 Rule I. — Given the intended speed of a ship in knots ; to find the least 
 length of the after-body necessary, in order that the resistance may not 
 increase faster than the square of the speed : take three-eighths of the square 
 of the speed in knots for the length in feet. To fulfil the same condition, 
 the fore-body should not be shorter than the length of the after-body given 
 by the preceding rule, and may with advantage be one and a half times 
 as long. 
 
 Rule II. — To find the greatest speed in knots suited to a given length 
 of after-body in feet, take the square root of two and two-third times that 
 length. 
 
 Rule III. — When the speed does not exceed the limit given by Rule II., 
 to find the probable resistance in lbs. : measure the mean immersed girth 
 of the ship on her body plan ; multiply it by her length on the water-line ; 
 then multiply by 1 + 4 (mean square of sines of angles of obliquity of stream 
 lines). The product is called the augmented surface. Then multiply the 
 augmented surface in square feet by the square of the speed in knots, and 
 bv a constant coefficient ; the product will be the probable resistance in 
 lbs. 
 
 Coefficient for clean painted iron vessels, . . 0*01 
 
 ,, ., coppered vessels, . . 0*009 to 0*008 
 
 ,, moderately rough iron vessels, . 0*01 1 and upwards. 
 
 Rule Ilia. — For an approximate value of the resistance in well-designed 
 steamers, with .lean painted bottoms, multiply the square of the speed 
 in knots by the square of the cube root of the displacement in tons. For 
 different types of steamers the resistance ranges from 0*8 to 1*5 of that 
 given by the preceding calculation. 
 
 Rule IV. — To estimate the net or effective horse-power expended in pro- 
 pelling the vessel, multiply the resistance by the speed in knots, and divide 
 by 326. 
 
 Rule IVa. — To estimate the gross or indicated horse-power required, 
 divide the same product by 326, and by the combined efficiency of engine 
 and propeller In ordinary cases that efficiency is from 0*6 to 0'625 (Rankine, 
 Rules and Tables). Marine engines to-day have a combined efficiency of 
 0*65 to 0*70, and some even higher. 
 
 Although the method here proposed has been found to give much more 
 accurate and reliable results than those obtained by the older plans, it is 
 open in practice to two very strong objections. First, it is necessary to 
 have an accurate plan of the ship from which to measure the dimensions 
 required ; and second, it is difficult in actual practice to measure accurately 
 the angles of obliquity of stream lines, and the calculation requires more 
 time than can generally be devoted to the purpose. Often the horse-power 
 requisite to drive a ship at a certain speed must be calculated at the time 
 the lines are being got out, and it would be too late to wait for a plan of 
 the ship before getting some idea of the power. Again, the size and fineness 
 of a ship cannot be finally decided upon until the weight of machinery is 
 roughly known ; and as this will depend on the power, it is necessary to 
 approximate to it on very rough and ready information, for which rough 
 and ready rules are more suitable than the more refined ones. Hence, the 
 rules based on immersed midship section and displacement could be con- 
 
32 
 
 MANUAL OF MARINE ENGINEERING. 
 
 veniently used to obtain that approximation, and the power calculated 
 accurately from the augmented surface afterwards. 
 
 Dr. Kirk's Analysis. — A method of analysing the forms of ships, and 
 calculating the Indicated Horse-Power, was devised by the late Dr. A. C. 
 Kirk, of Glasgow, and met with much favour on all sides. It is often 
 used by shipbuilders on the Clyde and elsewhere for comparing the results 
 obtained from steamers with those obtained from others, and likewise to 
 judge of the form and dimensions of a proposed steamer for a certain speed 
 and power. 
 
 The general idea proposed by him is to reduce all ships to so definite 
 and simple a form that they may be easily compared ; and the magnitude of 
 certain features of this form shall determine the suitability of the ship for 
 speed, etc. As rectangles and triangles are the simplest forms of figure, and 
 more easily compared than surfaces enclosed by curves, so the form chosen 
 by him is bounded by triangles and rectangles. 
 
 The form consists of a middle-body, which is a rectangular parallelepiped, 
 and the fore-body and after-body prisms having isosceles triangle for bases ; 
 in other words, it is a vessel having a rectangular midship section, parallel 
 middle body, and wedge-shaped ends, as shown in fig. 10. 
 
 This he called a block model, and is such that its length is equal to that of 
 
 Fig. 10.— Kirk's Analysis. 
 
 the ship, the depth is equal to the mean draught of water, the capacity equal 
 to the displacement, and its area of section equal to the area of immersed 
 midship section of the ship. The dimensions of the block model may be 
 obtained by the following methods : — 
 
 Since A G is supposed equal to H B, and D F equals E K, the triangle 
 A D F equals the triangle E B K, and they together will equal the rectangle 
 whose base is D F and height A G. Therefore, the area A D E B K F 
 equals E K X A H. The volume of the figure is this area multiplied by the 
 height K L. Then the volume of the block is equal toKLxEKxAH. 
 But K L X E K is equal to the area of mid section, which is by supposition 
 equal to the area of immersed midship section of the ship, and the volume 
 of the block is equal to the volume displaced by the ship. Hence, 
 
 Or, 
 Now 
 
 Displacement X 35 = immersed midship section X A H ; 
 A H = displacement X 35 -r- immersed midship section. 
 HB = AB— AH, and A B = the length of the ship. 
 
 Therefore, the length of fore-body of block model is equal to the length of the 
 ship, less the value of A H as found above. 
 
DR. KIRKS ANALYSIS. 33 
 
 Again, the area of section K L x E K is equal to the area of immersed 
 midship section, and K L is equal to the mean draught of water. Therefore, 
 
 E K = immersed midship section -r- mean draught of water. 
 
 Dr. Kirk also found that the wetted surface of this block model is very 
 nearly equal to that of the ship ; and as its area is easily calculated from 
 the model, it is a very convenient and simple way of obtaining the wetted 
 skin. In actual practice, the skin of the model is from 2 to 5 per cent, in 
 excess of that of the wetted skin of the ship ; for all purposes of comparison 
 and general calculation, it is sufficient to take the surface of the model. 
 
 The area of bottom of this model = E K X A H. 
 
 The area of sides = 2xFKxKL = 2(AB— 2HB)xKL = 2 
 
 (Length of ship — 2 length of fore-body) x mean draught of water. 
 
 The ar ea of sides of ends =4xKBxKL=4 V H B 2 + H K 2 X 
 K L = 4 >/Length fore-body 2 -f half breadth of model 2 x mean draught 
 of water. 
 
 The angle of entrance is E B L ; E B H is half that angle ; and the 
 tangent EBH = EH-HB. 
 
 Or, tangent of half the angle of entrance = half the breadth of model 
 -T- length of fore-body. 
 
 From this, by means of a table of natural tangents, the angle of entrance 
 may be obtained. 
 
 The block model for ocean-going merchant steamers, whose speed is from 
 15 knots upwards, has an angle of entrance from 24 to 15 degrees, and a 
 length of fore-body from 0*3 to 036 of the length. 
 
 For that of ocean-going steamers, whose speed is from 12 to 15 knots, the 
 angle of entrance is from 30 to 24 degrees, and fore-body from 0*26 to 0*3. 
 
 Rule. — For angle of entrance of " block model " — 
 
 Angle in degrees = 70 —=— 
 
 L is the length of ship in feet, S is the speed in knots. 
 
 Dr. Kirk measured the length from the fore-side of stem to the aft-side of 
 body-post on the water-line. This is an unnecessary refinement when screw 
 steamers alone are being compared, as then the length may be taken as that 
 " between perpendiculars." However, when small or moderate size screw 
 steamers are being compared with paddle-wheel steamers, it may be neces- 
 sary to measure in this way. 
 
 (2) The old Admiralty rules are as follows* : — 
 
 (a) Indicated horse-power = D« X S 3 -r C. 
 
 (b) „ ,, = area immersed with section X S 2 ■*■ K. 
 
 * If D x be the displacement in pounds, 8 t the speed in feet per minute, R the resist- 
 ance in foot-pounds per minute, A the constant, then 
 
 R = D^ X Si* X A. 
 
 Multiply both sides of this equation by S v then 
 
 R x S, = D,' x Si* x A. 
 
 Now R ,< S x is the work done in overcoming the resistance R, through a distance S u 
 and is, theuefore, the power required to propel D x at a speed S lf and if B is the efficiency 
 
 o 
 
34 MANUAL OF MARINE ENGINEERING. 
 
 D is the displacement iu tons, S the speed in knots, C and K are coefficients 
 determined from previous practice. Their value varies with the size of the 
 ship and the speed — that is, C and K will be less for a long ship than a short 
 ship, the speed being the same — and if the length is the same, the value will 
 be greater at the slower speed than at the higher. 
 
 The above two rules were, for many years, the only ones used by ship- 
 builders in determining the necessary power for a given speed. Their 
 partial accuracy depended on the fact that the wetted skin varies very nearly 
 with the displacement in ships of somewhat similar form,* and that the 
 proportions of steamships were such that the wetted skin varied nearly with 
 the area of immersed section. Their usefulness depended on the information 
 in the hands of the user, and on his discretion in choosing values for C and 
 K. These rules are, in experienced hands, a good check on the newer methods, 
 and can be used by themselves with fewer data than are required when rules 
 based on wetted skin, etc., are employed. Actual values for C are given 
 in Tables vii., viii., ix., etc., on pages 41 to 48, deduced from the per- 
 formances of ships on trial trips made with every care ; but in choosing 
 values discretion must be exercised that the ship for which a calculation 
 is to be made is somewhat similar in form, size, and speed, to the one whose 
 constants are selected. The values on Table iv. are made to suit all conditions. 
 
 The value of C may be taken as approximately = 140 X \/L -5- \^S v. 
 Table iv. gives them calculated in this way. 
 
 (3) Horse-power by calculation from wetted skin. 
 
 This is a simple and efficacious method, and one giving very satisfactory 
 results in practice. It is based on the assumption that so long as a ship is 
 
 * Note. — Let L be the length of edge of*a cube just immersed, whose displacement 
 is D and wetted surface W. Then 
 
 D = L 3 or L = 3 Jl), 
 
 and 
 
 W = 5xL» = 5x( \fD)K 
 That is, W varies as D. 
 
 of the machinery and propeller combined, so that B X I.H.P. is the effective horse- power 
 employed in propelling, then 
 
 33,000 (B X I.H.P.) = D, 3 x S^ X A 
 I.H.P. = (D x 3 x S^) X 
 
 33,000 B 
 
 Now, it is more convenient to express the displacement in tons and the speed in knots ; 
 so that if D and S be substituted for D x and Sj, D being equal to D l 4- 2240, and S = 
 (Si X 60) ~- 6080 s= S x — 101-33, it involves the introduction of other constant quanti- 
 ties, which do not, therefore, alter the expression, so that the whole of these constants 
 may be replaced by a single constant, C, which will express them. Therefore 
 
 I.H.P. = D§ * S '. 
 
 D being the displacement in tons ; S the speed in knots ; and C the coefficient. 
 
 It was also supposed that the resistance would bear a direct relation to the area of 
 section transverse to the direction of motion, as this would be the measure of the 
 channel swept out by a ship ; hence the following rule : — 
 
 T area of immersed midship section X S 3 
 T.H.P. = ^- , 
 
 K being also a so-called constant, but really only a coefficient. 
 
DR. KIRKS ANALYSIS. 35 
 
 not over-driven — that is, the speed does not exceed that appropriate to her 
 form — the power will vary as the cube of the speed with machinery whose 
 efficiency is not less than 0'9. and propellers suitable to the conditions, both 
 as to diameter and area of blade surface ; an allowance of 5 I.H.P. for each 
 100 feet of wetted skin at 10 knots is a fair basis for calculation, and easily to 
 be remembered. It is obtained by supposing that the resistance per square 
 foot of clean painted bottom at that speed should not exceed 1 lb. Dr. 
 Froude found, with the " Alert," having a coppered-bottom, it was about 
 1£ lbs., but a modern ship with a smooth steel surface coated with varnish 
 paints will not cause so much resistance as old copper sheathing on a wooden 
 ship. The author has come to the conclusion, from a careful examination 
 of the trial results of a large number of ships, that it is frequently less than 
 1 lb., and that 1 lb. is a fair all-round allowance. The resistance, then, is 
 100 lbs. for 100 square feet, at the velocity per minute of 101*3 feet. 
 
 Net horse-power = 100 X 1013-3 -=- 33,000 = 3*07. 
 
 Taking the net horse-power ae 62 per cent, of the gross, the 
 
 I.H.P. = 5-0 nearly. 
 
 If the efficiency, as is the case nowadays with high-class reciprocating engines, 
 is 70 per cent., then the 
 
 I.H.P. = 4-386 per 100 square feet at 10 knots. 
 
 The same remarks apply in this case as to the former, as to the variation 
 in value assignable due to the influence of length on speed ; hence a suitable 
 value to each case may be calculated as follows : — 
 
 Rate of I.H.P. per 100 feet wetted skin at 10 knots = 8*5 >/S -=- \/L. 
 
 If, then, the allowance for 10- knot basis is Q, then 
 
 /S\ 3 
 Gross I.H.P. = ( — J x Q for a speed of S knots. 
 
 For Example 1. — Find the I.H.P. required to drive a twin-screw steamer 
 500 feet long at a speed of 23 knots, whose wetted skin is 40,000 square feet. 
 
 Here 85 v/23 -=- VoOO = 5' 10. 
 
 Allowance for 23 knots = ( ^q) X 5*1 = 62*05. 
 
 Total I.H.P. = 62-05 X 400 = 24,820. 
 
 Example 2. — How much I.H.P. will be necessary to propel a steamer 
 300 feet long at a speed oi 21 knots, the wetted skin being 13,500 square 
 feet, and the displacement 1,950 tons. 
 
 In this case Q = 8"5 V2T - V300 = 5-63. 
 
 Allowance for 21 knots = (~) X 5-63 = 52' 2. 
 
 (a) Total I.H.P. = 52"2 X 135 = 7,042. 
 
 By Admiralty methods, C = 140 X x/300 - 1/21 = 212. 
 
 D« = 156. S« = 9,261. 
 
 rr * , TUD 156 >< 9 > 261 *Q1A 
 
 (b) Total I.H.P. = 212" = 6 ' 814 
 
36 MANUAL OF MARINE ENGINEERING. 
 
 Example 3. — A cargo steamer 400 feet long lias a displacement of 7,500 
 tons, a wetted skin of 29,500 square feet, what power ia required to propel 
 her at 12 knots ? 
 
 Here Q = 8'5 #12 -s- V400 = 4-38. 
 
 C = 140 X ^/400 -^ V 12 = 274. 
 
 Ds = 383. S 3 = 1,728. 
 
 /12\ 3 
 Allowance for 12 knots = ( ^J X 4'38 = 7'56. 
 
 (a) Total I.H.P. = 295 X 7'56 = 2,235. 
 By Admiralty method — 
 
 (b) Total I.H.P. = 383 **' 728 = 2,397. 
 
 Example 4. — A yacht 230 feet long is required to steam 15 knots, her 
 displacement is 1,100 tons, and the wetted skin 8,800 square feet, what 
 I.H.P. should she develop ? 
 
 Here Q = 8*5 #15 + J/2S6 = 5:37. 
 
 C = 140 x v ; 230 -=- VTB = 221. 
 D§ = 107. S 3 = 3,375. 
 
 (15\ 3 
 ~j X 85 = 1,594. 
 
 (h\ 107 X 3,375 
 
 (o) » = 22l = ' 
 
 It must not be forgotten, however, that these rules for speed and horse- 
 power apply only to ships working under trial trip conditions, and the results- 
 shown in the following schedules are from trials made under the same cir- 
 cumstances — viz., in practically smooth water, free from eddies and cross 
 currents, and when the wind and weather are moderate ; further, the ship 
 herself is in the best trim and condition, her bottom clean and fresh-painted, 
 and, finally, the water is deep enough to preclude the influence of the sea 
 bed seriously affecting her progress. The latter condition is an important 
 one, and has had only in late years the consideration it demands, although 
 it was patent to every one having to do with steamships long ago that in 
 shoal water there was a marked diminution in speed. Now, however, we 
 know, from the careful observations of many nautical authorities, that 
 the influence continues, though in a lesser degree, when the ship is in deeper 
 water than was formerly thought necessary for successful trials. 
 
 These things being so, the following rule should be observed if it i& 
 intended that the ship shall be able to attain the legend speed under some- 
 what unfavourable conditions of wind, weather, and condition of wetted skin, 
 there must be provided a margin of power beyond that sufficient under trial 
 trip conditions. This margin should be one of power, and not one of speed, 
 for by the latter method the conditions are by no means satisfactory. It 
 was, and is still, a common practice for the owners of express steamers to 
 
PROGRESSIVE TRIALS. 
 
 37 
 
 specify a trial trip speed considerably in excess of the speed desired on service ; 
 the hull is, as a consequence, of much finer form, with the corresponding lack 
 of capacity for carrying deadweight. Moreover, the power and consequent 
 weight of machinery is far in excess of what is necessary to fulfil the real 
 requirements of the service. 
 
 For example, suppose a cross-channel steamer is required to perform 
 her service in fair average weather at 20 knots. The owners specify that 
 the trial speed is to be 22 knots, their real intention being to have a 10 per 
 
 cent, margin. 
 
 Now, as a concrete example, suppose she is 350 feet long X 40 feet beam X 
 12 feet draught water, having a displacement of 2,644 tons, the following 
 table of comparison of her with what she might have been had a 20 per cent, 
 margin of power been provided for contingencies. It will be seen that the 
 latter ship has 100 tons greater displacement, and her machinery, with 
 the 20 per cent, excess power, is 120 tons lighter, and the cost will be about 
 £7.000 less. 
 
 
 
 s.s. A, in accord- 
 
 s.s. B, for same 
 
 
 
 ance with Spectfled 
 
 Service and 
 
 
 
 Requirements. 
 
 Scheduled Speed. 
 
 Length, . . . . . 
 
 . ft., 
 
 350-0 
 
 350-0 
 
 Beam, • 
 
 . ft., 
 
 40-0 
 
 40-0 
 
 Mean draft water, . . 
 
 . ft., 
 
 12-0 
 
 120 
 
 Prismatic coefficient, . . . 
 
 • 
 
 0-601 
 
 0-623 
 
 Displacement, . . . . 
 
 . tons, 
 
 2,644 
 
 2,744 
 
 Area immersed mid section, 
 
 . sq. ft., 
 
 440 
 
 440 
 
 Maximum I.H.P., 
 
 • • 
 
 9,000 
 
 7,800 
 
 I.H.P. for 20 knots, . 
 
 . a 
 
 6,322 
 
 6,488 
 
 Maximum speed — Maximum I.H.P 
 
 •» • 
 
 22-36 
 
 21-2 
 
 Weight of machinery, . 
 
 . tons 
 
 900 
 
 780 
 
 Difference in cargo capacity, 
 
 . 
 
 . • 
 
 220 
 
 Increased consumption of coal per 
 
 100 miles, 
 
 
 13 cwt. 
 
 It will be observed in examining the schedules that, whereas the merchant 
 ships are generally somewhat over-driven on trial for their forms, the naval 
 ships are often finer than demanded by the legend or even actual speed, 
 consequently they will attain their legend full speed under somewhat un- 
 favourable circumstances. 
 
 The high speed of torpedo boats and destroyers depends almost wholly on 
 their lightness of both hull and machinery, which enables them to do with so 
 small a displacement that they literally skim the water, and the resistance 
 per square foot of wetted skin is consequently comparatively small. Unless 
 small boats are made to float at a very light draught they cannot be driven 
 at high speeds, and all experiments with fast river steamers on the Clyde 
 and elsewhere have shown the decided advantage of light draught. The 
 effect on such ships of the depth of water in which they move is also consid- 
 erable, and very interesting data have been given already showing this. 
 
 Progressive Trials. — In modern ship-trials information is usually sought 
 both as to the power required for the highest speeds and (what is equally 
 important) for lower speeds, as such knowledge gives the means of gauging 
 
38 
 
 MANUAL OF MARINE ENGINEERING. 
 
 the efficiency of ship and engines, jointly and separately, and is useful in 
 ship designing. 
 
 The system of examination is as follows : — Let P x P 2 P 3 be the power 
 developed in obtaining the speeds $! S 2 S 3 in knots with R x R 2 R 3 revolutions 
 per minute. Take a line AN as a base line (fig. 11) ; on it take points B, C, 
 and D, so that A B, AC, AD are proportional to S t S 2 S 3 ; at the points 
 B, C, D erect ordinates, B b, C c, D d, so that they are proportional to P x 
 P 2 P 3 . Through the points b, c, d draw a curve, which is called the curve 
 of power, or curve of I.H.P., and it is such that if an ordinate be drawn through 
 any other point, X, on the line A N, the part X x intercepted will measure 
 the power corresponding to the speed measured by A X. If the curve ia 
 accurately drawn, it will be found that it does not pass through the point 
 A, but above A, at a distance A a ; this would signify that when the engine 
 was indicating the power measured by A a the ship would not move, and so 
 
 
 
 
 ./ 
 
 / 
 
 
 
 
 
 2 
 o 
 
 
 / 
 
 /i 
 
 
 
 e 
 
 
 / 1 
 
 >^ r\-- 
 
 > 
 
 J 
 
 ' 
 
 
 ..-~ 
 
 \ 
 
 \ \ 
 
 t 
 
 
 Soeed in knots. B ) 
 
 I t 
 
 ! i 
 
 ) H 
 
 Fig. 11. 
 
 A a is the amount of power required to overcome the resistance of the 
 machinery and propeller at starting, or rather, when not propelling the ship, 
 and hence A a is said to represent the initial friction of the machinery. 
 
 Curve of Revolutions. — A curve of revolutions is constructed in a similar 
 way, by taking points r x r 2 r 3 on the ordinates, so that Br x , Cr 2 , Dr 3 are 
 proportional to R x R 2 R 3 . When the slip is constant the curve of revolutions 
 becomes a straight line. 
 
 Curve of Slip. — The slip may be shown by a curve whose ordinates are 
 proportional to the percentage of slip at the speeds S t S 2 S 3 (v. fig. 12). 
 
 Examination of Curves will show — (1) what indicated horse-power revo- 
 lutions and slip correspond to any speed intermediate to those observed ; 
 (2) the efficiency of the engine at its lowest possible speed; and from it an 
 idea may be formed of its general efficiency, and a comparison made with 
 other engines ; (3) the efficiency of the ship, as tested by the rate of increase 
 
SEA PERFORMANCE OF STEAMERS. 39 
 
 of power for speed, which is seen by the form of the curve towards the higher 
 speeds — if it begins to mount upwards suddenly it is certain that the resist- 
 ance has there begun to increase abnormally ; (4) that, if the curve is one 
 fairly following the law of resistance increasing as the square of the speed, 
 an estimate may be made from it of the power requisite to drive a similar 
 ship at speeds higher than the highest observed curves, or lower than the 
 lowest ; (5) any sudden rise in the slip alone indicates the propeller to be 
 defective in either diameter or surface, or both. 
 
 Another method of expressing the results of progressive trials is by 
 setting out A B, A C, A D proportional to S x 3 S 2 3 S 3 3 , and erecting ordinates, 
 etc., as before. If the indicated horse-power throughout varies as the cube 
 of the speed, the " power curve," or line drawn through the points b, c, d, 
 will be a straight line ; and if the power increases at a higher rate than the 
 cube of the speed at any point, the line will again assume the curved form. 
 The advantages of this plan over the one before described lie in the fact 
 that a straight line is more easily drawn than any curve, that any deviation 
 from a straight line is more easily detected than that of one curve from 
 another, and that the production of a straight line is less liable to error 
 than a curve, so that the interception of Aa is less open to error than by 
 the previous method. Of course, the curves of slip and revolutions cannot 
 be examined so well by this latter method as by the former, and it is only 
 the " power curve " that should be analysed in this way. 
 
 The values of the different " constants," rates, etc., for ships found from 
 calculations made from the results of carefully conducted trial trips, are 
 more reliable than those got by taking averages when employed in calcula- 
 tions for proposed ships. Tables vii. to xiii. give the values of constants, etc., 
 as obtained from the performances of some well-known ships of various 
 types and sizes. 
 
 Sea Performance of Steamers. — That the engines may work economically, 
 both in consumption of coal and stores, as well as in wear and tear, it is 
 advisable to run them at such a speed that they develop about 80 to 90 per 
 cent, of the maximum power. Steamship owners do not always care to 
 pay for 20 per cent, more power than is requisite to drive their ships at the 
 speed intended, but there can be little doubt that it is true economy in the 
 end to do so. For short voyages there is not the necessity for this reserve 
 of power, and very fast steamers could not afford to carry the weight entailed 
 by such an excess of power beyond the actual requirements ; but ordinary 
 sea-going steamers making long runs can, as a rule, easily do this without 
 much sacrifice, and the wisdom of such a course would be shown by the 
 saving in working expenses at the year's end. 
 
 Although, as a rule, trial trips are made honestly, and what the engine 
 has done on the day of trial it can easily be made to do again, still, with the 
 limited staff available for continuous service when the ship is at sea, there 
 cannot be that attention devoted to the working parts which was bestowed 
 by the staff of the manufacturer ; and the application of water to the bear- 
 ings and brasses to prevent heating, which is often deemed absolutely neces- 
 sary by sea-going engineers when the engine is running at full speed, cannot 
 but affect those parts prejudicially, and is a poor substitute for an attendant. 
 Unless, then, the engine is run at a power somewhat below that of the trial 
 trip, either a larger staff of engineers and attendants must be employed, or 
 
40 MANUAL OF MARINE ENGINEERING. 
 
 the wear and tear may be appreciable. It is true, on the other hand, that 
 some engines will develop more power after a voyage or two than was obtained 
 on' their trial trip, due to the polishing of the rough surfaces of the guides 
 and cylinder walls, and to the general " smoothing " of all the rubbing sur- 
 faces ; but it is also true that even such engines should not be run for length- 
 ened periods at their maximum power. The improvements in design and 
 the employment of better materials for guides and bearings, however, admit 
 of modern engines being worked at high speeds with less risk than was 
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 and the almost perfect balancing of the four-crank engine, together with 
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 weather and emergencies is in every ship highly desirable. 
 
 Such remarks are, however, scarcely applicable to the turbine, inasmuch 
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 it is at high speeds that it is most economical, and in the mercantile marine 
 is fitted to express steamers whose service demands full power. Moreover, 
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RELATION OF POWERS AND DISPLACEMENTS. 
 
 49 
 
 TABLE XIV. — Relation op Powers and Displacements.* 
 
 
 No. L 
 
 No. 2. 
 
 No. 3. 
 
 No. 4. 
 
 - t 
 
 No. 5. 
 
 Length in feet, 
 
 280 
 
 300 
 
 360 
 
 435 
 
 600 
 
 Breadth in feet, 
 
 35 
 
 43 
 
 60 
 
 69 
 
 71 
 
 Mean draught in feet, 
 
 13 
 
 16} 
 
 23| 
 
 24* 
 
 26£ 
 
 Displacement in tons, 
 
 1,800 
 
 3,400 
 
 7,400 
 
 11,000 
 
 14,200 
 
 LHP. for 20 knots, 
 
 6,000 
 
 9,000 
 
 11,000 
 
 14,000 
 
 15,500 
 
 I. H.P. per ton of displace- 
 
 
 
 
 
 
 ment, • 
 
 3 3 
 
 2-65* 
 
 1-48 
 
 1-27 
 
 109 
 
 The following horse-powers were required to drive cruisers Nos. 4 and 
 5 in the above table at the speeds named : — 
 
 
 No. 4. 
 
 No. 5. 
 
 10 knots, 
 
 1,500 I. H.P. 
 
 1,800 LHP. 
 
 12 „ 
 
 2,500 ,, 
 
 3,100 „ 
 
 14 „ 
 
 4,000 „ 
 
 5,000 „ 
 
 16 „ 
 
 6,000 ,, 
 
 7,500 ,, 
 
 18 „ 
 
 9,000 „ 
 
 11,000 „ 
 
 20 „ 
 
 14,000 „ 
 
 15,500 „ 
 
 22 „ 
 
 23,000 ,, 
 
 23,000 „ 
 
 The frictional resistance of clean painted surfaces varies about as the 
 1*83 power of the speed, but resistance due to wave making may vary very 
 widely, since it is dependent on form. The total resistance of " Destroyers" 
 has been found to vary as follows : * — 
 
 nearly as speed " 
 „ speed 8 
 „ speed 3 ' 3 
 „ speed 27 
 ,, speed 2 
 practically as speed 183 
 
 and the resistances other than frictional vary as follows : — 
 
 Up to 1 1 knots, - - - as speed 2 
 
 At 12} to 13 knots, ,, speed 3 
 
 ,, 14^ knots, * - „ speed 4 
 
 „ 18 „ - - - ,, speed (more thau 5th P° wer > 
 
 „ 24 „ - - - „ speed 2 
 
 and at higher speeds as still lower powers of the speeds. 
 
 The relation of the frictional to the total resistance is * : — 
 
 Up 
 
 to 11 knots, 
 
 At 
 
 16 
 
 
 i) 
 
 18-20 
 
 
 » 
 
 22 
 
 ft) 
 
 »> 
 >> 
 
 25 
 25-30 
 
 9 9 
 
 
 " Destroyer." 
 
 Cruiser. 
 
 At 12 knots, 
 ,. 16 ,, 
 
 20 
 »> 23 ,, 
 „ 30 „ 
 
 80 per cent. 
 70 „ 
 nearly 50 ,, 
 
 45 „ 
 
 90 per cent. 
 85 
 nearly 80 , , 
 over 70 ,, 
 
 - If the coefficient of friction be doubled (as it might easily be with a foul 
 bottom), the maximum speed of the " Destroyer " would fall fully 5 knots, 
 and that of the cruiser would be reduced to 19 knots. See Tables vii. to xiv. 
 
 * Sir We White, British Association Address, 1899. 
 
50 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Progressive Trials should be made with all ships when at the measured 
 mile, and it should be remembered that for practical purposes it is more 
 important to know the power, revolutions, slip of propeller, etc., at speeds 
 
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 less than the maximum than those at the utmost speed ; it is especially 
 important that the ship shall be tried at as low a speed as possible consistent 
 
EFFECT OF DEPTH OF WATER. 
 
 61 
 
 MAPLIN 7 4- rATHOMS — — 
 
 SKELMORLIZ 40 •• 
 
 Fig. 13. — Effect of Depth of Water on Performance. 
 
 Speed Trials of H.M. Torpedo-boat Destroyer " Cossack " at Maplin and Skelmorlie. 
 270'x 26'x 93' draught Displacement, »36 tons. 14,000 S.H.P. 
 
52 
 
 MANUAL OF MARINE ENGINEERING. 
 
 with obtaining accurate observations; and that between it and full speed 
 there should be one or more trials at intermediate speeds. There should 
 be three consecutive runs made on the mile for each rate of speed, with 
 the steam, vacuum, and revolutions kept as steady as possible during the 
 whole time occupied in doing them. The mean speed should be calculated 
 in the usual way — that is, if the speeds observed are at the rate of x, y, and 
 z knots per hour — 
 
 m , x -f 2y -f- z 
 irue mean speed = f 
 
 The revolutions should be taken from a 
 
 counter, 
 
 ES 
 
 the number i 
 
 DQr 
 
 mile, 
 
 4500 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 / 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 C 
 
 > 
 
 
 
 4000 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 *-8| 
 
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 ClIRVEA 
 
 
 
 
 
 
 4500 
 
 
 trials of 400 ton Destroyer 
 on a depth of 40 feet assum- 
 
 
 
 
 
 
 
 
 
 1 
 
 
 i J 
 
 f 
 
 •V- 
 
 
 1000^ 
 
 
 
 
 
 
 
 / 
 / 
 
 
 
 V 
 
 '** 
 
 
 
 be 62 per cent of the Indica- 
 ted Power. 
 - —curvtb is deduced from the results 
 of the Model Experiments 
 
 1 000 
 
 
 
 
 
 
 / 
 
 / 
 
 <\i 
 
 i 
 
 
 
 
 
 500- 
 
 
 
 
 
 / 
 
 
 
 
 
 
 
 
 
 ATA DEPTH OF 30FEET 
 
 CURVEC is simTlar to Curve B but 
 
 500 
 
 
 
 
 
 It 
 
 ' V 
 
 
 
 
 
 
 
 
 
 
 ATA DEPTH OF 
 
 4ir 
 
 EET. 
 
 
 1 
 
 O 
 
 -f 
 
 
 1 
 
 
 
 
 
 
 ' 
 
 
 , 
 
 
 
 
 
 
 
 
 
 1 
 
 Fig. 14. — Harold Yarrow's Experiment. 
 
 Curves of Comparison between Model Experiments of 400 tons at depths of 30 and 45 feet, and 
 Actual Trials with a Destroyer of 400 tons at a depth of water of 40 feet. 
 
 this, when divided by the time taken on it, will be the true rate of revolu- 
 tion per minute. The results of horse-power, revolutions, slip of screws, 
 
 ])§ x S 3 
 and value of C = T „ p as a measure of efficiency for each speed should 
 
 be set up as ordinates, with the speed in knots per hour as abscissce : a 
 curve drawn through their upper ends of each set will indicate all that is 
 desired to know of the ships' performance. If, however, the ship's resistance 
 has been ascertained by model experiments or otherwise, a curve of E.H.P. 
 can be inscribed, and the curve of propulsive efficiency will then be obtained 
 
EFFECT OF DEPTH OF WATER. 
 
 53 
 
 rup 
 
 by inscribing the values of t ' tt *t>* as ordinates for it. 
 
 Fig. 12 is such a 
 
 I.H.P. 
 
 diagram, as it is highly desirable to have for all important ships, showing, as 
 it does, clearly the performance of the " Lusitania " on the measured mile. 
 Being a turbine-driven ship, the power is practically Brake Horse-Power, 
 being that obtained by observing the torque on the propeller shafting. It 
 is usual to speak of this as Shaft Horse-Power. 
 
 - The Effect of Depth of Water on the speed of steamships is somewhat 
 
 erratic, as may be seen by carefully examining the curves of performance 
 
 of several ships and models at trials made from time to time for this purpose. 
 
 Sir Philip Watts exhibited those of H.M.S. " Cossack," obtained as the 
 
 <M0TS 
 
 Fig 15. — Curve of Effective Horse-power and Speed with Various Depths pf Water. 
 
 Model Experiments by Harold V arrow, I.N. A. 
 
 results of her trials at the Maplin Sands, where the water is comparatively 
 shallow (7*4 fathoms), and those carried out at Skelmorlie, where the watei- 
 is deep — viz., 40 fathoms — and the influence of the bottom only felt by 
 the largest ships at high speeds. It will be seen that, at what may be called 
 the critical speeds of this ship, there were changes of trim corresponding 
 to the changes in revolution and torque at about 18 knots; these were 
 of a violent nature at about 20 knots, while at 26*5 knots things became 
 normal again ; at speeds above this the power required was actually less 
 1*1 the shallow water than in the deep {v. fig. 13). 
 
 Figs. 14 and 15 are equally interesting, as being the results of special 
 
54 
 
 MANUAL OF MARINE ENGINEERING. 
 
 S1V09 03ZIS linj JO U3M0d 3SH0H 3AU33JJ3 
 
EFFECT OF DEPTH OF WATER. 55 
 
 trials made by Mr. Harold Yarrow for the same purpose of finding the effect 
 of depth of water on fast ships. Fig. 14 shows the curves of comparison 
 between the model experiments and those made with the actual ship. The 
 ship herself acted in much the same way as did the " Cossack " in 7-4 fathoms ; 
 her critical point was 18 knots, but became normal at about 23 knots. 
 
 The North German Lloyd Company, of Bremen, had an interesting 
 series of experiments made by Herr Popper, and fig. 16 shows the results 
 of two sets of them, each being made with the boat and her model ; the 
 effects in both cases are even more striking than the former ones, inasmuch 
 as the changes are more emphatic and pronounced. 
 
 The full accounts of all these trials are given in the Transactions of the 
 Institution of Naval Architects, and may be studied there with advantage. 
 
 Dr. D. W. Taylor's formula for ascertaining the least depth of water in 
 
 which a ship should undergo her speed trials for a satisfactory performance 
 
 10 x d x s 
 is as follows : — Minimum depth of water in fathoms = j , where 
 
 d is the draught of water of the ship, I its length between perpendiculars, 
 and s the speed in knots. 
 
 Example. — The minimum depth of water for the trials of a cruiser whose 
 draught is 26 feet, the length 500 feet, and the speed 25 knots. 
 
 n ., 10 X 26 X 25 1Q . ,. 
 
 Depth = — ^r = 13 fathoms. 
 
 500 
 
56 MANUAL OF MARINE ENGINEERING. 
 
 CHAPTER III. 
 
 MARINE ENGINES '. THEIR TYPES AND VARIATIONS OF DESIGN. 
 
 The marine engineer, when dealing with design and construction, is faced 
 with, and has to solve, problems much more complex and involved than 
 those corresponding ones familiar to builders of land engines. Moreover, 
 he is hampered by circumstances and limitations quite unknown to the latter. 
 
 The space occupied by, and the weight of, the machinery of the ship 
 is limited at all times ; in the case of the cargo ship each ton of weight means 
 a ton less cargo on which freight is payable, and in that of the express steamer 
 and warship where the power is larger in proportion to the size of the ship, 
 both weight and space are of great consequence, and generally quite extremely 
 limited. The design, therefore, must be such to allow of inclusion in the 
 machinery space allotted, while leaving sufficient room to permit of accessi- 
 bility to all parts as required for working and overhauling ; the weight is 
 as strictly limited to that share of the displacement provided by the. naval 
 architect in his design. In bygone years more than one good ship failed 
 to comply with the conditions prescribed for her by her designers from a 
 miscalculation of the weight of machinery, or the adding to it by the engine 
 builders without regard to the consequences. To the fast paddle steamer it 
 was fatal, for, in addition to the extra displacement and wetted skin, there 
 was an increased immersion of float, which reduced the efficiency of the 
 propeller immensely. The large engines found in central electric light and 
 power stations on shore are secured on massive concrete foundations which 
 never move, whereas those on board ship, while being strong, are not massive, 
 and may and do move about in a way most trying to them and their bed- 
 plates. Not only does the engine move forward with the ship in the line 
 of the course she pursues, but it is liable to inertia stresses due to the acceler- 
 ation and retardation of the velocity of the ship from various causes, and 
 when pitching and rolling the angular motion is often considerable; so that 
 the inertia stresses thus set up are even more serious. It has not often 
 happened that a marine engine was torn away from its bed, but all of them 
 are liable to such an accident if ample provision is not made for such a 
 contingency. 
 
 In warships very special provision was formerly always made for the 
 shocks that would be set up when ramming the enemy ; this, however, is 
 a nautical manoeuvre no longer contemplated by naval commanders, as 
 it was apt to be more fatal to the rammer than the rammed, so that for this 
 and other reasons ramming is no longer spoken of. It remains, however, 
 as a contingency common to all ships, for both naval and mercantile ships 
 are liable to run on a rock or other massive obstruction, and even to ram 
 another ship accidentally. It is not desirable that the displacement of the 
 engines shall follow such a catastrophe. 
 
VARIOUS TYPES AND DESIGNS OF ENGINES. 57 
 
 The marine engineer has also to produce an engine that may be depended 
 on to keep running without a stoppage for an indefinitely long time, and 
 at a uniform speed, as it is highly desirable that there should be no slowing 
 down of the engine except when so desired by those navigating the ship. 
 Slowing down or stopping the engines at a critical moment might mean 
 the loss of the ship ; in the case of a naval ship, it might mean her capture, 
 or even the loss of a battle, which would decide the fate of a kingdom. It 
 is, therefore, of the very first and highest importance that the marine engine 
 shall be free from every extraneous fitting which might cause temporary 
 derangement, and itself should be so carefully designed, manufactured, 
 fitted, and cared for as to preclude the possibility of compulsory stoppages. 
 
 Further, in both naval and mercantile ships, the whole of the machinery 
 must be practically noiseless when in motion, and the engines free from 
 vibration, which would spoil the gunnery of the one and the comfort of the 
 passengers in the other ; even the auxiliary machiner)' must comply with 
 these conditions of absence of noise and vibration. Finally, since there is 
 a limitation to the weight of fuel which can be carried, and it is desirable 
 that it shall last over as long a voyage as possible, it is necessary on that 
 ground, as well as for the sake of economy in cost, that the consumption 
 of it be as low as possible consistent with a satisfactory compliance with the 
 conditions already insisted on as essential. 
 
 The marine engineer enjoys one advantage over his brother on shore ; 
 he has an unlimited and cheap supply of cold water, whereby he may con- 
 dense as much steam as he desires, and work with as high a vacuum as his 
 apparatus can produce and maintain. And if it be an advantage, and no 
 doubt it is in many ways, his engine varies in velocity as its load varies, 
 instead of running at constant velocity whatever be the load. 
 
 Various Types and Designs of Engine, of which in the early days of steam- 
 ships there was a very large variety employed, and although some of them 
 were developments of those common and successful for land service, the 
 greater number were evolved to satisfy the conditions imposed on the engineers 
 of the day, many of them showing considerable originality in form as well 
 as ability in design ; some of them display features which indicate that their 
 originators possessed a technical knowledge, for which they have not been 
 always accorded credit. On the other hand, not a few had inherent defects, 
 which, while not being apparent in the model state or in engines of small 
 power, were very evident in the larger engines, and soon caused their early 
 dismissal to the scrap heap ; such defects were generally due to a want of 
 technical knowledge, or to the misapplications of the empiric formula? of 
 the day, which were the rough and ready pilots of these earlier times. 
 
 The engineer of to-day is not required to waste time and energy in differ- 
 entiating the claims of many types before coming to a conclusion as to what 
 will suit the circumstances of his particular case, for time and experience 
 have decided much of this for him, and so now for a paddle steamer it is 
 almost the universal rule to choose the direct-acting compound engine, either 
 horizontal or nearly so, and for the screw the inverted form of the same. 
 He may, however, have to debate with himself or his advisors whether his 
 is a case for turbines pure and simple or for reciprocators ; he may also 
 compromise the matter by having reciprocators at the boiler end of the 
 installation, with low-pressure turbines at the condenser end, as is the case 
 
58 
 
 MANUAL OF MARINE ENGINEERING. 
 
VARIOUS TYPES AND DESIGNS OF ENGINE. 
 
 5$ 
 
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€0 MANUAL OF MARINE ENGINEERING. 
 
 now in many large ocean-going ships of high speed {v. fig. 44). When the 
 power is not very great a Diesel oil engine may be adopted, and with smaller 
 power even a semi-Diesel type will prove attractive. For small craft of low 
 or moderate speed the paraffin engine is popular and convenient. Petrol, 
 however, should be avoided on shipboard except in very special cases. 
 
 Although there is very little variety in the type of engine now employed 
 for paddle-wheel ships and screw ships, it is well to know why particulai 
 ones have survived, while others have died out, and in order to appreciate 
 the selection, some knowledge of the virtues and vices of those others now 
 discarded is desirable as well as instructive. 
 
 Paddle-wheels were the first practical form of propeller employed foi 
 steam navigation, and, as has already been stated, the earliest ship had a 
 horizontal direct double-acting engine. To-day this form of engine survives 
 as the fittest for the purpose of both side wheel and stern wheel paddle-ships 
 of all sizes ; in the case of the former it is, however, inclined to the horizontal 
 position, and called the diagonal or inclined engine. In this form it was 
 patented by Marc Isambard Brunei, the famous engineer, in 1822, who 
 claimed as his invention two inclined cylinders, the cranks at right angles, 
 the piston-rod ends fitted with roller guides, and the weight of the pistons 
 relieved by spring supporters on the extremities of the beam head ; the engines 
 were to be governed by means of a stream of water pumped through an 
 orifice. The condenser was to be formed of an assemblage of pipes, which 
 collectively formed a spacious chamber ; they were to be connected together 
 with a set of smaller pipes, and the whole placed in an iron reservoir. This 
 must be admitted to be a very ingenious and comprehensive invention, con- 
 sidering it included roller guides, balanced pistons, a cataract governor, 
 and last, but not least, a surface condenser. The roller guide never took 
 on for large engines, and a cataract governor seems almost to have met 
 with the same fate, otherwise Brunei's invention may be said to have sur- 
 vived as the fittest for the purpose of marine propulsion with paddle-wheels, 
 just as the vertical direct-acting inverted first adopted by the Thomsons, 
 of Glasgow, has for screw propeller work. 
 
 The inclined direct-acting engine (figs. 19 to 22) is the engine of to-day, 
 as being the one complying best with the conditions ruling on board such 
 ships as are still propelled by side paddle-wheels for the following reasons : — 
 
 (1) The design is simple and without complications or make-shifts of 
 any kind, hence less liable to derangement and breaks down. 
 
 (2) It is as light and cheap in construction as any other, and may be 
 arranged so that its framing helps to stiffen a lightly built ship rather than 
 to over-strain her. 
 
 (3) Great flexibility of design is possible ; the engine may have as many 
 cylinders as are desired within the limits of beam, and may be compound 
 or simple without any difficulties or objections. 
 
 (4) The stroke of piston may be long ; in fact, there is practically no 
 reasonable limit to it. 
 
 (5) The steam pressure may be as high as that in screw ships without 
 any disadvantages. 
 
 (6) The shaft may be at any desired height from the ship's floor, and 
 so permit of a small diameter of wheel with a corresponding increase in number 
 of revolutions if desired. 
 
PADDLK-WHEKI.S. 
 
 61 
 
 On the other hand, it has the defects of all horizontal engines, sucli as 
 heavy pistons running on the cylinder liners or bodies either unsupported or 
 inadequately so, whereby they are worn barrel shape in course of time ; there 
 is also the friction due to this which retards the movement. Further, the 
 momentum of the pistons and rods in connection therewith causes a pulsa- 
 tion fore and aft-ways in the ship, which in light river craft is sometimes 
 
 Fig. 19. — Diagonal Compound Paddle Engines (J. Brown & Co., Clydebank). 
 
 by no means agreeable to the passengers ; this is especially the case in some 
 of the older river steamers having only one cylinder, the more modern steamers 
 having three cylinders (fig. 22) are practically free from this defect, and in 
 those with two cranks at right angles it is not so serious, but it is yet observ- 
 able, whereas with vertical cylinders it was absent, although they had a 
 
€2 
 
 MANUAL OF MARfNE ENGINEERING. 
 
 vibration of theif own, and the steeple engines had the objectionable pulsa- 
 tion also. But, as a rule, the paddle steamer in old days was a more com- 
 fortable one in which to sleep than were the screws. 
 
 Of the inclined types, there may be one or more cylinders side by side, 
 as shown in fig. 19 ; or, when the ship is narrow, and from this or other cause 
 & limitation in athwartships space, the design of fig. 20 is a good one, and 
 
 Fig. 20. — Diagonal Compound Paddle Engines. 
 
 works quite satisfactorily. This latter permits of more deck room amid- 
 ships, and space available for passengers, etc., and generally gives a clear 
 deck fore and aft in even small ships. 
 
 The fore and aft space taken up by the diagonal engine is of course very 
 considerable, and formerly in some classes of ships would be of great disad- 
 vantage, but in modern paddle-steamers this is of no consequence, and is 
 
PADDLE- YVHEKLS. 
 
 63 
 
 more than compensated for by the advantages in other ways, not the least 
 of which is the ease with which a paddle-wheel of small diameter can be 
 adopted, whereby high revolutions are rendered possible. This, together 
 
 with the long stroke of piston which can be permitted with it, allows of a 
 very high rate of piston speed, amounting in some large steamers to as much 
 as 750 feet per minute with a low-pressure piston 108 inches diameter, the 
 
64 
 
 MANUAL OF MARINE ENGINEERING. 
 
BEAM ENGINES. 
 
 65 
 
 .revolutions in this ease being 52 ; this and other examples of these engines 
 may be studied on Table xiii. 
 
 On the American rivers the stem-wheel steamers of shallow draught have 
 the horizontal engine in general use ; a remarkable example of it is found in 
 the tugboat "Sprague," of 2,200 tons displacement ; her two engines have 
 cylinders 28 inches and 63 inches diameter, and a stroke of piston of 12 feet 
 driving a wheel 40 feet diameter and 40 feet long. At 16 revolutions the 
 I.H.P. is about 2,500. 
 
 Beam Engines, by their family likeness to those on shore, show then- 
 descent from them ; and the survival of type locally is illustrated in their 
 
 Fig. 23. — American Steamer Beam Engine. 
 
 •case as it is in that of the direct-acting variety. The second practicable 
 steamboat constructed was Fulton's p.s. " Claremont," built by him in 1807, 
 and used for service on the River Hudson ; she was fitted with an ordinary 
 overhead beam engine supplied by Boulton & Watt, from their Soho Foundry 
 at Birmingham, England. Probably, as a consequence, this type of over- 
 head beam engine soon came into general use in America, and continued to 
 be the favourite one till quite modern times, and may be still seen on service 
 
66 
 
 MANUAL OF MARINE ENGINEERING. 
 
 to-day with cylinders of enormous size with " drop " valves and wooden 
 framework, etc. (fig. 23). Indeed, it is to this latter fact that it continued 
 30 long to hold its own, for wooden structures and wooden connecting-reds 
 properly braced and bolted were permissible with such engines, and were, 
 and are, the peculiar product of the people inhabiting a new country where 
 wood is more plentiful and more easily manipulated than iron. Never- 
 theless, the machinery of a large Fall River steamer constructed on this 
 method was a splendid example of engineering genius, displaying as it did, 
 the adaptation of what was found to hand to the needs of the problems. 
 
 Side Lever Engines are a form of beam engine, and was the type developed 
 in this country from the beam idea. The third practicable steamboat was 
 the P.s. " Comet," already alluded to and shown in fig. 4. This ship was 
 
 Fig. 24.— Engine of the "Comet," 1811-12. 
 
 installed by Bell, her owner, with a single-cylinder engine having a beam 
 at each side, as shown in fig. 24, connected to the piston-rod cross-head and 
 the-crank pin by rods. A modification of this design was adopted many 
 years ago by the builders of tug boats, especially by the Tyneside builders, 
 and called the Grasshopper engine. Such engines are still in use on tugs,. 
 and are exceedingly suitable for them. They are very cheap in construction, 
 have a very long stroke of piston for such shallow ships, the racking action 
 when in motion is consequently comparatively slight, and is taken by the 
 keelsons, the stiff est part of the hull ; when only one cylinder is employed 
 there is in practice no " dead " point — that is, the crank can be moved by 
 the pressure on the piston from any position in which it may have stopped. 
 This latter quality is due to the position of the connecting-rod with respect 
 
*7 
 
 Fig. 25. — " Side Lever " Engine. 
 
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 Fig. 26. — Engines of Thames Steamer " Regent," 1816 (Maudslay). 
 
68 
 
 MANUAL OF MARINE ENGINEERING. 
 
 to the levers when the piston is at the end of its stroke, and aided by the 
 slight amount of " play " in the brasses. This class of engine is capable of 
 working satisfactorily when in such a state of disrepair that would in any 
 other form of engine prove dangerous ; it is also not exacting in the matter 
 of attention from the attendants. 
 
 The most important development of the side lever engine was, however, to 
 
 Fig. 27. — Oscillating Engine — Section through Trunnions. 
 
 be found in the sea-going steamers of the early half of last century, where it 
 might have been seen in large varieties both of size and design. It was the 
 means whereby the Cunard and some other important companies established 
 the reputations of their ships by regular passages and freedom from breaks 
 down. Fig. 25 is a good example of such an engine, and shows all its salient 
 features. It continued to be the most popular type to the end for sea-going 
 
OSCILLATING ENOINKS. 
 
 69 
 
 ships, and the " Scotia," the last of the Cunard paddle ships, had such 
 engines with two cylinders, each 100 inches diameter and 12 feet stroke, 
 and developing 4,950 I.H.P. Fig. 26 shows a modified form, inasmuch as 
 the lever is of the bell crank type. 
 
 Oscillating Engines were first suggested by Trevithick ; they were used 
 originally by Dr. Goldsworthy Gurney about 1827 for driving his steam 
 motor cars. In the same year Joseph Maudslay patented the use of such 
 
 Fig. 21a. — Oscillating Engine — Section through Valve-Boxes. 
 
 cylinders, and in 1830 William Church took out a patent for their applica- 
 tion to driving paddle-wheel shafts. The firm of Maudslay, Sons & Field, 
 however, developed the oscillating marine engine, and until they quarrelled 
 with the Admiralty over some repairs to one, had the sole supply of them. 
 The patronage of the Admiralty was transferred to John Penn & Son foi 
 their further supply, and since then this type ha6 been always associated 
 
70 MANUAL OF MARINE ENGINEERING. 
 
 with this firm. It was, moreover, brought by them to a very high state of 
 efficiency, and employed largely on river and cross-channel services, as well 
 as in all kinds of ships in the Navy. 
 
 " The Great Eastern " steamship had four oscillating cylinders, 84 inches 
 diameter and 145 feet stroke ; but perhaps the largest and most interesting 
 example was that of the p.s. " Ireland," 1,952 tons, built by Messrs. Laird 
 for the City of Dublin S.P. Company, for service between Holyhead and 
 Kingston, which, with two cylinders 102 inehes diameter and 8'5 feet stroke, 
 7,000 I.H.P. was developed, and a speed of 23 knots was attained. 
 
 The oscillating engine has survived to the present time, as they may 
 still be found in common use on dockyard tugs, and on river craft throughout 
 the world ; and although there have been compound engines of this kind 
 with cylinders of considerable size for steam pressure of 60 lbs., the type 
 is not so good for the higher pressures or for triple engines as the diagonal. 
 Figs. 27 and 27a show them to be simple in design, having only one rod 
 from piston to crank-pin, requiring no guides, and generally are free from 
 complications, except that the valve-gearing may be considered somewhat 
 complex. They are very light ; occupy little space, and permit of a fairly 
 long stroke, even in somewhat shallow ships. In the case of the " Great 
 Eastern," they were placed diagonally two to each crank-pin, and this design 
 has been followed down to modern times for the British Dockyard Tugboats, 
 where each wheel can be worked independently by a pair of cylinders. 
 
 It is not necessary to dwell on the other types of paddle engine, as they 
 have had their day and disappeared, except to say that the Steeple Engine, 
 as shown in fig. 28, was a favourite one with some Clyde engineers, and it 
 was the type adopted by Messrs. Laird when refitting the p.s. "Violet" 
 with triple-compound engines of 4,070 I.H.P. , in which the L.P. cylinder 
 was 108 inches diameter and 6 - 5 feet stroke. The space occupied by this 
 form of engine was small, and generally permitted of a fairly long stroke, 
 but the thrust of the connecting-rod caused a tilting action on the engine 
 framing, which, unless well provided for by bracing, severely strained the 
 entablature, etc. 
 
 Vertical Direct-acting Engines were used by some makers, and with the 
 old sea-going ships with their good depth of hold, and a radial wheel of large 
 diameter, a fairly long stroke could be obtained ; but it was always some- 
 what short compared with that of other types of engine. 
 
 Twin-cylinder Engine of Maudslay & Field (fig. 29) is now chiefly inter- 
 esting as being the one fitted in the first screw ship of the Navy, H.M.S. 
 " Rattler." It was geared to the screw shaft, so that the latter made four 
 revolutions to one of the engine, whose cylinders, four in number, were 
 40 inches diameter, and the piston stroke 4 feet M 
 
 Screw Engines were at first, as in the case of H.M.S. " Rattler," of the 
 same type as used for paddle-wheels (v. fig. 29), with spur and pinion gearing 
 connecting them to the screw shaft, and what, perhaps, was of more import- 
 ance, the necessary revolutions for the screw were got thereby without sub- 
 mitting the engines to such speeds as they were then not fit to bear. Gearing 
 was eventually discarded, and direct-driving adopted, both for safety and 
 economy. The higher speed of engine permitted of the use of smaller ones, 
 which were lighter and cheaper ; and as tooth-gearing involved considerable 
 friction as well as frequent repairs, further economies were affected by 
 
SCREW ENGINES. 
 
 71 
 
 abolishing it. To-day, however, we are witnessing a revival of wheel-gearing, 
 to permit of the use of turbines in cargo steamers where a small high revolu- 
 tion screw is not permissible, and in ships generally when turbines of high 
 revolution and consequent economy may be employed. It is, of course, under 
 
 Fig. 28. — Steeple Engine. 
 
 much more favourable circumstances that this experiment is being carried 
 out, for in this case the torque of the engine shaft is absolutely uniform, instead 
 of being highly variable ; the pinion is driving instead of being driven ; and, 
 moreover, there are two, one on each side of the spur-wheel ; finally, the teeth 
 
72 
 
 MANUAL OF MARINE ENGINEERING. 
 
 arifflfr. 
 
 of the cog-wheels are to-day double helical machined to exact shape and size, 
 instead of being common toothed of rough cast metal or ol wood that, while 
 tough, was not particularly hard (v. fig. 45). The efficiency of the helical 
 wheel connection is 98" 5 per cent., which is very high, and so far satisfactory, 
 as to permit of high-speed reciprocators, as well as turbines for the drivers. 
 
 The Horizontal Direct-acting Engine was the product of Humphrys & 
 Tennant latterly, and in an older form of Boulton & Watt for ships of the 
 Navy ; they and one or two other engine builders adopted this type also for 
 the mercantile marine. Although itself now superseded, it displaced and sur- 
 vived the other kinds of horizontal screw engine once so popular on shipboard, 
 and itself only succumbed when it was found that protection against shot 
 could be afforded to a vertical engine in a warship without inconvenience 
 or serious cost. Till then some horizontal type, which was well below the 
 water-line, was considered absolutely necessary, and as the direct-acting 
 kind of engine suited the higher steam pressures of modern time better than 
 the others, it prevailed ; moreover, with twin screws a longer stroke of piston 
 
 was permissible than was the case 
 when only half the breadth of the 
 ship was available for the cylinder 
 and rods. Very early on Boulton & 
 Watt made a horizontal four-cylinder 
 oscillating engine that was fairly 
 successful, and admitted of a longer 
 stroke than the ordinary kind. Ex- 
 perience with the "Simoom" does- 
 not seem, however, to have inspired 
 sufficient confidence in them to cause 
 many repeats. 
 
 Such direct-acting engines possess 
 the same good features as the direct- 
 acting paddle engine had when placed 
 horizontally or nearly so ; they like- 
 wise have the same objectionable 
 features in the way of heavy pistons 
 rubbing on cylinder sides, effect of momentum of the moving parts to cause 
 stresses on the hull, and general vibration notwithstanding balance weights 
 on the cranks ; and, finally, the general want of accessibility to some parts, 
 when running, and to others, such as the pistons, etc., when requiring 
 overhaul. 
 
 Trunk Engines, invented and introduced by John Penn & Son into the 
 Navy in 1849 on board H.M.S. " Arrogant," are of the horizontal type and 
 direct-acting, inasmuch as the rod connects the piston direct to the crank 
 pin, as is the case in the oscillating engine only. Fig. 31' shows such an engine 
 and how the connecting-rod is enclosed in a hollow cylinder or trunk attached 
 to the piston, and with a similar trunk in rear whereby the rod end and 
 gudgeon could be fitted and examined ; it also served to ventilate these 
 important parts, and prevent them from overheating. Both in the mer- 
 cantile marine and Navy they were, in their day, favourite engines ; they 
 were splendidly made and finished, and worked even at comparatively high 
 revolutions, and latterly with steam of 60 lbs. pressure quite satisfactorily. 
 
 Fig. 29.— Twin-Cylinder Engine. 
 
TIU'NK ENCINKS. 
 
 73 
 
 Jet Condenser 
 Y\a. 30.— Engines of P.S. " C. D.," 6,370 tons. Lake Service. U.S.A. 
 Cylinders 92 ;, -62"-92" diam. X 102" stroke. 7,000 I.H.P. at 28 revs. 
 
74 
 
 MANUAL OF MARINE ENGINEERING. 
 
 There was, however, considerable heat loss by radiation from the trunks, 
 and condensation in the cylinders by their cooling action ; the friction at 
 the stuffing-boxes was also somewhat of a drawback, being often considerable, 
 and could be easily made so great as to slow down the engines. 
 
 rfes 
 
 Fig. 31.— Trunk Engines. 
 
 Fig. 31a. — Return Connecting-Rod Engineo. 
 
 Their screw was generally " left-handed," so that the thrust of the con- 
 necting-rod when going ahead was upward ; but when at full speed it was 
 considerably in excess of the weight of the piston, so that the rubbing 
 was sufficient to " barrel " the upper side, and when in stern gear the 
 thrust and weight were combined in doing the same thing to the lower 
 side 
 
RETURN CONNECTING-ROD ENGINE. 
 
 75 
 
 The Return Connecting-rod Engine is shown in fig. 31a, by which it will 
 be seen that it is a steeple engine laid horizontally with the two piston-rods 
 prolonged to the gudgeon cross-head instead of into a banjo-frame. This 
 
 Fig. 32. — Three-Crank Triple-Expansion Engine (Naval). 
 
 type was introduced by Maudslay & Field to the Navy, and used by them 
 till the vertical engine displaced them. Several other builders also adopted 
 this type, inasmuch as therebv a much longer stroke was possible with ample 
 
79 
 
 MANUAL OF MARINE ENGINEERING. 
 
 room behind the cylinders for examination and repair, and all the working 
 parts were quite accessible when moving. An attempt was made latterly 
 
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 to support the pistons by fitting a substantial tall rod with a slipper support 
 and guide at the back ; this arrangement was only a palliation, and did not 
 prevent rubbing of the cylinder walls at bottom. With the vertical engiue 
 
VERTICAL DIRECT-ACTING ENGINES. 77 
 
 there is no absolute need of such a fitment, and although the tail rod through 
 their cylinder covers tends, no doubt, to steady the piston, and when the 
 ship is rolling to prevent excessive side play. 
 
 The Vertical Direct-acting Engines. — Figs. 32 and 33 are now the universal 
 type of screw engine employed throughout the world for both war and mer- 
 chant ships ; its advantage over others are so obvious as to need no further 
 ■comment, after having pointed out the defects of those others whose com- 
 petition lasted. This type was first employed by the Thomsons, who founded 
 the business now so extensive and successfully carried on by John Brown & 
 Co. at Clydebank. In the mercantile marine it quickly found favour, and was 
 soon adopted by most builders of marine engines for it, and is still largely 
 used for generating electricity on shore. For naval purposes it was deemed 
 unsafe formerly, as having the vital parts of the machinery above the water- 
 line, therefore much exposed to shot and shell in a way not obtaining with 
 the horizontal engine. This objection was, of course, from the military point 
 of view, and formed a bar to their use long after the introduction of armour 
 plates had provided a means for their protection if naval designers had so 
 desired, and as was eventually adopted, and still prevails. 
 
 In the mercantile marine, not only is the foundation or bed-plate of 
 ■cast iron, but the columns supporting the cylinders, both front and back, 
 and forming guides for the piston-rod ends, are often of cast iron. In the 
 Navy, for lightness and strength, and in a measure to resist concussion, the 
 columns are of forged or cast steel, and the foundations themselves also of 
 steel. Figs. 32 and 33 show two different makes of engines, the one for the 
 Navy, the other for an express steamer. Fig. 32 is that of an engine with 
 ■cast-steel foundation and front column with cast-iron back columns. In fig. 
 33 the columns on both sides are of cast iron, as also are the foundations. 
 In fig. 34 is seen an example of a wrought-steel structure of extreme lightness, 
 and yet so rigid and strong as to permit of 380 revolutions per minute being 
 run quite satisfactorily. The cost of the latter, and such structures as seen 
 in fig. 35, compared with either of the others, but more especially with the 
 engines of fig. 33, is very high, but, notwithstanding this, the saving in 
 weight, combined with the perfect rigidity, warrants the use of similar struc- 
 tures for the support of the cylinders of express steamers of quite large size. 
 Some builders of mercantile engines prefer these wrought-iron or steel columns 
 for the front of the engines of all kinds (v. fig. 36), as they permit of freer 
 access to the moving parts ; moreover, when secured in the foundation by a 
 single nut and to the cylinders in the same way, so that these columns can 
 be turned down from a plain rolled bar, they are even cheaper than cast- 
 iron ones. 
 
 In comparing the engine design and practice of to-day with that of past 
 generations, it must not be forgotten that the earliest engineers had to make 
 almost every part of cast iron ; for there were no steam hammers, and the 
 forgings made by the tilt hammer and the rolled-iron bars of that period 
 were of very limited size and form. Steel could be had only in very small 
 pieces, and was very dear. Modern engineers have an unlimited supply of 
 splendid mild steel, and now even reliable high tensile steel can be obtained 
 at quite moderate costs. Forgings, rolled bars, and plates of huge size can 
 be obtained cheaply from various sources of supply, and quickly compared 
 with the time required twenty-five years ago. Our predecessors knew not 
 
78 
 
 MANUAL OF MARINE ENGINEERING. 
 
 the advantages of " dumping," or of steel castings, nor did they enjoy such 
 white metal and stroug zinc bronzes as we know them. 
 
 Arrangement of Cylinders in the modern marine engine is a much more 
 important problem than was the case when there were two cylinders only. 
 
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 parts, as well as to determine what is the best sequence for the flow of steam 
 
ARRANGEMENT OF CYLINDERS. 
 
 79 
 
 through the system from the boilers to the condenser to attain the greatest 
 steam efficiency, together with the minimum amount of "piping for that 
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 the designer, especially when the engines are of such great power as to require 
 excessively large cylinders if their number is to be as limited as they may be, 
 and are generally just as found in engines of smaller powers. Large cylinders 
 
80 
 
 MANUAL OF MARINE ENGINEERING. 
 
 can be, of course, manufactured, and large pistons, being now of solid cast 
 steel, are no longer so heavy and cumbersome as to be almost impossible 
 
 of handling with the engine-room appliances . Moreover, all engines to-day- 
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ARRANGEMENT OF CYLINDERS, 
 
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82 MANUAL OF MARINE ENGINEERING. 
 
 to cylinders, liners, and walls, or difficulties due to the position of the cylinders 
 whereby the use of mechanical means for getting them in and out of place 
 was precluded, besides rendering them difficult to examine. But the risk 
 of casting large cylinders is still great ; the cost of replacing such an one, 
 if damaged, great, and perhaps, what is more important in these days of high 
 piston speed, there is always the impossibility of designing the low-pressure 
 ones with ports of sufficient size for a flow of steam slow enough to be con- 
 sistent with high efficiency. The area of port possible for a cylinder with 
 slide valves varies practically as the diameter, inasmuch as the length of port 
 axially does not increase rapidly with increase in the diameter. The volume 
 of steam passing through it varies as the square of the diameter. When 
 piston speeds seldom exceeded 500 feet per minute, ports with a sufficient 
 area for very large cylinders could be designed easily enough ; but with a 
 speed of piston of 1,000 to 1,200 feet it is impossible to provide them large 
 enough for moderate flows and with passages small enough for moderate 
 " clearance." 
 
 With the Diesel and other oil engines, where the initial pressures are very 
 high, while the mean pressures are quite low, the latter condition requires 
 a large total cylinder capacity and the former small units ; hence such engines 
 have as many as twelve cylinders per engine when of the large horse-power 
 required on shipboard for good speed of ship. 
 
 Even in the steam engine of to-day the boiler pressure, and consequently 
 the initial pressure in the H.P. cylinder at starting is as much as 230 lbs., 
 while the mean pressure of the compound cylinder system is not very much 
 more than that obtaining when it was a fifth. In the Navy, with water-tube 
 boilers, the pressure is often even greater still, but the referred mean pressure 
 in such ships is much higher than in the mercantile marine when running 
 full speed. 
 
 With a Compound Engine the least number of cylinders is of course two, 
 which may be in line axially with their pistons attached to a common rod, 
 and operating on one crank by one connecting-rod (v. fig. 75). Such engines 
 were at one time used in certain cargo steamers, and thought well of at the 
 time, in large measure owing to the small space occupied by them and their 
 comparative cheapness. More commonly the two cylinder compound engine 
 had them side by side, each operating on a separate crank, the one opposite 
 the other when the steam expanded direct from the H.P. to the L.P. cylinder, 
 or set at right angles with a receiver between the cylinders into which the 
 H.P. exhausted, and from which the L.P. took steam. In some few cases 
 the cranks were set at a different angle, generally about 105°, so as to favour 
 the timing of exhaust from H.P. and flow to the L.P., and to reduce thereby 
 the variation in pressure in the receiver. These latter engines were, how- 
 ever, somewhat unhandy in starting and reversing, and eventually were 
 given up in favour of the cranks at 90°, and having a somewhat larger 
 receiver between the cylinders. 
 
 The Three-cylinder Compound Engine, each cylinder with a separate 
 crank and connections, was deservedly a favourite one for large powers ; 
 for not only was there with it the advantage of having the L.P. cylinders 
 of moderate size, but a much higher initial pressure could be thereby main- 
 tained in them, consequently with the minimum amount of drop between 
 H.P. and L.P. cylinders ; a fair division of the work between the three cranks 
 
TWO-CRANK SCREW ENGINES. 83 
 
 was made, and they were usually set at angles of 120 3 . This engine was 
 a favourite for electric generation on shore stations with large, units. 
 
 Four-cylinder Compound Engines enabled the same subdivision of the 
 L.P. pressure member of the system without resorting to an additional 
 crank by dividing the H.P. member, and placing one H.P. cylinder tandem 
 with one L.P., each pair side by side operating its own crank, which was at 
 90° with the other. This type of engine was adopted by Messrs. Maudslay, 
 Sons & Field in the celebrated White Star steamers, " Britannic," etc. ; 
 moreover, it was the method used by many engineers as a cheap and ready 
 way of compounding the old expansive engines after the superiority of the 
 compound system was assured about 1870. 
 
 Six-cylinder Compounds, each pair of H.P. and L.P. being tandem and 
 side by side, each operating on a separate crank, as with the four-cylinder 
 system, of which it is an extension ; this design was adopted for the " City of 
 Rome," as a further subdivision of cylinder due to her then large power. 
 
 With a Triple-expansion Engine the least number of cylinders is, of course, 
 three, as with a quadruple expansion it is four. It does not, however, follow 
 that each must have its own separate crank, as is the rule with the single- 
 acting oil engine ; and, although it is common practice now to do so with 
 the marine steam engines, it was a few years ago not an unusual thing to find 
 a triple-expansion engine with two cranks and the cylinders as in fig. 73, 
 Nos. 1, 2, and 3. Moreover, the quadruple engine when first placed on the 
 market as a competitor with the triple had only two cranks (fig. 73, No. 1) ; 
 this, as a matter of fact, was at that time placed to its credit as a means 
 whereby it occupied no more space than a common two-cylinder compound, 
 and less than a three-crank triple or compound engine. When, however, 
 the demand for a non-vibrating engine was insisted on by the Admiralty, 
 and much desired by all interested in the passenger service, the Yarrow- 
 Schlick-Tweedy system of balancing the four-crank quadruple engine gave 
 it such an advantage over the three-crank triple that makers of the triple 
 engine had to adopt the four-crank arrangement, and with two L.P. 
 cylinders, either as Nos. 3 or 4 in fig. 73, to obtain like advantages. 
 
 A Single-crank Engine is seldom or never seen to-day, except in quite 
 small and cheap launches, or harbour service boats. Even single-crank 
 tandem compound engines, which at one time were much in favour with 
 one or two shipowners and some few engineers, are no longer made. Single- 
 crank paddle engines are also a thing of the past. 
 
 Two-crank Engines still survive in paddle steamers of moderate power 
 with compound cylinders (v. fig. 19) ; with the larger power compound 
 engines three cranks are favoured (v. fig. 21), as each of the two L.P. cylinders 
 are of moderate size, and the ratio of maximum to mean torque is lower than 
 with two cranks, consequently the movement of the wheels is less jerky, 
 and a smaller shaft possible. Triple-expansion paddle engine; are invariably 
 of the three-crank type, as shown in fig. 22. 
 
 Two-crank Screw Engines are still made with compound cylinders for 
 quite small powers, as in tug boats, steam launches, and other small craft ; 
 they may be adopted with advantage still in cases where prime cost, weight, 
 and space occupied are of more importance than fuel consumption. With 
 triple-compound cylinders such engines are now very seldom, if ever, made, 
 but with quadruple-compound cylinders something like those shown in fig. 75. 
 
84 
 
 MANUAL OF MARINE ENGINEERING. 
 
 the two-crank arrangement is quite a good one for smaJl craft, especially 
 those having water-tube boilers for the &ake of quick raising of steam, as in 
 yachts or their launches. 
 
 Three-crank Screw Engines (v. figs. 32 and 36) still continue to be the 
 common and, to a great extent, the favourite form in the mercantile marine, 
 when the power is not great, and the service either cargo-carrying or com- 
 bined with a passenger service of not high class, inasmuch as it is cheaper, 
 occupies smaller space than the four-crank, and can be balanced sufficiently 
 
 Fi^ 37. — Eive-CMiiKler Quadruple-Expansion Engines 01 S.s. 
 
 Marine Engine Works, West Hartlepool). 
 
 Inchdune " (Central 
 
 well, if thought necessary, by simple and comparatively inexpensive means, 
 so as to cause little or no inconvenience from vibration. 
 
 The Four-crank Engine (figs. 33 and 33a) is, however, more easy to balance, 
 and the balancing, when done, is more perfect ; the quadruple system permits 
 of the use of higher boiler pressures with higher efficiency and, therefore, 
 with advantage ; on this triple-expansion system, however, there are two 
 L.P. cylinders, with such advantages as arise from their smaller size. 
 
 It may be taken, then, that as a rule for express passenger ships and 
 
HEAVY OR CRUDE OIL ENGINE. 85 
 
 warships, the four-crank engine is a better one to adopt ; for very small ships, 
 or where the engines are small, the three-crank engine has advantages which 
 may outweigh those which would favour the four-crank in a general problem. 
 
 Five-crank Engines (fig. 37) have been made to a limited extent with 
 cylinders on the quadruple system, thus having two L.P. cylinders. It has 
 been claimed for them that there is a superiority in the matter of balancing 
 together with a steam consumption as low as any other. Such an engine 
 may be employed with advantage when, owing to size, it is desirable to have 
 two L.P. cylinders ; but for small power it seems an unnecessary expense 
 with considerable complication ; moreover, the mechanical efficiency cannot 
 be so good as that of a three-crank engine of the same power. 
 
 Six-crank Engines for steam have been used to a very limited extent 
 by marine engineers. Fig. 38 is a good example of such an engine of very 
 large power. Oil engines have now come into use on ships of considerable 
 power, so that six cranks are common ; they are generally in two pieces, 
 and so coupled that No. 1 and No. 6 cranks are in line, as are also Nos. 2 and 
 5 and Nos. 3 and 4, each pair being at angle of 120° with the others. With the 
 two-stroke cycle they may be in sequence at 60° angles. 
 
 Eight-crank Engines will also be used largely, as they are already for 
 internal combustion systems, when the power is very great, since there is 
 a decided limit to the size of the cylinders of such engines. It may be that 
 steam engines will also be made with these numerous cranks to compete 
 with such engines and turbines. 
 
 Of Oil Engines there are three kinds used on shipboard — viz., the petrol, 
 the paraffin, and the crude or heavy oil engine. 
 
 (1) The Petrol Engine, requiring as the fuel supply the light volatile oil 
 of that name, having a flash point from 80° to 100° F., is not admissible on 
 board ordinary passenger ships, from the danger attending the carrying and 
 storing such a highly inflammable liquid. It is, however, used extensively 
 on launches and other small craft for coasting work or on rivers and small 
 lakes, and is as efficient and convenient in them as in motor vehicles on shore. 
 The ease of starting a cold engine is always a strong recommendation for 
 this oil. 
 
 (2) The Paraffin Engine using a refined light oil obtainable almost every- 
 where, and safer to carry, use, and store than petrol, inasmuch as it is not 
 nearly so volatile, and its flash point is considerably higher — viz., 120° to 
 150° F. — is employed for many purposes now, as it is almost as easy to start 
 as the former and runs quite as well. Both it and the petrol engine work 
 on the usual Otto or four-stage cycle, and require ignition by an electric 
 spark, which may be produced by a secondary battery or by a small dynamo 
 worked by the engine itself, called a magneto. 
 
 (3) The Heavy or Crude Oil Engine, which is the desideratum for ship- 
 board, on account of its comparative safety, uses oil as fuel whose flash point 
 is over 200° F., and not volatile enough to be used in the same way as the 
 light oil engines. Moreover, as the oil is unrefined, or else refuse, special 
 treatment has to be accorded to it for the different varieties of fuel used. 
 Texas, Batoum, and similar oils practically free from bitumen may be used 
 with comparative ease with suitable carburetters, etc., and with suitable 
 engines and care will run sufficiently well as to be used for driving dynamos 
 for power purposes, and most engines of this kind can be trusted, therefore, 
 to drive propellers of ships quite well for considerable periods if there is no 
 
86 
 
 MANUAL OF MARINE ENGINEERING. 
 
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 stoppage at any time sufficiently long to permit of the fooling of the cylinders, 
 etc. "With a large number of the oils obtainable at other parts of the world 
 the crude oil contains varying amounts of bituminous matter, which not only 
 
THE SEMI-DIESEL ENGINE. 87 
 
 ■causes excessive deposits of tarry matter, but what is v/orse, a formation of 
 coke, hard and refractory, which, if deposited in the cylinders or among the 
 valves, will cause serious damage to them. 
 
 The Griffin and some other engines were designed to run with the gas ox 
 -vapour distilled from crude heavy oils and their residuals by the high tem- 
 perature of the exhaust gas ; this vapour, mixed with air, was drawn into the 
 cylinders as in the ordinary gas engine and exploded there by a spark. The 
 tarry matter and pitch were in this way thrown down in the distiller and 
 •excluded from the internal parts of the engines, so that they remained clean 
 and free from grit. The early promise of these engines was unfortunately 
 not maintained as fully as was desirable, and after the Diesel engine had 
 proved so successful it displaced them. The cylinders, etc., were kept suffi- 
 •ciently cool by the usual water- jacketting. 
 
 The Diesel Engine,* which is now largely used for ship propulsion, differs 
 fundamentally from other oil engines, inasmuch as it draws in a charge of 
 air only and compresses it highly, generally to about 500 lbs. per square 
 inch, when it becomes so hot as to ignite with certainty the spray of oil injected 
 into it at the commencement of the active stroke ; moreover, the ignition 
 is gradual instead of instantaneous, and the pressure is practically that at 
 which compression ceased, so that there is no shock due to explosion ; ex- 
 pansion commences before combustion is complete, and continues to the 
 end of the stroke, when the products escape through the exhaust valve, being 
 driven out by the piston of the four-stroke cycle and by a blast of fresh air 
 in the two-stroke cycle. Almost any oil will do, but generally heavy crude 
 oils freed from highly volatile constituents, or residuals of oils from which 
 petrol, paraffin, and lubricating products have been extracted, are used ; in this 
 country tar oil and shale oil residues are home products which can be used 
 with satisfactory results, their flash point is about 220° F. 
 
 The Semi-Diesel Engine, which also uses similar fuel, works with less 
 compression, about 150 lbs. or 60 per cent, of the initial pressure on ignition, 
 and consequently requires a hot plate or bulb on which the spray impinges 
 to effect it. It works on the two-stroke cycle, and is scavenged by air pumped 
 into a receiver by the imderside of the piston. It is fairly efficient, and much 
 used on small ships with cylinders up to 16-5 inches diameter. 
 
 The Diesel engine, like the other oil engines, was single-acting, working 
 on the four-stage cycle ; air only is admitted on the first descent of the 
 piston, on its return it is compressed, sometimes to the extent of 40 atmospheres, 
 so that even with means for cooling the cylinder the temperature is very 
 high. Just as the piston is about to make a second descent the necessary 
 supply of oil is sprayed into the cylinder top by means of a jet of air com- 
 pressed higher than that in it ; the finely pulverised oil at once ignites, and 
 burns during the early part of the stroke, and so maintains the pressure 
 attained by compression (v. fig. 101) ; at the end of the stroke the exhaust 
 valve is opened, and the piston on returning to the top scavenges the cylinder 
 — that is, drives out the products of combustion. It again descends, drawing 
 in air alone as before. 
 
 To compete with the steam engine, especially on shipboard where 
 weight and space are of importance, Dr. Diesel, and those acting with him 
 at Nuremburg, devised the double-acting two-stage cycle engine, whereby 
 an explosion takes place at each end at every revolution, so that its activities, 
 so to speak, are equal to those of the steam engine. With the two-stage cycle, 
 
 * Vide Appendix A 
 
88 MANUAL OF MARINE ENGINEERING. 
 
 whether the engine be single or double-acting, the modus operandi is the same, 
 and as follows [v. fig. 51) : — The cylinder is charged with air alone, as before, 
 but now it is forced in under quite moderate pressure by a special pump 
 worked by the engine, much in the same way that the air pump of a steam 
 engine is worked [v. fig. 68). The piston compresses, and the oil is injected 
 quite as before, so that when it descends for the first time it is filled with 
 products of combustion. An exhaust port opens at the end of the stroke, 
 and by another the admission of air is made in sufficient abundance to 
 thoroughly scavenge it, and leave it filled with pure air. In the case of 
 the double-acting engine the exhaust port is at the middle of the cylinder, 
 and the piston sufficiently deep to act as the valve for each end to close it 
 before compression, and keep it closed during explosion and expansion. 
 This, of course, means a very long cylinder, and a very hot one, too, when 
 there is an explosion every revolution at each end. Both it, the piston, and 
 piston-rod will require to be water-cooled somehow. 
 
 The oil consumption of the Diesel engine when in good working order 
 is very low, under half-a-pound per horse-power developed, and in some 
 cases it has been as low as 0-38. This latter is equal to 0475 of best Welsh 
 or 0-51 Newcastle coal, or less than half that of a turbine or quadruple- 
 expansion reciprocator, but much more lubricant is expended. 
 
 It is estimated that of the total heat of combustion in the cylinder of the 
 Diesel engine 40 per cent, is usefully employed, iO per cent, passes away with 
 the gases at exhaust, and 20 per cent, is absorbed by the cooling water. 
 The cycle of operations in its cylinders may be followed in fig. 51, Nos. 1, 2, 3, 
 and the engine itself with its various pumps in fig. 50a, and in Appendix A. 
 
 The Reversal of Oil Engines is accomplished by using one or more of the 
 cylinders as an air engine, supplying it with compressed air carried for the 
 purpose in small craft, or accumulated by the oil engine when running, and 
 stored for the purpose as it is in larger ones (o. fig. 50a). 
 
 Reversal of the Propeller, when driven by oil engines, can be accomplished 
 by means of wheel-gearing, as in a motor car, or, better still, by twisting 
 the blades sufficiently to bring the base or part of them near the boss to 
 a transverse position of no pitch, when the outer part and tips will have 
 reverse pitch sufficient for navigation purposes. This was done in various 
 ways, and one similar to the method adopted by the late Mr. Bevis for modi- 
 fying the pitch is quite a successful one. In this case, instead of a nut on 
 the shaft, there is a sleeve, which is free to slide on the shaft, and is carried 
 round on it ; it is grooved as is a sliding clutch, and the lever operates in the 
 same way as in a clutch gear. In head gear the two blades of the propeller 
 are of true pitch and the usual form to be highly efficient when at normal 
 work. Since in stern gear it is only the outer portion which is effective, 
 it is desirable that the tips be made fairly broad. 
 
 Turbine Machinery in its various forms is now common in express steamers 
 of all sizes and services, as well as in the ships of the British and some foreign 
 navies, where it has quite displaced the reciprocating engine. The Parsons 
 turbine has long been a favourite, and the success of this instrument, both 
 on land and shipboard, is due so very largely to the genius of that gentleman 
 that it will ever be associated with his name. The Curtis, Zoelly, Rateau, 
 and other turbines are fitted to ships with results equalling, if not excelling, 
 those already recorded by the Parsons, the Brown-Curtis being very efficient. 
 
 The turbine must have, of necessity for high efficiency, a great peripheral 
 
TURBINE MACHINERY. 89 
 
 speed ; the diameter of rotoi or the revolutions must, therefore, be great, 
 but for very large power direct-driven the latter cannot be. In the case of 
 the s.s. " Lusitania " (fig. 39), the L.P. rotor is 140 inches diameter, with 
 the last set of blades 22 inches long, and running on bearings 33 inches 
 diameter at 194 revolutions per minute on trial, and about 185 on 
 service, and weighing 120 tons. For smaller powers the rate of revolution 
 is much higher, and in case of a warship of 5,000 II. P. per shaft the rate of 
 revolution is as much as 500, and higher, even up to 700 revolutions, in 
 smaller ships of 4,750 H.P. per shaft. Such revolutions necessitate not only 
 a screw of small diameter, but one of such very small pitch that the pitch 
 ratio is so low that the efficiency of the propeller is lower than in competing 
 ships. The slip ratio, however, is wonderfully small in the turbine-driven 
 ship, taking all these things into account, being only 15-3 per cent, in the 
 " Lusitania," and seldom over 25 per cent, in others where revolution speed 
 is not very excessive. Formerly Sir C. Parsons fitted more than one screw 
 to each shaft, but without the success he anticipated ; the German Admiralty 
 also tried the same method for improving the propeller efficiency of the 
 turbine-driven ship, " Lubeck," with the same disappointment ; conse- 
 quently for such a ship a single screw of moderately small diameter and large 
 disc ratio is the rule. That is, the screw is somewhat larger than would be 
 fitted, if the regard of the designer were limited to turbine efficiency only. 
 As a matter of fact, in all cases of this kind the choice of screw and all that 
 pertains to it is governed by combining the efficiency of propeller and gene- 
 rator, as will be seen later on.* The turbine, when of a power exceeding 
 1,000 S.H.P., is superior to the best reciprocator in steam consumption 
 per unit of power ; the turbine of any size has a higher mechanical efficiency 
 than a reciprocator of equal power ; it occupies about the same amount 
 of floor space as the ordinary triple and quadruple engine, but is of less 
 height, so that much of it can go under deck, consequently the engine hatches 
 can be very much smaller for them than the reciprocator. The consumption 
 of lubricants is less and fewer attendants are required in the engine-room 
 when on service. The difference in weight is trifling, but the prime cost 
 and the repair and wear and tear account of the modern make of turbine 
 compares favourably with that of the average reciprocator. On the other 
 hand, the steam efficiency of the turbine falls off on reduction of load, and 
 since the marine turbine suffers reduction in velocity when the speed 
 of the ship is reduced, the fall in efficiency is considerable ; so much so, 
 indeed, that in the case of the " Lusitania," where the consumption of steam 
 per S.H.P. per hour of the turbines alone was only 12-77 lbs. at the full speed 
 of 25-4 knots per hour, it was as high as 21-23 lbs. at 15-77 knots. The 
 corresponding coal consumptions, which, however, of course included that 
 due to the requirements of the auxiliary machinery, were 1-46 lbs. and 2-76 
 lbs. respectively per S.H.P per hour. With the reciprocator there is no such 
 rapid increase in coal consumption at the lower powers ; on the contrary, 
 the rate is lower at moderately reduced speeds than at full, and at quite 
 low speeds is not very high. For example, on the trials of H.M.S. " Achilles " 
 the steam consumption at full speed of 23'275 knots, with the engines 
 developing 24,000 I.H.P., was 19-9 lbs. per I. H.P. .per hour, and the coal 
 2-03 lbs., while at 14-6 knots it was only 16-95 lbs. of steam and 1-88 lbs. 
 of coal; moreover, at 21-58 knots and 16,000 I.H.P. the coal consumption 
 
 * The employment of gearing has solved the problem of screw and turbine efficiency 
 {vide Appendix A). 
 
90 
 
 MANUAL OF MARINE ENGINEERING. 
 
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ADMIRALTY TURBINE TEST. 
 
 91 
 
 was 1'85 lbs. only. Her engines were four-cylinder four-crank triple- 
 expansion reciprocators. 
 
 The Admiralty, to test the Turbine, caused to be carried out some exhaustive 
 comparative trials with H.M.S. " Amethyst," fitted with Parsons turbines, 
 and her sister ship, the " Topaze," having the four-cylinder triple-compound 
 reciprocators. These ships are each 360 feet long, 40 feet beam, and 14*5 feet 
 mean draught of water ; their displacement is 3,000 tons, and wetted skin 
 about 16,100 square feet, with block coefficient of fineness of 0*503. Their 
 boilers were similar in all respects, and their trials were conducted on exactly 
 
 13,000 
 
 $>I2 2,000 
 10 1, 000 
 
 20 21 22 23 
 
 16 17 16 19 
 Scale of Speed. 
 
 Fig. 40. — Trials of "Amethyst " (Turbines) and Sister Ships (Reciprocators) 
 
 25 Knots 
 
 was exerting 
 
 the same lines, with the result that, whereas when the " Topaze 
 the maximum power of her reciprocating engines, 9,868 I.H.P., the speed 
 attained was 22'103 knots, the turbines of the "Amethyst" developed a 
 S.H.P. equivalent to about 14,200 I.H.P., which drove her at a speed of 
 23*63 knots. Fig. 40 gives the curves of power consumption, etc., of the 
 two ships and their sister ships fitted with reciprocators. 
 
 The Experiments made in the U.S. of America with the cruiser " Bir- 
 mingham," fitted with triple-compound reciprocators, and the " Salem," 
 having Curtis turbines, is equally interesting. These ships are otherwise 
 
92 
 
 MANUAL OF MARINE ENGINEERING. 
 
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COMBINATION OF TURBINES WITH RECIPROCATORS. 
 
 93 
 
 alike, and 420 feet long, 47 feet beam, 16*75 mean draft, the displacement 
 4,700 tons, and the wetted skin about 22,000 square feet. They are of 
 very fine form, the block coefficient being 0*50. The Curtis turbines drove 
 the " Salem " at a speed of 25-947 knots, with 19,200 S.H.P., and a coal 
 consumption of 201 lbs. per S.H.P. per hour (equivalent to 181 per I.H.P. 
 of a reciprocator). The reciprocators drove the "Birmingham" at 24*325 
 knots speed, with 15,540 I.H.P., consuming 1*92 lbs. of coal ; at 12*23 knots 
 with 1,600 I.H.P., the consumption of coal of the " Birmingham " was 
 2-89 lbs., against the 2-68 lbs. of the "Salem" at 11*93 knots with 
 1,360 S.H.P. 
 
 Experiments were made by the German Government with the cruiser 
 '* Lubeck," of 3,170 tons displacement, 341 feet long, 43 feet beam, and 
 16" 4 feet draft of water, driven by turbines operating on four shafts, each 
 shaft having two screws at a speed of 23 knots, and the twin-screw sister 
 ship, " Hamburg," having reciprocating engines ; but in this case the superi- 
 ority of the turbine was not demonstrated, inasmuch as it took 14,158 H.P. 
 to attain a speed of 23*16 knots with the "Lubeck," and only 11,582 for 
 23*17 knots with the reciprocators. Doubtless, however, this was in no 
 small measure due to the screw arrangement, which was a bad one, and 
 probably the propeller efficiency was very low. Since those experiments 
 were made, the German Admiralty have followed the lead of the British, 
 and now are having turbines of kinds fitted in their warships. 
 
 Table XlVa. summarises the above. 
 
 The following table is interesting and instructive, showing, as it does, 
 the comparative steam consumptions of the " Amethyst " and " Topaze " 
 at speeds varying from 10 knots to 22, including the steam used by the aux- 
 iliaries. It must be borne in mind, when considering the same, that the 
 engines of the " Topaze " were not designed primarily for economy, conse- 
 quently their consumption of steam per I.H.P. is very high compared with 
 that of the quadruple-compound engines of the mercantile marine, such as 
 fitted in s.s. " Saxonia," constructed by John Brown & Co., and tested by 
 the Admiralty Boiler Committee, where it was found to be 14*5 lbs. per I.H.P. 
 when on service. At the same time, the thermal efficiency of the " Topaze's " 
 engines was by no means bad considering the conditions under which they 
 worked, for her steam consumption in main engines only was but 16*91 at 
 a speed of 20 knots and 15'45 at 18 knots : — 
 
 TABLE XV.- 
 
 Water Consumption per I.H.P. per Hour of H.M.S. 
 
 on Progressive Trials. 
 
 " Amethyst " and " Topaze 
 
 Speed in Knots. '. 10 
 
 11 
 
 12 
 
 13 
 
 14 
 
 15 
 
 16 
 
 17 
 
 18 
 
 19 
 
 20 
 
 21 
 
 22 
 
 1 
 H.M.S. " Amethyst" (Turbs.), 29-3 
 
 H.M.S. "Topaze" (Recipros.), 238 
 
 260 
 22-0 
 
 237 22 
 20-6 19-6 
 
 20-4 
 18-8 
 
 19-0 
 18-4 
 
 179 
 18-3 
 
 168 
 184 
 
 159 
 187 
 
 1.5-2 
 19-2 
 
 14-7 
 198 
 
 14-3 
 20-5 
 
 140 
 
 21-4 
 
 Combination of Turbines with Reciprocators seems to be the most likely 
 development of marine steam machinery in the future, and it has already 
 
94 MANUAL OF MARINE ENGINEERING. 
 
 been adopted successfully by Denny & Co., also by Harland & Wolff 
 in the Transatlantic steamships recently built by them for the White Star 
 Company and others. Since the triple-expansion engine is more economic 
 than is the boiler end half of a turbine, while the other or condenser end of 
 the latter is much more economic than the reciprocator, and can make good 
 use of a high vacuum, such a combination of the two is obviously a fitting 
 one, and a desirable thing. Further, since at slow and cruising speeds the 
 turbine is not economical, while the triple-compound engine is very fairly 
 so, these latter engines can be employed by themselves to propel without 
 using the turbines by exhausting direct to the condenser. Fig. 41 shows 
 an arrangement proposed by Parsons for dealing with a three-screw 
 ship ; here each wing screw is driven by a four-cylinder triple-expansion 
 engine arranged to exhaust either to its own condenser or to a low-pressure 
 turbine operating a central screw and exhausting to the same condenser. 
 Fig. 42 is an example by the same gentleman for a four-screw arrangement, 
 in which each inner screw has its own triple-compound engine exhausting 
 direct to its own condenser, or to a low-pressure turbine operating a wing 
 screw, and exhausting to the condenser. That is, in each case a turbine 
 is interposed between the L.P. cylinder and the condenser, in order that 
 the steam from it at 10 lbs. pressure absolute may be usefully employed 
 in expanding down to the pound or even less pressure of the 
 condenser. 
 
 Messrs. Denny Brothers, of Dumbarton, fitted the s.s. " Otaki " with 
 triples and low-pressure turbines, and demonstrated that the gain over the 
 arrangement with triple engines in the sister ship, s.s. " Orari," was as much 
 as 17 per cent. Careful experiments with triple-expansion engines at electric 
 power stations on shore show that fully 15 per cent, more power is obtained 
 with the same consumption of steam if a low-pressure turbine is interposed 
 in this way between it and the condenser. 
 
 The following are the figures given by Com. Wisnom, of Denny's, from 
 the trials of the above two steamers : — 
 
 The Performance of s.s. " Otaki," having two sets of ordinary triple- 
 expansion engines, each driving a wing propeller as in a twin-screw ship, 
 and both exhausting to a low-pressure turbine driving a propeller on the 
 middle line as in a single-screw ship; built by Denny Bros., of Dumbarton, 
 for the New Zealand S. Co., and sister ship to the twin-screw steamer " Orari," 
 which was, however, 4*5 feet shorter. 
 
 The " Otaki " has a displacement of 9,900 tons on 27'5 mean draught, 
 and her principal dimensions — 
 
 Length between perpendiculars, „ . . 464*5 feet. 
 
 Beam moulded, 60*0 „ 
 
 Depth ,, 34 - „ 
 
 24"5"-39"-58" 
 Each of her reciprocating engines has cylinders - ~-^ Q „~ , while each 
 
 24'5"-4r3"-69' y 
 of those of the sister ships, " Opawa " and " Orari," are j^„ . The 
 
 turbines of the " Otaki " has a rotor 90 inches diameter. 
 
COMBINATION OF TURBINES WITH KECIPKOCATOKS. 
 
 »5 
 
 n 
 
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96 
 
 MANUAL OF MARINE ENGINEERING. 
 
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PERFORMANCE OF S.S. " OTAKI.' 
 
 97 
 
 The following is a comparative summary of results of trials at 146 knots 
 speed on the measured mile : — 
 
 Name of Ship. 
 
 E.H.P. 
 
 I.H.P. 
 
 Propulsive 
 Coefficient 
 
 Water Consumption 
 
 per Hour. 
 
 Total. 
 
 Per E.H.P. 
 
 Pei I.H.P. 
 
 3-screw s.s. " Otaki " (turbo- 
 
 recipro.), 
 2-screw s.s. " Orari " (reci- 
 
 pros.), .... 
 
 3,350 
 3,210 
 
 5,880 
 5,360 
 
 Per cent. 
 57 
 60 
 
 73,300 
 88,300 
 
 21-9 
 27-5 
 
 13-7 
 16-5 
 
 Gain per cent, in " Otaki," 
 
 •• 
 
 •• 
 
 
 17 
 
 20 
 
 17 
 
 TABLE XVI.— Measured Mile Trials of s.s. " Otaki," 
 October 31, 1908, on 20 Feet Mean Draught. 
 
 
 Mean of 
 
 Mean of 
 
 Mean of 
 
 Mean of 
 
 
 A Runs. 
 
 B Runs. 
 
 C Runs. 
 
 D Runs. 
 
 Total horse-power, being I.H.P. (recipros.) -f 
 
 
 
 
 
 S.H.P. (turb.) 
 
 6,857 
 
 5,348 
 
 4,704 
 
 3,282 
 
 Mean speed, .... knots, 
 
 1502 
 
 14-28 
 
 13-83 
 
 12-52 
 
 Revolutions, recipros., .... 
 
 103-5 
 
 97-9 
 
 93-5 
 
 83-4 
 
 „ turbine, . . t 
 
 224-5 
 
 209-7 
 
 197-2 
 
 1721 
 
 Total water consumption per hour, . lbs., 
 
 82,000 
 
 67,300 
 
 60,200 
 
 44,600 
 
 „ „ „ per H.P. „ 
 
 11-95 
 
 12-6 
 
 12-8 
 
 13-6 
 
 Mean absolute pressure at H.P. cylinder, „ 
 
 193 
 
 178 
 
 166 
 
 135 
 
 „ „ turbine inlet, „ 
 
 9-5 
 
 7-62 
 
 6-76 
 
 5-0 
 
 Vacuum at exhaust end of turbine, 
 
 28-1 
 
 28-2 
 
 28-4 
 
 28-5 
 
 ,, on condenser gauge, 
 
 28-2 
 
 28-4 
 
 28-3 
 
 28-5 
 
 Temperature of sea water, . . . F.° 
 
 56 
 
 56 
 
 56 
 
 56 
 
 „ circulating discharge, . 
 
 70 
 
 67 
 
 70 
 
 70 
 
 „ hot well, .... 
 
 72 
 
 70 
 
 73 
 
 74 
 
 Steam consumption based on the I.H.P. of 
 
 
 
 
 
 s.s. " Orari " by tanks, .... 
 
 13-66 
 
 13-7 
 
 13-8 
 
 13-07 
 
 As measured by pumps per I.H.P. per hour, . 
 
 14-12 
 
 14-1 
 
 14-3 
 
 15-2 
 
 Figs. 43 and 44 show the general arrangement of the machinery of the 
 s.s. " Olympic," built by Harland & Wolff for express service between 
 England and New York. She is one of the largest ships, and has worked 
 with satisfactory results ; she is 860 ft. long, 92-75 ft. beam, and 32-5 draft 
 of water, displacement 50,000 tons, and I.H.P. 54,800 ; propelled by three 
 screws, the two wing ones worked by reciprocating triple-compound engines, 
 each having four cylinders, 54, 84, 97, and 97 inches diameter, and 63 inches 
 stroke, and the middle by low-pressure Curtis turbine, taking the steam 
 supply from the L.P cylinders of the reciprocators (v. fig. 44). 
 
 A Development with existing Single-screw ShiDS could be made by fitting 
 two low-pressure turbines abaft the old triple engine, each operating a wing 
 screw, and both exhausting to the old condenser ; this would probably 
 
 7 
 
98 
 
 MANUAL OF MARINE ENGINEERING. 
 
ENGINE-ROOM OF R.M.S. "OLYMPIC." 
 
 99 
 
 
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 be quite as good for cargo and " mixed " steamers as pulling out the old 
 engines and fitting two turbines geared to the original screw shafting (tig. 45), 
 
TOO 
 
 MANUAL OF MARINE ENGINEERING. 
 
 as done by Sir C. Parsons in the s.s. " Vespasian." Fig. 46 shows such an 
 arrangement proposed by this gentleman ; it is one quite easily carried out. 
 and although it may be questionable if it is worth doing to an ordinary cargo 
 boat, it certainly would be quite a good thing for very many of the combined 
 cargo and passenger steamers designed . for long voyages at speeds from 
 12 to 15 knots with single screws. The gain of 13-7 per cent, in power shown 
 by him (in the schedule later on) is very material, and the alternative 15 per 
 cent, of fuel important in all places, but very much so where coal is 20s. to 
 30s. per ton delivered on board the ship in ordinary times. 
 
 Oil Engines of moderate power are usually of the well-known single-acting 
 type (figs. 17 and 49), having trunk pistons connected direct to the crank-pins 
 
 Fig. 45. — Wheel-Gearing of s.s. " Vespasian " (Parsons). 
 
 by ordinary connecting-rods and valve-gear wheel driven ; they seem to work 
 very well at steady loads, and are handled by compressed air. Increases 
 of power are obtained by increasing the numbers of cylinders, so that it is 
 not at all uncommon to find eight cylinders in line operating on one line of 
 shafting with a screw at the outer end. There is, of course, in this way 
 no limit to the number of cylinders, and probably the cost of increase by 
 this method is no greater than if the greater power were obtained with 
 cylinders of large size*, sometimes cylinders are placed in tandem axially — 
 that is, in rear of or above each cylinder is another operating with it on 
 the same connecting-rod. 
 
OIL ENGINES. 
 
 101 
 
 Trouble, however, is experienced 
 sometimes with these trunk pistons 
 when of large size, so that recourse is 
 had now to the piston-rod type when 
 the cylinders increase in diameter ; at 
 present 30 inches diameter* is con- 
 sidered about as large as should be 
 made for marine engines on the Diesel 
 system, where the initial pressure is 
 exceedingly high, especially if com- 
 pared with the mean. It is sometimes 
 as high as 40, but is generally 35 
 atmospheres in these engines at 
 commencement of the stroke, but, 
 owing to the high compression 
 effected, there is little or no shock 
 at commencement of stroke. This 
 engine, however, notwithstanding such 
 initial pressures, has become popular 
 in this country, as it had been for 
 some time on the Continent, where 
 it has proved successful when work- 
 ing with that comparatively safe 
 fuel heavy oil with a high flash point 
 or residuals. 
 
 The Design of Oil Engines was at 
 first very similar to that followed for 
 land engines. Now, however, the 
 tendency is to conform to the marine 
 practice found to be the best for 
 steam engines ; their builders also 
 adopt the enclosed type with forced 
 lubrication, and so obtain good and 
 safe running at high rates, of revolu- 
 tion. Fig. 17 is a good example of a 
 Thornycroft 100 H.P. marine petrol 
 engine, the reversing of which is made 
 by means of a clutch gear, which 
 does well enough in small craft, 
 being similar to the method by 
 which motor-driven vehicles are re- 
 versed. 
 
 The Diesel Engine is generally 
 designed to work on the usual 
 four-cycle system, and its modus 
 operandi is as already described 
 ('■. figs. 50 and 50a), but at- 
 tached to it is a three-stage air 
 compressor, which not only supplies 
 * Vide Appendix A 
 
102 
 
 MAM7AL OF MARINE ENGINEERING. 
 
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TWO-CYCLE DIESEL ENGINE. 10 
 
 9 
 
 the air for injecting the oil, but provides a means of working the engine and 
 reversing it when required by using the oil cylinders as in an air engine, 
 when, by special means, it is put out of action as an oil engine. The engines 
 on the two-cycle system have also one or more low-pressure pumps (v. fig. 
 50a) to supply the air for scavenging and filling the cylinders. In some 
 marine designs these pumps are worked by levers, and like the ordinary 
 air pumps of a steam engine. The high rate of compression of the 
 air prevents shock on exploding the mixture of air and oil vapour, as would 
 be the case if the load came on suddenly, as it does in the ordinary engine 
 when there is no such compression. The ratio of maximum to mean 
 pressure (4-31) is exceedingly high in all such engines, and consequently 
 the rods, framing, crank shafts, etc., must be large by comparison with 
 those of steam engines, it follows that the weight of these engines per horse- 
 power is very great, and not much less than that of a steam installation 
 including the boilers ; on the other hand, the consumption of oil fuel is only 
 about 40 per cent, that in an oil-fired steamship. 
 
 The double-acting cylinder in which explosives take place on both sides 
 of the piston have yet to be proved equal to continuous service ; doubtless 
 the pistons must be in that case water-cooled, and the stuffing-boxes most 
 carefully made and maintained to be satisfactory. To compete with large 
 powers the double-acting engine is desirable, but it still remains for engineering 
 skill and resource to get over these practical difficulties. In time, too, the 
 reversing may not be effected by quite such clumsy and indirect means as 
 prevail at present with gearing or subsidiary compressed air arrangements. 
 Fig. 49 is a sectional view of a Diesel engine as made by Mirrlees 
 Bickerton Company for small power on the four-cycle system, and 
 non-reversible. 
 
 Double-acting Diesel two-cycle engines of considerable power have been 
 made in large sizes on the Continent. Here the cycle with its compression 
 and combustion follows its course as in the single-acting engine, but in 
 this case on both sides of the piston, as in fig. 48. The power developed 
 is thus practically doubled in a given size of cylinder ; or the same power 
 obtained with about half the cylinder capacity. There are the usual practical 
 objections to this system of overheating both pistons and cylinders, 
 although they are water-cooled, and the difficulty with stuffing-boxes exposed 
 to such high temperatures. Moreover, to obtain such high compression 
 the clearance must be very small, for with 35 atmospheres at each end it 
 must be less than 3 per cent, of the cylinder capacity, and consequently 
 the stroke clearance about 2 per cent. — that is, with an engine of 10 inches 
 stroke it is not more than ^u lnca at eacn en( J> an d must not vary materially 
 from this at any time. 
 
 The stuffing-box difficulty may be overcome by having hollow rods, 
 through which water is passed to the pistons, and kept in circulation as 
 Dr. Kirk did with steam for heating the L.P. pistons of the early compound 
 engines made by J. Elder & Co. for the Navy. It is, however, very 
 doubtful if the two-cycle double-acting engine will be satisfactory for even 
 intermittent running, and there is reason to think it is unlikely to be so for 
 continuous work on shipboard in large sizes. 
 
 The Two-cycle Diesel Engine differs from the four inasmuch as the 
 fresh air is admitted above the piston when it is at the bottom of the stroke 
 
104 
 
 MANUAL OF MARINE ENGINEERING. 
 
 a 
 
 
 o 
 
 u 
 
 s 
 
 "3 
 
 a 
 
 -a 
 '3 
 
 50 
 
 3 
 3 
 
 after combustion, the products having begun to escape at six-sevenths the 
 stroke are by it ejected and replaced, the scavenging orifice closing again 
 
THE FUEL CONSUMPTION. 
 
 105 
 
 one-seventh from the bottom ; the piston continuing its stroke compresses 
 it to the end when the oil is sprayed in as before. In this way an impulse is 
 made at every revolution, instead of at every other one as in the " Otto " 
 cycle. This still further increases the power developed by a unit of cylinder 
 capacity, but it likewise increases the heat production and the difficulties 
 that arise from high temperatures. 
 
 The Fuel Consumption of these engines is generally less than 04 pound 
 of oil per horse-power, and in very favourable cases as low as 0-348 lb. 
 Taking 10 lbs. of steam as a very good performance for a turbine, and 16 lbs. 
 
 ft 
 
 Fig. 49. — Single- Acting " Mirrlees-Diesel " Marine Oil Engine, 
 of steam per pound of oil when burned to be produced in a good boiler, then 
 Consumption of oil per hour of a turbine = 10 -f- 16, or 0*625 lb. 
 
 If the consumption of the " Lusitania " be taken as 12*77 lbs. in the turbines 
 alone, and 14*46 lbs. the total for all purposes, the oil fuel consumption 
 will be 0*798 and 0*903 lb. per S.H.P., equal to 0*766 and 0*867 per I.H.P. 
 respectively. 
 
 Taking, however, the consumption of the " Otaki " with the combined 
 turbo-reciprocators as 12 lbs. of steam per hour, the oil fuel for her would 
 
106 
 
 MANUAL OF MARINE ENGINEERING. 
 
 be 0*75 lb. Although this consumption of oil fuel is considerably greater 
 than that of the oil engine, the amount of lubricating oil in the Diesel is 
 
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 very much higher ; in fact, excessive compared with that of a reciprocating 
 steam engine of equal power. 
 
 When oil engines are used on board sea-going ships, it is necessary for 
 
THE FUEL CONSUMPTION. 
 
 107 
 
 them to have power for steering and other purposes generated by an inde- 
 pendent oil engine and electrically distributed. The whistle may be blown 
 
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 CS 
 
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 with stored compressed air from the main engine compressor, so that there 
 are no insuperable difficulties now these engines can be relied on to stop, start, 
 
108 
 
 MANUAL OF MARINE ENGINEERING. 
 
 and reverse quickly, and to run for a week on end without stopping to clean 
 carburetters, etc., and certainly without stopping unexpectedly at awkward 
 times from this latter cause. Mr. Westgarth thought it prudent to provide 
 a small steam generator for such purposes above-named in the ship fitted 
 by him with oil engines, so that there might not be so many novelties at 
 one time. 
 
 Gas Engines on board ship were pretty much of the same general design 
 as the oil engine, but they require a gas producer, etc., to supply them with 
 fuel, which adds to their weight and the space occupied by them. More- 
 over, as Prof. Vivian Lewis very properly pointed out, " the suction plant 
 suffers from the limitation that before it can achieve commercial success 
 afloat a form of generator and scrubber, occupying small space, must be 
 devised, in which bituminous coal can be used as the fuel to be gasefied, 
 J and the gas supplied freed from all tar vapour. ... I am not aware 
 
 1. 
 
 Charged with Compressed 
 Air, and Fuel being ad- 
 mitted. 
 
 Expansion completed and 
 Piston about to open 
 Exhaust. 
 
 End of Down Stroke. 
 Scavenge Valve open. 
 Clearing out. 
 
 Fig. 51. — Two«Cycle Oil Engine (Diesel System), showing the Three Stages. 
 
 that it has yet been successfully done. The mechanical troubles of caking 
 and arching of the fuel in the generator can be overcome, but many years' 
 experience of efforts to decompose or get rid of tar vapour has impressed 
 me with a great respect for the difficulties of the problem and a perfectly clean 
 gas, absolutely free from tar vapour, is the first essential for success with the 
 gas engine." 
 
 All this is very true, though it must be a matter of regret, seeing that 
 this country abounds in coal, but has very little oil. For this reason alone 
 the adoption of the oil engine in British waters is scarcely politic, either in 
 warships or cargo-boats. 
 
STEAM TURBINES. 109 
 
 CHAPTEK IV. 
 
 STEAM USED EXPANSIVELY. 
 
 In the Reciprocating engine work is done by the elastic force of the steam 
 acting on the pistons, and pressing them forward on their strokes against 
 the back pressure behind them during its expansion from the time it enters 
 the cylinder to the time it is allowed to escape to the free atmosphere in the 
 case of a non-condensing or to the condenser of a condensing engine. 
 
 From the time of entering to the time of cut-off expansion is taking place, 
 though it is slight and often not appreciable ; after cut-off it is real, considerable, 
 and quick ; it is continuous to the end, and the rate is expressed by the ratio 
 of the capacity of the L.P. cylinder to that portion of the H.P. which was 
 filled with steam at cut-off. This is nominally the rate, and would be really 
 so if the admission valve were large and wide open till it closed and cut-off 
 made suddenly, as is the case with the Corliss or " drop " valves. In practice, 
 however, at the exact point of cut-off with the ordinary slide valves the 
 steam is wire-drawn down considerably below the pressure in the valve 
 chamber, which latter may be taken as the initial pressure. 
 
 With Steam Turbines steam is admitted continuously instead of inter- 
 mittently as in reciprocators ; the expansion commences immediately it 
 enters the nozzles or their equivalents, and may be wholly carried out in one 
 nozzle, and the whole velocity due to the fall from boiler pressure to exhaust 
 pressure generated at once, and the kinetic energy expended on the blades 
 of a single rotor, as shown in fig. 52. The well-known De Laval turbine 
 is on this principle, and consequently runs at a very high rate of revolution, 
 since the velocity acquired by steam in falling from 160 lbs. to 3 lbs. absolute 
 (vacuum 24 inches) is 3,660 feet per second, and the peripheral velocity of 
 rotor for this flow, if the efficiency is to be good, must be not much less than 
 1,800 feet ; for a turbine of this kind with a rotor 38 inches diameter, the 
 rate of revolution should be 180 per second, or 10,800 per minute ; as a matter 
 of fact, a De Laval turbine of 300 H.P. has a rotor 30 inches diameter running 
 at 10,600 revolutions per minute. 
 
 The Expansion may be in Stages, however, and fig. 52a shows how this 
 may be accomplished in a rudimentary way by causing the first rotor to run 
 in a chamber where the pressure is below that at entry, but is somewhat 
 above that at the final exhaust to condenser. The velocity of flow will then 
 be much less than that stated, as the drop will be less (say to 25 lbs. abso- 
 lute instead of 3). The steam from that chamber will pass into and through 
 auother nozzle, where further expansion takes place with a renewal of velocity 
 to the steam, thereby giving it further kinetic energy to be expended on 
 the blades of a second rotor. In all cases of turbines, the rate of expansion 
 is expressed by the ratio of the initial pressure to that at exhaust — that is, 
 Pi + Vo = r. 
 
 In the above example r — 160 -s- 3, or 533. 
 
110 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Moderately Moist Steam expands in accordance with Boyle and Marriott's 
 law — viz., whereby the pressure varies inversely as the volume — that is, 
 pr = c. 
 
 Then, 
 
 mi 1 + hyp. log r 
 lhe mean pressure = p x — ■ — — — . 
 
 The hyperbolic logarithm of a number may be found by multiplying the 
 common logarithm of that number by 2 "302585. 
 
 There is a simplicity in this rule that commends it to the practical mind, 
 and as steam in marine engines is usually fairly moist, it may generally 
 be used to solve with sufficient accuracy the every-day problems connected 
 
 ~ . . ;:,y$L 
 
 Fig. 52. -Turbo-Mofcoe 
 (i>e Laval system). 
 
 Steam from 
 boiler 
 
 Fir. 52a. — Compound Turbo-Motot 
 (Do Laval system). 
 
 with the marine engine. Such an expansion at constant temperature is called 
 isothermal. 
 
 The terminal pressure = 
 
 _ Pi 
 
 Therefore, 
 
 YD 
 
 (1) Ratio of mean pressure to terminal pressure = -— . 
 
 .„ - Pi 
 
 (2) Ratio of terminal pressure to mean pressure = 
 
 rp m 
 
 (3) Ratio of maximum pressure to mean pressure = — . 
 
 (4) Dry steam is assumed to expand pv" = constant. 
 
STEAM TURBINES. 
 
 Ill 
 
 The following table will be found useful, and contains the multipliers 
 for ascertaining the mean pressure of steam when expanding on Boyle's 
 jaw, as well as when expanding adiabatically. 
 
 TABLE XVII. — Steam used Expansively. 
 
 
 l 
 
 rp m 
 
 _£l 
 
 El 
 
 Pm 
 
 £»* 
 
 /' 
 
 r 
 
 P\ 
 
 rp n 
 
 Vm 
 
 Pi 
 
 20 
 
 0-050 
 
 4-00 
 
 0-250 
 
 5-00 
 
 0-1998 
 
 0-186 
 
 18 
 
 0-055 
 
 3-89 
 
 0-256 
 
 4-03 
 
 0-2161 
 
 0-200 
 
 16 
 
 0-062 
 
 3-77 
 
 0-265 
 
 4-24 
 
 0-2358 
 
 0-220 
 
 15 
 
 0-066 
 
 3-71 
 
 0-269 
 
 4-05 
 
 0-2472 
 
 0-230 
 
 14 
 
 0-071 
 
 3-64 
 
 0-275 
 
 3-85 
 
 0-2599 
 
 0-242 
 
 13-33 
 
 0-075 
 
 3-59 
 
 0-279 
 
 3-72 
 
 0-2690 
 
 0-254 
 
 13 
 
 0-077 
 
 3-56 
 
 0-280 
 
 3-65 
 
 0-2742 
 
 0-258 
 
 12 
 
 0-083 
 
 3-48 
 
 0-287 
 
 3-44 
 
 0-2904 
 
 0-271 
 
 11 
 
 0-091 
 
 3-40 
 
 0-294 
 
 3-24 
 
 0-3089 
 
 0-292 
 
 10 
 
 0-100 
 
 3-30 
 
 0-303 
 
 3-03 
 
 0-3303 
 
 0-314 
 
 9 
 
 0-111 
 
 3-20 
 
 0-312 
 
 2-81 
 
 0-3552 
 
 0-340 
 
 8 
 
 0-125 
 
 3-08 
 
 0-321 
 
 2-60 
 
 0-3849 
 
 0-370 
 
 7 
 
 0-143 
 
 2-95 
 
 0-339 
 
 2-37 
 
 0-4210 
 
 0-408 
 
 6-66 
 
 0-150 
 
 2-90 
 
 0-345 
 
 2-30 
 
 0-4347 
 
 0-417 
 
 6-00 
 
 0-166 
 
 2-79 
 
 0-360 
 
 2-15 
 
 0-4653 
 
 0-450 
 
 5-71 
 
 0175 
 
 2-74 
 
 0-364 
 
 2-08 
 
 0-4807 
 
 0-466 
 
 500 
 
 0-200 
 
 2-61 
 
 0-383 
 
 1-92 
 
 0-5218 
 
 0-506 
 
 4-44 
 
 0-225 
 
 2-50 
 
 0-400 
 
 1-78 
 
 0-5608 
 
 0-540 
 
 4-00 
 
 0-250 
 
 2-39 
 
 0-419 
 
 1-68 
 
 0-5965 
 
 0-582 
 
 3-63 
 
 0-275 
 
 2-29 
 
 0-437 
 
 1-58 
 
 0-6308 
 
 0-616 
 
 3-33 
 
 0-300 
 
 2-20 
 
 0-454 
 
 1-51 
 
 0-6615 
 
 0-648 
 
 3-00 
 
 0-333 
 
 2-10 
 
 0-476 
 
 1-43 
 
 0-6993 
 
 0-670 
 
 2-86 
 
 0-350 
 
 2-05 
 
 0-488 
 
 1-39 
 
 0-7171 
 
 0-707 
 
 2-66 
 
 0-375 
 
 1-98 
 
 0-505 
 
 1-34 
 
 0-7440 
 
 0-733 
 
 2-50 
 
 0-400 
 
 1-01 
 
 0-523 
 
 1-31 
 
 0-7664 
 
 0-756 
 
 2-22 
 
 0-450 
 
 1-80 
 
 0-556 
 
 1-24 
 
 0-8095 
 
 0-800 
 
 2-00 
 
 0-500 
 
 1-69 
 
 0-591 
 
 1-18 
 
 0-8465 
 
 0-840 
 
 1-82 
 
 0-550 
 
 1-60 
 
 0-626 
 
 114 
 
 0-8786 
 
 0-874 
 
 1-66 
 
 0-600 
 
 1-51 
 
 0-662 
 
 1-10 
 
 0-9066 
 
 0-900 
 
 1-60 
 
 0-625 
 
 1-47 
 
 0-680 
 
 1-09 
 
 0-9187 
 
 0-913 
 
 1-54 
 
 0-650 
 
 1-43 
 
 0-699 
 
 1-07 
 
 0-929:. 
 
 0-926 
 
 1-48 
 
 0-675 
 
 1-39 
 
 0-718 
 
 1-06 
 
 0-9405 
 
 0-940 
 
 When steam expands in accordance with the law p v = constant, the 
 curve drawn through the extremities of ordinates representing the pressure 
 at any position of the piston is a hyperbola. The mean height of such a 
 system of ordinates is found by the formula given above ; this mean height 
 will represent the mean pressure. 
 
 The mean pressure may be obtained without the aid of logarithms, by 
 resorting to arithmetical calculation of the ordinates, and finding the mean 
 by the method usually followed with indicator diagrams. 
 
 Example. — Initial pressure 80 lbs., rate of expansion 5. Suppose the 
 length of stroke divided into ten equal parts by points 1, 2, 3, . . . 9, 
 10. The cut-off is !. or two-tenths the stroke. 
 
112 
 
 MANUAL OF MARINE ENGINEERING. 
 
 The pressure at commence of stroke is 
 
 1 tenth 
 2 
 3 
 4 
 5 
 6 
 7 
 8 
 9 
 10 
 
 >> 
 »» 
 »> 
 
 ii 
 ii 
 
 80 lbs. 
 
 80 „ 
 80 .. 
 
 f of 80 or 53-33 
 4000 
 32-00 
 26-66 
 22-86 
 20-00 
 17-78 
 16 00 
 
 2. 
 
 * 
 2 
 
 T 
 i! 
 
 ii 
 
 5- 
 2 
 J 
 
 )> 
 )) 
 »> 
 )1 
 II 
 >> 
 )) 
 
 J>» - 
 
 _ 1 
 
 ( — *- + 80 + 80 + 53-33 + 40 + 32 + 26-66 + 22-86 + 20 + 17-78) 
 
 = 42-063 lbs. 
 
 By reference to Table xvii., with 5 for the rate of expansion. ^~ = - 5218, 
 
 and for 80 lbs. pressure p m = 41*744 lbs., or about 0*75 per cent, less than 
 that given by the summation above, the excess of which is clue to the small 
 number of points of observed pressure. 
 
 Graphic Method. — Professor Rankine proposed * a geometric method of 
 ascertaining the mean pressure, which is of interest and value, and which 
 first appeared in the Engineer. Draw a straight line, A B, of definite length, 
 
 A B 
 
 produce it to A C, so that A C = —r-. Through 
 
 A draw A D at right angles to CAB. With C as 
 centre, and C B as radius, draw the arc of a circle, 
 
 D A 
 
 cutting AD at D. Then if =r-g; is the rate oi 
 
 expansion — = -r-^z. 
 Pl AB 
 
 To suit this diagram for actual use, A B should 
 
 be taken of such a length as is convenient for scale 
 
 Fig. 53. measurement, say 4 inches ; A D should be divided 
 
 into ten parts and subdivided into quarters ; through 
 
 the divisions faint lines should be drawn parallel to A B. Scales should be 
 
 constructed 4 inches long, suitable for the usual pressures coming under 
 
 consideration. 
 
 The object of columns 3, 4, in Table xvii., is that mean pressure and 
 initial pressure may be easily determined from terminal pressures, when 
 the rate of expansion is known ; or when initial and mean pressures are 
 known, the rate of expansion may be found. Column 5 is given to show 
 the relation between maximum and mean pressures at the various rates of 
 expansion. 
 
 Adiabatic Expansion. — When steam is dried by slight superheating, so 
 that it is surcharged with heat, and is capable of very considerable expan- 
 sion without liquefaction taking place, it expands according to the law of 
 perfect gases, and then 
 
 ♦Rankine, Rules and Tables p. 291. 
 
STFAM TURBINES. 113 
 
 r " may be found by extracting the square root of — four times. 
 
 r-*-VVVVF. 
 
 Column 6 gives the value of — as calculated from the above formula 
 
 T\ . 
 It will be seen that, except at very high rates of expansion, there is no very 
 great difference between the ratio as calculated by this method and by the 
 method for moderately moist steam. 
 
 Clearance. — In practice the mean pressure in the cylinder is very materi- 
 ally affected by what is called clearance. However accurately the engine is 
 constructed, there is always at the commencement of the stroke a space 
 between the piston and cut-off valve, made up of the part of the cylinder 
 between the piston and the cover or cylinder end, and the passage between 
 valve face and cylinder ; this is called the clearance. Supposing this space is 
 equal to one-tenth of the capacity of the cylinder, and the cut-off is at two- 
 tenths the stroke, the effective cut-off is not two-tenths, but something more, 
 due to the fact that the expansion of a volume of steam equaling three-tenths 
 the capacity of cylinder is being effected, instead of that of a volume of two- 
 tenths. This practically amounts to making the cylinder 10 per cent, longer, 
 and cutting off at three-tenths the stroke without clearance. It is, there- 
 fore, customary to speak of the clearance as equal to a certain fraction 
 of the stroke. This, however, must be distinguished from the lineal clearance 
 or distance of the piston from the cylinder ends while at the extreme 
 limits of its stroke. This should be expressed always as a definite length, 
 and not as a fraction of the stroke ; it may be, and often is, called stroke 
 clearance, while the space is volume clearance. 
 
 To allow for the effect which the clearance will have when steam expands 
 in a cylinder, let r be the nominal rate of expansion as before, and r 1 be the 
 actual rate allowing for clearance, c the clearance as a fraction of the cylinder 
 capacity. Then 
 
 1 
 
 1 r J l + • /AN 
 
 - = , — and r, = r ■ ... (A) 
 
 7-,1-f-c * 1+cr y 
 
 - + c being the volume of steam at cut-off between the piston and the cut-off 
 
 valve, and which expands to the volume 1 + c at the end of the stroke. If 
 there is no cushioning of the steam before admission, then the whole of the 
 
 space - + c must be filled at each stroke with fresh steam. 
 
 Then the real mean absolute pressure will be 
 
 V'm ~ C iP\ ~ Pm) • - ' • ( B ) 
 
 p' is the mean pressure obtained by means of Table xvii., the actual rate of 
 
 8 
 
114 
 
 MANUAL OF MARINE ENGINEERING. 
 
 expansion being taken, and p 1 is the absolute initial pressure. If, however, 
 there is sufficient cushioning to fill the clearance space with steam at the 
 initial pressure, the volume of steam used at each stroke will be only that 
 
 swept by the piston at cut-off and equal to — . 
 
 Compression or Cushioning. — There will be an increase of back pressure 
 caused by this cushioning, and its effect on the mean pressure is as follows : — 
 
 Let p' m be the mean absolute pressure due to the effective cut-off — ; p the 
 
 absolute back pressure ; c the clearance ; and p l the absolute initial pressure 
 as before. 
 
 Fig. 54. 
 
 Fig. 54 is the indicator diagram of an engine working under these condi- 
 tions. A B is the stroke, A C the clearance, E F the nominal cut-off. and 
 
 C B 
 D F the effective. The actual rate of expansion is therefore -^-= C D 
 
 D F 
 
 represents the initial pressure and H K the back pressure. Cushioning com- 
 mences at K with pressure p , and at A the pressure is p v 
 
 The figure C D E H K is the diagram due to the expansion of steam of a 
 volume equalling c at a pressure p, to a pressure p M so that the rate of its 
 CK 
 
 expansion is 
 
 DE. 
 
 Now p x x D E = p x C K, and, therefore, 
 ft CK 
 
 Po DE" 
 Since p x and p are both known the rate of expansion is known, and by 
 referring to Table xvii.the mean pressure p" m due to this rate of expansion is 
 found. 
 
 Ihe area H E F G = area CDFGB - (area CDEHK + area H K B) 
 CDFGB = p' m x (AB + AC) = p' m {\ - c). 
 
 HKB =» (AB-AK) = Po (l-( c - Pl -4 
 
 \ Po ' 
 
 „ CDEHK = P \(CK) =p'J&c). 
 
 
 „ H E F G = p m x 1, or the effective mean pressure. 
 
HJ * 
 
 STEAM TURBINES. 115 
 
 Therefore, the effective mean pressure 
 
 -/.<i*«-{i-.(*-i)}*-*.(£-.) 
 
 = P« X « (P'n. ~ Po) + C Pl (l - P y\ 
 
 = fa'*. " Po) (1 + C) + Oft (l - £=). - - - (C) 
 
 General Effect of Clearance and Cushioning. — Let p', the absolute initial 
 pressure, be represented in Fig. 54 by C D, p the back pressure by B L, A B 
 the length of stroke, A C the clearance c, A K the compression a;,EF the 
 nominal cut-off, r the nominal rate of expansion, r x the real rate of expan- 
 sion, &c., (fee, as before. p' m the mean pressure due to an initial pressure ;/, 
 and a rate of expansion r r ; p m the real mean pressure of N E F G L H. As 
 before 
 
 1 + c 1 + c /1X 
 
 r, = = r- . (1) 
 
 1 1 1 + cr 
 
 - + c 
 r 
 
 Since the steam at point K is shut up in a space x + c, and is compressed 
 
 OS "4" C 
 
 into a space c, the rate of compression is ; and the pressure after com- 
 
 OC "I - C 
 
 pression at N is p e = p , and represented by ANorCM; let p c m be 
 
 c 
 
 the mean pressure of the figure MNHKC, which is that due to a pressure 
 
 x + c x + c 
 ■ p , and a rate of expansion -. 
 
 The area NEFGLH = CDFGB - DENM - MNHKC - KHLB. 
 
 NEFGLH = Pm x 1. 
 CDFGB = p' m (1 + c). 
 
 DENM = (p' - Po ^-p) c = (p'~ Po ) c - p x. 
 
 MNHKC = p' m (x + c). 
 KHLB = p (1 - x). 
 
 Therefore, 
 
 P m = P' m (1 + c) - {(p' - p ) c - p x} - p° m (x + c) - p (l-x) 
 = p' n (1 + c) - p'c - p (1 - 2x - c) - pi (x + c). - (2) 
 
 Example I. — To find the effective mean pressure in a cylinder having a 
 clearance space equal to one-seventh its capacity, the initial pressure 80 lbs. 
 absolute, the back pressure 10 lbs. absolute, and the nominal cut-off -g- the 
 stroke : 
 
 CD If no compression 
 
 r = 5 ^-±-i = 5 x » = 3-33. 
 1+5 ia 
 
 By reference to Table xvii.— - = 0-6615 for a rate of expansion = 333 
 
H6 MANUAL OF MARINE ENGINEERING. 
 
 Then 
 
 p' m = 80 x 0-6615 = 52-92 lbs., 
 and 
 
 The effective mean pressure = 52-92 - 10 = 42-92 lb* 
 
 (2) If full compression to 80 lbs. : Here rate of compression 
 
 r ■ «&•« 
 
 ° Po 
 Therefore, 
 
 Pi r p e m 
 
 80 8 
 
 = 3-08 (Table xvii.) 
 Po Pi 
 p' m = 52-92 lbs. as before. 
 
 Then the effective mean pressure by formula (C), p. 115, 
 
 = (52-92 - 10) (1 + |) + so (l - 3-08). 
 
 = f x 42-92 - ^j^ = 25-28 lbs. 
 
 If there was no clearance, the effective mean pressure would be 41-74 - 10, 
 or 31-74 lbs. 
 
 The s'eam used in the case (2) is the same as if there had been no 
 clearance, and as the effective mean pressure was only 25-28 lbs., there is a 
 loss due to clearance of 6-46 lbs., or 20 per cent. In case (1) the quantity of 
 steam used is ^| the volume of the cylinder per stroke, or one-seventh of the 
 volume in excess of the quantity with no clearance, so that with this increase 
 of steam, if there was no clearance and the rate of expansion 5, there should 
 be an increase in the work done, and that increased work will be to the work 
 done by the smaller quantity of steam as 12 is to 7. 
 
 The equivalent mean effective pressure is then X T 2 of 31-74, or 54-41, as- 
 against 42*92 lbs. which was obtained, showing a loss of 11-49 lbs., or 21 
 per cent. 
 
 The example given is a very extreme case, and such as would be rarelv 
 found in practice. The effect of clearance in the high-pressure cylinder of a 
 compound engine may be seen in the following : — 
 
 Example II. — The nominal rate of expansion is 2, the initial absolute 
 pressure 90 lbs., and the absolute back pressure 22\ lbs., the clearance being, 
 one-ninth the capacity of the cylinder. 
 
 (1) No compression : 
 
 1 + i 
 r = 2^4 = 1-82. 
 
 1 + f 
 
 By reference to Table ix., — = -8786 for a rate of expansion of 1-82. 
 
 Then ,) m = 90 x 0-876 = 79 lbs. 
 
 Mean effective pressure = 79 - 22-5, or 56-5 lbs. 
 
 The equivalent mean pressure when \ + \ or \\ of the volume of the 
 cylinder of steam is used will be y- of 5368 lbs., or 65-61 lbs., showing a 
 loss by clearance of 13-88 per cent. 
 
 (2) If full compression to 90 lbs. : 
 
 Here - 1 = 4, which is the rate of compression ; so that 
 
 ?JL«^J?= 2-39. • - - Table xvii. 
 P Pi 
 
MEAN PRESSURE IN A COMPOUND ENGINE. 117 
 
 The effective mean pressure by formula (C), p. 115. 
 
 = (79 - 22-5) (l + ») + ^ (l _ 2-39) 
 = 02-77 — 13-9 = 48-87 lbs. 
 
 Thus showing a loss of 5*81 lbs., or 10*8 per cent. only. The loss from the 
 clearance in a compound is not so serious as in the expansive engine, and 
 in the H.P. and M.P. cylinders of a triple or quadruple engine is of no conse- 
 quence, as the steam in the former (which has passed from the high-pressure 
 cylinder without giving out its full work), will do more work in the medium- 
 pressure and low-pressure cylinders ; whereas, with the expansive engine, the 
 ■exhaust steam passes direct to the condenser at a higher pressure than if there 
 is no clearance. Further, since the cut-off in an expansive engine is much 
 earlier than in a compound, and the clearance from practical considerations is 
 very much the same, the ratio of clearance to volume at cut-off will be much 
 higher in the former than in the latter. Considerable loss is, however, ex- 
 perienced in compound, triple, and quadruple engines if the clearance in the 
 low-pressure cylinder is large, therefore in that it should be as small as possible. 
 
 The beneficial effect of cushioning is seen in both the preceding examples, 
 but its value is greater still when the cut-off in the high-pressure cylinder is 
 somewhat earlier, as may be seen by the following : — 
 
 Example III. — The cut-off in the cylinder of example (2) is altered to 
 ^ the stroke, so that the nominal rate of expansion is 4. 
 
 (1) No compression : 
 
 14-1 
 r — \ ~ 9 — 3 
 
 Then p m = 62*1 lbs. 
 
 and the effective mean pressure = 62T — 22*5 = 29'6 lbs. 
 
 The equivalent mean pressure due to the amount of steam used is now 
 " of 31 '185, or 45 lbs., thus showing a loss of 12 per cent. 
 
 (2) If full compression to 90 lbs. 
 
 The effective mean pressure by formula (C) 
 
 = (62T - 22-5) *■£- + ^ (1 - 2-39) 
 = 44 - 13-9 = 30-1 lbs. 
 
 Thus showing a loss of T085 lbs., or 3*4 per cent. only. 
 
 The economy effected by working with a considerable amount of cushioning 
 is. therefore, very appreciable, and experience has proved the correctness of this. 
 
 In actual practice, however, it only happens that so much cushioning can 
 be effected as to fill the clearance space with steam of pressure equal to that 
 entering in the H.P. cylinder of a triple or quadruple engine which exhausts 
 steam of high pressure ; but still even what is conveniently obtained materially 
 adds to the economic working of the engine. It must not, however, be over 
 looked that the effective mean pressure is considerably reduced by cushioning. 
 
 Mean Pressure in a Compound Engine. — If the effective mean pressure in 
 the high-pressure cylinder of a compound engine be divided by the ratio of 
 capacity of low-pressure to that of the high-pressure cylinder, the quotient 
 represents the mean pressure necessary to do the same work in the low- 
 pressure cylinder as is effected in the high-pressure cylinder. If this be 
 added to the effective mean pressure in the low-pressure cylinder, the sum 
 will be the mean pressure necessary to obtain from the low-pressure cylinder 
 
118 MANUAL OF MARINE ENGINEERING. 
 
 alone, the whole work done by both cylinders, and may be called the equiva- 
 lent mean pressure. If there be no loss of mean pressure, owing to drop in 
 the receiver, or other cause, this equivalent mean pressure will be the same 
 as the effective mean pressure obtained by the steam expanding in one 
 cylinder at the same rate as the total expansion effected in both cylinders of 
 the compound engine. In the two-cylinder receiver form of compound 
 engine, there is sometimes a considerable fall in pressure from the release 
 point to the exhaust, owing to the low pressure maintained in the receiver, 
 or to the late cut-off in the H.P. cylinder. 
 
 il) Two-cylinder receiver compound engine. 
 .<et p i be the initial pressure, p Q the back pressure in the low-pressure 
 cylinder, p r the pressure in the receiver and back pressure in the high- 
 pressure cylinder ; R the ratio of cylinder capacities; r the total rate of 
 expansion, r : the rate of expansion in the high-pressure cylinder, and r„ that 
 in the low-pressure cylinder ; p' m the mean pressure due to an initial 
 pressure, p 1 , and a rate of expansion, r x : p" m , the mean pressure due to an 
 initial pressure p r , and a rate of expansion r„. P m is the mean pressure due 
 to a rate of expansion r, and an initial pressure p l : 
 
 The effective mean pressure in the high-pressure cylinder is then 
 (p'm - Pr) \ and that in the low-pressure cylinder is (p" m - p ). 
 Also _ 1 + hyp. log. r 
 
 *- ~ Pi ~ ' 
 
 Pm=Pl 
 
 Pm = P, 
 
 1 + hyp, log. r T 
 » 
 
 1 + hyp. log. r.. 
 
 r 2 
 
 Since the work performed in the engine is supposed to be equally divided 
 between the two cylinders, 
 
 Pm ~ Pr= R(j>"m ~ Pdh " " " M 
 
 But if there be no loss due to " drop," and the mean pressure in the high- 
 pressure be referred to the low-pressure cylinder, then 
 
 ^i* + (/. - Po) = P- " Po- 
 
 By suostituting the value of (p' m - p r ).ot (1) in the above 
 
 P"m - PO - <*. ~ P0)h \ 
 
 and R } - - (2) 
 
 P'm ~ Pr " ( P m - Po)-2 I 
 
 Let x be the efficiency of the system, so that (1 - x) is the proportion of 
 lo.^s due to drop. Then 
 
 P'm - Po = x ( p ™ - Po)h | 
 inc R > • - - (3) 
 
 P'm -Pr = x(?m " P ) 3 ) 
 
 To find the actual mean pressures when there is loss due to " drop," the 
 value of x must be determined ; this may be done by substituting the value 
 of p' m and p" ms found by the preceding formulae ; but an approximate value 
 may be found by determining the value of p r in equation (3) ; from the 
 "j,lue thus found calculate p" m , and refer the mean pressures of both 
 
F m - Po = 12-36 + ~ 
 
 MEAN PRESSURE IN A COMPOUND ENGINE. 119 
 
 cylinders to the low-pressure cylinder. If (P' m - p Q ) be the equivalent mean 
 pressure thus found, then, approximately, 
 
 Pm ~ Pq . 
 
 P m " Po " W 
 
 Example. — To find the mean pressure in a compound engine using steam 
 of 90 lbs. absolute pressure, the total rate of expansion being 7, the ratio of 
 the cylinder capacities 3-5, and the back pressure 4 lbs. 
 
 P m = 90 x 0-421 = 37-89 lbs. - - Table xvii. 
 
 r = 7 + 3-5 = 2. 
 Then 
 
 p' m - 90 x 0-8465 - 76-18 lbs. - - Table xvii. 
 
 Pr = 76-18 - ^ (37-89 - 4) = 16-88 lbs. 
 
 _p r r_ 16-88 x 7 _ 
 r * ~ p, ~ 90 " l 313 ' 
 
 jT m = 16-88 1 + hyP- log. 1-313 = ^ 
 
 1 O I o 
 
 That is, the effective mean pressure in high-pressure cylinder is 
 76-18 - 16-88, or 59-3 lbs., and that in low-pressure cylinder is 16-36 - 4, 
 or 12*36. Referred to low-pressure cylinder alone, 
 
 59-3 
 7 5~ 
 
 P m _ p Q = 37-89 - 4 = 33-89 lbs. 
 Therefore, 
 
 --ana- °' 865 - 
 
 Then, actual effective mean pressure I _ . 865 (37 . s9 _ „, _ u , 65 lbs 
 in low-pressure cylinder J x '•* 
 
 And the actual pressure in receiver) _ 7fi , fi ?jL/9Q-3\ 24-88 lb* 
 is then ....-( 2 ^ 
 
 (2) 2'^e three- cylinder receiver compound engine, having two low-pressure 
 cylinders. Ratio of each low-pressure cylinder to the high-pressure 
 . R 
 
 In this case only one-third of the work is done in each cylinder. Then 
 
 R 
 
 Pm ~ Pr = -S (Pm ~ P )> " (1/ 
 
 and as 
 
 " " R ^ + {j> " m ~ Po) = Pm " * 
 Then 
 
 P"m - 7>0 = I ( P ™ ' Po) ) 
 
 and /9\ 
 
 V m " Vr = -o (Pm " 7Po)J 
 
120 MANUAL OF MARINE ENGINEERING. 
 
 Also the actual values 
 
 P"~ ~ Po = |( p «- Po) x \ 
 
 R >--•(») 
 
 p' m - p r =j(P n - p )x\ 
 
 Example. — To find the mean pressures in a three-cylinder compound 
 engine (having two low-pressure cylinders), using steam of 90 lbs. absolute 
 pressure, the total rate of expansion being seven, the ratio of the combined 
 capacity of the low-pressure cylinder to that of the high-pressure being 3-5, 
 and the back pressure 4 lbs. 
 
 P m = 90 x 0-421 = 37-89 lbs., 
 r, = ^ = 7 -*- 3-5 = 2, 
 
 p' m = 90 x 0-8465 = 76-18 lbs., 
 
 p r = 76-18 - ~ (37-89 - 4) - 36-64 lbs., 
 
 ]> T r 36-64 x 7 
 
 r 2 =^=^7^ =2-85, 
 
 f n= 36-64 1 + teft 2 ' 85 = 26-4 
 
 Then the mean effective pressure in the high I -,. , D o^» r> , nn * . ,, 
 
 i- j f = 7o*lo - dbo4 = 39*54 bs. 
 
 pressure cylinder J 
 
 Then the mean effective pressure in the low ( a „ . n ., . .. 
 
 1- i > = *o*4 - 4 = 22-4 lbs. 
 
 pressure cylinder J 
 
 Then 
 
 P'*, - Po = 22-4 + i*||* = 32-7 lbs. 
 
 Then 
 
 x = 32-7 -*- 33-89 = 0965. 
 
 Then actual effective mean ) or 
 
 pressure in high-pressure > = -5- (37*89 - 4) x 0*965 = 38-15 lbs. 
 cylinder- - - - ) 
 
 Then actual effective mean ] 9 
 
 pressure in each low- V = ^(37-89 - 4) x 0-965 = 21-8 lbs. 
 pressure cylinder - - j 
 
 (3) The three-cylinder compound continuous expansion engine, having one 
 high-pressure, one low-pressure, and one medium-pressure cylinder, generally 
 called triple-compound. 
 
 K is the ratio of low-pressure to high-pressure cylinder ; R : the ratio of 
 low-pressure to mean-pressure cylinder; p the initial pressure, &c., as before. 
 
 p' m the mean pressure due to expansion, r v and pressure p', 
 
 P m it ■>■> >» **2> >> P > 
 
 P m J> » |] P3, ,, P . 
 
 p" is the pressure in the receiver between high-pressure and mean- 
 pressure cylinders, p"' that in the receiver between mean-pressure and 
 low-pressure cylinders. 
 
MEAN PRESSURE IN A COMPOUND ENGINE. 
 
 121 
 
 Then effective mean pressure in high-pressure cylinder = p' m - p", 
 „ „ mean-pressure „ = p" m - p", 
 
 low-pressure „ = p'" m - p 
 
 But 
 
 Then, if there is no loss due to drop, 
 
 p' m ~ P = R (P'" m - P°) and p" m - p" = R x (p'" m - p»). 
 
 (i) 
 
 Therefore 
 
 Ri 
 
 R 
 
 P - p° 
 P' m -P°=^^- 
 
 p m -p"'=^(P n -p°) 
 
 Pm-p-^Pn-p*) 
 
 r-) 
 
 This is true when there is no loss from " drop," but, as in practice there 
 is generally some loss from this cause, an approximation must be found in a 
 similar way to that for the two-cylinder compound engine. 
 
 The cut-off in the high-pressure cylinder will be, as before, 
 
 1 R 
 
 The cut-off in mean-pressure cylinder in order to maintain a pressure, />", 
 
 in the receiver between it and the high-pressure cylinder can be found in the 
 
 same way as before. Since R is the ratio of low-pressure to high-pressure 
 
 . R . 
 
 cylinder, and Rj that of low-pressure to mean-pressure cylinder, will be 
 
 the ratio of the mean-pressure to high-pressure cylinder, then 
 
 R 
 
 and 
 
 p tx p 
 
 r* X R^ ~ K 
 
 ± = h«K. 
 
 R 
 
 1 
 
 X — • 
 
 p 
 
 1 \ p 
 
 Substituting the value of — . Then, — = Rj -V- ■ 
 
 r \ r 2 P r 
 
 The cut-off in low-pressure cylinder to maintain a pressure, p"\ in the 
 receiver between it and the mean-pressure cylinder, 
 
 1 1 
 
 Ri p" 
 
 l 
 
 x — • 
 
 1 1 p' 
 
 Substituting the value of — . Then, — = -^-. 
 
 r 2 r 3 p r 
 
 But since the terminal pressure in the low-pressure cylinder will be 
 that due to an initial pressure, p, and a rate of expansion, r. Then 
 
 E. 
 r 
 
 P 
 
 or — = 
 
 p r 
 
122 MANUAL OF MARINE ENGINEERING. 
 
 Therefore. 
 
 R 
 
 Cut-off in high-pressure cylinder = — 
 
 T 
 
 „ mean-pressure „ = R x ^~ \ - - (3) 
 
 p r 
 
 „ low-pressure „ = -4tt— 
 
 To avoid any lengthy or elaborate calculations, a result sufficiently 
 accurate for practical purposes may be obtained by assuming a value 
 for x, and using it only in the first step of the calculation. This value 
 will vary from 1*0 to 0-9 in well-proportioned engines of this class, when 
 the steam pressure is not less than 120 lbs. absolute, and the rate of 
 expansion not less than 10 times. 
 
 Example.*— To find the mean pressures in the three-cylinder continuous 
 expansion engine, using steam of 120 lbs. absolute pressure, and expanding 
 12 times. The ratio of low-pressure to high-pressure cylinder being 6, and 
 of low-pressure to mean-pressure cylinder, 2 "5 ; the back pressure in low- 
 pressure cylinder being 4 lbs. : 
 
 Assume x = 0*9. 
 
 P m = 120 x 0-2904 = 34-85 lbs. ) T „ ,. 
 p' m = 120 x 0-8465 = 101-58 lbs. f laDle 
 p' m - p " = • (34-85 - 4) x 0-9 = 55-53 lbs. 
 
 Therefore, 
 Now, 
 
 Then, 
 
 p" = 101-58 - 55-53 = 46-05 lbs. 
 
 - = 2 " 5 x ^nl 20 to = ' 5U > orr 2= 1-838. 
 r 2 46-05 x 12 
 
 „ 1 + hyp. log. 1-838 ' , 
 
 v . - v rsss = 38 lbs - 
 
 If the work performed in the second cylinder is to equal that done in the 
 first, then 
 
 p" m - p'" = ^!x 55-53 = -^ x 55-53 = 2314 lbs. 
 
 Then, 
 
 p" = 38 - 23-14 = 14-86 lbs. 
 
 7 3 = 14-86 2Q x 12 i or r,» 1-486. 
 
 P'" m = p" L^Z2_^I^ = 13 . 96 lb, * 
 
 p'" m - p = 13-96 - 4 = 9-96 lbs. 
 
 Therefore, the mean pressures are 55-53 lbs., 23-14 lbs., and 9-96 lba. 
 Referred to the low-pressure cylinder, 
 
 F m - Po = 9-96 + *£g + 55 ^ = 28-471, 
 P m - p Q = 30-85. 
 
WIREDRAWING OF STEAM. 12$ 
 
 Therefore, 
 
 x _ 28H1 _ . 923 
 " " 30-85 ° 9i6m 
 
 So that if the work is exactly equally divided between the cylinders, then — 
 Mean pressure in low-pressure cylinder 
 
 = ^-^> 0923 = ^ x 0-923 = 949 lbs. 
 
 Mean pressure in mean pressure cylinder 
 
 = 5i (P m - p ) 0-923 = 2 3 5 x 30-85 x 0923 = 23 72 lbs. 
 
 Mean pressure in high-pressure cylinder 
 
 R 
 
 = -3 (P» - i> ) 0-923 = 4 x 30 " 85 * 0923 = 5694 lbs. 
 
 Actual Mean Pressure in Practice. — In the preceding pages, the mean 
 pressure spoken of is only such as would be obtained from a perfect engine 
 in which steam is dry and expanding at a constant temperature, and as such 
 is what may be called the theoretical mean pressure. In an actual engine, 
 however carefully designed, manufactured, and worked, there will be certain 
 causes of loss of pressure, so that the actual indicator-diagram will show a 
 mean pressure considerably less than that due to the initial pressure and 
 the rate of expansion, allowing that during expansion work has been done. 
 
 The following are the principal causes of loss of pressure in the cylinder 
 of a marine engine : — 
 
 (1) Frictional Resistance in the Stop-valves on the Boiler and Engine, 
 and in the Pipes connecting these. — If the initial pressure is taken as that 
 in the valve case, of course this particular loss does not affect the indicator- 
 diagram at all. If the stop-valves are opened to the extent of one-quarter 
 of their diameter, and the steam-pipe is fairly straight and short, and of 
 sufficient diameter, so that the flow of steam at any point does not exceed a 
 velocity of 8,000 feet per minute : the loss of pressure at the valve-case will 
 be very slight, and not exceed 2h per cent. If the capacity of the valve-case 
 is nearly equal to that portion of the cylinder filled at the cut-off point, the 
 loss will be still less, as the case then acts as a reservoir in which steam is 
 stored between the cut-off and admission periods. 
 
 (2) Friction or Wire-drawing of the Steam during admission and cut-off.— 
 This is one principal cause of loss of pressure in most marine engines, and is 
 generally due to defective motion of the valve-gear, combined with small 
 steam ports and passages. If the opening to steam during admission is 
 small at the most, and the valve closes slowly, large passages and ports are 
 of no avail ; and, on the other hand, if the passages are too contracted, there 
 will be considerable loss of pressure in the cylinder, however efficient the 
 valve-gearing may be. But the slow and limited motion of the ordinary 
 slide-valve itself is the most serious obstacle to the obtaining of good diagrams. 
 The slow opening of the valve causes no loss, as the piston speed is low at 
 that period. A perfect valve should open wide enough to allow the steam 
 to pass at a velocity of 8,000 feet per minute, and remain open until cut-off, 
 which should take place quickly ; the valve should remain closed until very 
 nearly the end of the stroke, when it should open quickly and wide to exhaust ; 
 
124 MANUAL OF MARINE ENGINEERING. 
 
 the slow closing to exhaust, and sjow opening to lead, are of no consequence, 
 and cause no practical loss. The loss of pressure from these causes with 
 engines having common slide-valves, and the ordinary link-motion, is con- 
 siderable ; especially is this the case when steam is cut-off early in the stroke, 
 by setting the main valves with very little lead, and having only single ports. 
 As has already been stated, however, the steam becomes superheated by the 
 friction ; it is, therefore, a little more efficient during expansion than it 
 would otherwise be. 
 
 When cut-off is effected by means of special valve-gear, or by a separate 
 cut-off valve, the pressure at cut-off is very little below that in the valve-casing, 
 and sometimes very nearly equal to it ; when effected by the ordinary single- 
 ported slide-valve and link motion, the pressure at cut-off is sometimes as much 
 as 15 per cent., and seldom less than 7^ per cent, below that in the valve-case. 
 
 (3) Liquefaction during Expansion, due partly to the cooling action of the 
 walls of the cylinder and the passages, is a frequent source of loss of pressure, 
 and this was especially so in expansive engines working with moist steam 
 in unjacketed cylinders ; it is observable also, though in a smaller degree, 
 in all compound engines working under similar circumstances. 
 
 (4) Exhausting before the Piston has reached the end of its stroke, although 
 conducive to the good working of a fast-running engine, will show a loss of 
 pressure in the indicator-diagram. The loss from this cause is, however, more 
 imaginary than real ; but it must not be forgotten that the I.H.P. will be there- 
 by less, which is important when the I.H.P. is deemed essential in the contract. 
 
 (5) Compression and Back Pressure due to " Lead " also tend to reduce 
 the mean pressure of the diagram when compared with the theoretical mean 
 pressure. But these are both essential to the good working of an engine, 
 and (as has been shown in a previous part of this chapter) compression tends 
 to balance the loss due to clearance. 
 
 (6) Friction in the Ports, Passages, and Pipes, between cylinder and. 
 cylinder and condenser, produces a loss of pressure, and, although not large 
 when the velocity through them does not exceed 6,000 feet per minute, 
 sometimes amounts to 2 or 3 lbs. in badly designed engines. 
 
 (7) Clearance has been shown to serve to increase the mean pressure 
 'beyond that due to the nominal rate of expansion, and therefore cannot be 
 reckoned as a source of loss, unless the equivalent cut-off is taken to obtain 
 the rate of expansion. 
 
 It will be seen, then, that the actual mean pressure expected to be 
 deduced from the indicator-diagram of an engine depends very much on 
 the proportion and arrangement of the cylinders and their valves, etc., and 
 in calculating the expected mean pressure from the theoretical mean pressure, 
 due allowance must be made in each individual case. 
 
 If the theoretical mean pressures be calculated by the methods laid 
 down in this chapter, and the necessary corrections made for clearance and 
 compression, the expected mean pressure may be found by multiplying the 
 results by the factor in the following Table xviii. 
 
 If no correction be made for the effects of clearance and compression, and 
 the engine is in accordance with general modern practice, the clearance and 
 compression being proportionate, then the Theoretical Mean Pressure may 
 be multiplied by 0"96, and the product again multiplied by the proper factor 
 in Table xviii., the result being the expected mean pressure. 
 
CLEARANCE. 
 
 TABLE XVIII. 
 
 125 
 
 Particulars of Enoink. 
 
 (1) 
 
 (2) 
 (3) 
 
 (4) 
 
 (5) 
 (6) 
 
 Expansive engine, special valve-gear, or with a separate cut-off 
 valve, cylinders jacketed, ....... 
 
 Expansive engine having large ports, etc., and good ordinary 
 valves, cylinders jacketed, ....... 
 
 Compound engines, with ordinary slide valves, cylinders jacketed, 
 and good ports, etc., ........ 
 
 Compound engines as in general practice in the merchant service, 
 with early cut-off in both cylinders, without jackets and expan- 
 sion valves, ......... 
 
 Triple- and quadruple-compound engines, with ordinary slide- 
 valves, good ports, unjacketed, moderate piston speed, 
 
 Fast-running engines of the type and design usually fitted in war- 
 ships, and express with fast-running engines, .... 
 
 Factor. 
 
 0-94 
 
 
 0-9 
 
 to 0-92 
 
 0-8 
 
 to 0-85 
 
 0-7 
 
 to 0-8 
 
 0-65 to 0-75 
 
 0-6 
 
 to 0-7 
 
 Example. — To find the expected mean pressure in the cylinders of a 
 marine engine using steam of 60 lbs. absolute pressure, the rate of expansion 
 4, the clearance equal to one-tenth of the cylinder, and the pressure in the 
 condenser 2 lbs., the valve-gearing specially adapted for an early cut-off, and 
 the ports, passages, etc., of ample size ; compression commences at § of 
 the stroke. The cylinders are jacketed. 
 
 The effective rate of expansion is 
 
 . 1+0-1 
 r, = 4 x : — = 3-143. 
 
 1+4x0-1 
 
 P' = 
 
 0-25 + 01 
 01 
 
 x 2 = 7 lbs. 
 
 and the rate of expansion 3-5. Then, 
 
 „ _ 1 + hyp. log. 3-5 . , ,, 
 p o m = 7 x J * s =4-5 lbs. 
 
 = 60 
 
 3-5 
 
 1 + hyp, log. 3-143 
 3-143 
 
 - 41 lbs. 
 
 Expected mean pressure - 41 (1 + 0-1) - 60 x 01 - 2(1 + 0-5 - 0-1) 
 
 - 4-5 (-25 + 0-1) = 35-3 lbs. 
 
 If the effects of clearance and cushioning be neglected, the mean pressure 
 = 60 X 0*5965 — 2, or 33-8 lbs. This is less than the result obtained by 
 the more accurate calculation in this case, because the cushioning is small 
 for so low a back pressure when compared with the clearance. 
 
 The mean pressure in practice will be found now by multiplying 35*3 lbs. 
 by 0-94, and is therefore 33*18 lbs. 
 
 Example. — To find the expected mean pressure in the cylinders of a 
 compound engine using steam of 100 lbs. absolute pressure, the cut-off in 
 both high-pressure and lew-pressure cylinders being at half stroke ; the 
 clearance in both cylinders is equal to one-tenth of their net capacity ; the 
 pressure in the condenser is 2 lbs. ; the cylinders are jacketed, and the ports. 
 etc., of ample size, no expansion valves. Compression commences at f the 
 stroke. Cvlinder ratio 4. 
 
126 MANUAL OF MARINE ENGINEERING. 
 
 Here the effective rate of expansion = 2 ^ - — - = 1-83 
 
 r 1 + 2x0-1 
 
 The theoretical pressure in the receiver = - — — x — = 27-3 lbs. 
 
 l-o2 4 
 
 The expected pressure in receiver =27-3 x 0-85 = 23*2 lbs. 
 
 The steam is compressed in high-pressure cylinder to 
 
 p. = i + tV x 23-2 = 81-2 lbs. 
 
 1 + 1 
 
 The rate of compression = * — ^^ = 3*5. 
 
 its 
 
 The mean pressure due to a rate of expansion 1-83, and an initial pressure 
 
 of 100 lbs. 
 
 , nA l + hvp. log. T83 n _ 
 = 100 ", ., = 87 lbs. 
 
 The mean pressure due to a rate of expansion 3*5, and an initial pressure 
 of 81-2 lbs. 
 
 = 81 . 2 l + hyplog.3-5 = 521bs 
 3-5 
 The theoretical mean pressure in high-pressure cylinder 
 
 =- 87 (1 + 0-1) - 100 x 0-1 - 23-2 (1 - t-5 - 0-1) - 52 (0-25 + 0-1) 
 = 58-22 lbs. 
 
 The expected mean pressure in high-pressure cylinder 
 
 = 58-22 x 0-85 = 49-5 lbs. 
 
 The mean pressure due to a rate of expansion 1-83, and an initial pressure of 
 23-2 lbs. 
 
 = 23-2 L±_^yP' lo S- 1^ 3 = 20-2 lbs. 
 1-83 
 
 The mean pressure due to a rate of expansion 3-5, and an initial pressure of 
 7 lbs. 
 
 = 4-5 lbs. 
 
 Then theoretical mean pressure in low-pressure cylinder 
 
 = 20-2 (1 + 0-1) - 23-2 x 0-1 - 2 (1 - 0-5 - 0-1) - 4-5 (0-25 + 0-1) 
 = 17-5 lbs. 
 
 And the expected mean pressure in low-pressure cylinder 
 
 = 17-5 x 0-85 = 14-87 lbs. 
 
 Example. — To find the expected mean pressure in a compound engine as 
 fitted formerly in merchant steamers ; the cylinders are unjacketed, the boiler 
 pressure 80 lbs. (95 lbs. absolute) ; the cylinder ratio is 3-5, and the cut-off, 
 effected by common slide-valves, is at half-stroke in the high-pressure 
 cylinder, and 0-6 the stroke in low-pressure cylinder. The clearance in both 
 cylinders is one-twelfth the cylinder capacity. Compression takes place in 
 the high-pressure cylinder when the piston is 0*2 of its stroke from the end, 
 and in the low-pressure cylinder at 0-3. 
 
 Efficiency in this case taken at 0-75. 
 
CLEARANCE. 127 
 
 The effective rate of expansion in high-pressure cylinder 
 
 = 2 : V = 1-86. 
 1 + T? 
 The theoretical pressure in the receiver 
 
 = IW * 3*TcF6 = 24 * 3 lbs 
 
 The expected pressure in the receiver 
 
 = 24-3 x 0-75 = 18-23. 
 
 The fate of compression in the high-pressure cylinder 
 _ 0-2 + 0083 
 " 0-0*3 
 
 and the steam is compressed to 18-23 x 3*4, or 62 lbs. The mean pressure 
 due to a rate of expansion of 1-86, and an initial pressure of 95 lbs. 
 
 = 95 * + hyP- frg- 1*6 = 80 lbs , 
 1 "ob 
 
 The mean pressure due to a rate of expansion 3-4, and an initial pressure of 
 
 62 lbs. 
 
 = 62 L±*!h**J+ _ 40 lbs. 
 3*4 
 
 Then theoretical mean pressure in high-pressure cylinder 
 
 = 80(1+tV)-95x t V,- 1 8-23 (1 - 0-4 - T \)- 40(0-2-^) 
 = 58 lbs. 
 
 And the expected mean pressure in high-pressure cylinder 
 = 58 x 0-75 - 43-5 lbs. 
 The back pressure in the condenser is 2 lbs. 
 
 The effective rate of expansion in low-pressure cylinder 
 
 , 1 
 
 1 * + T2 
 
 - JL x ■ l \ = 1-58. 
 
 0-6 J_ l_ 
 
 + 0-6 12 
 
 The rate of compression in low-pressure cylinder 
 
 0-3 + 0-083 
 
 = 0-083 = 4-6 ' 
 
 Steam is compressed in low-pressure cylinder to 4-6 x 2, or 9-2 lbs. 
 
 The mean pressure due to a rate of expansion 1 -58, and an initial pressure of 
 18-23 lbs. 
 
 = 18 . 23 x 1 + hyP- Jog- 1*8 . 16 . 8 lbs. 
 1 "Oo 
 
 The mean pressure due to a rate of expansion 4-6, and an initial pressure 
 
 9-2 lbs. 
 
 = 9-2 x L±hyP^Li'_6 = 5 lbs. 
 4-6 
 
 Then theoretical mean pressure in low-pressure cylinder 
 
 = 16-8 (1 + T V) - 18-23 x T V - 2 (1 - 0-6 - ^) - 5 (0-3 + T V) 
 = 14-13 lbs. 
 
 And the expected mean j/resfure in low-pressure cylinder 
 = 14-13 x 0-75 = 10-6 lbs. 
 
128 MANUAL OF MARINE ENGINEERING. 
 
 A Practical Method of Estimating the Expected Mean Pressure in a com- 
 pound engine may be followed by first calculating the theoretical mean 
 pressure due to the total rate of expansion, subtracting from it the back 
 pressure in condenser, and dividing by 2 for a two-cylinder engine, by 3 for 
 a triple, and by 4 for a quadruple engine. The result is the mean pressure 
 in the low-pressure cylinder, when there is no loss from " drop." Multiply 
 this by the factor in Table xviii., and again multiply the product by 0'8, and 
 the result is the expected mean pressure when the ivork is equally divided between 
 the two cylinders. 
 
 Example. — To find the mean pressures expected in the cylinders of a 
 compound engine, using steam of 90 lbs. absolute pressure, and expanding it 
 six times ; the ports being of ample size, and the cylinders jacketed ; the 
 cut-off in high-pressure cylinder effected by an expansion-valve, and the 
 pressure in the condenser is 3. 
 
 The mean pressure due to a rate of expansion of 6, and initial pressure of 
 
 90 lbs., = 90 X -4653 = 41'8 lbs. 
 
 The effective mean pressure . . . = 41*8 — 3 = 38 - 8 lbs. 
 
 The effective mean pressure in L.P. cylinder = 19*4 lbs. 
 
 The expected mean pressure in L.P. cylinder = 19*4 X 0'9 X 0'8 = 13'9 lbs. 
 
 If the ratio of the cylinder is 3 5, then 
 
 The expected mean pressure in H.P. cylinder = 13*9 X 35 = 48*7 lbs. 
 
 Graphic Method. — It has been assumed in this chapter that steam expands 
 in accordance with Boyle and Marriott's law — viz., pv = constant — on account 
 of its convenient form and easy calculation. As a matter of fact, the results 
 obtained in this way are sufficiently accurate for practical purposes, although, 
 of course, they are not scientifically accurate, for steam does not expand in 
 practice at a constant temperature as is assumed in Boyle's law, and, more- 
 over, when expanding in practice, it is doing work, and therefore giving up 
 some of its heat with a corresponding reduction in pressure. 
 
 The simpler and more easily-worked methods of obtaining the division of 
 power in compound engines are graphic — i.e., by means of geometrical 
 diagrams based on the same physical facts as before. 
 
 Draw a straight line AC (fig. 55) to represent the capacity of the low- 
 pressure cylinder, including its clearance, which latter is represented by A B. 
 
 Draw an ordinate, A K, representing the absolute pressure of the steam. 
 
 AC 
 Take a point, D, on A C, so that -r^- is the total rate of expansion required 
 
 in the system. 
 
 Then draw the expansion curve D Q' C". 
 
 Take a point F' in K D, so that KF' represents the clearance of the 
 high-pressure cylinder. 
 
 Draw F' F parallel to A K. 
 
 Now, K D' represents the capacity of the high-pressure cylinder, including 
 its clearance at the point of cut-off in that cylinder. 
 
 F D 
 
 Then, take a point L in A C, so that ^t" = rate * cu t-° n m tue high- 
 
 r Li 
 
 pressure cylinder, and draw L L' parallel to A K. 
 
GRAPHIC METHOD. 
 
 129 
 
 A L will, therefore, represent the capacity of the high-pressure cylinder, 
 including its clearance A F. 
 
 Now, take point Non AC, so that AN represents the clearance in the 
 medium-pressure cylinder. 
 
 Take a point, V, in A C, and a point, Q, in A C, so that A Q represents 
 
 the volume of the medium-pressure cylinder, including its clearance, A N, 
 
 N V 
 
 = the rate of cut-off in the medium-pressure cylinder. 
 
 and 
 
 NQ 
 
 Draw V V and Q Q' parallel to A K. 
 
 Take a point, S, in AC, so that =-— - the rate of cut-off in low pressure 
 cylinder. 
 
 Draw S S' parallel to A K, and S' R parallel to A C. 
 
 Then P" D' L' L" F" is the theoretical high-pressure diagram. 
 
 N' V Q' Q" N" is the medium-pressure diagram, and B' S'C'CB the low- 
 pressure diagram. 
 
 KF D' 
 
 AFND3L 
 
 o S 
 
 Fig. 55. — Theoretical Diagram for Triple 
 Expansion. 
 
 BL O S C 
 
 Fig. S.'xj.— Diagrams as in Practice, from 
 Theoretical Expansion Diagram. 
 
 In actual practice the indicator diagrams differ from the theoretical ones, 
 for reasons already given.* They may, however, be inscribed within the 
 theoretical diagram of the mean pressure, as shown in fig. 55a, which has been 
 drawn in accordance with the method prescribed above, so that F" D'F" is the 
 high-pressure diagram, VQN the medium, and SC'CB the low. In prac- 
 tice, the area of the actual diagram of the high pressure is 77 to 80 per cent, 
 of the theoretical. The actual medium-pressure diagram is 70 to 73 per cent, 
 of the theoretical, and the low pressure 55 to 60 per cent. 
 
 In fast-running engines the percentage is, of course, less than that 
 obtained from the diagrams of slow-moving engines. 
 
 If it is found by this means that the power is not sufficiently evenly 
 divided, the cuts-off in the medium-pressure and low-pressure cylinders must 
 be modified. For example, if the power in the low-pressure cylinder is too 
 small, the cut-off point, S, must be moved nearer to A, so that the figure. 
 B' S' C C E, is enlarged at the expense of the medium-pressure diagram. 
 N'V'Q'Q-N". 
 
 * For a refinement in method of construction, vide Appendix, 
 
130 
 
 MANUAL OF MARINE ENGINEERING. 
 
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136 
 
 MANUAL OF MARINE ENGINEERING. 
 
 CHAPTER V. 
 
 STEAM USED AFTER EXPANSION — TURBINES. 
 
 The Turbine is essentially a Velocity Machine, inasmuch as the velocity alone 
 of the steam is the active principle, and pressure has no part whatever in 
 producing its movement. There are rotary machines of various kinds, in 
 which pressure does produce motion, they are, however, not turbines. 
 
 Turbines are of Two Elementary Kinds— the one works by means of the jet 
 of steam issuing at high velocity from a generating nozzle acting impulsively on 
 the blades or vanes of a rotor wheel secured on a shaft, as shown in fig. 51, 
 thus causing it to turn round at a high rate of revolution. Such a machine 
 was invented by Branca in 1630, and it is analogous in hydraulics to the 
 
 undershot-mill wheel, and to the more 
 modern and more refined Pelt on wheel ; 
 since it is operated by the impulse of 
 the steam particles on the vanes, it is 
 known as an impulse turbine. The 
 other kind is known as the reaction 
 turbine, inasmuch as its rotor is caused 
 to move in the direction opposite to 
 that which a jet of steam issues tan- 
 gentially from its circumference. Such 
 a machine moves in consequence of the 
 reaction of the flowing steam, and was 
 the well-known steam engine of Hero 
 of Alexandra. 130 B.C. ; it is analogous 
 to Barker's mill in hydraulics. 
 
 The Modern Reaction Turbine has, 
 as a rule, no actual nozzles, but their 
 equivalent is provided by making the blades of a special section and setting 
 them at such an inclination, as shown in fig. 56, from, these channels the 
 steam issues at so fine an angle of inclination that the component of the 
 reaction resolved tangentially is large, and causes it to turn. The guide 
 channels or the equivalent of nozzles, through which the steam is led into 
 the rotor, may be formed in the same way as shown in this figure. 
 
 The Modern Impulse Turbine also may have, and generally has, a series of 
 expanding passages, the equivalent of nozzles formed by the blades in a similar 
 way, or they are formed in a casting and fixed to the circumferential part 
 of the stator, or fixed disc, attached to the turbine casing, as in fig. 57. which 
 is an example of passages made in that way. 
 
 Combination Turbines. — For many reasons the efficiency of the pure 
 impulse and pure reaction turbine is too low for commercial success, so that 
 all modern turbines are so constructed that both the elementary processes 
 of impulse and reaction occur within each pair of stator and rotor sets of 
 blades. Their construction has developed along two characteristic lines, 
 in which the impulse or reaction process predominates respectively. In 
 
 Reaction Turbine. 
 
COMBINATION TURBINES. 
 
 137 
 
 what is now called the impulse turbine expansion of steam and generation 
 of velocity take place solely in the fixed nozzles ; the issuing jet acts directly 
 on the leading half of the rotor blades, giving up part of its energy to them, 
 and being then deflected through an angle of nearly 180° gives up the remainder 
 by reaction. There is no change of pressure within the cell containing the 
 moving blados, nor in the moving passages, except what is due to friction. 
 
 Fig. 57. — Impulse Turbine Compounded for Pressure and Velocity. 
 
 Consequently the nozzles may be isolated or in groups (v. fig. 59). The 
 reaction turbine, as now made, is not unlike the above, but in it there is pro- 
 vided a further drop in pressure, and the corresponding generation of velocity 
 in the moving passages. In this case, since there is a drop in pressure in 
 these moving passages, the fixed passages or nozzles must occupy the whole 
 of the annulus, in order to avoid excessive leakage by short-circuiting. 
 
138 
 
 MANUAL OF MARINE ENGINEERING. 
 
 The Shape of Passages, and consequently that of blade section, depends 
 on the drop of pressure required. If that drop is greater than 7^ to p., = 
 0"58 p v the passage should converge and then diverge ; if p., is greater than 
 - 58 Pi, convergent only. Thus in the example given (fig. 56) of simple 
 turbines, large drop is provided for, and the passages are, therefore, con- 
 vergo-divergent. Marine turbines being compounded by numerous stages, 
 and having to run at low velocity, require convergent passages. The reaction 
 turbine, having for convenience a similar ratio of expansion for each fixed 
 and moving row of a given diameter, will have similar blading for both. 
 
 Compound Turbines of both types may be multiplied in series, limiting 
 the drop in pressure for each stage to a portion of the total " head." This 
 is called compounding for pressure. This compounding of the impulse turbine 
 may be supplemented by another kind, in which the velocity of the nozzle 
 jets is expanded in more than one row of moving passages, the velocity of 
 
 57a. 
 
 Fig. 576. 
 
 the latter being less than in the uncompounded stage. This is called com- 
 pounding for velocity. Fig. 57 shows diagrammatically an impulse turbine 
 compounded for pressure in three stages, each being compounded for velocity 
 once. The velocity changes of the steam are shown in fig. 576, and it may be 
 noted that the only function of the row of fixed blades between the moving 
 rows is to deflect the stream. 
 
 An impulse turbine compounded for pressure may have a series of rotor 
 discs with blades set transversely and interposed between a series of stator 
 discs having guide blades forming a set of nozzle-like channels. If these 
 blades are all of the same height, gradual decrease in pressure through the 
 series is provided for by starting with only a few passages in the first stator 
 disc, the number in the next disc is larger, and a further increase made in the 
 third, and so on, the increase being such at each step that the steam expands 
 in accordance with the natural law. 
 
COMPOUND TURBINES. 139 
 
 Expansion, however, may be provided for in quite another way ; the 
 length of stator and rotor blades may be gradually increased in length, so 
 that variation in area of the anulus is such as to permit of the steam expanding 
 in the same way while it follows its course through the machine. In either 
 case there is at each drop in pressure a corresponding generation of velocity 
 at each stator, and a delivery of kinetic energy to each rotor disc, so that 
 the expansion is continuous step by step from the initial pressure at entry 
 to that at exhaust, and the total work done distributed through the whole 
 of the stator discs. 
 
 The object of this system of compounding is to reduce the mean velocity 
 of flow to a minimum, so that the rate of rotation of the machine may be 
 reduced down to practicable limits. So far as marine work is concerned, 
 those limits must be such that a reasonable efficiency may be realised with 
 the propellers revolved direct by the turbine. 
 
 The impulse turbine is not necessarily compounded for pressure ; it may 
 be compounded for velocity a sufficient number of times to use up that gene- 
 rated in one drop from maximum to minimum pressures. The number of 
 velocity stages, as they are called, depends on the velocity of the blades 
 relative to the initial velocity of the jets. On account of the losses by friction 
 and eddies this type is not commercially efficient ; if this were not so they 
 would displace all other types for large sizes, inasmuch as they would be of 
 less size, lower cost, and freer from the mechanical troubles so common in 
 this instrument. 
 
 The turbine, however, may be treated in just the same way as a compound 
 or triple-compound engine by providing means whereby there are two or 
 three separate stages in the operations of the machine, in each of which 
 there is at commencement a partial drop of pressure with a corresponding 
 generation of velocity, followed by the gradual drop in velocity and the 
 imparting of kinetic energy by the usual steps above described ; at the 
 end of the first stage, and before beginning of the second, there is another 
 drop in pressure due to the passage through the second set of guide and 
 expansive nozzles of larger capacity than the first ; the steam acquires 
 thereby a fresh velocity, and gives up its energy step by step through the 
 second stage ; and so on in a similar way through a third of similar con- 
 struction, until the steam exhausted of practically all its potential energy 
 enters the condenser. 
 
 Fig. 57a shows diagrammatically the fall in steam pressure in a three- 
 stage compound-impulse turbine, and fig. 57& the corresponding generation 
 of velocity and its gradual extinction in the two steps in each stage, as shown 
 in fig. 57. In practice there are usually more than two steps in each stage 
 of a turbine, and on shipboard it happens usually that there is a high-pressure 
 turbine driving its own propeller, and exhausting to one and generally two 
 low-pressure turbines, each with its own line of shafting and propeller. There 
 may be, however, three turbines in series, high, medium, and low pressures, 
 as in a triple-reciprocating engine, each with its own line of shafting and 
 screw. These arrangements are necessary for marine purposes that the 
 revolutions may be as low as possible to admit of screws of such a reasonable 
 diameter as to be serviceable and at the same time efficient. Fig. 58 shows 
 a section of the Liungstrom turbine with its two interlaced rotors, which mo\e 
 in opposite directions and irreversible Fig. 59 shows in section a Curtis L.P. 
 
140 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Fig. 58. — Ljungstioni Turbine (Longitudinal Half-section). 
 
CURTIS TURBINE. 
 
 ,,Ll£ ... ._, 
 
 141 
 
h: 
 
 MANUAL OF MARINE ENGINEERING. 
 
 and astern-going turbine for a warship. Figs. 59a and 59b give the section 
 and elevation of the Zoelly turbine, as used in the German Service. 
 
 The Design of Screw for a Turbine Ship must first of all be determined 
 from the conditions imposed by the ship and her service, and the turbines 
 designed to suit the revolutions af which such screws can 
 be run consistent with good efficiency and safety. 
 
 Fig. 59a.— Zoelly Marine Turbine. 7,000 H.P. at 650 r.p.ni. 
 
 
 Fig. 59ft.— Marine Turbine of 7,000 H.P. (Zoelly.) 
 
 The Efficiency of the Turbine * must, therefore, be determined at varying 
 revolutions near to that rate at which the propellers may be run, and a 
 curve plotted in the way usual to express it. The efficiency of the screw 
 decided on by considering the above-named conditions should be plotted 
 in a similar way and the curve formed, after which the combined efficiency 
 can be calculated and expressed by its curve, and from it the designer can 
 definitely and finally decide the exact speed of revolution at which the 
 machines shall run. 
 
 " It is only with direct-driven turbines that these considerations require to be made (r. Appendix A). 
 
SEA EXPKRIENCES WITH TURBINES. 143 
 
 Power developed by a Turbine cannot be obtained direct from it. as that 
 of a reciprocator is by means of the indicator, so some other means of finding 
 it has to be adopted. When driving a dynamo to generate electrical current 
 it is easy to determine the power exerted by a turbine by means of the 
 electrical instruments used for measuring quantity and intensity of current. 
 From data obtained in this way turbine efficiency has been deduced so that if 
 the weight of steam used can be accurately determined, the power output 
 can be calculated. To-day, however, marine engineers have a better means 
 of gauging the capacity of the instrument driving the propeller in the torsion 
 meter than even the indicator ever was. 
 
 Prof. Rateau's Formula for Steam Consumption is as follows . — 
 
 • • L - , ,, , 16-2 — 2-05 log P 
 ^team consumption in a turbine = 2* 1-3 -| , p . -• 
 
 P is the initial pressure absolute. jj is the terminal pressure absolute. 
 
 The Sea Experiences with Turbines in the cruiser " Chester," as related 
 by Lieut. Yates, U.S. Navy, are interesting as well as instructive. This 
 ship is fitted with Parsons' turbines of 19,000 S.H.P., driving four screws ; 
 she attained a speed of 26*52 knots on trial. The lieutenant begins by 
 emphasising the extreme importance of warming up the turbines thoroughly 
 before attempting to run them, otherwise there is great danger of stripping 
 the blades. This operation takes much longer than with a reciprocator, 
 and should be from 3 to 3i hours when possible ; the turbine can be, of 
 course, warmed quicker, but it is not advisable to do so. 
 
 Water in the turbine causes no apprehension, and quickly disappears. 
 " Whipping " of the rotors occurs in the H.P. turbine in bad weather, causing 
 it to vibrate and groan slightly, and when the astern-going turbine is operating 
 there are tremors. 
 
 Manoeuvring was not difficult, for as many as 85 signals from the bridge 
 were responded to in 50 minutes. It is found that the turbines do not 
 adapt themselves to the new conditions imposed on them consequent on 
 changes of speed made during manoeuvres at sea. 
 
 A drop in boiler pressure from any cause generally is followed by wet 
 steam, and the increase in blade friction due to it ; the loss of speed is thereby 
 aggravated ; and a drop of 1 inch of vacuum caused a loss of speed of half 
 a knot. Great care is necessary with the bearings, for if they get warm 
 water cannot be applied as with the reciprocator ; there is nothing for it 
 but to slow down, and sometimes to stop altogether, for a little heating 
 causes sufficient expansion to make things much worse very quickly. The 
 oil served, therefore, should be most carefully examined and kept in a high 
 state of efficiency, and the oil used should be of the best and fittest for such 
 service. Hot thrust bearings were not uncommon, due to the fine adjustments 
 necessary to them. The wearing down of the bearings also is the cause of 
 much trouble, even though it be as little as x^Vo °f an inch, for 
 then the gland packings leak and make the engine-rooms unbearable if the 
 ventilation is not really good. If the propeller blades are injured, and even 
 slightly deformed, there is a marked effect on the performance ; but no 
 cavitation was observable in this ship. Seeing, however, that the screws 
 were quite small in diameter and well immersed, it is not very astonishing. 
 
 The mechanical efficiency of these turbines was very satisfactory, for 
 
144 
 
 MANUAL OF MARINE ENGINEERING. 
 
 ———PI 
 
 Fig. 00. — Fottinger Torsion Metes 
 
 ••us 
 
 Fig. 60a. — Fottinger Torsion Meter Diagrams (Reciprocator). 
 
PR. FOTTINGERS TORSION METER. 1^5 
 
 when there was a small leak at the main stop valve of one, it kept turning 
 at 100 revolutions per minute, while the vacuum held good ; and with steam 
 shut off and no vacuum another shaft ran at half the revolutions of those 
 then engaged in propelling the ship, due to " drag" on its propeller. 
 
 Torsion Meters are now used on all trials of marine turbines, and by 
 the engineers in charge of the machinery. As the name signifies, they are 
 the measurers of the torsion or twist of a shaft — that is to say, they register 
 the relative angular' movement of two transverse circular planes at a definite 
 distance apart by a line in a plane through the axis, thus noting the points 
 on their circumference and their distance apart when the shaft is subject to 
 torsion. Suppose two circular discs to be keyed on a shaft 10 feet apart ; 
 they are of considerable diameter, so that any small angular movement gives 
 an appreciate circumferential one. If each has a radial line marked on its 
 face so that these lines will be the traces of a plane through the shaft's axis 
 when transmitting no pow r er, then when transmitting power they will be no 
 longer in line one with the other, for the axial plane through the one line 
 will be at an angle with that through the other. Moreover, experiment 
 has shown that the angular movement will be in proportion to the amount 
 of torque and, therefore, is a measure of it. Now, suppose one of the discs 
 to have a light metal cylinder fitted to it, of sufficient length to come close 
 to the other disc without actually touching it, and a line is marked on this 
 cylinder longitudinally to indicate the position of the disc's radial line, any 
 torsion of shaft will be seen at once by its displacement past the marking 
 on the other disc. In a general way this displacement could not be observed 
 when the shaft is revolving, and certainly could not possibly be noted ct 
 the speed of revolutions of turbines. 
 
 Amlser's Torsion Meter, however, is constructed on this principle, and 
 its indications are easily and clearly seen and read by the ingenious method 
 of causing an electric spark to give an instantaneous momentary illumi- 
 nation of the index as it passes the eye, which produces the effect of an 
 apparent stoppage of revolution, and admits of the reading of it quite easily. 
 
 Dr. Fottinger s Torsion Meter is a mechanical apparatus, so arranged 
 that it can register its movements on a sheet of paper, as does the common 
 indicator. In this case (fig. 60) there is a mechanical connection with the two 
 discs and sleeve by means of a system of compound levers, the outer end 
 of the last one having a pencil or tracing pin, which rests on a sheet of paper 
 laid on a fixed cylinder surrounding the shaft. If no power is being trans- 
 mitted a plain line is made on the paper as the shaft revolves ; when power 
 is being transmitted by a reciprocating engine a wavy line some distance 
 from that base line is traced (v. fig. 60a), the waviness being due to the 
 variation in twisting moment during each revolution of such an engine. With 
 a t in bine under load there is no such variation in smooth water, consequently 
 the meter traces another plain line parallel to the first, and its distance from 
 it is the measure of the torque. The accuracy of this instrument depends 
 very much on that of the mechanism, and seeing that what it rests on is in 
 motion all the time the wear on it due to this will not improve it in that 
 respect. Nevertheless, effecting, as it does, its own register, makes it a 
 very useful instrument for observation, more especially is it the case when 
 applied to reciprocators. There are, however, other instruments in general 
 use which have very little or no mechanism, and such as there is 
 
 10 
 
146 MANUAL OF MARINE ENGINEERING. 
 
 permits of no doubt to arise as to the accuracy of the results derived 
 from its use. 
 
 The Hopkinson-Thring Torsion Meter has the great advantages over 
 any apparatus produced hitherto that it requires only a very short length 
 of shaft, and gives a direct reading of the torsion on the shaft. The zero 
 point of the apparatus also can readily be ascertained by "barring" the 
 shaft. The instrument offers also a simple means of determining the friction 
 of the shafts themselves. 
 
 The principle of the apparatus designed by Professor Hopkinson and 
 Mr. Thring is a differential one, and consists in the observation of the twist 
 between two adjacent points on the shaft by means of two beams of light 
 projected on to a scale from a fixed and a movable mirror. The beam pro- 
 jected on the scale by the fixed mirror is taken as the zero point, whilst 
 the beam projected by the movable mirror indicates the amount of torque 
 on the shaft. Both mirrors revolve with the shaft, but even at moderate 
 
 Fig. 61. — Hopkinson and Turing's Torsion Meter Mounted Complete on a Shaft. 
 
 speeds the reflections appear as continuous lines of light across the scale, and 
 there is, therefore, no difficulty in taking readings. 
 
 The torsion meter is shown in fig. 61, mounted complete on a shaft, whilst 
 a diagrammatic arrangement of the complete apparatus is shown in end 
 elevation and plan in fig. 61a. A collar, A, clamped to the shaft of which 
 the torque has to be measured, is provided with a flange projecting at right 
 angles to the shaft and an extension. 
 
 A sleeve B (fig. 61a), provided with a similar flange and extension at one 
 end, is clamped at its further end on to the shaft in such a manner that its 
 flange is close to that on the collar A, whilst its extension overlaps that of 
 the collar A. on which it is supported to keep it concentric. Both the collar 
 and sleeve are quite rigid, and it is, therefore, obvious that when the shaft is 
 twisted by the transmission of power, the flange on the sleeve B will move 
 relatively to that on the collar A, the movement being equal to that between 
 the two parts of the shaft on which these fittings are clamped. This move- 
 ment is made visible by one or more systems of torque mirrors mounted 
 
THE HOPKINSON-THRING TORSION METER. 
 
 147 
 
 between the two flanges, which reflect a beam of light, projected from a 
 lantern, on to a scale divided in a suitable manner on ground glass. 
 
 Each system of torque mirrors consists of a mounting, pivoted top and 
 bottom on one or other of the flanges, in which two mirrors are arranged 
 back to back. This mounting is provided with an arm, the end of which 
 is connected by a flat spring to an adjustable stop on the other flange. Any 
 relative movement of the two flanges will turn the torque mirror, and thereby 
 
 ^-TORQUE MIRROR 
 
 GLASS S&PXZr f^p 
 
 Elevation. 
 
 -zero mirrors 
 
 Fig. (lire.— Hopkinson and Thring's Torsion Meter. 
 
 cause the beam of light to move on the scale, the deflection produced being 
 directly proportional to the torque applied to the shaft. 
 
 With the arrangement described, a reflection will be received from each 
 mirror at every half revolution of the shaft ; but where the torque varies 
 during a revolution (as with reciprocating engines), a second system of mirrors 
 may be arranged at right angles to the first system, so that four readings 
 can be taken during one revolution ; or, if two scales are used, eight readings. 
 
148 MANUAL OF MARINE ENGINEERING. 
 
 Fig. 61a shows how the beam of light reflected by the mirror when in its 
 highest position passes through the upper part of the scale ; while the second 
 reflection will occur when the mirror is in the position occupied by the zero 
 mirror, the beam of light passing through the lower part of the scale. The 
 position of the torque mirror in plan is such that the reflected beam strikes 
 the scale to the right of the zero line, but when the shaft has made a further 
 half revolution, the reflected beam from the other mirror will strike the scale to 
 the left of the zero line. Obviously the deflection on both sides should be equal. 
 
 The fixed mirror is attached to one of the flanges (in fig. 61a to the flange 
 of the sleeve B). This must be adjusted so that the beam of light reflected 
 from it is received at the same point on the scale as those from the movable 
 mirrors when there is no torque on the shaft. To facilitate the erection and 
 adjustment of the apparatus, the box containing the scale and carrying the 
 lamp is fitted with trunnions, so that it can be inclined as required. 
 
 If the position of the apparatus becomes altered relatively to the scale 
 owing to the warming up of the shaft or from other causes, this is indicated 
 immediately to the observer by an alteration in the position of the zero as 
 reflected by the fixed mirror. Hence, the zero can be adjusted by moving 
 the scale so that its zero coincides with the reflection from the fixed mirror. 
 It will be obvious that it is not necessary to move the scale, as the mean of 
 the two readings will be the same. It will readily be understood that a move- 
 ment of the torque mirrors can only occur through a relative, movement of 
 the two flanges, so that vibration of the shaft or of the ship will not influence 
 the readings in any way. 
 
 The constant of the instrument — viz.. the factor which, when multiplied 
 cr divided into the product of the torsion-meter reading and the revolutions — 
 gives the horse-power, may be calculated within 2 or 3 per cent., if the section 
 of shaft within the instrument is uniform. But a direct calibration of the 
 shaft with the instrument in position before the former is put into the ship 
 should be made. This is effected easily by applying a known twisting couple. 
 It is no inconsiderable advantage of this instrument that a direct calibration 
 is established between the torsion-meter deflection and the torque on the shaft. 
 
 In the Bevis-Gibson Torsion Meter, which is largely used in every-day 
 work, light is also used to indicate the angular movement. In this case 
 however, the discs are quite a part, and perforated with small narrow slits 
 in such a way that when the shaft is running without load the light of an 
 electric lamp behind the one disc can be seen through a sighting instrument, 
 called a " finder," behind the other disc — that is, the lamp slit, its disc slit, 
 the other disc slit, and the " finder " slit are all in line every time that part 
 of the discs pass, when the shaft is running without load, so that on looking 
 through the finder the flash of light is seen every time the slit passes the 
 finder ; and since the revolution is rapid, it appears as a continuous illumi- 
 nation. When power is being transmitted, the shaft is twisted ; the slits 
 get out of line, and no light is seen through the " finder." The latter, how- 
 ever, can be moved through an angle by means of a finely adjusted screw, 
 etc., until the light is again picked up when looked at through the " finder " 
 disc's slot at the time the lamp's disc's slot is passing the light slit. The 
 angular displacement of the " finder " corresponds with the angular move- 
 ment of the shaft, and the shaft horse-power can be determined thereby in 
 the usual way. 
 
COLLIE'S TORSION METER. 
 
 U9 
 
 Collie's Torsion Meter (tig. 62) is also a mechanical contrivance of con- 
 siderable ingenuity, by which the angle of twist is caused to be registered 
 by a pointer and dial, not unlike the Bourdon pressure and vacuum gauge 
 used on the L.P. valve box of a compound engine. In this arrangement 
 two counter shafts are connected at the middle by the coarse-threaded end 
 of one entering the threaded sleeve turned by and free to slide on the end 
 of the other. Each shaft is driven independently by a Renald chain geared 
 up so that it runs about three times to one of the main shaft. If the main 
 shaft is not transmitting power, the sleeve simply revolves, and the pointer 
 
 =3= 
 
 =E3= 
 
 =3= 
 
 MAI* SHAFT 
 
 Fig. 62. — Collie's Mechanical Torsion Meter. 
 
 of the gauge is motionless in mean position. As soon as power is trans- 
 mitted the shaft twists and one counter shaft moves in advance of the other, 
 and forcing the sleeve to slide as a consequence of its screwed end moving 
 on the male end of the other ; the pointer is thus caused to turn round so as 
 to indicate the exact angular displacement of the length of shaft between 
 the two driving wheels. Both these two meters, as, indeed, must all 
 mechanical ones, depend on the accuracy of manufacture and the state of 
 repair for the value of their indications ; it is evident that the magnification 
 of the errors will be practically at the same rate as the magnification of the 
 
150 MANUAL OF MARINE ENGINEERING. 
 
 indications. It was natural, therefore, for scientists and inventors to turn 
 to the other means of indications and magnification, which had provided 
 the means for measuring other forces with a delicacy that no mere mechanicaL 
 contrivance can achieve. The mirror magnification of radius and conse- 
 quently of the arc used in telegraphy no doubt suggested itself to others as it 
 did to Prof. Hopkinson and Mr. Thring. 
 
 The Denny- Johnson Meter differs from the others, inasmuch as linability 
 of the points on the discs or otherwise is indicated by sound, instead of light, 
 as transmitted to a telephone receiver. When the two points or projections 
 in this case on the disc are in line, they are so close to two fixed projections, 
 as to make virtually an electrical connection. When torsion takes place 
 the connection is broken — that is, the meeting of the projections do not 
 synchronise — and it is only by displacement of one of these fixed points 
 equal to the relative circumferential movement of the point on the disc that 
 restores synchronising contact with the consequent production of sound in 
 the telephone. The amount of this movement is registered in the ordinary 
 simple way, and from it the horse-power is estimated. 
 
 Shaft Horse-power transmitted by a shaft can be calculated from the 
 torque as follows : — 
 
 T is the twisting moment or torque in inch-pounds. 
 
 R, the revolutions made per minute by the shaft. 
 The work performed) _ 2rxT _ 0-5236 TxR _ TxR 
 
 per revolution j ~~W ° T U *°^ b ' *>•*•*— 33,000 ~~ 63,000 
 
 The torque can be calculated from the angle of twist or torsion by means 
 of the following formulre : — 
 
 a is the arc at a radius r of the angle of torsion 9, - = fi ; I is the length 
 and d the outer and d 1 the inner diameter of shaft. 
 
 „ 10-2 xTxI,. r > 6 a 
 
 P = M.X.cZ *" ( Rankme >- 360 = 2Vr ° C = ft? 
 
 m , . ' x 2r. mt _ 10-2 xTx! 
 
 That is p = -^- = 5W Then 5y - = - M x # > 
 
 /• , fl __ 584 X T ><Jl for solid r \ ft - 584 T x l \f or hollow 
 
 or(i.) U-- UrXd 4 | shafts, U ' M M,(^-(/ x 4 )/ shafts. 
 
 That is T - /B84-xl ' 
 
 M,. is the modulus of stiffness or rigidity of the material, which for steel' 
 generally is 10 to 12 millions. With steel shafts of best make, experiments 
 have shown the value of M,. to be 11,750,000 for solid, and 12,150,000 for 
 hollow. In every-day practice. 11,250,000 is taken for solid steel shafts : — 
 
 Ti t " Xl|4x 11,250,000 X d* 
 
 Then T = m} = 19,264 — • 
 
 Substituting this value of T in the formula for S.H.P. 
 
 a tt -n m eta* x di R ° x d * x R n 1 tt- a 
 
 S.H.P. = 19,264 -j- X 6 3^ = -3^^-T "olid shafts. 
 
 S.H.P. = ^*_T_^^ f or hollow shafts. 
 3-17 X I 
 
ORDINARY STEAM liNOINE INDICATOR 
 
 151 
 
 It may be observed that it is usual to experiment in the workshops with 
 the shafts of every ship to ascertain beforehand what amount of torque is 
 necessary to produce a degree of angular movement in a definite portion of 
 its length ; from such experiments the modulus of rigidity is ascertained. 
 It is then easy to construct a diagram from that particular shaft, by which 
 the power may be read off for any angular movement indicated by the torsion 
 meter, and the revolutions at the time of observation. Fig. 63 is such a 
 diagram as to need no explanation. , • 
 
 It was held at one time that the end pressure on a ship's shaft due to 
 thrust seriously affected the register of torque ; it was supposed until lately 
 to affect it to the extent even of 3 per cent. ; but more recent investigations 
 by Dr. Hopkinson with more sensitive instruments seem to have removed 
 the impression, and that practically end thrust may be disregarded now as 
 a factor in the calculation of shaft horse-power. 
 
 fuii line Curve Si'Ows Cranh effort Curse deduced from indicator Diagrsm. i-etters corresponding position ct KPCrenk 
 Dotted, ., ,. .. „ obtained , Light Tors ion /deter. Numbers , „Si'oCa in Tonic' texrOtlt 
 
 MEAN TwiSTInO MOMiii DlAORAU fOR a RecipROCATino EnOINS. 
 
 h.t>. 
 
 IP 
 
 LP. 
 
 0' 30' 60' 90' Ob' ISO' 180' 210' 240' 770' 30f 330" 3d' 
 
 Fig. 63. — Crank Effort and Torsion Meter Diagram (J. Hamilton Gibson). 
 
 The Ordinary Steam Engine Indicator is often spoken of as a defective 
 instrument, and one not to be relied on to give accurate results. It is quite 
 true that it is not exactly a perfect one, and that the power registered by 
 means of it comes short of the actual amount developed in the cylinders, 
 but, at the same time, while admitting this, it should not be forgotten that 
 it renders another and a very good service to the engineer besides that of 
 giving him the power, and one that is quite as important to him. It shows 
 in a ready and rapid way whether the internal and unseen parts of the engine 
 are in good working order, and efficient for their service. It is to him what 
 the stethoscope is to a doctor of medicine. 
 
15*2 MANUAL OF MA1UNE ENGINEERING. 
 
 CHAPTER VI. 
 
 EFFICIENCY OF MARINE ENGINES. 
 
 The Efficiency of an Engine should be measured by the cost of the useful 
 work it does on service ; formerly the fuel burnt in the boiler from which 
 the steam was supplied was taken as the measure of cost, and so long as the 
 fuel was of standard quality and the efficiency of the boilers constant, it was 
 a fair and ready means of doing so. But coal from the same pit may vary 
 in calorific value even when freshly raised, and undoubtedly does so after 
 exposure ; the method of ascertaining the weight permits of inaccuracy, 
 although perhaps this is only slight ; the state of the atmosphere seriously 
 affects consumption, for it is obvious that cold moisture laden air will require 
 more fuel to raise it to the temperature of the furnace for combustion than 
 is necessary for an equal amount of dry warm air ; the human element like- 
 wise plays an important part in the production of steam, as may be evidenced 
 when the single boiler of a ship is barely large enough for the engine require- 
 ments ; in such a case a good and experienced stoker will keep the steam 
 pressure and supply steady, whilst an indifferent one will use more fuel, 
 and then fail to maintain full pressure. Further, atmospheric conditions 
 seriously affect the quality of the draught, so that while with a good breeze 
 the draught will be sharp and the combustion efficiently effected, on a hot 
 sultry day with no wind it will be bad, and the fires require much forcing. 
 Then, too, the condition of the boiler is not always the same ; the grate 
 bars may be faulty, the tubes dirty, and the inside with more scale at one 
 time than another ; but it is only fair to say that nowadays the scale should 
 never be so very great as to cause serious differences. Taking all these 
 things into account, it is obvious that the more correct method is to debit 
 the engine with the weight of steam supplied to it rather than with the 
 weight of coal. In old days, when there were no auxiliaries, but only the 
 main engines, there was no call to differentiate as there is now the steam 
 supplied from the boilers between that used by the main engine and that by 
 auxiliaries, etc. To-day the amount needed for other purposes than that 
 of driving the main engines is very appreciable, amounting to as much as 
 14'5 per cent, of the total used in the " Lusitania " when running at full 
 speed. There are in passenger ships centrifugal circulating and air and feed 
 pumps constantly going, but they are virtually a part of the main engines ; 
 steering engines, ash hoists, blowing fans, ventilating fans, refrigerators, 
 electric lighting, sanitary pumps, etc., are also working during a considerable 
 portion of the day, and bilge pumps, fire and wash-deck pumps, steam whistle, 
 etc., take a good share also ; besides all these in cold weather steam heating 
 will demand an extra expenditure of fuel, an important item in the North 
 Atlantic. 
 
VERTICAL ENGINES. 153 
 
 The main propelling machinery, therefore, must be debited only with 
 the steam it uses and that used by such of the auxiliary machinery as is 
 necessary to keep the main engines running. 
 
 With such feed and other pumps as are now employed in the engine- 
 rooms of important vessels, the water consumption can be closely measured, 
 as they are capable of acting as meters of the water they pass by simply 
 fitting them with counters to record the number of strokes they make ; the 
 chief engineer can thus quite easily calculate the actual amount of water 
 pumped from the main condenser and from the auxiliary condenser, and 
 record thereby the steam consumption of the machinery as a whole, or of 
 the main engines only. 
 
 To show the cost of the useful work done by marine engines used to be 
 somewhat difficult, and could be calculated only by using assumptions that 
 were always somewhat uncertain. That is, until the torsion meter was used 
 the work transmitted to the propeller was, except in some special exceptions, 
 practically guessed at ; now, however, thanks to these instruments, we can 
 measure what is called the shaft horse-power of both turbines and recipro- 
 cators, and thereby compare the outputs by reducing them to a common 
 denominator ; this being so : — 
 
 Water consumed per S.H.P. (shaft horse-power) is the measure now of 
 the efficiency of a marine engine, qud engine, without complications or 
 explanations. But by this method the general efficiency is shown without 
 any differentiating between the steam and the mechanical efficiency of 
 the system. 
 
 The Mechanical Efficiency, or the relative value of the engine as a piece 
 of mechanism, is measured by comparing the shaft horse-power with the 
 gross power generated, as shown by the indicator diagram ; 
 
 Mechanical efficiency of a marine engine, therefore, is S.H.P. -h I.H.P. 
 
 Mr. Denny found by using a torsion meter that the mechanical efficiency 
 of some quadruple engines made by his firm was as high as 94 per cent. 
 
 The late Mr. Mudd ascertained the power required to move certain triple- 
 compound engines standing in the erecting shop without screw shafting 
 was at working revolutions 45 I.H.P., or 5 - per cent, only of the gross power 
 indicated (900) when working at those same revolutions in a loaded ship. 
 This would show, then, the efficiency to be 95 per cent. ; but it must be noted 
 that in this case no allowance is made for the extra friction on the valves, 
 guides, etc., due to the greater pressure on them when at work at full power. 
 
 The older horizontal jet-condensing engines had comparatively a low 
 efficiency due to the general friction of the very heavy moving parts and 
 the resistance of the two air, two feed, and two bilge pumps worked by them. 
 There is good reason to believe that the mechanical efficiency of some of 
 them was seldom over 75 per cent. Latterly, however, the well-made hori- 
 zontal engines of Penns, Maudslays, etc., when running at the higher revolu- 
 tions, and developing much more power than •formerly, and having surface 
 condensers, had even 80 to 85 per cent, efficiency. 
 
 Vertical Engines with surface condensers and slow running had an effi- 
 ciency from 5 to 8 per cent, better than the horizontals of similar size, and 
 working under similar conditions. To-day naval machinery and that of 
 express steamers have an efficiency from 90 to 94 per cent, at full speed, 
 
154 MANUAL OF MARINE ENGINEERING. 
 
 when with only the air pumps driven by it, and of the triple- and quadruple- 
 reciprocating type. 
 
 The Mechanical Efficiency of a Turbine has been determined approxi- 
 mately by observing the electro-motive power required to revolve it at 
 working speed by means of a motor, and taking it as the mechanical loss 
 when working. Its efficiency is, of course, high, and from such observations 
 made with similar machines in service on land, it may be taken that the 
 marine turbine has a mechanical efficiency of about 95 per cent, in a general 
 way, and that with the largest ones it may have a somewhat higher, probably 
 96 to 97 '5 per cent. 
 
 The Mechanical Efficiency of the Reciprocating Engine is not so high as 
 this, although that of well-designed carefully made and balanced engines 
 of high speed will not be far short of the 95 per cent, when running in good 
 working order and free of all pumps, and well lubricated. Under these cir- 
 cumstances its efficiency is probably 92 to 94 per cent, at least. It is, how- 
 ever, somewhat misleading to express the mechanical losses as a fraction 
 of the Total Indicated Horse-Power, inasmuch as that varies closely with 
 the cube of the revolutions, while the losses vary more nearly as the revolu- 
 tions ; so that, although an engine may show an efficiency of only 80 per 
 cent, at 50 revolutions, it may be as much as 94 at full speed of 75 revolutions 
 
 /50\ 3 
 — that is, the loss is (,=v) X 20 — without any change in the adjustment of 
 
 a single part. 
 
 Mechanical losses really depend largely on the size of the engine. Now, 
 in a general way the Nominal Horse-Power expresses fairly well the size 
 of any engine, and, therefore, all other things being equal, it will be a suffi- 
 ciently accurate assumption that frictional losses are proportional to N.H.P. 
 Practice has demonstrated that about 70 per cent, of these losses will in the 
 ordinary marine engine vary directly with the revolutions within reasonable 
 limits, and further that the remaining 30 per cent, increase at a more rapid 
 rate ; in fact, in modern engines total losses roughly vary as the cube root 
 of the revolutions raised to the fourth power — that is, as R*. 
 
 It is obvious that the mechanical efficiency of any engine will vary from 
 time to time, and under some circumstance the variation may be very con- 
 siderable. It is also well known that at very slow rate of revolution the 
 apparent mechanical losses are uncertain, and always proportionately large. 
 The only losses that can be considered here are those which inevitably occur 
 with any engine when in quite a good state of repair and in really good working 
 order. 
 
 The Efficiency of Marine Engines, when well made, in this good state 
 of repair, and in good working order, should be approximately in accordance 
 with the following rule : — 
 
 «..;. , N.H.P. xR. . »«. 
 
 Junction horse-power = ■ — (// + # vR). 
 
 „«. . LHP. - F.H.P 
 Efficiency = - -yg-p 
 
 Nominal horse-power = D X S -s- K, 
 where D is the diameter of the L.P. piston, and S the stroke, both in inches. 
 
EFFICIENCY OF MARINE ENGINES. 155 
 
 For the two-stage compound engine K = 15 "0. 
 ,, triple-stage ,, K = 126. 
 
 ., quadruple-stage „ K = 10"5. 
 
 R the revolutions per minute. 
 
 For diagonal paddle-wheel engines with air pumps only X == 1*5 ; y — 10. 
 
 „ vertical screw engines, mercantile, with all pumps 
 
 connected, . . . . . . . x = 1*0; y = 8. 
 
 ,, light quick-running screw engines with all pumps 
 
 connected, . . . . . . x = 07 ; y = 7. 
 
 ., naval and express screw engines, centrifugal cir- 
 culating, . . . . . . . x = 0'6 ; y = 7. 
 
 ., naval and express screw engines, with only air pumps x = 0'5 ; y = 6 - 5. 
 
 ,, special naval screw engines, no pumps . . . x = 0'3 ; y = 6'0. 
 
 ,, ,, forced lubrication, no pumps. . . x = O'l ; y = 5*0. 
 
 Example 1. — A " tramp " steamer having an engine with cylinders 22 r 
 37, and 62 inches diameter and 45 inches stroke, which at 85 revolutions- 
 develops 2,300 I.H.P., what is the efficiency ? 
 
 Here N.H.P. = 62 x 45 - 12 6, or 221. 
 
 Frictional H.P. = — ^~- (8 + ^85) = 220. 
 
 Efficiency = (2,300 - 220) -j- 2,300, or 0-004. 
 
 Example 2. — A destroyer has two sets of engines, each having cylinders 19, 
 28 - 5, and 43 inches diameter with 18 inches stroke at 370 revolutions ; the 
 total T.H.P. is 4,200. 
 
 Here N.H.P. each engine = 43 X 18 - 12'6 = 61. 
 
 Friction H.P. = 6 -\^-° (6 + 0'3 ^370) = 184. 
 
 Efficiency = (2.100 - 184) + 2,100 = 0"913. 
 
 Example 3. — A paddle steamer having cylinders 56 and 110 inches diameter 
 and a piston stroke of 72 inches develops 7,150 I.H.P. at 50 revolutions, 
 what is the F.H.P. and efficiency ? 
 
 Here N.H.P. = 110 X 72 - 15, or 530. 
 
 F.H.P. = ^^° (19 + 1-5 V50) = 411. 
 
 Efficiency = 7,1 ^^ U = 0*942. 
 
 Example 4. — A naval ship has two engines, each with cylinders 43 and 
 69 inches, and two L.P. each 77 inches, the stroke being 42 inches. 
 
 Here the two L.P. cylinders are equivalent to one 109 inches diameter. 
 
 At 140 revolutions each engine develops 12,000 I.H.P. 
 „ 126 „ „ 8,000 „ 
 
 ,. 85 ,. „ 2,500 „ 
 
156 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Now, 
 
 N.H.P. 
 
 (<0 
 
 F.H.P. 
 
 
 Efficiency 
 
 (h) 
 
 F.H.P. 
 
 
 Efficiency 
 
 (c) 
 
 F.H.P. 
 
 
 Efficiency 
 
 109 X 42 h- 12-6 = 363 
 
 363 x HO 
 
 1.000 
 
 12,000 - 514 
 12,000 
 
 363 x 126 
 
 (7-0 + 0-6 x t/l40j = 514. 
 = 0-957. 
 
 1,000 
 
 8,000 — 460 
 8,000 
 
 363 x 85 
 
 (7 + 0-6 4/126) = 460. 
 = 0-942. 
 
 1,000 
 
 2,500 - 298 
 2,500 
 
 (7 + 0-6 £/85) = 298. 
 = 0-881. 
 
 Example 5. — A triple-compound engine with cylinders 18J, 28, and 40 
 inches diameter, the stroke 20 inches, and at 250 revolutions the I.H.P. 
 is 1,150. There are no pumps, and lubrication is forced throughout. 
 
 Here 
 
 N.H.P. = 40 x 20 - 12-6 = 63-5. 
 tion HP. = 63 1 ?J 5 ° (5 + 0-1 #250) = 89'5. 
 
 Efficiency = 
 
 1,000 
 1,150-8 
 
 1,150 
 
 , or 0-922. 
 
 Dr. Bauer states that from observations of and experiments with marine 
 engines made in Germany the efficiency of small ones with an indicated 
 horse-power not exceeding 50 is only 0'59, while those of 5,000 I.H.P. and 
 upwards it was as high as 0"91 ; also that the efficiency varied roughly from 
 059 to 0'91 in accordance with their power for other engines. He quotes 
 some examples of engines whose actual efficiency had been determined, and 
 they are as follows : — 
 
 TABLE XXV. — Efficiency of Some German Engines. 
 
 1 
 
 German Engine. 
 
 A 
 
 BCD 
 
 E 
 
 V 
 
 Indicated horse-power, 
 • Efficiency, 
 
 1,630 
 0-885 
 
 1,640 1,940 2,370 
 0-910 0-911 0-920 
 
 1 
 
 2,890 ! 4,500 
 0-911 0-935 
 
 Mr. Denny found by torsion meter trials with a reciprocating engine 
 that the efficiency at full power, 1,550 I.H.P., was 0'92 ; and in another 
 case with a power of 1,950 it was as high as 0'935. 
 
 The Results of Trials made with Triple-compound Engines designed for 
 electric generating are shown in fig. 64, and are very instructive, although 
 
TRIALS WITH TRIPLE-COMPOUND ENGINES. 
 
 157 
 
 not made under quite the same conditions as those obtaining with a marine 
 engine, for in these cases the load was an artificial one (brake), and somewhat 
 arbitrarily varied. 
 
 The following figures and diagrams show the results of some carefully 
 made trials with two triple-compound engines, by Messrs. Belliss & Morcom. 
 Birmingham. They are of their special enclosed type, having three cylinders 
 and three cranks supplied with steam at 150 lbs. pressure, and. exhausting 
 into a surface condenser whose pumps are operated by independent engines. 
 The whole of the pins, bearings, guides, and working parts are lubricated 
 with their special American mineral oil forced through them by a pump 
 constantly at work connected to the engine. It will be seen that the friction 
 per revolution, both with and without load, varies, and has two minimum 
 and two maximum values ; that at full speed the frictional H.P. per revo- 
 lution loaded is really only a little less than that when running free, whereas 
 at "200 revolutions, or half speed, there is the greatest difference between 
 them, that at full load being half that running free ; also that at dead slow 
 speed, say 25 revolutions, the friction per revolution, both light and loaded, 
 is nearly four times that of the loaded engine at 200 revolutions. 
 
 11" — 17" — 24" 
 Engine (No. 1401) ^y> 150 lbs. at Engine Stop-valve. 
 
 Table of Powers, <kc, Under 
 
 Load Con 
 
 ditions. 
 
 
 
 Revolutions. 
 
 166 
 
 300 
 
 350 
 
 400 
 
 I.H.P. 
 
 B.H.P., . . . . - . 
 Difi'erence or friction, H.P., . 
 I.H.P. per revolution, . 
 B.H.P. ,, ,, 
 F.H.P. ,, ,, ... 
 
 B.H.P. , . 
 
 i h p efficienc y> • • . • 
 
 248-2 
 
 240-4 
 7-8 
 1-496 
 1-45 
 •047 
 
 06-87. 
 
 384-15 
 360-5 
 2365 
 1-280 
 1-201 
 •079 
 
 03-9 7. 
 
 408-1 
 380-5 
 27-6 
 1-168 
 1086 
 •079 
 
 93-25 7. 
 
 429-4 
 400 
 29-4 
 1073 
 100 
 •073 
 
 93-15 7. 
 
 
 
 
 
 
 Friction Powers Under No-Load Conditions. 
 
 • 
 
 Revolutions. 
 
 56 
 
 102 
 
 113 
 
 213 
 
 313 
 
 365 
 
 Friction H. P., .... 
 ,, ,, per revolution, 
 
 -t 
 
 4-9 
 
 •0875 
 
 10-7 
 •1049 
 
 7-4 
 •0655 
 
 24-6 
 •1153 
 
 361 
 •1153 
 
 30-9 
 
 ■0847 
 
 The above figures are embodied in the diagram (fig. 64). 
 The no-load cards were taken non-condensing, but the load cards were 
 taken condensing. 
 
 This is probably the reason why the friction power shown at no-load 
 
158 
 
 MANUAL OF MARINE ENGINEERING. 
 
EXPERIMENTS WITH TORPEDO BOAT. 
 
 159 
 
 is greater than that at full-load. The efficiency under load condensing is 
 invariably slightly better than when non-condensing. 
 
 It is also interesting to note how rapidly the rate of friction per revolution 
 increased when the power was high and revolutions were quite small. The 
 efficiency of the larger engine at full load and highest revolutions was - 929, 
 and the friction losses were '047 I.H.P. per revolution at 170 ; 0"079 at 
 300 revolutions ; 0"078 at 350 revolutions, dropping to 0"073 at 400. These, 
 however, are engines having forced lubrication. 
 
 Mr. Yarrow's Experiments with a Torpedo Boat to ascertain the efficiency 
 of engines, propeller, and ship were made in a most careful and exhaustive 
 way, and are most interesting and instructive. They show that the mechanical 
 efficiency of the engines varied from - 766 at 9 knots to - 923 at 15 knots, 
 and that the mechanical losses varied at a higher rate than given by the 
 arithmetical progression of revolutions. The high efficiency of these quick- 
 running engines at full power when carefully manufactured and adjusted 
 is manifested by an inspection of the figures in the subjoined table. 
 
 TABLE XXVI. — Yarrow's Experiments on Efficiency. 
 
 Speed, • knots, 
 
 9-0 10-0 
 
 110 
 
 120 
 
 130 
 
 14-0 
 
 15 
 
 Ind. horse-power, . 38-5 49-5 
 Friction, etc., loss, 9-0 10-1 
 Efficiency, . . 0-766 f 0-796 
 
 67-1 
 11-7 
 
 0-827 
 
 99-0 
 13-5 
 
 0-864 
 
 143-0 
 15-4 
 
 0-892 
 
 193-6 
 17-4 
 
 0-910 
 
 255-2 
 19-5 
 0-923 
 
 Mr. A. H. Tyacke, of Hull, made a series of trials with the engines of 
 two ships manufactured by Earles Co. to ascertain frictional resistance at 
 different speeds. In each case the engines were allowed to run without their 
 piopeiler shafting connected ; indicator diagrams were carefully taken at 
 various rates of revolution. At the same rates a set of diagrams were taken 
 with the ship running with propellers connected. The first ship had cylinders 
 12, 20, and 32 inches diameter, and a piston stroke of 21 inches. The second 
 ship had somewhat larger engines, the cylinders being 12|, 22, and 36 inches 
 diameter, and a piston stroke of 24 inches, and the highest revolutions only 
 111, which is small for such small cylinders. Under these circumstances 
 the efficiency is remarkably good, especially that of the latter. 
 
 TABLE XXVII. — Tyacke's Efficiency Experiments. 
 
 Revolutions per Minute. - 54 70 
 
 S8 90 
 
 106 
 
 Indicated horse-power, . . 54-6 87-0 180-9 223-6 
 Frictional horse-power, . 20-5 27-4 34-7 4(1-4 
 Efficiency (I.H.P. - F.H.P.)- \ Q ,^ 1 . 6g9 . 808 j . gl9 
 1 ri.l ., . . . . ) 
 
 1 , 1 
 
 306-0 
 46-2 
 
 0-849 
 
160 MANUAL OF MARINE ENGINEERING. 
 
 TABLE XXVIIa.— Tyacke's Further Experiments. 
 
 Revolutions per Minute, - 
 
 40 
 
 60 
 
 93 111 
 
 Indicated horse-power, 
 Frictional horse-power, 
 Efficiency (I.H.P. - F.H.P.) -f- I.H.P., . 
 
 38-80 
 1415 
 0-635 
 
 82-20 
 22-39 
 
 0-728 
 
 322-00 
 40-70 
 0-873 
 
 480-60 
 47-05 
 0-902 
 
 Steam Efficiency is quite as necessary as mechanical efficiency to a success- 
 ful marine engine, and has to be as carefully considered. First of all, the heat 
 conditions under which it works are the raising of a pound of water from the 
 temperature of the hot-well to a pound of steam with a sensible temperature, 
 and a latent heat corresponding to the boiler pressure ; during the cycle of 
 operations on the engine it is reduced again to water of the original tempera- 
 ture. Its efficiency then will be measured by comparing the heat given up 
 during the cycle when doing work with the total heat. If T is the tempera- 
 ture of steam at P pressure, L its latent heat, and t the temperature and I the 
 latent heat due to the pressure in the condenser — that is, the total heat of 
 evaporation at boiler pressure minus the heat of the feed-water is the amount 
 debited, and that amount less the total heat of evaporation at the pressure in 
 the condenser is to the credit account — the efficiency is, therefore, 
 
 Efficiency of steam 
 
 (T + L - t) -{I) T + L - * - I 
 
 T +L-t 
 
 T + L-« 
 
 The numerator of this fraction is the greatest amount of heat possible to 
 be used with steam of the pressure, and multiplying it by 772 will give the 
 maximum amount ci mechanical work possible in an engine working under 
 the conditions imposed. Dividing the product by 33,000 will give the horse- 
 power. 
 
 For example, marine engine condensers may now be relied on to produce 
 a vacuum of 28 inches, so that the back pressure is 1 lb. The total heat 
 at this pressure is 102° sensible, together with 1,042° latent, or 1,144°. At 
 100 lbs. pressure absolute the total heat is 327°+ 884°, or 1,211° ; at 150 lbs. 
 it is 358°+ 862°, or 1,220°; at 200 lbs. it is 381°+ 845°, or 1,226°, while at 
 250 lbs. it is 401 c + 831°, or 1,232°. 
 
 The heat required to change a pound of water at 102° temperature to a 
 pound of steam at these various pressures will be 1,109, 1,118, 1,124, and 
 1,130 B.T.Us. ; in each case the rejection is 1,042°. 
 Hence at 100 lbs. pressure the heat available for power is 1,109°— 1,042°, or 67°. 
 
 The efficiency of the steam = 67 -J- 1,109, or 0*0604 — that is, 6 per cent. 
 At 150 lbs. pressure the heat available is 1,118° — 1,042°, or 76°. 
 
 The efficiency of the steam = 76-4- 1,118, or 0*068 — that is, 6*8 per cent. 
 At 200 lbs. pressure the heat available is 1,124° - - 1,042°, or 82°. 
 
 The efficiency of the steam = 82-4- 1,124, or 0*073— that is, 7*3 per cent. 
 At 250 lbs. pressure the available heat is 1,130° — 1,042°, or 88°. 
 
 The efficiency of the steam = 88 -*- 1,130, or 0*078 — that is, 7*8 per cent. 
 In actual practice steam of 200 lbs. pressure enters the engine at a tempera- 
 ture of 381° F., and leaves it at the condenser at about 120°, so that it gives up 
 in the engine 261°, which multiplied by 772 and divided by 33,000 is 6*1 H.P. 
 
STEAM EFFICIENCY OF RECIPROCATORS. 
 
 10 1 
 
 The triple-expansion engine of the mercantile marine consumes 16 lbs. 
 per horse-power-hour, or 0-267 per minute. In such an engine a pound of 
 steam produces 3-75 I.H.P. 
 
 Hence the steam efficiency of the engine is 3*75 -f- 6-1, or 0-615. 
 
 An engine using steam at 250 lbs. pressure absolute receives it at a tem- 
 perature of 401°, and rejects at 120°, giving up to the engine 281°, which is 
 equivalent to 6-58 H.P. 
 
 A quadruple engine under these conditions consumes 14-5 lbs. of steam 
 per I. H.P. -hour, or 0-242 lb. per minute, consequently a pound of steam 
 produces in it 4-13 I.H.P. 
 
 Hence its steam efficiency is 4-13 -f- 6-58, or 0-628. 
 
 A turbine using steam of 200 lbs. pressure and rejecting at 81° will absorb 
 300°, equivalent to 7-07 H.P. 
 
 Assuming it to consume only 12 lbs. of steam, the power generated in 
 by a pound of steam is 4-99 H.P. 
 
 The efficiency of this turbine is 4-99 -*- 7-07, or 0-706. 
 
 The Steam Efficiency of the Best Turbines is as high as 72 per cent, when 
 of large size, and under favourable circumstances ; that of turbines of good 
 make and over 2,000 shaft horse-power can be taken at 64 to 66 per cent. 
 With superheated steam those on shore of large size can be depended on to 
 show an efficiency of 72 to 75 per cent. 
 
 The Steam Efficiency of Reciprocators of large size, good design, and 
 good construction, 60 to 63 per cent, is satisfactory. 
 
 The following table gives the maximum amounts of work generated 
 theoretically by a pound of steam during admission and expansion : — 
 
 TABLE XXVIII. — Maximum Work done by 1 lb. of Steam Press. p v expanding 
 Adiabatically to Press. p 2 , and exhausting to Condenser at 1 lb. Press. 
 
 Initial pressure, 100 lbs. 125 lbs. 
 
 150 lbs. 
 
 175 lb3. 
 
 1* 
 
 200 lbs. 
 
 250 lbs. 
 
 Terminal p 2 . 
 
 B.T.T7. 
 
 HP. 
 
 B.T.U. 
 
 H.P. 
 
 B.T.U. 
 
 H.P. 
 
 B.T y. 
 
 n.p. 
 
 B.T U. 
 
 H.P. 
 
 B.T.U. i H.P. 
 
 1-0 lb. 
 
 288-0 
 
 6-80 
 
 303-0 
 
 714 
 
 314-5 
 
 7-42 
 
 324-0 
 
 7-64 
 
 332-5 
 
 7-84 
 
 348-0 
 
 8-21 
 
 20 lbs. 
 
 286-8 
 
 6-76 301-8 
 
 7-09 
 
 313-3 
 
 7-26 
 
 322-4 
 
 7-61 
 
 331-3 
 
 7-81 
 
 346-8 
 
 8-18 
 
 30 .. 
 
 277-2 
 
 6-54 ! 292-6 
 
 6-90 
 
 309-2 
 
 717 
 
 313-5 
 
 7-40 
 
 322-2 
 
 7-60 
 
 337-2 
 
 7-95 
 
 4-0 .. 
 
 266-2 
 
 6-28 281-9 
 
 6-65 
 
 294-0 
 
 6-93 
 
 303-2 
 
 7-15 
 
 3120 
 
 7-36 
 
 328-2 
 
 7-74 
 
 5-0 ., 
 
 258-0 
 
 6-09 ■ 273-5 
 
 6-45 
 
 285-0 
 
 6-72 
 
 295-0 
 
 6-96 
 
 304-0 
 
 717 
 
 320-0 
 
 7-55 
 
 60 .. 
 
 249-9 
 
 5-90 266-9 
 
 6-27 
 
 278-0 
 
 6-55 
 
 287-9 
 
 6-80 
 
 297-4 
 
 7-01 
 
 313-9 
 
 7-40 
 
 70 .. 
 
 243-9 
 
 5-75 
 
 259-3 
 
 6-11 
 
 271-9 
 
 6-42 
 
 281-8 
 
 6-64 
 
 290-8 
 
 6-86 
 
 307-8 7-26 
 
 80 .. 
 
 237-5 
 
 5-61 
 
 253-5 
 
 5-96 
 
 265-5 
 
 6-26 
 
 276-8 
 
 6-53 
 
 285-5 
 
 6-73 
 
 302-5 7-13 
 
 9-0 „ 
 
 2320 
 
 5-47 
 
 248-0 
 
 5-85 
 
 260-5 
 
 6-14 
 
 271-3 
 
 6-40 
 
 280-7 
 
 6-62 
 
 297-0 
 
 7-00 
 
 10 „ 
 
 226-9 
 
 5-35 
 
 2421 
 
 5-71 
 
 254-0 
 
 6-00 
 
 265-5 
 
 6-26 
 
 275-9 
 
 6-5) 
 
 292-1 
 
 6-89 
 
 11 „ 
 
 2221 
 
 5-24 
 
 237-1 
 
 5-59 
 
 249-8 
 
 5-89 
 
 260-7 
 
 615 
 
 270-5 
 
 6-38 
 
 287-1 
 
 6-77 
 
 12 ,. 
 
 217-0 
 
 512 
 
 233 
 
 5-49 
 
 245-5 
 
 5-80 
 
 256-1 
 
 6 04 
 
 265-4 
 
 6-26 
 
 283-0 ' 6-67 
 
 13 „ 
 
 212-2 
 
 5-00 
 
 228-6 
 
 5-39 
 
 240-7 
 
 5-69 
 
 251-6 
 
 5-93 
 
 261 1 
 
 6-16 
 
 278-7 6-57 
 
 14 ., 
 
 207-4 
 
 4-90 
 
 224-1 
 
 5-28 
 
 236-4 
 
 5-60 
 
 247-2 
 
 5-83 
 
 257-3 
 
 6-07 
 
 274-4 6-47 
 
 15 
 
 2040 
 
 4-81 
 
 220-0 
 
 519 
 
 2320 
 
 5-47 
 
 243-5 
 
 5-74 
 
 253 
 
 5-97 
 
 270-0 <6-37 
 
 10 „ 
 
 200-6 
 
 4-73 216-6 
 
 511 
 
 228-6 
 
 5-39 
 
 239-6 
 
 5-64 
 
 249-9 
 
 5-89 
 
 266-4 6-28 
 
 17 „ 
 
 197-4 
 
 4-65 213-3 
 
 5 03 
 
 225 
 
 5-30 
 
 236 
 
 5-56 
 
 246-5 
 
 5-81 
 
 262-6 6-19 
 
 18 „ 
 
 194-3 
 
 4-58 210-1 
 
 4-95 
 
 220 
 
 5-23 
 
 232-9 
 
 5-49 
 
 243-3 
 
 5-74 
 
 259-9 6-11 
 
 19 „ 
 
 191-3 
 
 4-50 207-2 
 
 4-88 
 
 219-4 
 
 517 
 
 230 
 
 5-42 
 
 240-5 
 
 5-67 
 
 257-3 
 
 6 06 
 
 20 „ 
 
 188-4 
 
 4-44 204-4 
 
 4-82 
 
 217-3 
 
 5- 12 
 
 228 
 
 5-38 
 
 238-0 
 
 5-61 
 
 255-4 
 
 6-02 
 
 25 „ 
 
 1760 
 
 4-15 191-3 
 
 4-51 
 
 2(15-0 
 
 4-83 
 
 215-9 
 
 5-09 
 
 225-4 
 
 5-31 
 
 243 5-73 
 
 30 „ 
 
 164-8 
 
 3-89 180-0 
 
 4-24 
 
 194-5 
 
 4-59 
 
 205-0 
 
 4-83 
 
 214-9 
 
 5-07 
 
 232-0 
 
 5-47 
 
 35 .. 
 
 L55-2 
 
 3-66 169-6 
 
 4-00 
 
 184-2 
 
 4-34 
 
 L95-2 
 
 4-60 
 
 205-2 
 
 4-84 
 
 .).)■>. o 
 
 5-24 
 
 40 „ 
 
 146 -2 
 
 3-45 L60-5 
 
 :(-so 
 
 175-1 
 
 413 
 
 I86 
 
 4-40 
 
 .196-0 
 
 4-62 
 
 21 30 
 
 5-02 
 
 
 
 
 
 
 
 
 
 
 
 
 11 
 
 
162 MANUAL OF MARINE ENGINEERING. 
 
 Steam Efficiency may also be ascertained by referring to Table XXVIII., 
 where it will be seen that a pound of steam expanding adiabatically from 
 200 lbs. absolute to 1 lb., when it enters the condenser, can theoretically give 
 332-5 B.T.U., equivalent to 7-84 horse-power. 
 
 A Turbine practically works on these conditions with a steam consumption 
 of about 11-5 lbs., or 1 lb. of steam is good for 5-24 H.P. In this case, then — 
 The efficiency = 5-24 + 7-84, or 66-8 per cent. 
 
 A triple-compound engine using steam at 200 lbs. pressure absolute ex- 
 panding to 8 lbs. (nominal expansion rate 25), and exhausting at 1 lb. to the 
 condenser, requires 15 lbs. of steam per H.P.-hour, so that 1 lb. is good for 
 4 H.P., and theoretically under these conditions develop 285-5 B.T.U., equi- 
 valent to 6-73 H.P. (v. Table XXVIII.). In this case— 
 
 (a) Efficiency = 4 -4- 6-73, or 0-594, or 59-4 per cent. 
 But 1 lb. of such steam if expanded to its full extent is shown to be good foi 
 7-84 H.P., then— 
 
 (b) Efficiency = 4-4- 7-84, or 0-51, or 51-0 per cent. only. 
 A Quadruple Engine working with steam at 250 lbs. absolute, expanding 
 to 8 lbs., or nominally 31 times, exhausts to the condenser at 1 lb., consumes 
 14 lbs. of steam per H.P.-hour, so that 1 lb. is good for 4-286 H.P. But 1 lb. of 
 steam is theoretically good for 302 -5 B.T.U. or 7 -13 H.P. under these conditions- 
 Then efficiency = 4-286 -f- 7-13 = 0-601, or 60-1 per cent. 
 Taking the full value of 1 lb. of such steam as 8-21 {v. Table XXVIII.)— 
 (b) Efficiency = 4-286 -=- 8-21 = 0-52— that is, 52-0 per cent. only. 
 
 High Pressure : its Advantages and Disadvantages. — Before the intro- 
 duction of the compound engine for marine purposes, the boiler pressure had 
 been as high as 60 lbs. in quite large steamers, in H.M. service, with the 
 non-condensing engines of 200 N.H.P. ; some of these were fitted in certain 
 battle-ships during the 1855 Russian war ; but after the compound engine 
 secured the confidence of all classes of steamship owners, that pressure was 
 very much exceeded with beneficial results. 
 
 Prior to the general use of the triple-compound engine 90 lbs. was a very 
 common boiler pressure, and many ships' boilers were made for a working 
 pressure of 100 lbs., and a few for as high as 110 lbs. The triple-compound 
 engine itself was for a long time worked with steam of 150 lbs. ; then 165 lbs. 
 became a fashionable pressure, to be soon superseded by 175 lbs. ; and 
 now even 200 lbs. is general, although the economic gain with triples by 
 going from 150 to 200 lbs. is questioned by some engineers who have made 
 careful observations of all the conditions involved in the change.* 
 
 The objection to still higher pressures is rather of a practical nature, 
 but can be safely overcome, since steam superheated to 600° F. may now be 
 use*l. Steam at a pressure of 250 lbs. absolute has a temperature of 401° F., 
 or nearly that of the melting-point of tin. It will, therefore, affect the 
 condition of some of the metals with which it comes in contact, rendering the 
 surfaces brittle and in a bad condition to withstand severe rubbing. More- 
 over, common unguents are vaporised, and the walls of the cylinders become 
 too hot to condense their vapour when exposed to very high temperatures ; 
 but the heavy mineral oils now used, however, have a boiling point of 
 
 •The N.E. Coast Inst. E. and S. Standard Specification of 1917 gives 180 lbs. for cargo steamers 
 "it!) triples. 
 
FRICTION OF THE PISTON. 163 
 
 700° F. The difference in expansion of different metals is so considerable, 
 that the utmost care must be exercised in the design and manufacture of the 
 cylinder, etc., to prevent racking, which causes leakages and breakages. The 
 pressure necessitates considerable thickness in all cast-iron work ; and 
 neglecting all considerations of weight or cost, this alone constitutes a source 
 of objection and danger, inasmuch as the sudden exposure of thick masses 
 of this metal to high temperatures on one side only is sure to distort, and 
 very likely to fracture it. The liability to leakage is, of course, greater 
 with the higher pressure, independent of temperature, and the danger to 
 the attendants, from even small explosions, is very much increased. Again, 
 in order to expand steam at this pressure, so as to obtain from it the maximum 
 efficiency, it will be found necessary to use a series of cylinders ; and although, 
 as will be shown later on, the engine with four stages is not without virtue, 
 a larger number will not commend itself to engineers generally for engines 
 of the smallest and largest sizes. In practice, it has been found that an 
 engine can be worked with steam of a pressure less than that usual, but 
 superheated to the temperature corresponding to the higher pressure, and 
 yet be more economic and efficient than a similar one working with that 
 higher pressure of steam. 
 
 Efficiency of the Engine as a Machine. — The marine engine suffers loss, 
 in common with all machines, from certain physical causes beyond the 
 absolute control of the most skilful designer, and engineers can only aim at 
 mitigating the evil, without entirely overcoming it. The chief cause of loss 
 of energy is, of course, friction (1) of the piston, (2) of the stuffing-boxes, 
 (3) of the guides and slides, (4) of the shaft journals, (5) of the valves and 
 valve-motion. Another source of loss is that from the resistance of the 
 pumps ; and, finally, the inertia of the moving parts, which have a reciprocal 
 action, as in the piston and rods, may be the cause of further waste of energy. 
 Unless the momentum is balanced, or the energy imparted at one part of 
 the stroke and stored in the heavy moving masses, is given out wholly by 
 the end of the stroke, a serious loss ensues, and the mechanism has to sustain 
 the strain of forces which might be otherwise usefully employed. 
 
 1 . The Friction of the Piston in the vertical engine in smooth water depends 
 on the pressure of the packing on the sides of the cylinder ; so that, if the 
 piston were solid and simply a good fit, there would theoretically be no 
 friction, and in practice none beyond that due to the viscidity of the unguent 
 and to the pressure on the sides of the cylinder from the rolling and general 
 motion of the ship. This is what is required in a piston, and that one is 
 most nearly perfect which is capable of moving steam-tight in the cylinder 
 with least pressure of the packing, and so approximates to the condition of 
 a solid one. Resistance due to this cause has been reduced to a minimum 
 in modern engines by care in manufacture and skill in design ; the cylinder 
 is now truly and smoothly bored from end to end, the metal, which should 
 be hard and close-grained, soon becomes polished and glazed, and in the 
 best possible condition for smooth working ; the packings of the piston are 
 metallic, and the methods of pressing out the packing ring such as to ensure 
 a uniform and evenly spread pressure of small magnitude. The loss from 
 packing the pistons too tightly may become very great, and too much care 
 cannot be exercised in attending to this most important part of the engine. 
 
 In horizontal and diagonal engines, the weight of the piston pressing 
 on the side of the cylinder sets up friction, and as the cylinder wears in 
 consequence more on the bottom than the top, it gets out of shape, thus 
 
164 MANUAL OF MARINE ENGINEERING. 
 
 necessitating more pressure on the packing-ring to maintain steam-tight 
 contact, and thereby increasing the friction. The arrangements necessary 
 for carrying the weight of the piston prevent all the better forms of piston 
 rings and springs from being adopted in horizontal engines ; so that the 
 friction of pistons alone renders the engine less efficient than the vertical 
 form. The most important improvement effected of late years, tending to 
 the better working of diagonal engines, has been making the piston solid 
 and of steel (v. fig. 90); in this way the weight has been considerably 
 reduced, and the strength increased very materially ; the prevention by this 
 plan is better than the cures attempted by guide-rods, etc. 
 
 2. The Friction of Stufling-Boxes. — Since the use of the higher pressures 
 of steam this has become a very important factor in the consideration of 
 engine efficiency. It may be extremely variable, as it depends so much on 
 the care and judgment of the attendants, that, however carefully these parts 
 may have been designed and constructed, their good working is entirely 
 beyond the control of the designer. The manufacture of good and reliable 
 metallic packing at reasonable rates has removed to a large extent this 
 difficulty, so that now the glands of the high-pressure and medium-pressure 
 cylinders are almost invariably fitted with some satisfactory form of metallic 
 packing ; and even those of the low-pressure cylinder are so treated in all 
 important steamers, especially those of large size, although some engineers 
 still prefer good vegetable packing for them. That the resistance may be 
 very considerable is proved by the fact that the old trunk engines could 
 be slowed down and nearly stopped by tightening the trunk glands, and in 
 the ordinary piston-rod engine the speed may be appreciably retarded by 
 the same process. The glands should never be so tight that the rods are 
 rubbed absolutely dry in passing through them. For the efficient working 
 of the engine it is better that a faint leakage of vapour should pass out with 
 the rod, as that is generally an indication that the packing is just tight 
 enough, besides which the moist vapour lubricates the packing and keeps 
 it soft when of vegetable composition. 
 
 3. The Friction of the Guides and Slides is now, perhaps, the least im- 
 portant in the everyday experience of the losses to which an engine is liable ; 
 as those parts may be said to design themselves, and are now of ample surface. 
 and they are in such a position as" to command the attention of the engineer ; 
 they are also, as a rule, easily lubricated. In most classes of engines, the 
 piston-rod guide is the chief one for consideration, and since, from the form 
 of the rod-end, it is nearly impossible to give too small a surface to the shoe, 
 it is seldom found, even in ill-designed engines, to give much trouble ; but 
 in certain forms of engine this is not always the case, so that care has to 
 be taken both in designing and attending to the guides. The maximum 
 pressure on the piston-rod guide of a marine steam engine is usually from 
 two to three-tenths of the load on the piston, and supposing the ratio of 
 maximum to mean pressure to be l - 50, and the coefficient of friction, under 
 good circumstances, probably not less than 0*05, then 
 
 _ . L , ., AA1 . ftni . load on piston 
 
 Resistance of guide = 01)1 to 0-015 X z-~- > 
 
 1 'OO 
 
 which means that from two-thirds to one per cent, of the energy of the piston 
 is of necessity consumed in overcoming the friction of this one guide. 
 
 Improvements in the manufacture of, and choice of suitable metals for 
 guides, have very sensibly diminished the loss from this cause ; besides 
 
LOSS FROM FRICTION. 1G5 
 
 which the pressure on the guides has been in modern engines reduced by- 
 making the connecting-rod longer in proportion to the stroke. When cast-iron 
 has become, by rubbing, glazed on the surface, there is no material better 
 for guides. However, as some engineers will not wait for, or do not trust to, 
 this state of metallic surface, but prefer the rubbing surface to be of a softer 
 nature than the rubbed, it is not unusual to find white metal fitted to the 
 " shoes." That some of the older engines were inefficient from loss at the 
 guides, is proved by the rapid wearing of the old bronze shoes ; the work 
 necessary to convert so many cubic inches of metal into powder being the 
 measure of avoidable loss at that point. 
 
 4. The Loss from Friction at the Shaft Journals, etc., may be also very 
 considerable, as the load on the piston is transmitted from the crank-pin 
 to them, in addition to that caused by the weight of the shaft itself and the 
 connections. The same may be said of the crank-pins, which have pressing 
 on them the whole of that load, in addition to the weight of the rods. This 
 friction may be very severe, especially in fast-running engines. It was 
 formerly held that friction was independent of velocity, so far as movement 
 through a fixed distance is concerned ; that is, if a body be moved through 
 10 feet, the friction is the same if the movement takes place in 1 second or in 
 10 seconds ; but if time be taken into account, the friction of moving the body 
 ten times over the 10 feet in 10 seconds is ten times that of moving it once 
 in 10 seconds. Since M. Morin made his experiments further investigations 
 have been most carefully made by Mr. Tower and others, which show that 
 friction does vary with the velocity, probably as the square root of the 
 velocity ; hence for modern piston speeds a larger allowance of surface is 
 required than formerly. In a marine screw engine making seventy revolutions 
 per minute, the friction is at least seventy times that of one revolution ; 
 and, consequently, if a paddle engine, having the same size of cylinders, 
 •and working with the same pressure of steam, makes only thirty-five revolu- 
 tions per minute, its friction of journals will be half that of the screw engine 
 per minute. As the first screw engines working without gearing were gene- 
 rally designed by men whose experience had been gained with the slower 
 working paddle engines, it is not astonishing to find that the bearings were 
 not always sufficient for the work on them, and perhaps the speed of the 
 rubbing surfaces prevented the lubrication from being so efficient as had 
 been the case previously, and so aggravated the evil. Again, the old paddle 
 engine and geared screw engine had cylinders of longer stroke compared 
 with their diameters than had the direct-working screw engine, and as the 
 diameter of the shaft depends on the area of piston and length of stroke 
 combined, while the pressure on the bearings is affected only by area of 
 the piston, the diameter of the shaft might remain the same, although the 
 size of the piston had been very much increased. Now, since most of the 
 old rules for length of journals took cognisance of the diameter of shaft 
 only, although the pressure on the journals might have been doubled, there 
 was only the same surface to take it. 
 
 For example, a paddle engine of 5 feet stroke might have the same 
 diameter of shaft as a screw engine of 2 feet 6 inches stroke, each having 
 the same cylinder capacity ; but the engine with the short stroke would 
 have a piston area twice that of the long stroke, and consequently with the 
 same steam pressure there would be double the load on the journals, and 
 tins with, generally, double the number of revolutions of the shaft. 
 
166 MANUAL OF MARINE ENGINEERING. 
 
 Friction and Surface. — It was formerly an axiom that friction was inde- 
 pendent of surface; but there must always have been limitations to such 
 a statement. It meant that so long as the moving body was supported 
 on a surface sufficiently large to prevent the unguent from being squeezed 
 out, the resistance due to friction was the same fraction of its weight, what- 
 ever the area. Practical engineers, however, never had unbounded belief 
 in this axiom. It is now clearly understood that when two surfaces are 
 separated by a stratum of oil, or other unguent, they do not touch one 
 another, and, therefore, each is sliding on the unguent. It is also generally 
 supposed that the particles of unguent assume a globular form in the process, 
 so that they form a kind of ball-bearing surface. Anyway, so long as the 
 unguent is maintained in position, the resistance is small and is constant. 
 It is, however, well known that when the same unguent is used with different 
 metals having apparently equally smooth surfaces, there is a difference in 
 the coefficient of friction, as the fraction is called ; and, further, that certain 
 metals will not work at all well on others — e.g.. soft steel on bronze. It is 
 evident, then, that either the metals affect the unguent, or they do not really 
 present equally good surfaces. It is known that oil acts chemically on 
 copper, and, therefore, on copper compounds, while it has no action on tin 
 and antimony. This may possibly account for the different behaviour 
 of white metal and bronze. 
 
 Modern experiments have shown that surface has its influence, even 
 within the safe limits ; that the coefficient is higher with very light loads 
 per square inch than heavier ones, so that when an engine is running quite 
 light, the percentage of loss from friction is higher than when running full 
 load. In the case of a marine engine running slow, the pressure per square 
 inch on the guides, etc., will be much less in consequence of reduced load; so 
 that the coefficient of friction will be, from that cause, somewhat higher. On 
 the other hand, as the engine is moving slower, and inasmuch as friction varies 
 as JV, the coefficient should be less, and so one may balance the other. 
 
 Tower found that at 90° F.. 
 
 . . v'V 
 
 Coefficient of friction = 20c -p— 
 
 Where V is the rubbing velocity in feet per minute, P the nominal pressure 
 in lbs. per square inch, c a coefficient depending on the lubricant, which, 
 for sperm oil, is "0014, rape oil "0015, mineral oil -0018, and olive oil '0019, 
 the oil supply being liberal and constant. 
 
 For example, the coefficient for a guide on which the pressure is. say, 
 100 lbs. per square inch, the piston speed 900, and on which mineral oil is 
 freely used, 
 
 F = 20 x -0018 ^j^ = -0108. 
 
 If the oil is only sparingly supplied, as with a syphon lubricator, the coeffi- 
 cient may be four times this, or '0432, which is very nearly what Morin 
 stated to be the best result with well lubricated surfaces. 
 
 Temperature also affects' the friction ; the coefficient is reduced when 
 the temperature rises so as to render the unguent fluid and not sticky. On 
 the other hand, further rises of temperature tend to make it too fluid and 
 
LOSS FROM THK PUMPS. 167 
 
 P 
 
 thin. With some mineral oils, raising the temperature from 60° to 120 
 caused the friction to be quadrupled. 
 
 The aim of the engineer must be, therefore, to prevent metallic surfaces 
 from coming into actual contact, for then the friction would be very severe, 
 and soon cause the surfaces to abrade, and even, as is seen sometimes in the 
 case of cast iron, to strike fire. The lubricant should be introduced at the 
 points where the pressure is least, in order that those where greatest may be 
 well lubricated. Friction at the journals, too, must always be a source of 
 anxiety, though not nearly so much so now as formerly. Crank-shafts are 
 more truly turned, although even now there is room for improvement ; the 
 foundations of the engines are more stiffly made, so that there is no springing 
 at the bearings ; the bearing surface is pro rata larger, and white metal is 
 universally fitted to both crank-pins and main bearings. 
 
 5. The Friction of Valve Motions and of the valves is very considerable 
 at all times, and may be severe when the valves are running dry. Even 
 when the pressure on flat valves is partly relieved by frames, etc., on their 
 back, the load on the rods was sometimes so great as to bend them. With the 
 large increase of piston speed of necessity have come larger valves. This, 
 with increased boiler pressure, forced most makers of marine engines to 
 revive the piston valve for the high-pressure cylinders, and latterly to fit 
 them to the medium-pressure and even to the L.P. cylinders. On the 
 other hand, a few eminent builders stuck to flat valves for all their 
 cylinders, and that, too, with a success that was somewhat remarkable, 
 but it was with lower pressures and slower rates of revolution than now 
 obtain. 
 
 Attempts have been made from time to time to use Corliss and other 
 quick-shutting " drop " valves instead of slide valves. So far the results 
 have been disappointing, and the increased efficiency due to quick cut-off 
 and small clearance which such valves and gear give to land engines is now 
 not yet attainable to the same extent with marine compound engines. But 
 the marine engine may have its efficiency very much improved by some 
 other form of valve which, while giving plenty of egress for the steam to 
 the condenser, very much reduces the clearance now obtaining in low-pressure 
 cylinders. 
 
 6. Loss from the Pumps. — In all engines, whether condensing or non- 
 condensing, there is a loss of efficiency from the feed-pump when it is worked 
 by the engine. It is true that the work done by it in forcing the water into 
 the boiler is stored energy, and, therefore, not lost, but its low efficiency 
 is a source of loss. It is now, however, the common practice for the feed- 
 pump to be worked by an independent engine. This is convenient, and 
 has advantages which will be discussed elsewhere, but it must not be over- 
 looked that the cost in steam is greater than by the old method, as the 
 independent engine consumes at least double the amount of steam per I.H.P. 
 that the main engines do. 
 
 To some extent the air-pump may be said to store energy, for it takes 
 water from under a pressure of, say, 2 lbs. per square inch, and places it at 
 the disposal of the feed-pumps under a pressure of 15 lbs. The chief work 
 of the air pump is to withdraw air and other gases from the condenser at the 
 pressure in it, and deliver them at atmospheric pressure. This, except as 
 a means to an end, is all lost energy, while the friction, etc., of this pump 
 adds to t»he loss, and so reduces the efficiency of the engine. 
 
168 MANUAL OF MARINE ENGINEERING. 
 
 Air-pumps, especially in warships, are sometimes worked by an inde- 
 pendent engine (v. figs. 124 and 127) ; the same objections apply to this pump 
 being dealt with in this way as to the feed-pumps, and there is not the corre- 
 sponding advantage in practice, although there might be if the pump were 
 regulated to run exactly to the requirements of the condenser. 
 
 From carefully made experiments with land engines, there is good grounds 
 for estimating the power absorbed in driving the air pumps of a marine 
 as varying from 5 per cent, of the gross I.H.P. in small engines to about 
 1 per cent, in large ones. If the air pump does not exceed one-twentieth the 
 L.P. cylinder, it will be only from 2'0 to 07 per cent. 
 
 In an engine of 1,500 I.H.P. the air pump will absorb 3 per cent, of the 
 power developed, while in one of 3,000 I.H.P. it will be T35 per cent., and 
 in one of 5,000 I.H.P. 0'82 per cent. 
 
 The circulating-pumps render nothing for the large amount of energy 
 put into them. and. viewed in this light only, would seem to be a great im- 
 pediment to the engine. As a means to a most valuable end it is otherwise 
 appraised ; as a necessary adjunct to the surface-condenser it has conduced 
 very much to the economy of the modern marine engine. Besides all the 
 energy lost in the driving of this pump the greatest waste of the steam engine 
 is taken by it, for its water conveys away all the latent heat required to 
 make the steam. It has been suggested that by turning the stream of water 
 towards the stern some propelling effect would be obtained, and some of this 
 loss be prevented. So far no engineer has gone out of his way to effect 
 this small economy. 
 
 The power absorbed in circulating the water through the condenser in 
 steam engine installations on land is heavy ; much more so than that on 
 board ship, where the condenser itself is commonly below the water-level, 
 and there is no " head " but that due to resistance in tubes, etc., to be 
 pumped against. The variation found in these land installations is, how- 
 ever, instructive, for in cases of 1,500 I.H.P. the circulating pumps absorb as 
 much as 10 per cent., at 3,000 I.H.P. stations it is 4'5 per cent., and at those 
 of 5,000 I.H.P.. 3"2 per cent. On shipboard, with the discharge valve at or 
 below the water line, 1'7 to 2 - 5 per cent, of the main engine power is sufficient 
 in the tropics, and about a half in temperate zones. In merchant ships, 
 with the discharge well above the water line, the loss is 2*2 to 3'0 per 
 cent, in the tropics. 
 
 7. Inertia of Moving Parts. — There should be no losses from this cause 
 if proper provisions are made, for all the energy stored in them in getting 
 them to their maximum motion should be usefully employed and absorbed 
 before coming to rest. In the older engines, as in many modern ones, con- 
 siderable loss arose from the fact that much of the energy stored in the 
 pistons and other heavy moving parts was not usefully employed, but was 
 wasted in vibrating the engine itself and the ship. This, however, is fully 
 dealt with in Chap. xxix. There will, of course, still be small losses due to 
 friction on bearings and guides from the inertia of the moving parts after 
 the engine has been balanced, for the balancing cannot bring all the forces 
 into one plane. 
 
 It will be seen by the foregoing that the energy lost in a marine engine 
 is not a definitely fixed quantity, and is dependent neither on design nor 
 on construction alone, but rather on the degree of care exercised by those 
 having the charge of it. That much may be saved by good and careful 
 
INERTIA OF MOVING PARTS. l(j ( J 
 
 designing and workmanship is evident, but, within certain limits, a good 
 engine may prove less efficient than a really inferior one from mere lack of 
 proper attention from those in charge, and especially if they are not supplied 
 with good and suitable lubricants. 
 
 The Losses due partly to Mechanical Defects and partly to Physical Causes 
 are those which cannot be classed as belonging to the engine as a machine, 
 nor to it as a heat engine simply. The most important of these is conse- 
 quent on the employment in its construction of materials having a high 
 power for conductivity of heat. The steam pipes were generally of copper, 
 and when of steel are comparatively thin, so that there is much loss of heat 
 from their surfaces. Careful covering with material having a low conductivity 
 does much to prevent this waste, but, with the high pressures now used and 
 consequent high temperatures, unless the lagging is very thick and the pipe 
 flanges well covered there is still great loss. The loss, too, is not limited 
 to the mere heat which escapes, for if saturated steam — that is, steam con- 
 taining as much water as it can — is robbed of any of its heat on its road to 
 the cylinders it deposits some of its water ; this water obstructs the free 
 passage of steam through valves and passages, and, till forced through the 
 escape valves or drain cocks, obstructs the pistons themselves. The 
 efficiency of the engine as a heat engine is also materially affected by the 
 presence of water in the cylinders. The losses are very manifest when 
 the speed of the engine is, as often happens, materially checked by water 
 coming with the steam. The gain from superheating the steam for the 
 marine engine arose, and will again be found to arise, largely from the 
 steam being quite dry rather than from the higher temperature at the 
 cylinder. 
 
 Liquefaction takes place on expansion in the cylinders in spite of steam 
 jackets and superheating, and a small amount of water to lubricate the 
 internal parts is a good thing, especially as the use of oils is very restricted 
 or wholly done away with nowadays. Notwithstanding, the cylinders should 
 be as carefully " lagged " as the steam pipes, and the covers and flanges 
 of the high-pressure cylinder should be as carefully covered as any other 
 part. 
 
 It is very doubtful if steam, during expansion in the fast-running marine 
 engine, can take any appreciable heat from the jackets ; that such jackets add 
 somewhat to the efficiency of the engine must, however, be admitted when 
 they are properly drained ; it is probable, therefore, that the gain is one of 
 mechanical efficiency, from the fact that the cylinders are freer of water with 
 steam jackets in use. 
 
 If this argument is based on fact, a jacket to the low-pressure cylinder 
 might be of some use, and even outweigh the loss arising from the reheating 
 of steam about to enter the condenser at exhaust. 
 
 Superheating the steam so that it might continue wholly in the vapour 
 state till it reaches the condenser has been the hope of some sanguine in- 
 ventors. Superheating the steam with heat that otherwise is lost is obviously 
 the most desirable thing as a means of economy. Superheating, however, 
 must have its limits, for in economising steam there may be great loss of 
 energy, as. for example, by excessive friction of the high-pressure piston, 
 valves, and rods from want of a lubricant when the steam is so dry as to 
 absorb it all in the vaporous state. As a matter of fact the steam may leave 
 
170 
 
 MANUAL OF MARINE ENGINEERING. 
 
 the marine boiler with a large amount of superheat, and even on reaching 
 the cylinder still have a fair margin, but the piston will not have moved 
 many inches before it is saturated steam. If water deposit in the high- 
 pressure cylinder is to be avoided, the superheater will require to have 
 special heat supply or its own fire, and unless very special means are 
 adopted for lubricating the piston and valve will work badly and cut the 
 surfaces. 
 
 Superheating went out of fashion on the advent of steam of 60 lbs. pressure, 
 partly due to the fact that its temperature, 307° F., was nearly as high as 
 superheating had then gone, and to the restrictions put by the Board of 
 Trade on superheaters and their somewhat rapid decay. 
 
 Metallic packing was unknown at that time, and the gain in economy by 
 the compound engine over the old surface condensing 30-lb. pressure ones 
 satisfied fully the enterprise of the good folks in those days of good freights 
 and cheap fuel. 
 
 By superheating the steam for electric light installations on shore great 
 reductions in consumption of fuel have been effected (fig. 64). No doubt 
 similar results may also be obtained by employing satisfactory superheating 
 apparatus on board ship, as shown below, but great care must be taken in 
 the design and construction of the piston and valves of the first cylinder 
 into which such steam enters, or the loss may exceed the gain in steam 
 efficiency. 
 
 Experiments with Superheated Steam in Modern Times* have been made, 
 and the results of some very interesting ones were given by Mr. Felix Godard 
 in a paper read at the Inst. Naval Architects. These trials were made with 
 the screw steamers " Garonne " and " Kance," each 300 feet long, 40 feet 
 beam, and 25 '5 feet deep. Mean draft of water 21 feet, gross register 2,700 
 tons. 
 
 The " Garonne " has no superheater, but her total heating surface is 
 equal to that of the " Kance." together with the surface in that ship's super- 
 heater. That is, the total surface exposed to heat is the same in each ship. 
 Their engines have cylinders 23 inches, 36 inches, and 59 inches diameter 
 and 42 inches stroke. 
 
 
 '■ Garonne." 
 
 "Ranee." 
 
 Boiler pressure, ..... 
 
 178 lbs. 
 
 177 lbs. 
 
 Steam temperature, .... 
 
 378 3 F. 
 
 518' F. 
 
 Heating surface of boilers, . . . 
 
 3,767 sq. ft. 
 
 2,982 sq. ft. 
 
 Superheating surface of boilers, . . 
 
 none 
 
 785 sq. ft. 
 
 Revolutions per minute, . . . 
 
 72-3 
 
 75-4 
 
 Indicated horse-power, .... 
 
 1,104 
 
 1,304 
 
 Coal consumed per l.H.P. per hour, 
 
 1-12 lbs. 
 
 0-90 11). 
 
 Coal per mile on service, 
 
 1 .VI lbs. 
 
 12(5 lbs. 
 
 Increase in power, .... 
 
 
 18-1 per cent. 
 
 Decrease consumption per I.H.P., . 
 
 . . 
 
 20-1 „ 
 
 Saving of fuel on voyage, 
 
 
 18-2 „ 
 
 Two other ships were fitted in the same way — that is, one, the 
 
 * Vide Appendix E for more recent experiences. 
 
EXPERIMENTS WITH SUPERHEATED STEAM. 
 
 171 
 
 
 * 
 
 V 
 
 
 Steam Consumptions of Various Engines 
 Using Superheated Steam. 
 
 A = Vertical compound fast running, 295 I.H.P., 
 
 
 17 
 
 
 
 t 
 
 vacuum 20 inches, boiler pressure 175 lbs. 
 
 V' 
 
 
 
 
 
 B = Three-crank compound slow running, 5500 
 
 
 
 Q 
 
 
 
 
 
 I.H. P. , vacuum 28 inches, boiler pressure 
 
 -C 
 
 
 
 
 
 150 lbs. 
 
 k 
 
 
 
 
 
 
 C = Triple-expansion, 670 l.H. P., vacuum 26 
 
 
 
 
 
 
 
 inches, boiler pressure 150 lbs. 
 
 O. 
 
 
 
 
 
 
 D = Triple-expansion high speed, 1925 I.H. P., 
 
 X 
 
 
 b 
 
 
 
 
 vacuum 26 inches, boiler pressure 183 lbs. 
 
 mi 
 
 || 
 
 \ 
 
 
 
 
 
 E = Triple-expansion, 3000 I.H. P., high speed, 
 
 ^ 
 
 \ 
 
 
 
 
 
 
 J 
 
 
 
 
 
 
 vacuum 26 inches, boiler pressure 180 lbs. 
 
 
 
 
 
 
 
 F = Triple-expansion slow running, 3000 I.H. P., 
 
 j 
 
 
 
 V* 
 
 
 
 
 
 boiler pressure 199 lbs. 
 
 1 
 
 15 
 
 
 
 
 
 
 
 
 2 
 
 
 sc 
 
 
 
 
 
 ^ 
 
 
 7 
 
 
 
 
 
 
 
 
 
 
 • 
 
 
 
 
 
 
 
 
 
 
 c 
 
 
 
 
 
 ^ \ 
 
 
 
 
 
 
 
 <J 
 
 
 
 
 
 
 
 
 
 
 
 k. 
 
 
 
 
 
 
 
 
 
 
 
 i» 
 
 14 
 
 
 
 
 
 
 
 
 
 
 
 a 
 
 
 
 
 
 
 
 
 
 
 
 > 
 
 
 
 
 
 
 
 
 
 
 
 1- 
 
 
 
 
 
 
 
 
 
 s 
 
 
 ^w _ 
 
 
 
 
 13 
 
 
 
 
 
 
 ^ N 
 
 * 
 
 t. 
 
 
 
 
 
 
 
 
 
 
 
 
 ^^ 
 
 s 
 
 V. 
 
 
 
 
 1 
 
 
 
 
 
 
 
 
 
 
 . s 
 
 V 
 
 
 
 
 i 
 
 
 
 
 
 
 
 
 
 \n 
 
 » > 
 
 t 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 v> 
 
 X 
 
 
 ^3 
 
 
 
 
 
 
 
 
 
 
 
 
 ^^^x- 
 
 V s 
 
 N S 
 
 
 N 
 
 b 
 
 
 
 
 
 
 
 
 
 
 
 
 
 . V 
 
 
 
 % 
 
 
 
 
 
 12 
 
 
 
 
 
 
 
 
 
 
 
 
 % 
 
 j^ 
 
 
 
 
 fj 
 
 
 
 
 
 
 
 
 
 
 
 ^ 
 
 ^^ 
 
 
 
 
 
 10 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 c 
 
 JC <0 *° 80 ,0 ° "0 MC IM ,80 3oo J3C J40 JtO J»0 ' #oo 
 
 Sci foerheof de c rees fob 1 
 Fig. 64 . 
 
172 MANUAL OF MARINE ENGINEERING. 
 
 " Gtiaclaloupe," had the same quantity of heating surface, 13.509 square feet, 
 as the heating and superheating surface together of the other, the " Peron." 
 The temperatures were 378° as before, and only 460° with the superheaters ; 
 the indicated horse-power was practically 6,585 and 6,750 with 88 revolutions, 
 the speed of the " Peron" being on service 16 - 95, against the 166 of the 
 "Guadaloupe." These ships were 430 feet long, and of 6.800 tons gross 
 register. 
 
 Another source of loss is the resistance due to the friction of the steam 
 in passing through the pipes, passages, and valves ; and although here again 
 there is not a total loss, still it is not compensated for in the way that the 
 engineer desires. As friction causes heat, so the friction of the steam along 
 the surface of the pipes and passages generates heat ; but since this heat is 
 not allowed to escape, it is taken up by the steam, and tends to superheat 
 it. The loss due to this cause in the steam-pipes is probably very small, 
 especially when the pipes are of such a size that the velocity of steam through 
 them is not excessive, and Mr. D. K. Clark found that it is inappreciable 
 when the velocity is not more than 130 feet per second with very dry steam, 
 and 100 feet per second with ordinary dry steam. The greatest loss is when 
 the steam has to pass through narrow orifices where the perimeter of such 
 orifices is large compared with the area, as is the case at the steam ports of 
 a cylinder ; this is called " wire-drawing " the steam, and there is always 
 a loss of pressure from this cause, even when the area through which the 
 steam passes is equal to that of the section of the pipe through which it 
 has previously passed. When the area is reduced, and the perimeter is large, 
 the loss is, of course, still greater, and hence the loss at the ports of a cylinder, 
 where the cut-off is early by means of a common slide-valve, is very con- 
 siderable, and may amount to as much as 10 per cent, of the pressure, unless 
 the ports are very large, or the travel of the valve exceptionally long. To 
 obviate such an evil, the double and treble-ported valves are used, and 
 other plans adopted, whereby increased area of opening to steam may be 
 obtained. If care is taken so as to avoid all unnecessary obstructions to the 
 passage of the steam in the pipes and stop-valves, and there be sufficient 
 opening of the port at the beginning of the stroke, the loss of initial pressure 
 should not exceed 2| per cent., and in some marine engines there is no appreci- 
 able fall of pressure from the boiler to the cylinders. There is also a loss 
 of energy when the steam enters the cylinders due to sudden change in 
 velocity, which will be from 150 feet to 15 feet per second in large engines, 
 and even greater in small engines, where the piston velocity is very much 
 less than 10 feet per second. This cannot be avoided in any way, as it is 
 practically impossible to increase the piston velocity to even one-half that 
 of the steam ; and it would be excessively inconvenient to increase the 
 area of ports, etc., so that the velocity of steam should more nearly approach 
 that of the piston. But as the loss from this cause is very slight indeed, 
 no extra cost expended in attempting to avoid it would meet with an adequate 
 return. 
 
 A considerable amount of heat is lost by the radiation from those parts 
 which are alternately exposed to the hot steam and to the atmosphere ; and 
 this was especially great in trunk engines, where the surface of the trunks is 
 very large, and, being hollow, of course have the inner surface giving off heat 
 as well constantly. Fortunately, the surfaces of the piston-rods, trunks, etc., 
 
ANOTHER SOURCE OF LOSS. 173 
 
 soon become highly polished, and so do not radiate the heat so quickly as 
 they would were they rough. This loss, too, cannot be avoided, or even 
 reduced to any appreciable extent. 
 
 Finally, there is the loss due to the heat conducted from the cylinders, 
 pipes, etc., to the other parts of the engine with which they are connected, 
 and which pass it away by radiation at their surfaces. 
 
174 MANUAL OF MARINE ENGINEERING. 
 
 CHAPTER VII. 
 
 ENGINES, SIMPLE AND COMPOUND. 
 
 Elementary Steam-Engine. — The steam-engine, in its most elementary form, 
 has only one cylinder, into which the steam is admitted at each end 
 alternately, so as to move the piston backwards and forwards, and having 
 performed its work is then allowed to escape into the atmosphere. Although 
 from certain causes this was not the form of steam-engine as first invented, 
 it is nevertheless the most simple one, and by taking it as the origin, the 
 genesis of the marine steam-engine can be better explained. As the engines 
 for marine purposes are still largely those having cylinders and pistons, it 
 will be unnecessary in this chapter to deal with any other forms. 
 
 Genesis of the Compound Engine. — The exhaust steam issuing from an 
 engine having a late cut-off and an initial pressure of, say, two atmospheres 
 (or 15 lbs. above atmospheric pressure), would attract the attention of an 
 observant engineer from the force with which it emerges from the exhaust- 
 pipe, and would naturally lead him to inquire how so great a waste of energy 
 might be avoided. It would be clear to him that there was sufficient 
 remaining in it to do useful duty after it had accomplished its work in the 
 cylinder. Being acquainted with the steam-engine of Watt, he would suggest 
 that, instead of allowing it to escape into the atmosphere, it might be con- 
 ducted to the cylinder of a condensing engine, which could work with a 
 steam pressure of one atmosphere, and while operating the piston of this 
 second engine, would cause no more back pressure in the first cylinder than 
 before. Such an arrangement would be a combination of a high-pressure 
 and a condensing engine, and hence it was called a compound engine. 
 
 The engineer to-day would suggest that the steam be taken to a low- 
 pressure turbine, as is often done, and the turbine exhaust to a condenser, 
 whereby a better use would be made of it than the Watt engine did. 
 
 The idea of a compound engine, however, was due to the genius of Hornblower, a 
 Cornish engineer, who in 1781 took out a patent, in which he claimed: — "I use two 
 steam vessels, in which the steam is to act, and which in other steam-engines are called 
 cylinders. I employ the steam, after it has acted in the first vessel, to operate a second 
 time in the other, by permitting it to expand itself, which I do by connecting the 
 vessels together, and forming proper channels and apertures whereby the steam shall, 
 occasionally, go in and out of the said vessels, etc." 
 
 Arthur VVoolf, in his patent taken out in 1804, states that " if the engine be con- 
 structed originally with the intention of adopting my said improvement, it ought to 
 have two steam vessels of different dimensions, according to the temperature or the 
 expansive force determined to be communicated to the steam made use of in working 
 the engine ; for the smaller steam vessel or cylinder must be the measure of the larger. 
 . . . The small cylinder should have a communication, both at its top and bottom, 
 with the boiler which supplies the steam. . . . The top of the small cylinder 
 should have a communication with the bottom of the larger cylinder, and the bottom 
 of the smaller one with the top of the larger, with proper means to open and shut 
 those alternately by cocks or valves, etc., . . . and both top and bottom of the 
 large cylinder should, while the engine is at work, communicate alternately with the 
 condensing vessel"." He proposed to use steam at a pressure of 40 lbs. 
 
EFFECTS OF INCREASE OF PRESSURE. 175 
 
 Expansive Engine. — If the observer, however, happened to be better 
 acquainted with the expansive force of steam than with the use of a turbine 
 and condenser, he would suggest that the steam should be cut off at such 
 an earlier part of the stroke as would ensure its pressure, at emission, being 
 only slightly above that of the atmosphere, and all available energy abstracted. 
 Such an engine would naturally be called expansive, in contradistinction to 
 the elementary engine, working without expansion. Any further attempt at 
 increased expansion would prove less fruitful, as the steam, expanded below 
 the pressure of the atmosphere, will fail to escape into it when opened to 
 exhaust ; besides which, the " back " pressure on the other side of the piston 
 would, during the latter part of the stroke, be greater than the forward. 
 Then it is that by connecting the exhaust pipe to the condenser, in which 
 the pressure would be 10 or 12 lbs. below that of the atmosphere a higher 
 rate of expansion, could be obtained. 
 
 Effects of Increase of Pressure. — If an engine is to work economically, so 
 far as steam is concerned, it has been stated that the terminal pressure, before 
 admission to the condenser, should be as low as possible consistent with good 
 working. 
 
 It is now easy to maintain a vacuum of 28 inches in a modern condenser, 
 but 25 inches was usual with the jet condensers. As some engineers may 
 still prefer to work their engines with only 24 to 25 inches of vacuum (for 
 the sake of obtaining warm feed-water), let it be assumed, for the sake of 
 argument, that 24 inches is the vacuum in the condenser. When the full 
 benefit of expansion is required, the terminal pressure should not exceed 
 7 lbs., which will be 3 lbs. above the back pressure. With a turbine ex- 
 pansion may go on till the difference is less than 1 lb. 
 
 In order to appreciate fully what is encountered in making advances in 
 boiler pressures, it will be well to compare two engines working under the 
 conditions set out above. Suppose these two engines to have each one 
 cylinder of the same diameter and stroke, the boiler pressure of the first 
 to be 2 atmospheres, or 30 lbs. absolute, that of the second 3 atmospheres, 
 or 45 lbs. absolute, the terminal pressure in both cases to be 7 lbs., and the 
 back pressure 4 lbs. The cut-off in the first will be ~ v of the stroke, or a 
 rate of expansion of 4*285, and the mean pressure with an initial of 30 lbs. 
 is 174 lbs. ; deducting 4 lbs. of back pressure, the effective mean pressure 
 will be 134 lbs. In the second engine, the cut-off will be T 7 T of the stroke, 
 and the rate of expansion 6'43, and the mean pressure with an initial of 
 45 lbs. is 20 lbs. ; deducting, as before, 4 lbs. for back pressure, the effective 
 mean pressure is 16 lbs. The effective initial pressures will be 30 — 4, or 
 26 lbs., and 45 — 4, or 41 lbs., respectively. 
 
 Since the cylinders are of the same capacity, and the terminal pressures 
 are the same, each engine consumes the same weight of steam ; but the 
 total heat of evaporation of steam from the temperature corresponding to 
 4 lbs., and at that corresponding to 45 lbs., is 1,130° F., while at that corre- 
 sponding to 30 lbs. it is 1,121° F., there will be, therefore, an expenditure 
 of fuel to obtain the steam at 45 lbs. slightly in excess of that at 30 lbs. As 
 this, however, amounts to less than 1 per cent., it may be neglected, and 
 the cost of the steam assumed to be the same in both cases. It will be seen 
 then that, with this advance of boiler pressure, there is an advance in mean 
 pressure, and the gain in power amounts to nearly 20 per cent. ; but the 
 
176 
 
 MANUAL OF MARINE ENGINEERING. 
 
 initial load on the piston of this more economic engine is 57 per cent, higher 
 than that on the more wasteful one, and consequently the rods, framing, 
 etc., must be 57 per cent, stronger ; moreover, the shaft will be increased 
 in size, and the cylinder and passages must be stronger. Altogether, then. 
 the engine will become heavier and more costly, as the boiler pressure is 
 increased, while the expenditure of fuel is less per horse-power. 
 
 To render the comparison strictly fair, it would be necessary to take two 
 engines of equal power, so that if the stroke of the pistons is the same, their 
 areas will be inversely proportioned to the mean pressures, and, conse- 
 
 1JV4 
 16 
 
 of 41, or 34'5 lbs. 
 
 quently, the initial loads will now be 26 lbs. and 
 which gives an excess of 31*3 per cent. 
 
 Progress Made by Early Marine Engineers. — The early marine engineers, 
 however, did not advance on these lines as a rule, for nearly the same rate 
 of expansion was observed at full speed with steam of three atmospheres 
 as had obtained with steam of atmospheric pressure ; consequently very 
 little benefit was derived in economy, compared with what might have 
 been the case had they done differently. The chief result accruing from 
 the increased boiler pressure was in practice the larger indicated horse-power 
 developed by engines of the same size as formerly, partly due to the aug- 
 mentation of mean pressures, and partly to the increased piston speed resulting 
 from them. 
 
 Engines of as much as 200 N.H.P., working with steam of 60 lbs. pressure, 
 were supplied to the Navy by Messrs. Penn and Messrs. Maudslay as early 
 as 1853 ; but a period of more than fiiteen years elapsed before the Admiralty 
 again employed such pressure in any larger ship than a gun-boat. 
 
 By the following table a comparison can be made of four typical simple 
 engines working with steam at different pressures, and their performance 
 under the varying conditions : — 
 
 TABLE XXIX. 
 
 Load on safet}' valves, 
 
 lbs. 
 
 30 
 
 45 
 
 60 
 
 60 
 
 Diameter of cylinder, 
 
 . ins. 
 
 50 
 
 50-8 
 
 511 
 
 38 
 
 Initial absolute pressure, 
 
 lbs. 
 
 45 
 
 60 
 
 75 
 
 75 
 
 Cut-off, 
 
 . stroke 
 
 6 
 
 Tff 
 
 tV 
 
 A 
 
 « 
 
 TO 
 
 Back pressure, 
 
 lbs. 
 
 4 
 
 4 
 
 4 
 
 4 
 
 Mean effective pressure, . 
 
 • >> 
 
 3677 
 
 35 66 
 
 3515 
 
 63-95 
 
 Terminal ,, 
 
 • >i 
 
 27-00 
 
 18 00 
 
 15 00 
 
 45-00 
 
 Maximum load on piston, 
 
 • »! 
 
 80.4S3 
 
 113,456 
 
 145,550 
 
 80,514 
 
 Mean ,, ,, 
 
 • »J 
 
 72, ISO 
 
 72,180 
 
 72,180 
 
 72,180 
 
 Ratio of max. to mean, . 
 
 • 11 
 
 1115 
 
 1572 
 
 2016 
 
 1-115 
 
 Weight of steam used, . . 
 
 • !> 
 
 893 
 
 609 
 
 515 
 
 S68 
 
 Examples (1), (2), and (3) show conclusively that by increasing the boiler 
 pressure and rate of expansion, so as to obtain nearly the same mean pressure, 
 the weight of steam used is considerably diminished and the terminal pressure 
 considerably reduced, but that there is' an increase in the maximum load 
 on the piston proportionate to the absolute pressure in the boiler, so that 
 the ratio of maximum to mean pressure of example (3) exceeds that of example 
 (1) by more than 80 per cent. 
 
RECEIVER COMPOUND ENGINE. 177 
 
 Although such an increase in weight of machinery as would be neces- 
 sitated by so great an increase in load, may not be of much importance in 
 some ships, in others it would be prohibitive ; for when the power required 
 for certain speeds of ship becomes large compared with the displacement, 
 it requires the utmost care in design to keep down the weight, so as to admit 
 of the engines being carried by the ship on the required draught of water. 
 For this reason, in actual practice, it was found advisable to use steam of 
 under 50 lbs. pressure (above the atmosphere) in very fast river steamers. 
 or even in high-powered steamers for Channel service of moderate size, on 
 account of the limited speed of piston obtainable with the paddle-wheel, 
 until, by means of forced draught, constructing the ship and engines as 
 much as possible of steel, and a special design of light compound engines, 
 pressures of 100 to 125 lbs. could be employed in such ships ; now in some 
 few cases triple-expansion engines using steam of 175 lbs. pressure are fitted 
 in paddle steamers of bigh speed. 
 
 Example (4) is given that the effect of two widely different boiler pressures 
 may be compared, when the rate of expansion is the same. The mean pressure 
 is, of course, much higher, and, but for the back pressure being constant, 
 would bear the same proportion to that at the boiler pressure of 30 lbs. as 
 the initial absolute pressures — viz., 5 to 3. The weight of steam used is 
 very little less than that of 30 lbs. pressure, and, owing to the reduction in 
 the size of the piston, the maximum load in this case is practically the same 
 as in example (1). The pressure at exhaust is exceedingly high, being 45 lbs. 
 absolute, or 30 lbs. above that of the atmosphere, so that it is capable of 
 doing considerably more work, if admitted into another cylinder of larger 
 size, than that of the first ; and even if admitted into one of the same 
 size (provided it finally exhausts into a condenser), more work will be 
 obtained from the steam than if it is allowed simply to escape into the 
 atmosphere. 
 
 Receiver Compound Engine. — Now, suppose that an engine working 
 under the conditions set out in example (4) (so far as pressure and cut-oft 
 are concerned) exhausts into a steam-tight space, so that there is back pres- 
 sure in front of the piston equal to the pressure in this receiver of the exhaust 
 steam ; and suppose, further, that the steam is taken away by anothei 
 cylinder from the receiver at the same rate as it is supplied by the cylinder, 
 there will then be a constant mean pressure maintained there. For the sake 
 of fixing the application to example (4), suppose, again, the pressure in its 
 reoeiver to be 30 lbs. absolute, then the mean pressure in the cylinder will be 
 67 # 95 — 30 = 37'95 lbs. only. Now, suppose a second and larger cylinder 
 to be supplied with steam from the receiver at such a rate that there is no 
 change of pressure in it (this being accomplished by so arranging the cut-off 
 in the second cylinder, that the weight of steam taken by it equals the weight 
 of steam exhausted from the first one) ; the cut-off may be determined from 
 the formula p v = const. ; so that if V be the volume of the second cylinder 
 and v that of the first, 45 lbs. the terminal pressure in the small and 30 lbs. 
 the initial in the large cylinder. Then, cut-off in the second cylinder = 
 
 — -^ = o X ^, and if the ratio of V to v is 3, the cut-off in the second cylinder 
 oU v ^ v 
 
 is h stroke, and the rate of expansion in it 2. 
 
 The mean pressure, with an initial pressure of 30 lbs. and a rate of 
 
178 MANUAL OF MARINE ENGINEERING. 
 
 expansion 2, is 25*38 lbs. ; and allowing for a back pressure of 4 lbs., the 
 mean effective pressure in the second cylinder is 21*38 lbs. 
 
 Since the area of the second piston is three times that of the first, the 
 work done in the second cylinder is equivalent to what might be done by 
 one of the same area as the first, with a mean effective pressure three times 
 as great, or 64*14 lbs. per square inch. It will be seen from this that the 
 total work done by the combined cylinders is the same as would be done by 
 the original cylinder with a pressure of 37*95 + 64*14, or 102*09 lbs. per 
 square inch ; hence we find that there is a gain of nearly 60 per cent, by 
 the introduction of the second cylinder. So far, the compound engine would 
 be undoubtedly more economical than the expansive engine, as exemplified 
 in examples (1), (2), and (4), but less so in this particular instance than 
 example (3). 
 
 Expansive and Compound Engines Compared. — To examine the relative 
 
 economy of a compound and of a simple expansive, engine, it is necessary 
 
 that they should both work with the same boiler pressure and the same 
 
 rate of expansion. Now examples (3) and (4) satisfied the first condition, 
 
 and if the second is also satisfied, then they may be compared. The rate of 
 
 expansion in example (4) is 1*666, and since the volume of steam in the 
 
 second cylinder at the end of its stroke will be three times that in the first 
 
 at the same period, the total expansion effected by both cylinders will be 
 
 3 x 1*666, or five times. The cut-off in example (3) was two-tenths the 
 
 stroke, and therefore its rate of expansion is five ; so that these two examples 
 
 may be compared as to the efficiency of the steam. The effective pressure 
 
 of the compound system may be referred to the large cylinder, in the same 
 
 way in which it was referred to the small one, and will be that Actually on 
 
 the large cylinder, together with that on the small one divided by the ratio 
 
 of their capacities ; hence, effective mean pressure referred to the large 
 
 / 37*95\ 
 cylinder is (21*38 -1 5 — ), or 34*03 lbs. per square inch. It will be seen 
 
 that this is 1*12 lbs. less than that obtained in the simple expansive engine, 
 and therefore a loss has occurred somewhere in the compound system. 
 
 Suppose, now, that the cut-off in the large cylinder is so altered that 
 the pressure in the receiver is 45 lbs., so that it receives steam at the same 
 pressure as that which exhausts from the small one ; in this case there will 
 be no " drop " in the pressure from commencement of exhaust to the end in 
 the small cylinder. 
 
 The mean effective pressure in the small cylinder is now 67*95 — 45, or 
 22*95 lbs. per square inch. 
 
 4-0 *w 
 The cut-off in the large cylinder = j^^n, or £ the stroke, which gives 
 
 3 as the rate of expansion. 
 
 With an initial pressure of 45 lbs. and rate of expansion 3, the mean 
 pressure is 31*5 lbs. ; allowing 4 lbs. for back pressure, the effective mean 
 pressure is 27*5 lbs. 
 
 Referred to the large cylinder the effective mean pressure of the system 
 
 22*95 
 is now 27*5 -\ 5 — , or 35*15 lbs. on the square inch, or exactly the same as 
 
 that obtained in the simple expansive engine. 
 
 Effect of " Drop " in the Receiver.— It is seen from the above, then, when 
 
DIRECT EXPANSION COMPOUND ENGINE. 179 
 
 no " drop " occurs there is no loss of efficiency ; but that when the pressure 
 in the receiver is less than the terminal pressure in the small cylinder, there 
 is somehow a loss of effective mean pressure. This arises from the steam 
 being allowed to expand from the small cylinder into the receiver without 
 doing work. But it is known that, when this takes place, the steam becomes 
 somewhat superheated ; for, inasmuch as the loss of pressure has occurred 
 without conversion into external work or loss of heat in any other way, it 
 must appear in some other form. Although this loss is not wholly recovered, 
 it must be to some extent reduced by the benefit which the steam derives 
 from the superheating in expanding in the large cylinder. 
 
 Division of the Work. — It will be also seen that, as the ratio of the 
 cylinders' capacity is 3, and the effective mean pressures 22*95 lbs. and 
 27 '5 lbs., the work done in the small cylinder to that in the large is as 22*95 
 to 27*5 x 3, or nearly 1 to 3*6 ; while in the former case it was as 37*95 to 
 21*38 x 3, or nearly 1 to 1*7. 
 
 Therefore, with an earlier cut-off in the large cylinder, more work is 
 developed in it than is the case when with a later cut-off ; moreover, with 
 this ratio of cylinders, in order to get the highest efficiency of the steam, 
 
 the ratio of the work done is as 1 to 3*6 ; and the initial pressure on the 
 
 75 4.5 
 
 large piston is 41, and on the small ^— — , or 10, as against 71 on the 
 
 expansion engine of equal size ; and even if the compound engine were 
 arranged with one cylinder above the other, the combined initial pressure 
 would be 41 -f- 10, or 51, as against 71 of the simple expansive. 
 
 Direct Expansion Compound Engine. — The compound engine may, how- 
 ever, work without any intermediate receiver, if the pistons are arranged 
 to move simultaneously, either in the same or opposite directions. To 
 consider this case — suppose the cylinders to be side by side, and the pistons 
 to move in opposite directions, as originally proposed by Woolf, so that 
 when the small piston has receded one-tenth of the stroke, the large one 
 has advanced by exactly the same amount, and the space between them 
 is 0*9 v + 0*1 V ; the volume of steam at commencement of exhaust is 
 v, and the pressure at that period, as before, 45 lbs. ; the volume at any 
 
 point of the stroke (va), or the space between the pistons 
 
 n ,, 10 — n n 
 
 = io v + -Io- t; = io (V - v)+?; ' 
 
 and since V = 3 v, space between the piston at n-tenths of the stroke 
 
 The pressure at this point = 45 -*- f^+lj. The pressures at every 
 
 tenth of the stroke from to 10 will be 45, 37*5, 32*14, 28*12, 25*0, 22*5, 
 20*45, 18*75, 17*3, 16*07, 15*0 ; the mean of which is 24*78 lbs. per square 
 inch. Deduct this from 67*95, and the effective mean pressure in the small 
 cylinder is 43*17 ; deduct 4 lbs. from 24*78 lbs., and the effective mean 
 pressure in the large cylinder is 20*78 lbs. 
 
 The effective mean pressure of this system, referred to the large cylinder, 
 
180 MANUAL OF MARINE ENGINEERING. 
 
 is now (20"78 -| o~)> or 35*17 lbs. per square inch, which is the same as 
 
 that obtained by direct expansion in one cylinder — example (3). 
 
 It will be observed, however, that the ratio of the power exerted by the 
 small cylinder to that exerted by the large one. is as 43" 17 to 3 X 20*78, or 
 1 to 144, being nearer an equal distribution of the work than in either of 
 the cases of the intermediate receiver engine. This latter result is caused 
 principally by the decreased back pressure in the small cylinder. 
 
 In actual practice there ar3 certain causes which materially modify the 
 results shown by both of these forms of compound engine, as illustrated in 
 the foregoing. 
 
 In the receiver compound engine, the pressure in the receiver is not 
 constant, because of its limited size ; the difference in the periods of exhaust 
 and admission, and the cushioning after cut-off in the L.P. cylinder by the 
 small piston cause considerable oscillation. 
 
 The direct expansion engine is only nominally without a receiver, as 
 the space between the small piston and the large one is often necessarily 
 considerable (v. fig. 75), from the size of the communication pipes and the 
 valve-box of the large cylinder. The valve of the L.P. cylinder cuts off 
 some time before the small one ceases to exhaust, causing cushioning in the 
 latter and in the spaces ; the small cylinder also commences to exhaust 
 before the large one can take steam. 
 
 Direct expansion compound cylinders have, however, been arranged so 
 that one cylinder communicates directly with the other, without any inter- 
 vening space, by placing the cylinders side by side, and causing the pistons 
 to operate on cranks set opposite one another (v. fig. 187). But such engines 
 have, for other reasons, proved generally unsuitable for propelling. 
 
 Requisites in the Marine Engine. — A marine engine must be (1) when 
 required readily started, stopped, and reversed ; (2) it should have a turning 
 moment or torsion as nearly uniform as possible ; (3) it must be able to work 
 continuously for long periods without stoppage from any cause ; and (1) 
 and lastly, it must be economical. 
 
 The first condition is a sine tjud non, and is generally fulfilled by having 
 two or more cylinders operating on cranks at suitable angles. 
 
 The second condition is absolutely necessary when weight is a serious 
 consideration, and is very fairly satisfied by the two or more cranks at proper 
 angles, and the cylinders so designed as to divide the work nearly equally 
 between them. 
 
 The third condition will depend very much on the variation in stress, so 
 that the engine with small initial pressure in each cylinder compared with 
 mean pressure, is more likely to fulfil it than one with large initial pressures. 
 
 The fourth condition, which is of the utmost importance to the ship- 
 owner, is very well complied with by the triple- and quadruple-compound 
 engines in all sizes, and by the turbine and combined turbine and reciprocator 
 in large sizes. 
 
 With two cylinders and cranks at right angles, there must inevitably 
 be some amount of "' drop," if the work is to be evenly divided when the 
 power is as great as usually developed by a marine engine at service speed. 
 The crank of the H.P. cylinder should lead — that is, should be in advance 
 of the other crank by 90°, or such other angle as it is deemed best to set 
 
THEORETICAL EFFICIENCY OF VARIOUS MARINE ENGINES. 181 
 
 :he cranks at. When this is the case, the small cylinder begins to exhaust 
 just alter the crank of the L.P. cylinder has got well over the centre, and 
 tends to maintain a constant pressure on the large piston through the earlier 
 portion of its stroke, and at cut-off the pressure in the receiver is not much 
 below its average pressure. If, on the other hand, the crank of the large 
 cylinder leads, exhaust takes place only a little before cut-off in the large 
 cylinder, and causes a hump in the indicator-diagram, showing an increase 
 in the amount of " drop," and that with no diminishing in the mean back 
 pressure in the small cylinder. Engines having the low-pressure engine 
 crank as the leading one were also generally unhandy. 
 
 The first triple-compound engines were, as a rule, designed so that the 
 high-pressure crank " led," or was in advance of the medium-pressure crank 
 by 120°, and the medium-pressure crank 120° in advance of the low-pressure 
 irank ; but in modern practice it is usual to fit the low-pressure crank in 
 advance of the medium-pressure crank, etc., as in this way there is less vari- 
 ation in temperature in each cylinder, although the load on the pistons is less 
 during the first half of the stroke and greater during the second half than is 
 the case when the high-pressure crank leads, and the engine is not appreciably 
 less *' handy " (v. fig. 164). 
 
 The first cylinder of a compound system is called the " high " pressure, 
 and the last the " low," from their association with the condensing and 
 non-condensing engine. For convenience in speaking of them, they are 
 designated by the initials H.P. and L.P. Hitherto, in the chapter, all com- 
 parisons of the compound with the simple expansive engine have been made 
 on the supposition that the expansive engine has only one cylinder of the 
 same capacity as the low-pressure one of the compound system. To render 
 the comparison perfectly fair, it will be necessary to take such cases as may 
 be found in actual practice. In triple-expansion engines the middle cylinder 
 is called the " medium-pressure," and sometimes the " intermediate," and 
 designated by the initials M.P. ; and in quadruple engines the third cylinder 
 is called the 2nd M.P. 
 
 Comparative Theoretical Efficiency of Various Marine Engines.— (1) A 
 single-ciflinder expansive engine : rate of expansion, 5 ; initial pressure, 80 lbs. ; 
 absolute back pressure, 4 ; area of piston, A. 
 
 Mean pressure . . = 41*76 lbs. 
 
 Effective mean pressure . = (41*76 — 4) = 37*76 lbs. 
 
 Effective initial load on piston = (80 — 4) A = 76 A lbs. 
 
 Efficiency of the system . =1*00. 
 
 (2) A single-crank tandem compound engine : rate of expansion, 5 ; initial 
 pressure, 80 lbs. absolute ; area, L.P. piston, A ; ratio of cylinders, R, 
 generally in practice, 4. 
 
 Effective mean pressure referred to L.P. piston . — (41*76 — 4) 
 
 = 37*76 lbs. 
 
 R 
 
 Terminal pressure in H.P., and initial pressure in L.P. = -= X 80 = 64 lbs. 
 
 o 
 
 A 
 
 Effective initial load on H.P. piston . . . = (80 — 64) - = 4 A. 
 
 L.P. „ ... =(64 -4) A = 60 A. 
 
 Efficiency of the system = 1*00. 
 
182 MANUAL OF MARINE ENGINEERING. 
 
 TotaJ load on crank is, therefore. 64 A, against 76 A with the single- 
 cylinder expansive engine. 
 
 As in actual practice there is invariably a drop, which will amount to as 
 much as 10 lbs. in an engine of this kind, the initial pressure in the low- 
 pressure cylinders being decreased by that amount, and that of the high- 
 pressure increased. So that, actually, the total loads will be as 56*5 to 76, 
 or a saving in the compound engine of over 25 per cent, of the load put on 
 the rods, framing, etc. ; and also enabling a large reduction to be made in 
 the diameter of the shafting. The engine will work much more steadily, 
 owing to the ratio of maximum to mean pressure being so largely reduced '. 
 and the handiness very much increased from the cut-off in the high-pressure 
 cylinder being so late as y^ the stroke. The friction of the two cylinders, 
 etc., is, however, considerably greater than that of the one, but this is set off 
 by the reduction in friction on the guides and journals ; and the friction on 
 the valve of the single cylinder exposed to high-pressure steam will be 
 more than the combined friction of the two valves, the small one of 
 which only is so exposed, while the expansion valve (which is necessary 
 to the single cylinder for so early a cut-off) will also increase its loss from 
 friction. 
 
 The compound engine compares more favourably with the simple expan- 
 sive when both have two cylinders and two cranks. Then each engine has 
 the same number of working parts of necessity, and the simple expansive, 
 besides having the usual slide valves, each of which is exposed to the boiler 
 pressure, has. an expansion valve to each cylinder, in order to effect so early 
 a cut-off. The following examples will show the results of the two systems 
 under the same circumstances : — 
 
 (3) A simple expansive engine having two cylinders, each of whose 
 
 A 
 
 pistons has an area of -^ inches ; rate of expansion, 5 ; initial pressure, 
 
 80 lbs. 
 
 Mean pressure . . . = 41*76 lbs. 
 
 Effective mean pressure . . = 41'76 — 4 = 37'76 lbs. 
 
 Effective initial load on piston . = (80 — 4) -^ = 38 A lbs. 
 
 •a 
 
 Effective mean load on both pistons = 37*76 x A lbs. 
 Efficiency of the system . . = TOO. 
 
 (4) A compound engine having two cylinders, the ratio of whose piston 
 area is 3, and the area of low-pressure piston, A ; rate of expansion, 5 ; 
 initial pressure, 80 lbs. ; pressure in receiver, 21 lbs. 
 
 The cut-off in high-pressure cylinder to effect this rate of expansion = i 
 or 0*6 stroke. 
 
 The cut-off in low-pressure cylinder to maintain 21 lbs. pressure in the 
 
 receiver = -^ 5- = 0*76 the stroke. 
 
 21 x 3 
 
THEORETICAL EFFICIENCY OF VARIOUS MARINE ENGINES. 
 
 183 
 
 Effective mean pressure in H.P. cylinder 
 
 T P 
 
 )J »J .Li. A . , r 
 
 Effective initial load on H.P. piston . 
 
 T P 
 
 >> »> AJ.A . ,, 
 
 Effective mean load on both pistons . 
 Efficiency of the system . 
 
 72-48 - 21 = 51-48 lbs. 
 = 20-32 - 4 = 16-32 lbs. 
 
 = (80 -21)g= 17-3 x A lbs. 
 
 = (21 - 4) A = 17-0 X A lbs. 
 
 = 51-48 X I" + 16-32 x A 
 
 = 33-48 x A lbs. 
 
 = 0-887. 
 
 It will be seen that the initial load on each of the pistons of the com- 
 pound engine is very nearly equal in this case, and is less than half that on 
 each piston of the expansive engine. The work done is very nearly equally 
 divided between the two cylinders, but falls short of that done by the expan- 
 sive engine by more than 11 per cent. 
 
 The compound engine, therefore, is nominally not so economic in steam, 
 but is subject to much lighter loads, and to less variation in load and tem- 
 perature. 
 
 To test the merits of the compound system carefully devised experi- 
 ments were made by the British Admiralty, and by the Government of the 
 United States, which show that, although the difference between the 
 coal consumed per I. H.P. in the two systems was not very great, the com- 
 pound engine, on the whole, is more economical. The best known of these 
 experiments was the trial between the gunboats " Swinger " and " Goshawk," 
 the latter having compound engines, with cylinders 28 inches and 48 inches 
 diameter and 18-inch stroke, the former expansive engines, having two 
 cylinders, 34 inches diameter and 22-inch stroke, and those of the sister ships 
 given below. In these and in many other cases it was demonstrated that 
 while on trial trips the coal and water consumptions were always less with the 
 compound engines, it was really the results of prolonged trials on service that 
 proved conclusively the real and practical superiority of the compound 
 system, which was not alone in consumption of fuel per I.H.P., but in 
 mechanical efficiency, whereby the consumption per voyage was marked, 
 and the reduction in wear and tear even more so. These arguments were, 
 of course, convincing to the shipowner. 
 
 Diameter of cylinders and stroke. 
 
 "Sheldrake." 
 
 "Moorhen.' 
 
 "Mallard.'' 
 
 
 
 
 
 Exp., 2 cyls., 
 
 Exp., 2 cyls., 
 
 Comp. 
 
 
 34" x 21". 
 
 34" x 21". 
 
 31" - 48" x IS". 
 
 1 
 
 Boiler pressure, lbs., 
 
 62-35 
 
 53-0 
 
 63 5 
 
 41 
 
 58-4 
 
 59 
 
 Vacuum, . - ins., 
 
 20 
 
 24 
 
 23-3 
 
 24-4 
 
 23-75 
 
 25'1 
 
 Revolutions per minute, .... 
 
 115-5 
 
 84-7 
 
 121-3 
 
 92-9 
 
 124 8 
 
 98-8 
 
 Speed of piston, - - feet per minute, 
 
 404 
 
 295 
 
 424 
 
 324 
 
 374 
 
 296 
 
 Indicated horse-power, .... 
 
 367 
 
 137 
 
 387 
 
 180 
 
 398 
 
 213 
 
 Speed of ship, - - . - knots, 
 
 9 251 
 
 7 232 
 
 9-634 
 
 7-899 
 
 9-894 
 
 8413 
 
 Water consumed per I. H.P. per hour, 
 
 21-5 
 
 25-4 
 
 201 
 
 24-6 
 
 1712 
 
 1766 f 
 
184 MANUAL OF MARINE ENGINEERING. 
 
 Further Comparison of Efficiency of Engines. — The compound engine with 
 three cranks and having two low-pressure cylinders and one high, possesses 
 advantages beyond those of the two-cylinder two-crank arrangement. It 
 was, no doubt, first chosen principally to avoid excessive diameter of low- 
 pressure cylinders for very large power, and next because of the uniformity 
 of the twisting moment on the shaft with the three cranks at angles of 120° 
 apart. But, further, in this engine the work can be fairly equally divided 
 between the cylinders without those disadvantages previously shown to 
 exist in the two-cylinder engine working under this condition. To divide 
 the work equally, only one-third will be allotted to the high-pressure cylinder, 
 and one-third to each of the two low-pressure cylinders, and this can be 
 done by maintaining a considerably higher pressure in the receiver than 
 obtains in the two-cylinder arrangement. The " drop," therefore, is con- 
 siderably less ; and since each low-pressure piston has only half the area 
 of th^t of the two-cylinder engine, the initial loads under these conditions 
 are not abnormally large. The increase in receiver pressure reduces the 
 initial load on the high-pressure piston, and hence its diameter may be 
 increased, so as'to get increased expansion in it and decrease in "drop" 
 without increasing the initial load beyond that on the high-pressure piston of 
 a two-cylinder engine of equal power. 
 
 (5) An engine having two low-pressure and one high-pressure cylinders 
 working on three cranks. — To fully appreciate these advantages, suppose 
 such a three-cylinder engine to be working under the same circumstances 
 as that of example (4), page 182, so that the area of each low-pressure piston 
 
 A A 
 
 is ^r, and of the high-pressure piston », the rate of expansion 5, and the 
 
 initial pressure 80 lbs. absolute. 
 
 As in the previous example the cut-off is 0'6 the stroke. Suppose the 
 pressure in the receiver to be 32 lbs. absolute, 
 
 DA V O'fi 
 
 Then the cut-off in L.P. cylinders = -^ ~- = 056 the stroke. 
 
 The effective mean pressure in the H.P. ) _ /70-48 — 321 -- 40"48 lbs 
 
 cylinder, . . . . . J 
 
 The effective mean pressure in each L.P. ) _ /oq-iq __ a\ — 24-19 lb- 
 
 cvlinder . . . . . J 
 
 \ 
 Effective initial load on the H.P. piston = (80 — 32) ~ = 16 X A lbs. 
 
 each L.P. „ = (32 - 4) ~ = 14 X A lbs. 
 
 The effective mean load on all three f = 40-48 X g + 2 x 24'19 X g 
 pistons .... 
 
 A , „ _ .„ A 
 
 W48 X 
 ( = 37-68 x A lbs. 
 
 Efficiency of the system . . = 0'998. 
 
 It is seen by this that the initial load on the high-pressure piston is 74 per 
 cent., and that on each low-pressure piston 17A_ per cent., less than the corre- 
 sponding loads of the two-cylinder engine of the same size ; that the gain of 
 efficiency of the steam is 124 per cent., if no allowance is made for possible 
 superheating of the steam on expanding into the receiver ; the " drop " in 
 
TRIPLE-EXPANSION COMPOUND ENGINE. 185 
 
 this case is (54 — 32) or 22 lbs., as against 33 lbs. in the two-cylinder engine. 
 It is seen, then, that theoretically this engine is nearly equal in steam effi- 
 ciency, both to the simple expansive, and to the compound direct-expansion 
 engine. On the other hand, the friction of three engines must be set against this, 
 besides the extra cost of manufacture and the space occupied by machinery. 
 
 Experiments in the Mercantile Marine with engines having cylinders with 
 ratios better suited to the steam pressure and other conditions demonstrated 
 much more emphatically the superiority of the compound engine in economy 
 of fuel and, perhaps, more decisively the other advantage of that system, 
 for it was soon found that the wear and tear and oil consumption of expansive 
 engines having a working pressure of 50 to 60 lbs. The two-cylinder and 
 three-cylinder receiver type of compound engine soon became popular, and 
 remained so until steel boilers could be made for such high pressures as 
 to require an extension of the compound system. 
 
 Triple-Expansion Compound Engine.— The compound engine having one 
 high-pressure, and one low-pressure, with a medium-pressure cylinder, is the 
 one most commonly found in use, if steam of over 120 lbs. pressure is to 
 be used economically. To ensure economy, the steam must be expanded 
 down to about 10 lbs. absolute, the initial loads on the pistons moderate, 
 and the " drops " not excessive. The low-pressure cylinder may be rather 
 smaller in size than that of the ordinary iit'o-cylinder compound arrangement, 
 due to the increased efficiency of the steam from the high rate of expansion, 
 and the greater referred mean pressure. 
 
 (6) For example : — To determine the particulars of a triple-compound 
 engine on this system to be equal to that set out in example (4), page 182, 
 the initial pressure being 127 lbs., and the rate of expansion 10. 
 
 The mean pressure in a single cylinder, with a cut-off at T ^ the stroke, is 
 41*91 lbs. ; deducting from this 4 lbs. for back pressure, the mean effective 
 pressure is 37'91 lbs., or nearly the same as that of example (5), page 184. 
 Suppose the cut-off in the high-pressure cylinder is 0"6 the stroke, then the 
 ratio of the high-pressure to low-pressure cylinder must be 6 to effect a rate 
 of an expansion of 10. The ratio of the medium-pressure cylinder to high- 
 pressure cylinder may be taken as f, and, consequently, the ratio of low- 
 pressure to mean-pressure cylinder is \J\ The pressure between the 
 high-pressure and mean-pressure cylinders is to be 50 lbs., the cut-off in the 
 mean-pressure cylinder will, therefore, be 0*61 the stroke. The pressure 
 between the low-pressure and mean-pressure cylinders is to be 21 lbs. ; the 
 cut-off in low-pressure cylinder must, therefore, be 0*6 the stroke. 
 
 The effective mean pressure in H.P. cylinder = (115 — 50) = 65 lbs. 
 
 MP. „ = (45-4-21) = 24-4 lbs. 
 
 L.P. „ = (19 — 4 = 15-0 lbs. 
 
 \ 
 The effective initial load on H.P. piston . = (127 — 50) '-. = 12'8S A lbs. 
 
 M.P. „ . = (50-21) A ^ = 12-1 A lbs. 
 
 L.P. „ . = (21 -4) A = "17-0 A lbs. 
 
 The effective mean load on all three J = (65 x g ) + <£4"4 x £ A) -f 15 A 
 
 pistons ) . ... 
 
 r I = 36 X A lbs. 
 
 Efficiency of the system . = ^rx = 0"949. 
 
lfJB MANUAL OF MARINE ENGINEERING. 
 
 It will be seen that in this case, owing to " drop," there is a loss of nearly 
 2 lbs., but the work is fairly divided between the three cylinders, and the 
 initial loads are by no means' high ; the drop from the high-pressure cylinder 
 is only 26 lbs., and that from the mean-pressure cylinder 9| lbs. This engine 
 then effects an expansion of 10, and is, therefore, a very economic one ; it 
 will have a very regular motion, and share in all the benefits of a three-crank 
 engine ; and the stresses on the working parts will be very moderate. 
 
 (7) To see how far this is true, it is only necessary to compare these results 
 with those from a three-crank engine having one high- and two low- pressure 
 
 cylinders, each L.P. having a piston area of -• 
 
 A 
 Suppose the high-pressure cylinder to have a piston area of ^, the cut-off 
 
 in the high-pressure cylinder to effect an expansion of 10, must be 0'4 of the 
 stroke ; the receiver pressure will be 42 lbs., and the cut-off in low-pressure 
 cylinder 03 of the stroke. 
 
 The effective mean pressure \ ._ ^ _ ^\ = 55 lbs 
 in the H.P. cylinder .J 
 
 The effective mean pressure ) _ ^y.y _ ^\ _ 23-7 lb« 
 in each L.P. cylinder . / 
 
 The effective initial load on { = (12 - _ ^ A = ^.^ x A ^ 
 the H.P. piston . . J 4 
 
 The effective initial load on ) = „ 2 _ 4) A 19 . Q x A lbg 
 each L.P. piston . . J v ' 2 
 
 The effective mean load on \ = fa A\ /^ x A\ = ^.^ ^ Albg 
 all three pistons . -J V 4/ \ 2/ 
 
 The initial load on the high-pressure piston is here 65 per cent, larger than in 
 the preceding case, and that on each low-pressure piston 56 per cent, larger 
 than on the mean-pressure piston, and llf per cent, above that on the low- 
 pressure piston ; but the efficiency of the steam is somewhat higher in the 
 latter case, the drop from the high-pressure cylinder being only 8'8 lbs. The 
 ratio of maximum pressure to mean in the high-pressure cylinder is 1*54, and 
 in the low-pressure cylinder P6, which are about the same as those of the old 
 expansive engine working with a boiler pressure of 45 lbs., and cutting-off at 
 - 3 of the stroke. It may not be always advisable to set the cranks at angles 
 of 120° in a three-crank compound engine ; their precise position should 
 depend on the power developed in each cylinder, and the relative twisting 
 efforts at any period. 
 
 The success of the triple-expansion engine is now so well assured, and 
 all doubts as to its efficiency and good working are so effectually dispelled 
 as to require no further discussion. It does not differ in any essential feature 
 from the ordinary compound engine, and its success was in no small measure 
 assured by the fact that most makers of the new type departed as little as 
 possible from their previous practice in its general construction. A few 
 years' experience demonstrated that the triple-expansion engine is more 
 economical than the ordinary compound engine ; that the wear and tear 
 
INCREASED PRESSURE OF STEAM. 187 
 
 is no more, but rather less, when three cranks are employed, than with the 
 two of the ordinary compound ; and that boilers of the common marine 
 design could be made to work satisfactorily at a pressure of 150 lbs. per 
 square inch, and even higher, while, with ordinary care, their durability and 
 good continued working are not less than those of similar boilers pressed 
 to 60 lbs. per square inch under similar circumstances. Speaking generally, 
 the consumption of fuel is 20 per cent, less with a triple-expansion engine 
 than with an ordinary compound engine working under similar circumstances. 
 That is, a triple-expansion engine, supplied with steam at 150 lbs. pressure, 
 uses 20 per cent, less weight of water per I.H.P. than an ordinary compound 
 engine supplied with steam at, say, 90 lbs. pressure, both engines being 
 equally well-designed, manufactured, and attended to. Also, that a triple- 
 expansion engine is more economical than an ordinary compound engine when 
 both are supplied with steam at the same pressure, for all pressures of 95 lbs. 
 and upwards, and especially so in the case of large engines. Hence, it may 
 be taken that the superior economy of the triple-expansion engine is due 
 to two causes, viz. : — (1) To the superior pressure of steam used and the 
 higher rate of expansion thereby possible ; and (2) the system whereby large 
 initial loads and large variations of temperature in the cylinders and large 
 " drop " in the receivers are avoided. 
 
 Increased Pressure of Steam is obtained by a very slight increase of con- 
 sumption of fuel, and the efficiency of steam rapidly increases with increased 
 pressure ; hence, steam of high pressure is more economical than that at 
 inferior pressures. For example : — 
 
 • (i.) The total heat of evaporation of 1 lb. from 100° and at 274° F. (corre- 
 sponding to 45 lbs. pressure absolute) is 1,097 thermal units. 
 
 (ii.) From 100° and at 320° F. (corresponding to 90 lbs. absolute) is 1,110 
 thermal units. 
 
 (iii.) From 100° and at 353° F. (corresponding to 140 lbs. absolute) is 
 1,120 thermal units. 
 
 (iv.) From 100° and at 377° F. (corresponding to 190 lbs. absolute) is 
 1,127 thermal units. 
 
 Suppose in each case the steam to be expanded to a terminal pressure of 
 10 lbs. absolute, the rates of expansion will then be 45, 9, 14, and 19 respec- 
 tively ; and the mean pressures corresponding to these initial pressures and 
 rates of expansion will be 25 lbs., 32 lbs., 36 lbs., and 39 lbs. respectively. If 
 the volume of a pound of steam varied exactly in the inverse ratio of the 
 pressure, these figures would represent the relative values of the efficiency of 
 the steam at the various pressures. But, taken exactly, the relative values 
 are 25, 333, 38'5, and 42*6, thus showing that a pound of steam at 90 lbs. 
 pressure is capable of doing 33 per cent, more work than a pound at 45 lbs. ; 
 a pound of steam at 140 lbs. pressure, 16 per cent, more than a pound at 
 90 lbs. ; and a pound at 190 lbs. pressure, 10"6 per cent, more than a pound at 
 140 lbs. pressure, or 28 per cent, more than at 90 lbs. pressure. In other 
 words, an engine using steam at 140 lbs. pressure should, apart from any 
 practical considerations, consume 16 per cent, less fuel than one using steam 
 at 90 lbs. ; and, again, that an engine using steam at 190 lbs. should consume 
 28 per cent, less fuel than one using steam at 90 lbs., and 10'6 per cent, less 
 fuel than one using steam at 140 lbs. 
 
 Looking to see how far practice agrees with these results, it is found that 
 
188 MANUAL OF MARINE ENGINEERING; 
 
 the ordinary compound engine, using steam at 140 lbs., is only a little more 
 economical than one using steam at 90 lbs., while the triple-expansion engine, 
 with steam at 140 lbs. pressure, gives an economy rather greater than theory 
 shows should be due to increased pressure. It follows, then, that there is some 
 other cause operating to produce the economic results shown in every-day 
 practice with this system, for there is now no question that the saving 
 in fuel effected by a triple-expansion engine, using steam and expanding 
 11 or 12 times, is about 20 per cent, of that used by an ordinary com- 
 pound engine of the same power, using steam at 90 lbs. and expanding 8 
 to 9 times. 
 
 It used to be maintained by the opponents of triple-expansion engines 
 that, if the ordinary compound engine is designed with a long stroke, it is 
 as economical. In order to see how far this is true by practice, it is sufficient 
 to examine the diagrams of the engines of the s.s. " Northern," whose 
 cylinders were 26 inches and 56 inches diameter and 60-inch stroke, used 
 steam at 130 lbs. absolute, and indicated 1,235 H.P., which show a con- 
 sumption of 154 lbs. of water per I.H.P. per hour ; and those of s.s. " Draco," 
 whose engines had cylinders 21 inches. 32 inches, and 56 inches diameter 
 and 36-inch stroke, using steam at 125 lbs. absolute, and indicated 618 H.P., 
 and showed a consumption of 14*1 lbs. per I.H.P. per hour. Or, again, 
 by comparing the performance of the " Draco " with that of the " Kovno," 
 whose cylinders were 25 inches and 50 inches diameter and 45-inch stroke, 
 using steam at 105 lbs. pressure absolute, and indicating 809 H.P., with 
 a consumption of 16"6 lbs. of water per I.H.P. In these cases, the consump- 
 tion of the " Kovno " was 17*73 per cent, in excess of that of the " Draco," 
 and 7"58 per cent, in excess of that of the " Northern " ; and the " Northern " 
 consumed 9*4 per cent, more than the " Draco." 
 
 It is, however, needless now to multiply cases, as it is a matter of common 
 observation that the saving in fuel is from 20 to 25 per cent., and it may 
 safely be taken that the triple-expansion engine, using steam at 165 lbs. 
 absolute pressure, uses 20 per cent, less fuel than an equally good ordinary 
 compound engine of equal power and working under similar circumstances, 
 using steam at 100 lbs. absolute pressure. 
 
 As the three-crank triple-expansion engine is now accepted as the most 
 suitable for marine practice generally, it is instructive to compare it. so far 
 as practical considerations are concerned, with the ordinary compound. 
 For that purpose, suppose, now, two engines are to be taken — viz., a triple- 
 expansion and an ordinary compound — to develop equal powers with the 
 same stroke of piston and same diameter of low-pressure cylinder. The 
 initial pressure in the one case is to be 100 lbs. absolute, and in the other 
 165 lbs. absolute. Let the number 14 represent the area of the low-pressure 
 piston in each case ; the mean pressure referred to the low-pressure piston 
 is to be 24, and the efficiency of the systems equal, so far as the steam is 
 concerned. The area of the high-pressure piston may then, without fear of 
 controversy, be taken as 4 for the ordinary compound, and 2 for the triple, 
 and the area of the medium-pressure piston of the triple as 5. If the referred 
 mean pressure is equally divided in each case between the cylinders, the 
 
 14 
 
 mean pressure in H.P. of the compound engine will be -j- X 12, or 42 lbs. ; 
 
 in the triple-expansion engine the mean pressure in the H.P. cylinder will 
 
TRIPLE-EXPANSION COMPOUND ENGINE. 189 
 
 14 14 
 
 be tt X 8, or 56 lbs. ; and in the M.P. cylinder -r- X 8, or 22'4 lbs., and 
 
 2 J 5 
 
 the following shows the relative work done, viz. : — 
 
 Ordinary Compound Engine. 
 
 High-pressure cylinder, 4 x 42 or 168 
 Low „ 14 x 12 or 168 
 
 Triple-Expansion Engine. 
 High-pressure cylinder, 2 x 56 or 112 
 Medium „ 5 x 22 4 or 112 
 
 Low ,, 14 x 8 or 112 
 
 That is, the average load on the rods, columns, guides, etc., is 50 per cent, 
 more with the ordinary compound engine than with triple-expansion. 
 
 To obtain a mean pressure of 24 lbs., with an initial pressure of 165 lbs. 
 absolute, and a pressure in the condenser of 2 lbs., the rate of expansion is 
 14, with an efficiency of 06 ; and with an initial pressure of 100 lbs. the rate 
 of expansion is 7. 
 
 On examining diagrams taken under these circumstances from actual 
 engines, the following is to be observed : — 
 
 Initial pressure in the high-pressure cylinder of the compound engine, 
 98 lbs. ; back pressure, 23 lbs. ; effective initial pressure, 75 lbs. per square 
 inch ; or load on the piston, 75 X 4, or 300. In the low pressure the initial 
 pressure is 2*2 lbs. ; back pressure, 4 ; giving an effective initial pressure of 
 18 lbs. per square inch ; or load on the piston, 18 X 14, or 252. 
 
 The initial pressure on the high-pressure cylinder of the triple-expansion 
 engine is 160 lbs. ; back pressure, 63 lbs. ; effective initial pressure, 97 lbs. ; 
 or the load on the piston, 97 X 2, or 194. In the medium pressure the 
 initial pressure is 70 lbs. ; and the back pressure, 21 lbs. ; effective initial 
 pressure, 49 lbs. ; or the load on the piston, 49 X 5, or 245. In the low 
 pressure the initial pressure is 18 lbs. ; and the back pressure 4 lbs. ; giving 
 an effective initial pressure of 14 lbs. per square inch ; or a load on the piston 
 of 14 X 14, or 196. Thus showing the loads in the case of the triple-expansion 
 engine to be much less, notwithstanding the higher pressure of steam 
 employed. 
 
 This, too,' is shown in actual practice by comparing the initial loads on 
 the engine of the s.s. " Northern," whose cylinders are 26 inches and 56 inches 
 diameter and 60-inch stroke, indicating 1,242 H.P., and supplied with steam 
 at 130 lbs. pressure absolute, with those of the triple -expansion engine of 
 the s.s. " Ariel," whose cylinders are 23 inches, 35 inches, and 60 inches 
 diameter and 57-inch stroke, indicating 1,527 H.P., and supplied with steam 
 at 165 lbs. pressure absolute. 
 
 The " Northern's " high-pressure piston sustains an initial load of 
 530 X 100, or 53,000 lbs. ; the low-pressure, 2,463 x 24, or 59,112 lbs. The 
 "Ariel's" high-pressure piston sustains 415 X 100, or 41,500 lbs.; the 
 medium-pressure, 962 X 60, or 57,720 lbs. ; and the low-pressure, 2,827 X 18, 
 or 50,886 lbs. — notwithstanding that the engines are larger and develop 
 nearly 25 per cent, more power with higher boiler pressure. 
 
 The more even distribution of pressure also very materially affects the 
 resistance of the slide-valves, and so tends in every way to reduce the losses 
 due to mechanical causes. 
 
 Experience has shown that the wear and tear of the triple-expansion 
 engine with three cranks is very considerably less than that of ordinary 
 two-crank compound engine of the same power and stroke, and no doubt 
 
190 MAVUAL QF MARINE ENGINEERING. 
 
 this is due to those causes already shown to exist with this class of engine, 
 as compared with the expansive engine. 
 
 The three-crank triple-expansion engine has, however, shown another 
 valuable quality, and one which may easily be surmised from the foregoing 
 reasoning — viz., that a much higher indicated power may be developed with 
 a low-pressure cylinder, equal in size to that of a common compound engine, 
 without any increase in the initial loads on the pistons. This may be shown 
 by comparing the performance of the engines of the " Eldorado," whose 
 cylinders were 26 inches, 40 inches, and 68 inches diameter and 39-inch 
 stroke, supplied with steam at 165 lbs. absolute pressure, with that of the 
 engines of the " Juno," having cylinders 35 inches and 69 inches diameter 
 and 39-inch stroke, supplied with steam at 100 lbs. absolute pressure. When 
 running at 72 revolutions, the " Eldorado's " engines develop 1,572 I.H.P., 
 and the " Juno's " 1,249 I.H.P., or 26 per cent, more power from the triple- 
 expansion than from the ordinary compound, although the low-pressure 
 cylinder was 3 per cent, smaller. The initial loads on the pistons of the 
 " Eldorado " are 54,060 lbs., 66,568 lbs., and 61,727 lbs., as against 69.264 
 lbs. and 71,041 lbs. on those of the "Juno." 
 
 Similar results can be shown with four-crank engines of various sizes ; 
 and to extend the question, it may be taken as approximately correct that a 
 referred mean pressure may be used in a triple-expansion engine 50 per cent, 
 higher than with an ordinary compound engine without any serious difference 
 in the stresses on the working parts and frame-work. It is for this reason 
 possible to manufacture a triple-expansion engine at the same price per 
 I.H.P. as an ordinary compound engine. The pcopelling efficiency of the 
 three and four-crank engine is especially noticeable when running at low 
 speeds, and it is no doubt on this account they were best for naval purposes 
 where so much cruising is done at comparatively slow speeds. They are 
 also capable of being worked at much fewer revolutions without stopping 
 on the centres than a two-crank engine, which is highly advantageous in 
 navigating intricate channels and docks and during a fog, as steerage power 
 is kept without much " way " on the ship. They are also, when well con- 
 structed and properly adjusted, almost noiseless, and cause little or no 
 vibration, which is an advantage in every ship, but especially in yachts and 
 passenger steamers. 
 
 The Compound Systems of Cylinders are now admitted on all hands without 
 any further controversy to be superior to the simple-expansive one, not- 
 withstanding that in the past at each of its stages of development this claim 
 has been contested by those who seemed unable to grasp the fact that 
 the determining factors in each particular controversy were practical rather 
 than theoretical ones, and much of the same nature as the one which caused 
 Watt to remove the condensation of the steam from the cylinder to the 
 condenser. Moreover, it was by practical tests rather than argument in each 
 of these stages that engineers were weaned from their love of the old to their 
 faith in the new system. It is unnecessary now to dwell on the subject, 
 or to recapitulate further the early experiments made to prove that the 
 compound engine was superior to the simple expansive, as measured by the 
 water consumption per I.H.P. ; or those later ones, whereby the compound 
 was shown to be inferior to the triple-expansion engine ; nor even to those 
 later still, when the advocates of a further subdivision claimed for their 
 
COMPOUND SYSTEMS OF CYLINDERS. 191 
 
 quadruple-expansion engine a distinct improvement on the good results 
 achieved by the triple. 
 
 So far no one has ventured on a quintuple stage system, although engines 
 with five cylinders and cranks have been made, besides which the increase 
 in boiler pressure now possible with cylindrical boilers has warranted such 
 an advance. Although the controversy has ceased, and a compound system 
 of cylinders of some kind is always found in a modern marine reciprocating 
 steam engine, it is well that the reasons for such an adoption should be clearly 
 stated and well understood, for after all it is somewhat in the nature of 
 promulgating a paradox to proclaim that if steam is expanded through all 
 the ports, passages and pipes of a compound system of four stages, the 
 economic result is better than if expanded in a single cylinder in the simplest 
 way possible. 
 
 The fact is that, theoretically, the compound system is wrong so far as 
 efficiency of steam expansion is concerned ; nevertheless, in practice, the 
 same conditions which caused the compound engine to triumph over the 
 expansive have established the superiority of the triple over the compound, 
 and given an advantage to the quadruple over the triple. The system with 
 its extensions have permitted the safe employment of higher steam pressures 
 with their higher efficiencies without making increases in the initial loads 
 on the rods, crank-pins, bearings, guides, etc., and, more important still, 
 without the large variations in temperature in the cylinders so objectionable 
 when steam expansion is to be effected without loss. 
 
 When the boiler pressure was 30 lbs. (45 lbs. absolute) the steam at entry 
 to the cylinder would have a temperature of about 270° F. ; during exhaust 
 the temperature in the cylinder would be only about 150° F., or a difference 
 during the cycle of 120°. The expansive engine using steam at 75 lbs. 
 absolute would have a range throughout the system of 155°, while in each 
 cylinder of the compound engine it will be only about 70°, allowing for the 
 usual drop of pressure between them. In the triple-expansion system using 
 steam of 195 lbs. absolute pressure, although the variation throughout is 
 about 230°, in each cylinder it is not much more than 70°. 
 
 In the case of a compound engine, the L.P. cylinder is about the same 
 size as the single cylinder of an expansive engine ; or, supposing the latter 
 to be a two-crank engine, each of its cylinders will be half the capacity of the 
 low pressure of the compound, and so each piston is half the area of that of 
 the L.P. piston. The initial pressure per square inch on the pistons of the two 
 engines will be the same if the boiler pressures are equal (say 60 lbs.) ; but 
 the area of the high-pressure piston of the compound would be only about 
 0*36 of the low-pressure, and consequently the ratio of forward load on it 
 would be under these conditions as 36 to 50 of that on each of the expansive 
 engine cylinders, but the back pressure of these latter would be only about 
 3 lbs., so that their effective initial load would be (60 + 15 — 3), or 72 lbs. 
 per square inch, while the high-pressure piston of the compound would have 
 a back pressure of about 22 lbs. absolute ; consequently the effective maxi- 
 mum load would be in this case (60 -f- 15 —.22), or 53 lbs. per square inch 
 only. The total loads will be then measured relatively as — 
 
 Compound, H.P. piston, . . 36 x 53 = 1908 
 Expansive, each piston, . . . 50 x 72 = 3600 
 
192 MANUAL OF MARINE ENGINEERING. 
 
 The L.P. cylinder of the compound engine will have a forward initial 
 pressure on it of about 18 lbs. absolute, and a back pressure of 3 lbs., the 
 effective initial pressure being thus 15 lbs. Then the following holds good 
 for comparison as being the maximum loads : — 
 
 On each of the cylinders of the expansive engine the 
 
 load may be taken as 50 X 72 = 3600 
 
 That on the HP. cylinder of the compound engine, 36 X 53 = 1908 
 
 That on the L.P. cylinder of the compound engine, 100 X 15 = 1500 
 
 But not only is the range of temperature greater with the cylinders of 
 the simple engine, but all the steam that enters must pass over a surface, 
 which, for 45 per cent, of the cycle time, is exposed to the cooling action of 
 the condenser ; moreover, those ports, passages, valves, cylinder ends, and 
 pistons have a surface three times the area of the corresponding parts of the 
 H.P. cylinder of the compound system, over which all its steam passes, and, 
 moreover, its exposure is only to that of the receiver, whose temperature 
 will be about 230°, as against 140° of the condenser. The difference will be 
 greater still with a modern condensing apparatus, providing a vacuum as 
 high as 29 inches, with a temperature during exhaust probably as low as 
 110°. Under such conditions the condensation on entry to the cylinder, 
 and during expansion in it, would be great ; and, besides, beyond the loss 
 of heat, the efficiency of the engine would be lowered by the impeding action 
 of the deposited water. Any re-evaporation that is possible would take 
 place too late in the stroke to be of any practical use in the cylinder of the 
 simple engine ; but in the case of the compound, however, it would be of 
 considerable use in the second cylinder, although no value may have accrued 
 from it in the H.P. cylinder where the transformation took place. The loss 
 by condensation at the valves, whether they be of the flat or piston type, 
 is more considerable in practice than might have been anticipated. The 
 inside or exhaust side of the L.P. valve of a compound, as of both of those 
 of the simple engine, is exposed always to the cooling effect of the condenser, 
 while the hot steam is on the other side; under these circumstances, and 
 inasmuch as the metal of these valves is comparatively thin, the condensa- 
 tion is really quite considerable ; but it is even more so in the simple 
 engine, as the steam is then of boiler pressure or nearly so, and consequently 
 of much higher temperature, so that transmission is the more rapid due to 
 the greater difference of temperature. 
 
 In fact, the difference of temperature is the all-ruling factor in this 
 as in some other engineering problems ; for, beside the above, on the practical 
 side there is always the additional risk of fracture of cylinder and valves, 
 and, in the case of a compound system, the straining action on it of the 
 cylinders, if they are cast or rigidly bolted together. The solution of the 
 latter difficulty, however, is easy, being simply to avoid rigid connections, 
 and provide that each cylinder shall be quite free to expand or contract as 
 circumstances may cause it. 
 
 Fig. 65 is interesting and instructive, showing, as.it does, the economic 
 process of expanding steam of high pressure through the four cylinders 
 of a quadruple-expansion engine made by Messrs. Richardson & Westgarth. 
 The stages can be clearly traced on the combined diagram, and the tem- 
 perature and pressure changes marked. 
 
COMPOUND SYSTEMS OF CYLINDERS. 
 
 193 
 
 o 
 
 'So 
 a 
 K 
 
 c 
 _o 
 
 "55 
 
 S 
 c8 
 n 
 
 M 
 O 
 
 '■si 
 
 to 
 
 — 
 S 
 
 s 
 
 
 pq 
 
 «3 >> 
 
 ft 
 
 h 
 O 
 
 c9 
 
 c 
 
 SO 
 
 ■LjOu/ jod sqf g/ */wj" 3jnsssJ,j 
 
 13 
 
194 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Fig. 66 is also instructive, as by it can be seen the various stages of 
 progress in the use of steam on shipboard, and the effect of adding a L.P. 
 
 uadrofit 
 
 At nqsghertc _Pr* insure 
 
 —~~1 
 
 Fig. 06. — The various stages in the use of Steam expansively in Marine Engines. 
 
 turbine to a compound cylinder engine studied, and how the attenuated and 
 apparently feeble steam at exhaust is capable of producing in a suitable 
 generator quite a large amount of power. 
 
HORSE -POWER. 195 
 
 CHAPTER VIII. 
 
 TTORSE-POWER — NOMINAL, INDICATED, AND SHAFT OR BRAKE. 
 
 When the steam engine began to replace other motors, it was soon found 
 necessary to introduce some unit by which its power could be expressed 
 without using such high numbers as foot-pounds ran to, as to place it beyond 
 the grasp of ordinary minds. As the engine was frequently taking the place 
 of horses to operate mining and other machinery, it was only natural that 
 the work performed by a horse should be taken as the basis for this unit 
 of measurement. The number of units of work performed by the most 
 powerful dray horses in a minute was found to be 33,000, so Watt chose this 
 as the unit of power for his engines, and called it " horse-power," and this 
 has continued to be the standard ever since, both for land and marine engines. 
 Watt found that the mean pressure usually obtained in the cylinders of 
 his engines was 7 lbs. per square inch. He had also fixed the proper 
 piston speed at 128 X v/stroke per minute, and his engines were arranged 
 to work at this speed, so that he estimated the power which would be de- 
 veloped when at work to be 
 
 Area of piston x 7 X 128 ^stroke -f- 33,000. 
 
 The power so calculated was called " Nominal," because the engine was 
 denominated as of that power, and in practice that power was actually 
 obtained. But when the boiler could be constructed so as to supply steam 
 above the atmospheric pressure, and the engine ran with more strokes per 
 minute than before, the power actually developed exceeded the nominal 
 power, and from the name of the instrument by which the pressure of steam 
 in the cylinder was obtained it came to be called the " Indicated " Power. 
 
 The discrepancy between Nominal and Indicated Power became in time 
 so great, that for all scientific purposes the former ceased to be of value. 
 It remains, nevertheless, in general use among manufacturers and users 
 of engines, because it better conveys the commercial value and size than 
 does the developed power ; for since the area of the piston is usually the 
 only variable in the expression, it follows that the size of the cylinder, and 
 therefore the size of all the other parts, must vary directly with the 
 Nominal Horse-Power. But since Indicated Horse-Power depends on three 
 functions — viz., area of piston, speed of piston, and the pressure of steam — ■ 
 the value may be changed by altering the value of one or more of these, which 
 alteration may be material without affecting the commercial value. For 
 example, an engine may be caused to run at a much higher number of revolu- 
 tions, even so as to double its Indicated Power, with hardly any additional 
 cost whatever in construction to the engine itself. 
 
 The Admiralty modified Watt's rule to suit the practice of the early 
 
196 MANUAL OF MARINE ENGINEERING. 
 
 days of steam-navigation, by substituting the actual«|>iston f.peed for the 
 arbitrary one, and so the old rule was for 
 
 Admiralty Nominal ) _ Area of piston X speed of piston X 7 
 Horse-Power J — 33,000 
 
 The Admiralty, however, dropped the use of the expression altogether, but, 
 before doing so finally, used it in the modified sense of being one-sixth of the 
 Indicated Power. 
 
 In the Mercantile Marine the rule for Nominal Horse-Power is by no 
 means uniform. Before the introduction of the compound engine and 
 increased boiler pressures, it was usual to allow 30 circular inches per N.H.P. 
 — i.e., the rule was- — 
 
 _ T TT _. (Diameter of cylinder in inches) 2 , , .. , 
 
 N.H.P. =- -^ X number oi cylinders. 
 
 For two-stage compound engines — D being the diameter of the low-pressure 
 cylinder, and d that of the high-pressure — 
 
 But neither of these rules took into account the length of the stroke 
 or the boiler pressure, although it was generally understood and' roughly 
 standardised. The general adoption of the triple-compound engine, however, 
 caused the question of stroke to be removed from the field of competition, 
 and there is now more uniformity in practice. Besides which, in some large 
 centres of marine engineering, all the makers have a standard set of sizes 
 of cylinders for N.H.P. to which they adhere. In order to meet difficulties 
 some engine-makers have adopted a rule for Nominal Horse-Power, based 
 on the capacity of the cylinder, and in so doing have nearly met the require- 
 ment on which the continuance of the expression depends — viz.. that it is a 
 measure of the commercial value of the engine. The power per revolution 
 depends on the capacity of the cylinder so long as the mean pressure is the 
 same ; but since small engines are usually worked at a higher number of 
 revolutions than larger ones, the power developed by the former will bear a 
 larger ratio to the Nominal Power than will be the case in the latter. 
 
 A simple and fair rule for N.H.P.* is deduced as follows :— D is the diameter 
 
 of low-pressure cylinder, d that of the H.P., d x that of the M.P., and d 2 that 
 
 of the second intermediate of a quadruple engine ; S is the stroke, all in 
 
 inches, and D X 0-65 is the standard length of stroke ; in a triple compound 
 
 /D\ 2 /D\ 2 /D\ 2 
 
 i^j = 2-7 ; (-7-1 =6 or thereabouts; in a quadruple compound ( -,-J =8; 
 
 * X * 
 
 J — 4; ( -7- J =2 or thereabouts. 
 
 N.H.P. (Triple) - — + ^ 
 
 N.H.P. (Quadruple) = 
 
 30 
 
 d 2 + d* + d. 2 + D 2 
 
 30^ 
 
 • The JJonril of Trade Kulc now in force is X.1I.1*. = (3 H 4 D- \/S'» x VP + 700. 
 
 H is the total heatins surface. D the diameter of L.P. cylinder, S the stroke in feet. an<! P the load 
 on safety valve in lbs. per square inch. 
 
ESTIMATED HOESE-POWER. 197 
 
 Any change of stroke should give a proportionate change of N.H P., 
 hence 
 
 N.H.P. = ^ + <*i 2 + D 2 - S 
 
 30 D x 0-65 
 
 Substituting the values of d, d-^ and d. 2 in terms of D, then 
 
 N.H.P. = — ^K — f° r a compound engine. 
 
 D x S 
 N.H.P. = for a triple-compound engine. 
 
 D x S 
 N.H.P. = " for a quadruple-compound engine. 
 
 Lloyd's N.H.P. — No Nominal Power, however, can be any guide to the 
 •capabilities of the engine, unless the power of the boilers is also in some way 
 expressed or understood ; and as it is not easy to imagine how the former 
 can be introduced into any expression which shall effect the latter, or vice 
 versa, the suggestions of Lloyd's Committee remained unfulfilled, but the 
 Register now contains a statement of the leading particulars of the boilers, 
 and for purposes of levying the fees for the Survey and Registration of 
 Machinery, Lloyd's Register employ the following rule, 
 
 Lloyd's nominal horse-power = p x z ( 1 — ), 
 
 where p is the boiler pressure; D is the diameter and S the stroke of L.P. 
 piston, both in inches; and H is the heating surface in square feet. 
 
 The value of x is 15 for the ordinary boiler with natural draught, but 
 with forced or induced draught x is 12. 
 
 The value of z is 0'34 when the boiler pressure is under 160 lbs. ; when it 
 is 160 lbs. per square inch or above z is 0*393. 
 
 That there is need of uniform practice in naming the power of engines, 
 is apparent to everyone having to do with steamships, and the Board of 
 Trade Department, which registers the power, has so far limited its efforts 
 
 in this direction to the old formula, N.H.P. = * J. . As this takes 
 
 60 
 
 no cognisance of stroke, it was never satisfactory ; the Department, however, 
 is still satisfied with it for its purpose, but it is surely time to find some other 
 rule which shall determine the rating of engineers and such other matters, 
 as well as be a fair indication of the power the ship possesses to propel her. 
 
 Estimated Horse-Power.* — As it is desirable that a power be named for 
 an engine which shall enable the lay mind to judge of its capabilities, pro- 
 bably the better plan would be to revert to the principle of Watt, who, as 
 has been shown, attempted to specify the power which the engine was actually 
 expected to develop ; such a rule, therefore, should give approximately 
 the Indicated Horse-Power. It would, of course, be far better to register 
 the I.H.P., but as it is not always possible to obtain this, the next best method 
 js to estimate it, and call it the Estimated Horse-Power, or E.H.P. 
 
 » N.E. Coast I. E. and S. recommend the following : — 
 
 E.H.P. = " .,' — . D in inches, S in feet. 
 
198 MANUAL OF MARINE ENGINEERING. 
 
 The following rule will give approximately the horse-power developed 
 at full speed by a two-stage, triple-, or quadruple-expansion engine made 
 in accordance with modern practice : — 
 
 «, xr t> D ' 2 x Jp x R x S . 
 
 ^ HP = 7^00 
 
 Where D is the diameter of the low-pressure cylinder, p the absolute pressure,. 
 R the number of revolutions per minute, S the stroke of piston in feet. 
 
 For Example. — To estimate the Indicated Horse-Power of an engine 
 having cylinders 30 ins., 48 ins., and 80 ins. diameter and 48-ins. stroke, 
 revolutions 75, and boiler pressure 165 lbs. 
 
 ^ xr t> 80 ' 2 X s/180 X 75 x 4 
 
 EHR = 7^00 
 
 = 3297. 
 
 Many other rules have been propounded for N.H.P., some of which are 
 ingenious, but impracticable, while others fail to give results of any value 
 whatever, so that neither class needs notice here ; but it may be mentioned 
 that when non-condensinn engines were more used in steamships than they are 
 at present it was found necessary to have a special rule for them, which was 
 
 D 2 x ^S 
 
 N.H.P. = 
 
 20 
 
 D being the diameter of the cylinder in inches, and S the stroke in feet. 
 
 Indicated Horse-Power may be defined as the measure of work done in 
 the cylinder of a steam-engine, as shown from the indicator-diagrams, and 
 only falls short of the actual work by such small losses as are caused by 
 the friction of the pin or pencil against the paper, the friction of its working 
 parts, and that in the pipes or passages connecting the indicator to the 
 cylinder. The latter discrepancy is by far the most important, and is some- 
 times serious in very long stroke engines, where the indicator pipe is several 
 feet long. The others, in the hands of a skilful operator, are not so serious, 
 certainly not in modern marine engines to the extent stated by Mr. Him, 
 who says he found the Indicated Horse-Power, owing to losses in the diagram 
 from the friction of the indicator, to correspond with the useful xvork done 
 by the engine. All the same, it should not be forgotten that with such an 
 instrument as the indicator, the nearer it is to the steam in the cylinder the 
 better. There should be no pipes, if possible, and, if any, they should be 
 fairly large. 
 
 The Indicator Diagram. — The diagram itself shows only the pressure 
 of steam acting on the piston at any and every part of its stroke ; but from 
 it may be calculated the mean effective pressure acting during that stroke, 
 and it is assumed that the particular diagram measured is only a sample 
 of what might have been taken at every stroke, so that the mean pressure 
 thus calculated is the force acting on the piston during the whole period 
 of its motion in which the power is taken — usually one minute. Hence, 
 Indicated Horse-Power = area of piston in inches X mean pressure in lbs. 
 per square inch X number feet travelled through by the piston per minute ■*■ 
 33,000. 
 
 This, of course, applies only to double-acting engines, as in single-acting 
 tMgines the pressure is acting only half the time on the piston, and hence. 
 
MEAN PRESSURE. 
 
 199 
 
 instead of taking the number of feet travelled through by the piston per 
 minute as the multiplier, — the length of stroke in feet X number of strokes 
 per minute should be substituted. 
 
 Mean Pressure. — The mean pressure is usually obtained by dividing the 
 indicator-diagram by a number of equidistant ordinates perpendicular to 
 the atmospheric line, and so placed that the distance of the first and last 
 from the extreme limits of the diagram is half the distance between two 
 consecutive ones ; the sum of their lengths, intercepted by the diagram, 
 divided by their number, gives the mean length, and this, referred to the 
 scale on which the diagram was drawn, will give the mean pressure. To 
 illustrate this : — Fig. 67 is an indicator-diagram whose length, A X, is, say, 
 5 inches, and taken with a spring requiring a pressure of 30 lbs. per square 
 inch to compress it 1 inch ; so that if M L is 2 inches, it represents a pressure 
 of 60 lbs. ; and if B L is 2| inches, it signifies that, at the point L, the pressure 
 on the piston was 2| X 30 lbs., or 75 lbs. per square inch above the line 
 A X, which, in this case, shall be the line of no pressure, and hence is 75 lbs. 
 absolute, or 60 lbs. above the atmospheric pressure. Now, for convenience 
 of division, let there be 10 ordinates enclosing 9 spaces — since there is to 
 
 ABODE 
 
 Fig. 67. — Indicator Diagram. 
 
 be a half space at each end, there will be in all equal to 10 spaces — so that 
 the distance between the ordinates is 5 inches -f- 10, or half an inch. Measure 
 off A B = } inch ; B C, C D, D E, etc., each = | inch, and at B, C, D, E, etc., 
 draw perpendicular lines, cutting the diagram at M L, N, etc., Y Z. Then 
 
 (M L + O N + etc +YZ)-rl0 = x inches, and a; X 30 is the 
 
 mean pressure of the diagram. 
 
 This diagram is from one side of the piston only, and, when one only is 
 obtainable, it is sometimes assumed to represent both, and the mean pressure 
 thus obtained used to calculate the power ; but it seldom happens, although 
 it is much to be desired, that the mean pressure is precisely the same on 
 both sides of the piston, consequently, any result obtained in this way is not 
 satisfactory. If the effective area of the piston is the same on both its sides 
 — that is, if there is the same area on which the steam acts to propel the 
 piston forward, on the one side as on the other — the mean pressure found 
 from the diagram taken from the one side, may be added to that found from 
 the diagram taken from the other, and divided by 2 to give the true mean 
 pressure per revolution. 
 
200 MANUAL OF MARINE ENGINEERING. 
 
 Professor Rankine showed that the diagram should be divided from 
 A to N bv ordinates equidistant apart, and the mean obtained by the 
 following rule : — Let n be the number of such divisions (usually 10), 
 b , 6j 6 2 6 3 . . . b n the length of the ordinates intercepted by the diagram : 
 then 
 
 Mean length = (-"-±-5? + b x +b 2 + etc., . . . + b n -i} -s- 
 
 n. 
 
 The chief objection to this is that, in actual practice b n would be always 
 without value, and b either without value, or so difficult to measure as to 
 cause differences of opinion as to its value. 
 
 Another, and a very ready way of obtaining the mean pressure, is by 
 measuring the area of the diagram by a planimeter, and dividing it by the 
 length A N, the result being the mean breadth as before, and this multi- 
 plied by the scale of lbs. of the spring will give the mean pressure. This 
 is, of course, the quickest plan, and the most accurate, as being mechanical ; 
 .and where many diagrams have to be calculated with despatch, it is very 
 advisable to have a good planimeter. Special instruments are now made 
 for this purpose. 
 
 In whatever way the mean pressure be measured, it forms the basis of 
 calculation of actual energy, or, as it has been called, Indicated Horse- 
 Power, and is therefore of the utmost importance, since most modern 
 formulae bearing on marine machinery and marine propulsion are based on 
 I.H.P. Hence, any error in taking the diagrams must lead to errors in 
 design from calculations by formulae based on false premises ; this should 
 always be borne in mind by the operator, on whose skill and care a good 
 and true diagram depends as much as on a good instrument. It would be 
 a very valuable quality in an indicator to be able to give the useful ivork 
 of the engine, as was stated by Mr. Hirn to be the case generally ; but it is a 
 quality which no such instrument can possess, inasmuch as, with the same 
 cylinder performance, there may be a great variety of actual performances 
 of the engine, depending on the efficiency of the various parts, and the indi- 
 cator only gives this cylinder performance. Could the effective power be 
 easily obtained as it is with electric generating engines, a great benefit would 
 be conferred on masine engineers in making calculations, etc., and in deter- 
 mining the best make and design of engine. The precise power absorbed 
 in overcoming the resistance of the working parts of a marine engine could 
 not be measured ; for if the engine be allowed to run without load, it is 
 not running in the same state as when running with its load ; and any diagrams 
 taken then cannot be taken as the loss from friction, etc., on the guides, 
 journals, pistons, and valves when running with the increased pressure on 
 them due to the increased pressure on the piston with the load. Since the 
 efficiency of an engine very much depends on the resistance on these parts, 
 any calculation or formula which excludes, or does not give due allowance 
 to this, is misleading. Hence, to assume that the power absorbed by an 
 engine in overcoming its resistance is measured by the power indicated 
 when running without load is not correct — as it is possible for an inefficient 
 engine to show a high efficiency when tested in this manner. It is. however, 
 true that a large portion of the resistance is the same, or practically the 
 same, when the engine is running at the same number of revolutions with 
 
SHAFT HORSE-POWER. 201 
 
 and without load. The frictional resistance of glands, pistons, pumps, 
 tunnel-shaft journals, high-pressure piston valve, and sundry small gear, 
 is the same per revolution at any speed, whatever be the load, or power 
 developed. The total resistance or non-useful work of engines, therefore, 
 probably varies nearly directly as the revolutions. 
 
 Shaft Horse-Power is a term now often used and well known to the marine 
 engineer, and is likely to be more so than the I.H.P., which hitherto has been 
 in such general use to express the performance of an engine. 
 
 Inasmuch as the indicator was of no service to the maker of turbines, 
 nor could any instrument which merely shows the pressure of steam be a 
 means of determining the power developed by a velocity machine, some 
 other method had to be devised for that purpose. When a turbine was 
 driving a dynamo it was easy to calculate the mechanical power of the driver 
 by the measure of the electrical output from a dynamo whose efficiency was 
 known. The brake horse-power of a turbine could be found by causing it 
 to operate on a water brake, and so for some considerable time that was the 
 only way in which turbine power was determined. It is true when there 
 were two sister ships, whose actual resistance was known, an approximation 
 to the power developed by the turbine was made by means of that of the 
 reciprocating engine in the other ship, as given by the indicator. This, 
 however, was not a satisfactory state of things, especially as I.H.P. is looked 
 on with suspicion, and that not without justification, for even with a slow- 
 running engine the human element is a factor involved in the accuracy of 
 the I.H.P., while with a very fast-running engine the diagram made by the 
 best of indicators handled by a man is influenced by the skill of the operator. 
 There is, however, an indicator, the product of Prof. Hopkinson's genius, 
 that does give a diagram which is free from suspicion, by which the actual 
 cycle of pressure in a cylinder may be viewed as it really is, and from it the 
 power may be determined. But it is as inapplicable to the turbine as the 
 Richards or other indicator for the reasons given above. 
 
 It was necessity, therefore, which stimulated invention and caused to 
 be brought forth the torsion meter, whereby the power transmitted by a shaft 
 can be determined by observing the angle of twist of a definite length of that 
 shaft. There are various kinds of such meters, all of which show a consider- 
 able amount of inventive genius, mechanical knowledge, and skill. In some 
 of them measurement is mechanical, made by a self-registering instrument, 
 in others the eye is employed, and in others the ear, to fix the amount of 
 distortion at any given time when the shaft is transmitting energy. It is 
 certainly most desirable to eliminate the human element if possible, but in 
 doing so at present other factors of an equally undesirable kind are intro- 
 duced (v. Chap. v.). 
 
 Shaft Horse-Power calculated as on p. 150 : — d is the diameter of a shaft, 
 I the length under observation, both in inches, T is the torque in inch-pounds, 
 6 is the angle of twist in the length I, and Mr is the modulus of rigidity, which 
 is about 11,750,000 in solid shafts and 12,150,000 in hollow ones. 
 
 584xTxL .. , , fa 
 
 6 in degrees = — ^ n — tor solid shafts 
 
 6 Mr X d i 
 
 = nr-pjj j-f- for shafts with a bore of diameter d,. 
 
 Mr (d 4 — d x 4 ) 
 
202 MANUAL OF MARINE ENGINEERING. 
 
 The Shaft Horse-Power at revolutions R = r= X tt^ = TxRt 63,000. 
 
 The formula in daily use is therefore (v. p. 150) 
 
 S.H.P. = ° d * X 7 R for solid and d (rf 4 - df)-^—. for hollow shafts. 
 Q X I y X t 
 
 Q is usually taken at 3-27 ; this assumes the modulus to be 11,250,000, 
 which is below what is generally found for steel shafts of very good quality. 
 
 Example 1. — What is the total H.P. of a ship having three shafts 
 6 inches in dameter revolving at 600 times per minute, and twisting 0*4° in 
 40 inches ? 
 
 S.H.P. of each shaft = M *%"*:«» = 2,378. 
 
 The total power of the ship is, therefore, 3 x 2,378, or 7,134 H.P. 
 
 Example 2. — A tunnel shaft is revolving 300 times per minute, and is 
 
 10 inches diameter ; the angle of twist in 40 inches is 0*33°, what power is it 
 
 transmitting ? 
 
 xj o tt Tj °' 33 X 10 >°°0 X 300 _ ... 
 Here S.H.P. = ^-^ = 7,645. 
 
 The angle of twist is at all times very small, seldom exceeding 1'3° in 
 120 inches of length at full power, so that the pair of discs of a torsion meter 
 are necessarily of large diameter to give accurate and fine readings, especially 
 as in many ships it is possible to use only a very short length of shaft. When 
 possible, however, it is certainly desirable to cover as long a portion. of a 
 tunnel shaft as possible. 
 
 The shafting of a ship, however, has another load on it besides that of 
 transmission of the power generated by the engine, for the thrust block is 
 usually close to the engine room, and consequently the whole of it from the 
 screw to the block has to resist the thrust. This may amount to as much as 
 to equal 20 per cent, of the torque force, but its influence on the twist is really 
 very slight, probably only about 1*5 per cent, at most ; generally it is so 
 small as to be negligible, as stated by Prof. Hopkinson, from a considerable 
 experience gained in testing his torsion meter. Each shaft should be and is 
 tested by levers and weights to ascertain its actual resistance to torque, and 
 to definitely determine the real and exact amount necessary to twist the 
 shaft through a definite angle. From the observations so obtained, it is 
 easy to make a diagram, to which reference may be made on trial when 
 the torsion meter is showing the value of d to obtain the corresponding S.H.P. 
 per revolution, so that by merely multiplying by the number of revolutions 
 per minute the full S.H.P. is determined. 
 
 It is, however, not sufficient for a marine engineer to know only the power 
 developed by the engine or turbine ; he must be acquainted equally well 
 with the power taken and delivered at every critical point of the ship and 
 machinery, so that he may make up a balance sheet which shall show on the 
 one side the maximum gross power the engine has given out, and on the 
 other side the disposal of the same, so that nothing is missing and unaccounted 
 for. In fact, the economies of engineering are of the very highest importance, 
 
THRUST HORSE-POWER. 203 
 
 and cannot be too carefully studied by all concerned from the designer to the 
 engineer in charge. Each and every part of the machinery of a ship must 
 be in such a state as to be working at the highest rate of efficiency, and how 
 to ascertain that rate exactly must be the constant care of those in charge 
 of it and responsible for its good and economic working. 
 
 Thrust Horse-Power is an expression that will be more often used in the 
 near future than in the past, inasmuch as means may be soon provided where- 
 by the actual thrust on a line of screw shafting will be as easily shown as the 
 pressure in the condenser now is. Inventors have for some time turned 
 their attention to devise some simple and easy method of doing this, and 
 one of them, Mr. Heck, fully disclosed to the members of the Institution 
 of Naval Architects (Transactions, 1909) two or three such methods that are 
 within the bounds of practical politics, while not quite satisfying engineers 
 that they are the ones to be finally adopted. They are all ingenious and 
 capable of giving results fairly free from inaccuracy. So far, these and some 
 others are based on the principle of showing, by means of a hydraulic ram or 
 its equivalent, the pressure per square inch in the chamber, or the total 
 load on the equivalent ram, and how to make allowance for the friction of 
 packings, etc. A more recent idea is to make the flange couplings hollow 
 and elastic, fit them together, water-tight, and fill the space between them 
 with water whose pressure is indicated by a gauge in the usual way. The 
 thrust block might also be used to give its own indications of thrust (as also 
 suggested by Mr. Heck), if it were mounted on roller bearings or suspended 
 on a stirrup and its thrust taken by a pair of hydraulic rams with chambers 
 connected, and the water in them acted on a gauge and spring or loaded 
 resistance of some kind, as the hydraulic brakes on a gun carriage. 
 
 The Gross Power of a Steam Engine is, of course, that generated in the 
 cylinders by the steam pressure on the pistons acting through the space 
 traversed by them. This is known to the marine engineer as the Indicated 
 Horse-Power. 
 
 Shaft Horse-Power is that transmitted through the tunnel shafting from 
 the engine to the propeller, and ie, therefore, the net product of the engine 
 or the gross power less that absorbed in moving the engine and its appur- 
 tenances, ealled the Friction Horse-Power. 
 
 The mechanical efficiency of the engine is, therefore, S.H.P. -4- I.H.P. 
 
 Brake Horse-Power is also that transmitted through the shaft to a resistance 
 capable of absorbing it just as the propeller does that from the marine engine. 
 In this case the brake not only takes the power, but indicates exactly the 
 torque or twisting moment from which the B.H.P. may be calculated. Brake 
 horse-power should and does coincide very closely with S.H.P. 
 
 The Thrust Horse-Power is that exerted by the screw in pushing the ship 
 forward, and is measured by multiplying the actual thrust in pounds by the 
 number of feet moved through by the ship in a minute and dividing by 33,000. 
 
 If S is the speed in knots per hour, and T the thrust in pounds, 
 
 ™ . v, T x S X 6,080 . 
 
 Thrust horse-power = 00 ^^ ^— = 1 X o h- 6Zb. 
 
 r 33,000 X oU 
 
 This measures the capacity of the screw as a propeller, consequently 
 
 The efficiency of the screw = T.H.P. h- S.H.P. 
 
204 MANUAL OF MARINE ENGINEERING. 
 
 Indicated Thrust is an expression introduced by Dr. William Froude ai 
 
 a measure of the thrust of a propeller. 
 
 I.H.P. X 33,000 
 
 Indicated thrust = - ., , , — , ,. 
 
 pitch of screw X revolutions 
 
 Tow Rope Horse-Power is that necessary to propel the ship at the required 
 speed if freed from what are called the augmented resistances, the chief of 
 which is due to the action of the screw itself on the sWrn of the ship to retard 
 her motion. If Tr is the tension on an imaginary tow rope from the ship 
 to the towing agency, and S the speed of the ship in knots per hour, then 
 
 Tr X S 6,080 Tr X S X 101-3 
 Tow rope horse-power = ^qq X -^ - = - -33^ 
 
 This is the net-horse-power of propulsion ; hence — 
 
 Tr X S 
 Propulsive efficiency = — ^ 5- I.H.P. 
 
 Net Horse-Power can be calculated, as already shown, by estimating the 
 skin resistance and the residuary resistance, adding them together, multiplying 
 the sum by the feet passed through in a minute, and dividing the product 
 by 33,000 lbs. 
 
 If Sr be the skin resistance and Rs be the residuary, both in pounds, and 
 S the speed of the ship in knots per hour, then 
 
 Net H.P. = < Sr + R „l * 101 3 . or Sr + R? 
 
 33,000 ' 326 
 
 Piston speed and revolutions enter largely into the calculation of horse- 
 power, and, therefore, important factors at every stage of a design. The 
 modern engineer does not permit himself to be hedged in and bound by the 
 arbitrary rules of former generations, nor indeed would they have been 
 to the extent they did had they had the benefits of the knowledge, experience, 
 and materials now enjoyed. Fertina lente was the policy of every progressive 
 as of every prudent engineer in early days, and may with advantage form 
 a portion of that of modern ones, who now benefit by the failures as well as 
 the successes of the pioneers of the profession. The original users of the screw 
 propeller cannot be accused of fearing high revolution of that instrument, 
 and had those who followed them so soon attempted to make engines which 
 could revolve at the speeds of these screws when directly coupled to them, 
 such wretchedly low efficiencies as exhibited in the machinery and propeller 
 of the "Greyhound " under Dr. William Froude's analysis would not have been 
 possible. 
 
 To-day both piston speeds and revolutions of engine are much higher 
 than prevailed with the old compound engine. The better distribution of 
 torque with triples and quadruples, and the balancing of the inertia forces 
 have enabled this to be done successfully, so that to compete with the turbine 
 still higher revolutions may be attempted in the future with engines specially 
 designed for it. 
 
 Experience has shown that heavy marine pistons may run safely at a 
 mean velocity of 900 feet per minute, and, in some instances, the pistons 
 
PISTON SPEED AND REVOLUTIONS. 005 
 
 of some large vertical engines in first-class cruisers have reached a velocity of 
 even 1,000 feet ; while higher speeds still have been attained in torpedo-boat 
 destroyers, whose pistons move, when run at express speed, at a velocity over 
 1,200 feet per minute. Although there is no difficulty in causing a piston to 
 move at even higher speeds than these, it is doubtful if there would be any 
 advantage in doing so, besides which the risk of causing serious damage to> 
 the cylinders, and precipitating a break-down without any warning, is very 
 great. There is no doubt that a well-fitted piston, moving in a smooth and 
 true cylinder at a speed of 1,000 feet per minute, will work well so long as 
 the rubbing surfaces receive some lubrication from the moisture of the steam 
 or the oil injected, and there is not the slightest fear of danger under these 
 circumstances ; but if, with a little priming, scum is carried into the cylinder* 
 and causes abrasion of the rubbing surfaces, an immense amount of mischief 
 may be done in a few seconds. Moreover, when the cylinders wear a little 
 out of shape from one cause or other, so that the packing-rings will have 
 lateral motion, the danger increases with the velocity of the piston. In the 
 Navy engines supplied with steam from water-tube boilers have no internal 
 oil lubrication beyond what passes in on the rod surfaces ; so that in destroyers,, 
 whose pistons are moving at a speed of 20 feet per second, there is only the 
 moisture from the steam to lubricate them. 
 
 Although the revolutions of a screw engine may be, within certain limits,, 
 as few or as many as the designer chooses, experience or prejudice has fixed 
 very closely in practice the limits beyond which it is not considered expedient 
 to go. In the days of the geared engine, the screw revolved three or four 
 times to one of the engine, and no objection was raised to the small screw 
 and the high number of revolutions ; later such a thing was deemed very 
 objectionable — on the ground of excessive speed of piston and excessive 
 friction in journals. The slow-moving engine was quoted as a proof of the 
 economy of slow piston speed and small friction without being a real founda- 
 tion for the argument. 
 
 The fine lines of the older steamships admitted of the small screw, which 
 was the accompaniment of the engine, by necessity geared. Bluff ships, as 
 now built for mercantile purposes, require a much larger screw for the same 
 power of engine and dimensions of hull than formerly obtained ; and it is not 
 to the slowness of the pistons that they owe their economy, but rather to the- 
 small number of strokes per minute made by them in turning the large screw. 
 
 An engine requires a certain power to be expended in moving it through 
 one revolution to overcome internal resistances ; if the number of revolutions; 
 is 80 per minute, this power will be double that at 40, and, roughly, will 
 vary directly with the revolutions. But the resistance of the propeller, caused 
 by friction of the water on the surface of the blades, will increase roughly 
 as the square of the revolutions, so that the power expended to overcome 
 this resistance at 80 revolutions is eight times that required at 40 revolutions. 
 If now the screw can be so altered with respect to pitch that, at 40 revolu- 
 tions, the same speed of ship is obtained as at 80 revolutions, the indicated 
 horse-power will be found to be considerably less ; and although the coal 
 consumed per I.H.P. will not be less, and may possibly be more than before, 
 the consumption per day will be considerably less. Now, although this 
 economy is co-existent wfth decreased piston speed, it is not due to it. 
 
 The object of a high rate of piston velocity is to decrease the piston area 
 
20t) MANUAL OF MARINE ENGINEERING. 
 
 and that generally for the sake of reducing the size of the engine. But an 
 increased velocity may be obtained either by increasing the stroke of piston, 
 or by increasing the number of revolutions ; if the former method is adopted 
 there will be no decrease in the size of engine ; but, on the contrary, an 
 increase in space occupied and in the weight. If a high piston speed is 
 obtained by a high number of revolutions, a smaller cylinder will suffice for 
 a certain indicated horse-power than if the same piston speed were obtained 
 by length of stroke alone. In other words, engines which are required to 
 develop a certain power in a minute will vary in size of cylinder inversely 
 as the number of revolutions per minute, all other things remaining constant; 
 and if the cylinders are of the same diameter, the stroke will vary inversely 
 as the number of revolutions. 
 
 The piston speed of many engines is gOA'erned entirely by circumstances 
 beyond the immediate control or will of the designer. An example of this 
 is the case of the paddle-wheel engine with vertical oscillating cylinders. 
 If the position of the shaft is determined by the structural arrangements of 
 the hull, as is often the case, then the diameter of the wheel is fixed, and the 
 speed of ship fixes the number of revolutions to be made by the wheel ; the 
 length of stroke of piston is limited by the distance from the centre of the 
 shaft to the floors or keelson of the ship. Further, if the engineer is free to 
 decide the position of the shaft, any attempt to increase the piston speed by 
 placing the shafting high is frustrated by the fact that, the higher the shaft 
 the larger will be the wheel, and consequently the fewer the revolutions. If 
 the engine is inclined, then the designer may fix the diameter of the wheel 
 to suit the revolutions which he deems most advisable, or he may fix the 
 position of shaft to suit the ship's structure, and still be free to choose the 
 stroke of piston. 
 
 Again, the horizontal engine had to be designed so as to accommodate 
 itself to the space allotted to it in the ship, which means that only a limited 
 length of stroke was permissible. The revolutions, however, in this case 
 could be varied considerably ; but there is, after all, a limit to the number ; 
 beyond this limit any increase will result in very little gain in speed, and a 
 very certain loss of efficiency. If the screw is of comparatively small diameter, 
 owing to the shallow draught oi the ship, a higher number of revolutions than 
 usual is absolutely necessary to project a sufficient mass of water back to 
 propel the ship forward with the necessary velocity ; and it is the medium 
 number, or that number at which the engine can be run without loss of 
 efficiency so as to obtain the maximum speed of ship that is so difficult to 
 decide, and which can only be determined with any degree of certainty by 
 experiment. 
 
 One great feature which places the vertical engine so much above all 
 the other forms of screw engine, as an economic and good working machine, 
 is its superior length of stroke. Power for power, the vertical engine always 
 has exceeded the horizontal in this respect ; and although in the practice of 
 the past there was no very great difference in this matter between the two 
 types, the tendency is now to make the stroke as long as is possible or con- 
 venient in the engines of all ships. 
 
 The advantages of the long stroke are due to the corresponding decrease 
 
 in piston area. Two engines of the same power, and working at the same 
 
 timber of revolutions, must have the same volume of cylinder ; or, to speak 
 
REVOLUTIONS. 207 
 
 more correctly, the pistons must sweep out the same volume if their efficiency 
 is the same. The crank-shafts will be of the same diameter, and the crank- 
 pins, also, practically of the same dimensions. Now the one with the long 
 stroke will have smaller pistons than the other, consequently the total pressure 
 on the pistons will be smaller — and, in fact, is inversely proportional to the 
 stroke ; consequently, the pressure on the guides, crank-pins, and journals will 
 vary in the same way, and the friction on them correspond also. The lateral 
 pressure of the piston packing rings will vary with the diameter, so that any 
 reduction in diameter will produce a corresponding reduction in the friction. 
 
 But, perhaps, so far as .economy in working is concerned, there is no more 
 important consideration than the reduction in clearance space effected in 
 the low-pressure cylinder by the reduction in piston area. The steam ports 
 will be nearly the same in section, whether the engine be long or short stroke ; 
 but the space between the piston and cylinder-ends is very considerably 
 reduced, and will vary inversely as the length of stroke, because the axial 
 distance of piston from the cylinder-ends is constant. 
 
 The Rate of Revolution oi Marine Engines * at full speed varies roughly 
 inversely as the square root of the nominal horse-power, which for this 
 purpose may be taken as N.H.P. = D x S v K. 
 
 D is the diameter of the low-pressure cylinder (or the equivalent, if there 
 are two) in inches. 
 
 S is the stroke of piston also in inches. 
 
 K is a coefficient of 15 for two-stage compound, 12-6 for triple-compound, 
 and 10-5 for quadruple-compound engines. Then — 
 
 Rate of revolution per minute . . . = Q -j- -y/N.H.P. 
 
 For the ordinary cargo boat, Q . . =1,200. 
 
 For express steamships, Q . . . = 1,800. 
 
 For naval and very fast express ships, Q . = 2,250. 
 
 Example (a). — The proper rate of revolution for an express steamship 
 having cylinders 30 inches, 45 inches, and 70 inches diameter X 42 inches 
 stroke. 
 
 Here N.H.P. = 70 X 42 -*- 12-6, or 233. 
 
 Revolutions per minute = 1,800 -e- V233, or 118. 
 
 Example (b). — Rate of revolution for a warship having engines 33 inches, 
 52 inches, 64 inches, and 64 inches diameter X 48 inches stroke. 
 
 Here D = y/2 X 64 2 , or 90-5. 
 
 N.H.P. = 90-5 X 48 - 12-6, or 345. 
 
 Revolutions per minute = 2,250 s- V345, or 121. 
 
 Revolutions. — Although there is a very considerable range for choice of 
 number of revolutions of the engines of most merchant steamers, there are 
 certain well-defined limits beyond which very few practical engineers go. 
 
 Very few screw engines are now worked below 75 revolutions per minute 
 when in good condition ; and it is at this speed that most of the engines of 
 the large mail steamers are kept running on the voyage so long as the weather 
 permits. The engines of warships, for two very good reasons, work at much 
 higher speeds. Their machinery must be light, and go into a small space, 
 so that it is necessary to make an engine of certain dimensions to suit these 
 
 * The N.E. Coast Inst. E. and S. recommend the following as the rule for rate of revolution of caruo- 
 ship engines when on voyage : — -, oa 
 
 N = 32 (S + 4) -=- S or ~ + 32. 
 
208 
 
 MANUAL OF MARINE ENGINEERING. 
 
 conditions, and cause it to develop the requisite horse-power by running 
 at a higher number of revolutions. The speed of a warship is much higher 
 in proportion to its size than is that of the merchant ship, while the draught 
 of water is no more, and often less. For these reasons the screw of the war- 
 ship is small for the power to be developed, so that even if large engines 
 were admissible to drive the screw, they would be of small advantage, as 
 they would have to move at a high rate. It will be seen, then, that small 
 fast-running engines are a necessity, and especially is this so with modern 
 warships, whether armoured or unarmoured. The latter must be as fine as 
 possible, and every ton of weight saved to obtain the very high speeds which 
 their service demands ; the former demands every sacrifice to save weight in 
 machinery, for the sake of adding it to the armour and armament. 
 
 Since a warship has so seldom to steam at full speed, and when she does, 
 it is only for a short period, the short-stroke fast-running engine is not so 
 very objectionable, and rigid economy is quite a secondary consideration in 
 war questions. 
 
 The following table gives the number of revolutions at which naval 
 engines are run on their trial trips : — 
 
 TABLE XXIXa. — Rates of Revolutions of Screw Propellers. 
 
 Type of Engine. 
 
 Description of Ship. 
 
 Power of each Engine to 
 
 a Screw. 
 
 Revs, per Minute. 
 
 Reciprocators, 
 
 Battleships and lst-class 
 
 
 
 
 
 cruisers, 
 
 19,500 to 15,000 IE 
 
 130 to 14u 
 
 >• 
 
 2nd-class cruisers, 
 
 4,500 to 6,500 
 
 >» 
 
 140 to 146 
 
 11 
 
 3rd-class cruisers, . . 
 
 3,500 to 5,000 
 
 ii 
 
 225 to 245 
 
 91 
 
 Scouts, 
 
 7,000 to 8,500 
 
 it 
 
 210 to 202 
 
 91 
 
 Destroyers, . ... 
 
 2,000 to 3,500 
 
 ii 
 
 360 to 400 
 
 »» 
 
 Torpedo boats. 
 
 . . 
 
 
 
 11 
 
 Large Atlantic expresses, 
 
 15,000 to 20,000 
 
 ii 
 
 76 to 80 
 
 11 
 
 Large mail steamers, , 
 
 5,000 to 15,000 
 
 ii 
 
 90 to 80 
 
 19 
 
 Large cross-channel 
 
 
 
 
 
 steamers, . . . 
 
 3,000 to 4,000 
 
 ii 
 
 175 to 150 
 
 >1 
 
 Small cross-channel 
 
 
 
 
 
 steamers, . 
 
 2,000 to 3,000 
 
 »> 
 
 200 to 175 
 
 » 
 
 Large twin-screw cargo 
 
 
 
 
 
 steamers, 
 
 2,200 to 4,500 
 
 19 
 
 80 to 90 
 
 M 
 
 Large single-screw cargo 
 
 
 
 
 
 steamers, . 
 
 2,500 to 4,500 
 
 l> 
 
 80 to 70 
 
 !) 
 
 Medium single-screw cargo 
 
 
 
 
 
 steamers, . 
 
 1,000 to 2,500 
 
 >> 
 
 90 to 80 
 
 9) 
 
 Yachts and small craft, . 
 
 250 to 500 
 
 Jt 
 
 150 to 250 
 
 Turbines, 
 
 Battleships, . . . 
 
 5,500 to 7,500 S.H.P. 
 
 320 to 300 
 
 »> 
 
 Large cruisers, 
 
 10,000 to 20,000 
 
 91 
 
 200 to 175 
 
 91 
 
 Second-class cruisers, 
 
 6,000 to 12,000 
 
 11 
 
 500 
 
 11 
 
 Scouts, 
 
 4,500 to 7,500 
 
 11 
 
 750 to 600 
 
 H 
 
 Destroyers, . 
 
 2,500 to 5,000 
 
 11 
 
 940 to 750 
 
 .. 
 
 Atlantic expresses, largest, 
 
 15,000 to 17,500 
 
 »» 
 
 190 to 165 
 
 91 
 
 Express steamers, large, 
 
 4,000 to 6,000 
 
 »» 
 
 300 to 190 
 
 »> 
 
 Cross-channel, large, 
 
 3,000 to 5,000 
 
 11 
 
 600 to 450 
 
 *» 
 
 Yachts and small craft, . 
 
 1 ,000 to 2,000 
 
 11 
 
 1,000 to 750 
 
 ,, geared, 
 
 Expresses, . 
 
 2,500 to 3,000 
 
 *1 
 
 300 
 
LENGTH OF STROKE. 
 
 209 
 
 The stroke of horizontal engines varied from 18 inches of the gunboat to 
 54 inches of the large armour-clad, and the vertical engines of the Navy vary 
 from 18 inches in the "destroyers" to 51 inches in the first-class cruisers 
 and battleships. Latterly, with the four-cylinder engines, the stroke has 
 been 42 to 48 inches in the large ships. 
 
 Length of Stroke. — For very many years there existed a standard scale 
 for the stroke of the vertical engine of the mercantile marine, and although 
 there was no written law which guided engineers in the choice of this important 
 dimension, it was so well known that only the diameter of the cylinders was 
 mentioned in speaking of the size of engine, and in most of the rules for 
 nominal horse-power used by manufacturing engineers in their dealings with 
 shipowners, no direct allowance was made for length of stroke. With the 
 walls of the cylinders only jacketed the best relation of diameter to stroke 
 was as 1 to 1*5. So the ratio of stroke to diameter of H.P. cylinder is usually 
 1-7 and the ratio of stroke to diameter of L.P. cylinder 0-9 to 0-6. 
 
 The following Table gives the stroke corresponding to the different 
 powers ; the one column giving the standard, and the other the stroke a3 
 existing in ordinary e very-day practice in the mercantile matin e : — 
 
 TABLE XXIXb. 
 
 N H P 
 
 Standard 
 
 Stroke, as in 
 
 N.H.P. 
 
 Standard 
 
 Stroke, as in 
 
 1 
 
 Stroke. 
 
 common practice. 
 
 Stroke. 
 
 common practice. 
 
 ! 
 
 20 
 
 12 ins. 
 
 15 ins. to 18 ins. 
 
 140 
 
 33 ins. 
 
 33 ins. to 42 ins. 
 
 30 
 
 15 ins. 
 
 18 ins. to 21 ins. 
 
 160 
 
 33 ins. 
 
 33 ins. to 42 ins. 
 
 40 
 
 18 ins. 
 
 20 ins. to 24 ins. 
 
 180 
 
 36 ins. 
 
 30 ins. to 45 ins. 
 
 50 
 
 21 ins. 
 
 24 ins. to 30 ins. 
 
 200 
 
 36 ins. 
 
 36 ins. to 48 ins. 
 
 60 
 
 24 ins. 
 
 24 ins. to 27 ins. 
 
 250 
 
 39 ins. 
 
 39 ins. to 54 in 6 !. 
 
 80 
 
 27 ins. 
 
 27 ins. to 30 ins. 
 
 300 
 
 42 ins. 
 
 42 ins. to 54 ins. 
 
 1U0 
 
 30 ins. 
 
 30 ins. to 36 ins. 
 
 400 
 
 45 ins. 
 
 45 ins. to 60 ins. 
 
 120 
 
 30 ins. 
 
 30 ins. to 39 ins. 
 
 500 
 
 48 ins. 
 
 48 ins. to 66 ins. 
 
 TABLE XXIXc. — Eevolutions and Piston Speeds according to Rules 
 recommended by n.e. coast inst. e. and s. for cargo steamers 
 on Service. 
 
 Stroke 
 
 Revolutions. 
 
 Piston Speed. 
 
 Stroke. 
 
 Revolutions. 
 
 Piston Speed. 
 
 Feet. 
 
 Per Minute. 
 
 Feet 
 
 Feet. 
 
 Per Minute 
 
 Feet. 
 
 1-50 
 
 117 
 
 352 
 
 3-25 
 
 72 
 
 464 
 
 1-75 
 
 103 
 
 368 
 
 3-50 
 
 69 
 
 480 
 
 2-00 
 
 96 
 
 384 
 
 3-75 
 
 66 
 
 496 
 
 2-25 
 
 89 
 
 400 
 
 4-00 
 
 64 
 
 512 
 
 2-50 
 
 83 
 
 416 
 
 4-25 
 
 62 
 
 528 
 
 2-75 
 
 78 
 
 432 
 
 4-50 
 
 60 
 
 544 
 
 3-00 
 
 75 
 
 448 
 
 4-75 
 
 59 
 
 560 
 
 14 
 
210 MANUAL OF MARINE ENGINEERING. 
 
 CHAPTER IX. 
 
 GENERAL DESIGN AND THE INFLUENCES WHICH AFFECT IT. 
 
 The General Design and Arrangements of Marine Engines are to-day charac- 
 terised by simplicity perhaps more than anything else, notwithstanding that 
 multiplicity of parts and connections gives the modern engine-room an air of 
 complexity that was wanting in the older ships. Forty years ago there 
 were, as a rule, only two cylinders to the main engines : both air and cir- 
 culating pumps, as well as the feed and bilge pumps, were worked from the 
 pistons, direct in the Navy, and in that way or by means of levers interposed 
 in the mercantile marine. There was, as a rule, only one auxiliary feed 
 pump, and in large ships one fire engine in addition, which did duty for a 
 bilge pump as well as for deck service when required ; merchant cargo ships 
 with ballast tanks had also the second auxiliary pump for emptying them, 
 and arranged to act for general purposes. Steam starting gears- were being 
 used generally with very large engines, but never with small ones. Steam 
 steering gears were equally rare, and there was, of course, no electric light 
 engine, nor any refrigerating plant. Distillers had been introduced for 
 giving fresh water as auxiliary feed to the boilers by Hall, of condenser fame, 
 thirty years before this period, but the idea had not taken on. There were, 
 however, in general use the well-known Normandy distillers for producing 
 drinking water ; every Naval ship had a set, and they were to be found 
 in the better-class passenger steamer. The mercantile engineer had, for the 
 exercise of his talents and the occupation of spare time, the care of the steam 
 winches, just as the naval engineer had in some ships the charge of the turret 
 and turntable engines and ammunition hoists, etc., outside his own domain. 
 Very few ships had more than one main engine or a steam launch or pinnace, 
 and no ship had air compressors. All these things have been added one after 
 another, until the care and anxiety of the chief engineer is no longer centred 
 and concentrated on the main engines, but now it is rather the very numerous 
 parasites of them and of the crowd of outside machinery — machinery so 
 necessary to the well-being of the modern warship and express steamer — 
 that gives him the most concern. But, in spite of all this, there is about 
 the main engines, especially if they be turbines, a simplicity that is com- 
 mendable under such circumstances. The day for fancy design and odd 
 arrangements of machinery has gone ; in their place there is now only a 
 stereotyped system that marks the final stage of natural selection and the 
 survival of the fittest. The engine of one modern engine builders differs 
 very little in essence from that of another ; the turbine, it is true, has its 
 variation in principle as well as in arrangement, but it, too, is gravitating 
 to the stereotyped design, although makers may still each have his own, 
 differing in form of details and fittings. 
 
DESIGN AND ARRANGEMENTS OF MARINE ENGINES. 211 
 
 The accepted and approved type of reciprocating engine is, of course, 
 the vertical inverted direct-acting one, whether it be worked by steam or by 
 the internal combustion of oil or gas. The cylinders are placed in line each 
 over its own crank, and, as a rule, they are separated one from another ; each 
 has its columns braced and secured so as not to be dependent on its neighbour 
 for stability or steadiness. The connections for steam transmission are by 
 pipes so fitted as to expand and contract freely under thermal or pressure 
 forces without acting or reacting on the cylinders and so cause them bad 
 alignment. At the same time the general structure of columns is braced 
 together, so as to give it, as a whole, additional stability against external 
 disturbance, as from the shock or strain of rolling, pitching, and collisions. 
 Small mercantile engines, however, are still made with the cylinders cast 
 or bolted together in pairs, but the tendency is in the direction of separating 
 all of them, especially when the stroke of piston and consequently their 
 length is greater than that usual with the high-speed land engines of equal 
 power. Each cylinder now operates on a single complete piece of crank 
 shaft running in two bearings coupled to the next cranks ; this is so in all 
 but very small engines (under 100 N.H.P.) or in larger engines, where weight 
 and space have to be cut down to the minimum, when the shaft is then in one 
 piece. The valves are now seldom placed on the cylinder outer sides, but 
 generally on the fore or aft side, so that the valve spindle is immediately 
 over the shaft, and consequently capable of being driven by the double 
 eccentrics and link motion direct without the interposition of weigh shafts 
 and levers. Steam reversing gears are fitted to all but the very smallest 
 engines, and are generally of the push-and-pull type introduced and still 
 supplied by Brown Bros., of Edinburgh, for the mercantile marine, while 
 the " all round " worm and wheel gear is most frequently used in the Navy. 
 The former is decidedly the handier, and can now be obtained at quite as low 
 a cost as the all round gear, but the latter can be used to turn the main engines. 
 Similar push-and-pull gears, of various designs, are employed to operate 
 the very large stop valves of the reciprocating and turbine engines, as they 
 <lo also the heavy change valves of the latter, whereby the turbine is reversed 
 by changing the incoming steam from the head-going to the stern-going 
 part of it, or when the exhaust steam from the L.P. cylinder is diverted to 
 a L.P. turbine. The circulating pump is now, in all but the small mercantile 
 engines, of the centrifugal type driven by an independent engine ; in all 
 sorts of fast-running engines the air pump is also, as a rule, absent, and it, 
 too, has its separate engine. With turbines where high vacuum is a necessity 
 for high economy of steam, the double-stage air pump of Mr. Weir is often 
 employed ; there is still a prospect of the centrifugal pump being employed 
 in series on this service, as it has been used for boiler feeding and delivering 
 •water generally against a high " head " on land. 
 
 The Admiralty long ago initiated, and the mercantile marine followed 
 in express steamers, the practice of removing the whole of the feed and bilge 
 pumps from connection with the main engines. The merchant ship, how- 
 ever, was really first in the employment of the independent feed pump with 
 automatic regulating gear, whereby its speed was only such as to give just 
 the necessary supply of feed demanded by the stokehold watch-keeper. 
 Now, in all important steamers, and certainly in all turbine-driven ships, 
 the feed bilge fire and general service pumps are absolutely separate and 
 
212 MANUAL OF MARINE ENGINEERING. 
 
 independent of one another ; moreover, they are generally in duplicate to 
 avoid the risk of stoppage of the main engines in case of the failure of an 
 auxiliary. 
 
 Formerly the Condenser was a huge cast-iron vessel, forming an integral 
 and important part of the engine structure or framing, to which it generally 
 gave a massive foundation for the guide columns, and contributed to the 
 stability of the framing (v. fig. 71), as well as a firm support for the bearings 
 of the pump beam's weigh shaft. With the advent of the three-crank engine 
 came the tendency to separate the condenser from the engine frames, 'and 
 as this kind of engine became lengthened out, and; furthermore, when there 
 came a four-crank engine to stay, it was no longer expedient to have it with 
 such long tubes as would be entailed with a condenser body the whole length 
 of the engine. Moreover, the addition of three and four sets of columns 
 or column feet on it made it a costly and somewhat risky casting, as well 
 as a heavy portion of the weight of the machinery. To-day the condenser is 
 generally a distinct and separate vessel placed close to the L.P. cylinders, 
 and of a form suitable to its own particular service, and in no way subser- 
 vient to any other such service as was formerly exacted from it when it formed 
 part of the engine superstructure (v. fig. 68). The cylindrical form was the 
 common one, on account of its natural strength and cheapness, but now the 
 demand of condenser experts have caused the cylindrical to be rejected for 
 one with the heart shape cross-section, notwithstanding that the general 
 requirements of such experts can be carried out sufficiently well in the cylin- 
 drical shell by fitting it with the form of special baffle directors (v. fig. 116), 
 or in the rectangular contraflo design of Mr. Morison (v. fig. 117). 
 
 The framework of the engine is now much as was usual formerly in it? 
 general design, but greater care is exercised to give both general and local 
 stiffness to every part, which, no doubt, is an improvement which tends tG 
 the better working of all engines in every way. 
 
 The Lubricating Arrangements are much more extensive and complete 
 nowadays than formerly obtained, and the necessity for a regular steady 
 and positive supply of oil to every bearing guide and pin is now recognised 
 and provided by means of small pumps worked by the engines, or from over- 
 head tanks, which deliver a steady stream at a pressure high enough to force 
 the lubricant into every part requiring it. 
 
 The Design of an Engine is influenced by External Causes, which are not 
 always seen or even acknowledged, but nevertheless are often in active opera- 
 tion all the time. There is even in the engineering and shipowning world 
 also that which is generally elsewhere called fashion, exercising its powers, 
 and imposing on the designer conditions with which it is often futile to struggle. 
 Even the better general and technical knowledge possessed by experts to-day 
 does not always preserve them from following a lead which is more or less of 
 a blind nature, and from starting out in directions quite contrary to the 
 convictions based on their experience, and for no other reason apparently 
 than because someone else has done so with all the appearance of success. 
 The trial and error system still prevails largely, in spite of the technical and 
 scientific training now so freely obtainable everywhere at quite low cost. 
 Model experiments are always interesting, often instructive, and, when 
 properly understood, may be good and safe guides for directing the pro- 
 ceedings of the work-day engineer, but such models, even when telling truth, 
 
DESIGN OF ENGINE INFLUENCED BY EXTERNAL CAUSF-S. 
 
 213 
 
 do not always tell the whole truth ; what they suppress often comes out as 
 a ghastly truth when the full-sized working engine is produced ; what was 
 a trifling matter and hardly tangible in the miniature becomes in the preat a 
 
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 terrible incubus, producing the most damning consequences ; what was a 
 pigmy in the small engine and quite easy of control becomes a very Franken- 
 stein" in the large one, irreducible, and carrying all before it, and spoiling 
 
214 MANUAL OF' MARINE ENGINEERING. 
 
 everything. It behoves everyone, therefore, to use models and model 
 experiments with great discretion. 
 
 The Supply of Materials exercises a powerful influence on the designer 
 
 of engines generally, but especially on him who has to cut down to a minimum 
 
 the weight of and space occupied by machinery in the way the marine engine 
 
 builder is compelled to do. When Brunei designed the " Great Britain " 
 
 there was no steam hammer in existence and no forge capable of making 
 
 so large a shaft as that required for the ship ; consequently he had to be 
 
 content with one of cast iron when at that stage ; fortunately for him, 
 
 Nasmyth invented just then the steam hammer, and made one in time to 
 
 produce a wrought-iron shaft, with which the first voyage of that ship was 
 
 made. Cast iron was largely used for all parts of marine engines, and designs 
 
 were made accordingly. When copper was boomed up to £80 per ton* and 
 
 even higher by an enterprising but somewhat short-sighted syndicate, the 
 
 equally enterprising and versatile marine engineer adopted steel and iron for 
 
 piping, and had other things made of cast steel which formerly had been 
 
 exclusively of bronze. Steel castings, doubtful and often unsatisfactory 
 
 as they may be, have served their turn in changing the design of many 
 
 parts of an engine, but what has effected the greatest departure from old 
 
 practices to that which prevails to-day is the cheap and unlimited supply 
 
 of excellent mild wrought steel. By its means the pressure of steam possible 
 
 in cylindrical or tank boilers has been raised from 100 to 240 lbs. per square 
 
 inch, and their possible diameter increased from about 14 feet to 18, while 
 
 their cost has been reduced very considerably. The internal plates of such 
 
 boilers used formerly to .be made of " Lowmoor quality," at a cost of £27 
 
 per ton and upwards ; similar plates in the superior metal (mild steel) and 
 
 of very large sizes can be purchased now for a third of that sum ; shell plates 
 
 of iron, 1| inches thick, could be bought, but they were narrow and not very 
 
 long ; moreover, if of the full area the mill could turn out, they were very 
 
 costly. Now, plates up to 50 feet long and 12-5 feet wide can be rolled, up to 
 
 a thickness of 1*8 inches ; shell plates are often If inches thick and 30 feet 
 
 long, so that one, or at most two, plates when single-ended are sufficient ; 
 
 circular-end plates are made up to 13 feet in diameter. Forgings of every 
 
 description and size can be made of this material, and, if needed, steel of a 
 
 higher tensile, say 40 tons ultimate, with quite a good amount of stretch, 
 
 can be obtained also in large size and at moderate cost ; for special purposes 
 
 at costs by no means prohibitive vanadium, nickel, or other high-class steels 
 
 are made, and supplied of a quality and fitness beyond reproach. 
 
 Rolled bars of excellent steel can be obtained of any diameter up to 
 15 inches, and square bars to 6 inches ; rectangular section bars can also be 
 had in various sizes up to 15 inches by 2 inches, and 11 inches by 2.V inches,, 
 so that steel caps for bearings, rod ends and other similar purposes can be made 
 in wrought, steel at a cost very little in excess of that of castings in iron. With 
 the self-hardening tool steel and high-speed lathes and machine tools now in 
 use, these rolled bars can be converted into bright ones at a trifling extra cost. 
 The framework of an engine in wrought columns and tie bars is, therefore, 
 now a much less expensive luxury than it was formerly, even for large engines. 
 Then, too, the other sectional steel now obtainable in such variety has per- 
 mitted designers to fashion engine beds and their framing in ways not possible 
 without it, whereby great saving in weight is effected, and inasmuch as little 
 
 * Now £150, due to war conditions. 
 
THE INFLUENCE OF TONNAGE LAWS. 215 
 
 * 
 
 or no pattern-making is required for such designs, it is an economic method 
 when the engine required is a special one with few or no repetitions of it 
 expected. 
 
 The method of obtaining sectional material (introduced by Mr. Dick) 
 by the extrusion of the zinc bronzes through dies has permitted of multiplying 
 varieties without necessitating the expense involved in cutting rolls. This 
 and the other ways in which these high-class tough, strong bronzes mav 
 be used have had no small influence also on the design of the smaller special 
 engines. 
 
 Such things permit of refinements in design not possible in the days 
 when engineers were limited to the choice of wrought iron of 20 tons ultimate 
 tensile strength, or, by paying a high price, 25 tons at most. Crank and 
 straight shafts of almost any size can be forged and machined at prices that 
 were impossible for even small ones a few years ago. Even cast iron has been 
 improved by selection and mixing, that without paying fancy prices for any 
 of the good brands, the tensile and bending tests of castings are equal to the 
 very best given by Fairbairn & Whitworth. Aluminium has not yet seri- 
 ously influenced the marine engine designer ; that it will in the near future 
 is certain ; its lightness alone will attract him, and probably alloys will be 
 found which, while not adding seriously to their weight, will improve their 
 strength and resistance to corrosion. 
 
 The Influence of Tonnage Laws on the hulls of ships is well known and 
 obvious, but it is not limited to them. It pervades the whole of a ship more 
 or less, and perhaps more so the machinery space than elsewhere. Latterly 
 that influence has been more potent and insidious in its effect on marine 
 engine design than believed to be possible years ago. Steamships have had 
 from the earliest days of their construction some consideration given for 
 the disadvantages under which they were worked compared with the sailing 
 ship. The space occupied by the machinery has always been deducted 
 from the gross tonnage, inasmuch as it could take no cargo : later on, in order 
 to encourage shipowners in making the spaces, not only habitable, but health- 
 ful for those in charge of, and those labouring at the machinery, special 
 allowances were made for the light and air spaces to engine and boiler rooms, 
 as well as for those rooms themselves. If the actual total of the spaces 
 which are allowed off the gross tonnage of a ship amount to 13 per cent, of it, 
 the actual deduction permitted, so as to arrive at the net or register tonnage 
 is no less than 32 per cent. In the days of low freights, costly fuel, etc., 
 it is highly necessary to keep the register tonnage down, as on it the ship is 
 taxed. Now, whereas in Naval ships the space allotted to machinery is 
 always small for it, as indeed it was formerly only too often the case in the 
 mercantile marine, until it was found that the enlarging of it enabled a con- 
 siderable saving in working expenses to be effected at little cost and with 
 inconsiderable drawbacks. Formerly with the two-cylinder engine often 
 of quite small power, but requiring large supplies of fuel, it was quite impos- 
 sible to arrange for a reduction of tonnage measurement so large as 13 per 
 cent, without a much too serious sacrifice ; nor, indeed, did then the exigencies 
 of the times press for such drastic measures to obtain it, consequently it was 
 only tug boats and express steamers having very large engines in proportion 
 to their size that enjoyed these liberal concessions, as indeed they alone 
 were intended so to do ; consequently to the majority of steamers there was 
 
216 MANUAL OF MARINE ENGINEERING. 
 
 * 
 
 no gain by making the engine and boiler rooms larger than required by the 
 bare necessities of the case, except that, until the amendment of the Merchant 
 Shipping Acts, light and air spaces were not measured into gross tonnage, 
 although they were included in the deductions from it ; consequently every 
 ton of space devoted to that purpose caused a virtual reduction of two tons 
 in the register tonnage. 
 
 Since the space occupied by the machinery could not be laden with cargo, 
 every foot of it was a foot less space in the cargo holds, it was made as small 
 as possible, and the engineer designer had to exercise his wit to devise the 
 most compact engine to occupy the least possible space. Hence the popularity 
 with shipowners of the single-crank engine, either with one cylinder or two 
 compound ones tandem ; of the diagonal engine with a cylinder in each wing 
 operating on a single crank (fig. 69), or with a third cylinder vertical, as in 
 fig. 70. The compound engine with one cylinder vertical and the other 
 horizontal, both operating on the same crank, found favour, as did also the 
 engine with one pair above them and the other pair of cylinders behind 
 operating on the same pair of cranks with levers, as revived by Mr. M'Alpine, 
 latterly with the idea to provide a naturally balanced engine. Besides these 
 typical instances, there were other ways in which the engine was treated by 
 the ingenious, as, for example, in the way of special valve gears, etc., to 
 reduce still more thereoy the space occupied by it (fig. 71). 
 
 A further effect of the tonnage law, however, was more serious, inasmuch 
 as it retarded the adoption of the better forms of triple- and quadruple-com- 
 pound engines, and encouraged the manufacture of those that were worse ; 
 the fitting of valve gearing of almost fantastic design, which displayed an 
 exaggerated inventiveness, while it gave endless trouble and anxiety to those 
 in charge of it, was entirely due to the desire to save space and reduce the 
 size of the engine-rooms. The early three-crank triple engines were often 
 treated in this way ; by shipowners they were objected to, although only 
 too anxious to benefit by the three-stage system, on the ground that they 
 required too much room ; by the engine builder they were subject to such 
 cutting and squeezing that the crank bearings and pins were reduced 
 past the just minimum, and the valve gears were obstructions, and pre- 
 vented the proper attention necessary to such working parts ; moreover, 
 they required such an expenditure of oil as to detract from the economy of 
 fuel (r. fig. 71). 
 
 The two-crank triple, which in its way worked and did good service, was 
 not so good an engine mechanically as the three, nor was the two-crank 
 quadruple so good as the present four-crank one, although it was better in 
 some ways than the two-crank triple. It was, however, due probably to the 
 tonnage question that the quadruple engine came so soon as it did into the 
 field of practical engineering, inasmuch as the claim that won it most attention 
 and patronage at the first was the small space the two-crank engine occupied 
 compared with that required for the three-crank compound. To-day all this 
 is changed ; every ship must have the 32 per cent, reduction ; and, further, 
 in some cases, since the law says that where the spaces which may be deducted 
 amount to over 13 per cent, of the gross, the rate of reduction shall be one and 
 three-quarter times the actual amount. In some very fully-powered ships the 
 actual spaces may amount to even 54 per cent, of the gross ; then the net or 
 register tonnage is really only 5 # 5 per cent, of the gross. There are, however, 
 
THE INFLUENCE OF TONNAGE LAWS. 
 
 217 
 
218 
 
 MANUAL OF MARINE ENGINEERING. 
 
THE INFLUENCE OF TONNAGE LAWS. 
 
 219 
 
 limitations set now so that the allowance may not amount to a public scandal, 
 as it did quite recently to the Dock Companies, Harbour Boards, etc. Now it 
 would seem as if an engine-room could not be too large, but, since the Board 
 
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220 
 
 MANUAL OF MARINE ENGINEERING. 
 
 of Trade will not permit it to be wholly measured off tonnage unless the size 
 of the engines themselves will warrant the space allotted, there is a direct 
 premium existing for making an engine as long as possible, and also to extend 
 it athwartship liberally likewise. Consequently the four-crank engine (fig. 
 
 72), triple as well as quadruple, is the favourite in all express steamers, 
 because ostensibly it is so much easier to balance. 
 
 The " Joys," " Marshall," and other special gear for driving valves placed 
 in front or rear of the cylinders are consequently no longer necessary and so 
 
THE INFLUENCE OF THE BOARD OF TRADE. 221 
 
 gone, and in their place the old and well-tried pair of eccentrics with their 
 link motion is revived in their best forms, so that they are now with ample 
 bearing surfaces, sufficient breadth of strap, etc. ; they are, moreover, quite 
 accessible. The condenser, for perhaps the same reason, sometimes is placed 
 in the wings and the pumps of various kinds scattered about, each with its 
 allotted and ample space. No longer are tbey and all other auxiliary engines 
 and machines squeezed into odd corners and niches, which were narrow 
 and inaccessible often, and always cramped and confined. 
 
 All this may not be altogether the result of the tonnage laws alone, but 
 it is pretty certain that if there were not such good allowances off gross tonnage 
 there would not be such liberal spaces, and if there were not roomy spaces, 
 then the engines would have to be cut to fit them, such as they might be. 
 
 The Influence of the Board of Trade on marine engine design and practice 
 in other directions has been, and is still, very great and, on the whole, bene- 
 ficial. The better construction and design of the cylindrical boiler was 
 largely due to the action taken by Mr. Thomas Traill and his staff officials of 
 the Board about 1873 in formulating Rules and Regulations, which, while 
 being somewhat arbitrary and often unnecessarily rigid in application, were, 
 nevertheless, one of the chief means whereby the marine boiler has been so 
 free from accident, slight and serious, for so long. The investigations carried 
 out by the late Mr. Peter Sampson when assistant to Mr. Traill brought to 
 light an enormous amount of useful information for the guidance of engineers 
 generally, as well as for their own, in drawing up the Rules. That they then 
 viewed all steel with a suspicion that did not seem warranted was the subject 
 of much regret by those who felt sure that good steel could and would be 
 produced at a cost which must, in course of time, drive wrought iron out 
 of the market, if it was given the same freedom as accorded to wrought iron 
 of every make and sort. Nevertheless, it must be admitted now that in thus 
 keeping all steel under strict surveillance, much that was bad was prevented 
 from coming into general use, and all makers of the material were thereby 
 compelled to exercise the greater care in manufacture and treatment, which 
 now permits of the greater freedom in its use. At the same time, there has 
 been, and still is, thoiigh in a lesser degree, the regret that more encourage- 
 ment is not given to those manufacturers who honestly strive to provide 
 good and safe material for engineers' general use, having virtues superior 
 to those of the common sort. In a general way, perhaps, low tensile steel 
 is safer than the higher kinds, but it does not, therefore, follow that no high 
 tensile steel can command the same confidence reposed in it that obtains with 
 the " best mild steel." In this respect their progress has been slow in the 
 mercantile marine. The high tensile steels are employed very sparingly 
 compared with what might have been the case had the restrictions on their 
 use been exercised in accordance with their real merits, instead of in com- 
 pliance with the policy of the " Board." It is, of course, freely admitted 
 that a Government Department has to be very discreet in its actions when 
 dealing with one manufacturer and another, especially seeing that the desire 
 is generally of all concerned for the practice by all officials to be uniform and 
 absolutely impartial ; that the treatment of all shall be on the same lines 
 without differentiation, baseless or otherwise. In the use of an old and 
 tried material like cast iron, however, there are not the same conflicting 
 causes for singular treatment by the Board of Trade and its officials ; yet, 
 
222 MANUAL OF MARINE ENGINEERING. 
 
 while at one time nearly the whole of the engine was made of cast iron, and 
 some of the most important parts are still formed of it, where it could be 
 squeezed out, it has been by the policy of both Admiralty and Board of Trade 
 of late years to do so. Yet this material has some distinct virtues which render 
 it not only a convenient but a safe one for the composition of certain parts. 
 Under tensile stress it stretches more than steel does up to the elastic limit ; 
 to resist compression it has no equal, so that some structures which are liable 
 when under load to be subject to a considerable amount of compression 
 are really better made of it than of steel. Notwithstanding this, however, 
 the tendency is to avoid its use, largely due to the policy and rules of these 
 Government Departments on the ground of its comparative brittleness and 
 liability to crack. The continued testing by expensive methods of the steel 
 material, both in the cast, forged, and rolled state, as now used in engine 
 construction, has tended to retard the use of it for a considerable period 
 after steel makers had found the means of producing it cheaper than wrought 
 iron. Since very little, if any, reduction in scantlings was claimed by engineers 
 when substituting steel for wrought iron in many parts, it was of little or no 
 consequence what its ultimate tensile strength was ; it was, however, impor- 
 tant, and is still necessary, to be assured that it is tough and ductile. This 
 could be quite certainly and satisfactorily ascertained by submitting samples 
 to a simple cold-bending test, and further to punching and upsetting tests, 
 such as applied to rivets. Had such simple and inexpensive means been 
 adopted, this superior material might have taken the place of common wrought 
 iron, and even of cast iron, in many places long ago. 
 
 The restrictions of Government Departments, as a rule, tend to stag- 
 nation in engineering practice, while fulfilling the function for which they 
 exist to protect the public from the dangers that might arise if greater freedom 
 were given to engine builders. But neither at the Admiralty nor at the 
 Board of Trade is there now raised that dead wall to rational and desirable 
 progress ; indeed, the Admiralty during the past two decades have become 
 almost too progressive for the economic working of those who serve them. 
 
 The Balancing of Engines and Avoidance of Vibration is a distinguishing 
 feature in latter-day marine engineering practice, and the advent of the 
 turbine on shipboard has rather accentuated the necessity for the exercise 
 of the art of balancing than abrogated it ; for, although that instrument 
 itself is free from the inertia defects of the reciprocating engine, it requires 
 nevertheless the most careful of balancing itself ; but to compete with the 
 turbine-driven steamer, that one having reciprocating engines must have 
 them so that they run with an absence of vibration as far as is possible. 
 Thanks to that veteran engineer and practical scientist, Dr. Otto Schlick. who 
 gave us so freely the benefit of his years of patient labour and careful research, 
 we understand now, as we did not before, the reason for and the causes of 
 all vibration in a steamship, and, better still, he had devised means for 
 correctly ascertaining them and curing their effects, for which we were more 
 indebted to him. Formerly it was not uncommon to attribute all the vibra- 
 tion of a screw ship to the action of the propeller ; that some of it might be 
 due to the momentum of the moving parts of the engines was, of course, 
 apparent to all engineers having even an elementary knowledge of dynamics ; 
 attempts of a feeble and tentative kind were made to check them by fitting 
 balance weights to the shafts opposite the cranks of the horizontal engine, 
 
BLADES OF A SCREW. 223 
 
 or by casting with the turning wheels balance weights of segmental form. 
 No doubt the inertia effects of the pistons and rods were to some extent 
 neutralised by these balance weights, but it would appear that what the 
 makers of these engines really aimed at was merely to balance the weight 
 of the crank-pin, arms, and connecting-rod ends rather than the horizontal 
 forces due to the movement of the pistons, etc. Balance weights were very 
 seldom fitted to vertical engines, notwithstanding that from these statical 
 considerations they were really the more needed in them than in the hori- 
 zontal. 
 
 Nowadays balance weights are freely used in all classes of engine and 
 for all sizes, but the application of them is effected with more discretion 
 and judgment than formerly prevailed, so that much better effects are 
 obtained with considerably less material and cost. The four-crank engine, 
 balanced on the system introduced by Dr. Schlick and perfected in this 
 country by Messrs. Tweedy and Yarrow, is without added weights, and now 
 in general use ; an almost perfect balance with an absence of vibration is 
 obtained by this system with the special arrangement of the angles of cranks 
 to suit the momenta of the moving parts. 
 
 Whether the engine be a three-, four-, or five-crank one, it must be satis- 
 factorily balanced in every warship and express passenger steamer, and even 
 in the cargo steamers, which may and often do convey passengers, it is 
 desirable to avoid unnecessary vibration. 
 
 That Vibration arises in part from the Action of the Screws is only too true 
 now as it was formerly, for the same causes exist in full force to-day not- 
 withstanding our better knowledge of the subject, thanks to Dr. Schlick, 
 for, after the engines have been most carefully and perfectly balanced, there 
 often is manifest a residual vibration or trembling, which, while it may be 
 quite local and not general, is nevertheless disagreeable. Moreover, the same 
 defects are observable in as pronounced a manner in turbine-driven steamers 
 notwithstanding the uniformity of torque and absence of any unbalanced 
 inertia forces. 
 
 There are various ways in which a screw may set up vibrations, all of 
 which, even the smallest, is capable of creating the evil if its application 
 is intermittent and regular and its periodicity coincides with that of the 
 natural vibration of the ship as a single structure, or of any part of it which 
 is free to vibrate. Under such circumstances of synchronism quite small 
 and insignificant forces are capable of producing enormous results, as is well 
 known to all engineers, and the most familiar illustration is perhaps that 
 of a disciplined force crossing a bridge, who, if walking in step, are liable to 
 set up dangerous oscillations in the most substantial of structures. 
 
 If the Blades of a Screw are of Different Pitch, especially at or near the 
 tips, the pressure on opposite blades is not the same, the propeller will be 
 running out of balance, and as each blade should take its due proportion of 
 thrust to run in perfect balance, such differences in pitch or blade surface 
 Avill easily cause vibration to be set up as the true centre of pressure will not 
 coincide with the shaft axis, but have an orbit of its own, the centre of 
 which is on that axis. The desire for the utter absence of vibration and the 
 economy of high efficiency have induced some engineers to have the blades 
 of propellers chipped and ground, so that, not only are the working faces 
 smooth and true to pitch throughout, but each blade is of equal pitch with 
 
224 MANUAL OF MARINE ENGINEERING. 
 
 the others. In spite of the time and cost of carrying out this refinement, 
 there is ample justification when the revolution is high and the ship a fast 
 express steamer, or for good gunnery. For a cargo steamer it is, of course, 
 quite desirable that the blades shall be of equal pitch, and, as far as possible, 
 by care in moulding, to have a uniformity, but there is not the same warrant 
 for chipping and grinding in their case. The same remarks apply to screws 
 that are damaged by bending or breaking of the tips. It is, however, aston- 
 ishing how badly damaged a screw may be and yet do its work fairly well. 
 
 The Screw is always working in Water disturbed more or less by the passage 
 of the ship herself, as also by the currents set in motion by the suck of the 
 screw. The former, known as wake currents, cause a difference of pressure 
 on the upper and lower blades of the screw ; the result is, therefore, pretty 
 much the same as in the case of a screw with a blade out of pitch, except 
 that in this case the real centre of pressure has no orbit but remains constant 
 outside the shaft axis. The upper or surface current in wake of a ship 
 has more forward motion than lower ones, and moreover is often quite a 
 distinct layer of water flowing over practically still water, each blade in 
 turn comes into it suddenly, and receives a sudden accession of load and a 
 tendency to retardation of motion ; if the period of blade stroke on this 
 stream coincides or synchronises with the periodicity of the ship, it will soon 
 set up quite sensible vibrations, and, if there is no break in the timing, they 
 may become quite violent. If the synchronism is not perfect they will 
 damp out and disappear, only to reappear later with equal violence. This 
 kind of disturbance causes horizontal vibrations, which were very pronounced 
 in ships having only two blades to a single screw ; it was, and is, still noticeable 
 with single-screw ships with four-bladed screws, especially when the tips are 
 broad. 
 
 The Proximity of Screws to Portions of the Hull is also a common cause 
 of vibrations, especially of the small or residual ones, that are often as trouble- 
 some as the larger ones. This is due to more than one agency. If the blade 
 tip passes closely to a fixed obstruction, such as the stern frame, or the hull 
 of the ship in the case of multiple screws, its resistance or thrust pressure 
 is for the moment changed, and as each blade in turn does so, there is pro- 
 duced a series of impacts, which in themselves may be slight and almost- 
 imperceptible, yet may in their collective action, when synchronising with 
 some massive portion of the ship, produce in it serious activities. There is 
 always a film of water of quite sensible thickness dragged on by friction 
 with the skin of the ship at a velocity not much inferior to the ship itself. 
 If, therefore, the blade tip is so close as to come into that envelope of inert 
 water from water that was flowing past the screw at the rate of 2,000 to 
 3,000 feet per minute, it would be almost as bad in its effect as if it struck 
 a solid, and the blow each time a blade tip passed through would be, and 
 probably is, very considerable, and the result is manifest on Dr. Schlick's 
 interesting pallograph diagrams. 
 
 Then such obstructions as the stern frame and brackets, spectacle pieces, 
 etc., act as brakes to the spiral currents from the screw tips, so that as each 
 blade passes near it there is the change in direction of flow, which, although 
 very slight in its effect, may in the cumulative form be very real. 
 
 The Power of Sucking in Air, so beautifully and clearly demonstrated 
 by Professor Flamm, and experience by each observer of the screw in 
 
THE INTRODUCTION OF STEEL CASTINGS. 225 
 
 everyday practice, may, and probably does, lead to a considerable amount of 
 vibration. It is well known that when air is drawn into the screw race there 
 is a diminution of thrust, and there may be, and often is, a momentary 
 cessation of it ; if this becomes periodic, and the effects cumulative, a most 
 unpleasant form of hull disturbance will be experienced. This air suction 
 occurs the more readily with screws whose blade tips are near the surface, 
 and even when there is a fair amount of immersion in still water the suction 
 or feed to the propeller when running fast produces a hollow or depression 
 of the surface, so as to render air "spouts" easy of formation. The same thing 
 happens when there is quite a gentle swell on, so that the heave and pitch 
 of the ship causes the blade tips to approach too close to the surface. 
 
 Cavitation may also be a cause of Vibration when the screws are quite 
 thoroughly immersed, especially when the screw is driven by a steam engine 
 with a torque of great variation, or with an internal combustion engine 
 with the same or worse defect, so that the propeller has sudden and severe 
 acceleration in angular velocity, whereby the pressure per square inch is 
 increased beyond the limit of good working, so that cavities are formed 
 by the reduction in pressure behind the blades, permitting of air separating 
 from the water and forming masses of bubbles, which, when liberated, may 
 have the same effect on the screw as the air drawn down from above by 
 " suction." 
 
 AH these Things may cause Vibration, and always are sources of loss of 
 efficiency. The former can be generally reduced to a minimum, if not alto- 
 gether damped out, by causing the engines to run at a rate of revolution 
 that precludes synchronism with any of the important masses of the hull. 
 The efficiency of propeller and hull can be little effected by such means, but 
 doubtless the hull that does not vibrate is likely to permit of more gross 
 power being applied to propulsion than one that does. 
 
 The Necessity for Perfect Balance of Engine and a quiet-running Screw 
 is imposed on the designer, and he must study the questions involved as 
 carefully as formerly he had to do those involving the safety and economic 
 running of the machinery on shipboard. 
 
 The Auxiliary Machinery must also be free from Vibration and unevenness 
 from the power to produce it ; for to-day, with turbines and the beautifully 
 balanced reciprocators, it would be grotesque to find the ship vibrating from 
 a donkey pump's unbalanced forces ; and now that air pumps, as well as 
 all other pumps, are separable from the main engines, it is incumbent on 
 their makers to supply them free from a vice, from which the main engines 
 have been eradicated. The makers of electric-power generating engines have 
 been compelled long ago to balance them so that they create no nuisance 
 to the neighbourhood in which they are situated by themselves vibrating 
 or causing any tremble in the neighbouring houses ; therefore, as a rule, the 
 makers of these engines who have had such shore experience can be trusted 
 to supply them on shipboard quite free from vice. 
 
 The Introduction of Steel Castings was hailed with delight by engine designers 
 thirty-five years ago, as it raised the hope that the weight of machinery 
 generally and the cost of many of its parts would be very materially reduced, 
 and that in other parts- the risks run consequent on the necessary employment 
 of cast iron, and even with cast bronze in their manufacture, would cease. 
 To-dav the engine designer is still living largelv in the same hope, inasmuch 
 
 15 
 
226 MANUAL OF MARINE ENGINEERING. 
 
 as the sound, perfect casting at moderate cost in a metal strong and tough 
 and capable of working at quite a high temperature remains yet to be pro- 
 duced in quantity and promptly. British steel founders can, and do, make 
 huge castings, which for ship work are invaluable ; they also supply engineers 
 with some very fine and useful ones, but they had to look to another country 
 for that soundness and uniformity in quality so much to be desired by those 
 who are responsible for and take the risks of such things. It is a matter oi 
 great regret that there lacks sufficient certainty in the home product to 
 encourage its more extended use for such parts of an engine as must be 
 sound or machined all over. This is a disappointment to those who have 
 looked year after year for the improvement so long hoped for. 
 
 The Use of Aluminium in engine construction has been somewhat delayed 
 from somewhat similar causes, in spite of the fact that the material was 
 sold at such a low price as to permit of its free use in engine construction. 
 Now, however, the very beautiful castings supplied for motor car machinery 
 must induce the marine engine designers to use them more freely in their 
 products. The metal now used is an alloy, and is only a trifle heavier than 
 pure aluminium ; the castings are sharp and sound in outline and clean in 
 surface ; the tensile strength is sufficient, being as good as ordinary bronze. 
 It will not, however, withstand the action of sea water as does bronze, so that 
 it cannot be used for the parts exposed to it, but there are very many other 
 parts which are quite free from contact with sea-water, and others where a 
 coat of paint will give sufficient protection. 
 
 Duralumin, a 90 per cent, alloy of aluminium introduced by Vickers, 
 Ltd., has such good qualities as soon to command the attention of the 
 marine engine designer, inasmuch as it is quite strong (28 tons tensile with 
 15 per cent, extension), and will resist shock ; it is very light (sp. gr., 2-8) 
 and malleable, and can be treated as the zinc bronzes are, by drawing out into 
 bars of various sections, also rolled into plates and sheets, and sold at quite 
 moderate prices, so as to be extensively used in the construction of air-craft, 
 motor cars, etc. ; it cannot, however, be used for castings as yet. 
 
 In the Future the Reciproeator will probably undergo more changes, 
 necessary to its success in competition with the. steam turbine ; and also that 
 it may hold its own with the oil engine, it may be necessary to follow some 
 of the features peculiar to that engine. For example, the oil engine at present 
 has certain limitations, which hedge it in completely. The steam engine 
 maker may impose on himself, if he chooses, some of them with advantage. 
 For instance, the huge cylinders of the present triple and quadruple engine 
 may give place to a multiplicity of smaller ones ; so small may they be indeed 
 that the low-pressure ones may have relatively such large ports and valves 
 as will enable them to benefit usefully by the high vacua now attainable in 
 condensers on board ship at a trifling extra cost. 
 
 Superheating of steam to an extent now practically impossible also may be 
 followed with complete success in small cylinders, including a marked reduc- 
 tion in the consumption of fuel. There may be even modifications in the 
 methods of reversing the engines, whereby there will be effected savings in 
 first cost, as well as economy in working costs The flexibility of the reci- 
 procating steam engine is a strong and marked feature in the eyes of the 
 marine engineer, especially of him who has to work and be responsible for 
 the machinery of a ship : it should count likewise with owners, underwriters, 
 
FUTURE OF THE REC1PROCATOR. 227 
 
 •and " all those who go clown to the sea in ships " as a fair set off for such 
 •savings in fuel as claimed, even though they be substantial ; especially 
 •shoidd this consideration weigh heavily in the counsels of citizens of a country 
 having an abundant supply of excellent coal but little oil, and that little 
 limited to one or two remote districts. To have to import food into a country 
 is regrettable, but to import fuel where it can be avoided is folly. If our 
 ships are to be largely fitted with internal combustion engines, or have boilers 
 capable of burning oil fuel only, its state in case of a war with a country 
 having a powerful fleet will be a perilous one, and lead to a worse disastei 
 than a temporary shortness of food. 
 
228 MANUAL OF MARINE ENGINEERING 
 
 CHAPTER X. 
 
 THE CYLINDER AND ITS FITTINGS. 
 
 The Cylinder is the most Important Part of the reciprocating engine, for on; 
 it every other part is more or less dependent. The capacity of the engine 
 for developing power — that is, for converting the energy of the steam into 
 mechanical work — depends on its size and efficiency. The efficiency of the 
 cylinder for this purpose will depend largely on the design of the various 
 ports, passages, and valves, while its mechanical efficiency will be affected 
 by design and the quality of the material and workmanship of the cylinder 
 and its parts, inasmuch as the losses in this member of the machine are chiefly 
 due to friction aggravated by thermal conditions not present in other portions 
 of mechanism. It is, therefore, very highly necessary that the utmost atten- 
 tion is bestowed in the first place on all the problems involved in the deter- 
 mination of size, the general design of it, and all its working parts, and the 
 greatest care in the manufacture, both in the workmanship and the selection 
 of materials. Finally, in the working of the engine, judgment as well as care 
 is necessary in fitting and adjusting them, and constant attention during 
 service, that the highest efficiency may be obtained. 
 
 The Size of the Cylinders can be calculated on first principles in quite a 
 simple way, but to determine the best size, or that most suitable for each par- 
 ticular case, requires something more ; it requires special consideration of 
 all the circumstances involved in the case. In a general way, it may be 
 taken for granted that the greatest economy, both in prime cost and in working 
 expenses, is to be attained, with the smallest capacity of cylinder in which 
 the maximum horse-power guaranteed by the builder can be developed. 
 There are, of course, exceptions to this as to other equally useful practical 
 rules, such exceptions must, of course, receive exceptional treatment. 
 
 The marine engine differs in many important respects from the land 
 engine of similar design ; in one respect the conditions of service are very 
 dissimilar, for whereas the land engine usually runs at a fixed rate of revolu- 
 tion when employed for continuous work, as in driving mills, dynamos, etc.. 
 the marine engine varies in revolution as the speed of the ship changes. 
 Further, the load on a land engine may, and often does, vary from nothing 
 to the maximum power with only a very slight, and that a temporary, vari- 
 ation in rate of revolution ; whereas in the marine engine the gross horse- 
 power developed varies roughly as the cube of the rate of revolution — that 
 is, if a marine engine is slowed down by throttling or " linking up " to half 
 the rate of revolution at full power, the horse-power will be only one-eighth 
 of it. Even a reduction of 10 per cent, in speed of ship and revolution 
 involves a reduction in power of no less than 27 per cent. Now, as the marine 
 engine, even in express steamers, is seldom worked at full power, that 
 
ECONOMY IN PRACTICE. 229 
 
 developed on service is generally well below the full capacity of the engine ; 
 the periods of full power development are few and of short duration, and 
 of such a nature is the service that demands them, that the cost of producing 
 the power is insignificant compared with that of the general working, so that 
 if extravagant by comparison with the ordinary expenditure, it is, never- 
 theless, quite warranted. In other words, so long as the increase in power 
 is obtainable on demand, the cost of it is of little consideration ; conse- 
 quently the efficiency under those conditions may be quite low, if by these 
 temporary sacrifices the gain in efficiency under the normal conditions of 
 service is thereby attained and substantial. High efficiency during the long 
 periods of service is, therefore, the first consideration of the designer, as well 
 as that of the purchaser of the engine, and just as electrical engineers on 
 shore demand engines which can on an emergency, and for a short time, 
 develop something like 20 per cent, more than their normal maximum output, 
 and call it " overload," so the marine engineer should be content with such 
 overloads on the same terms for trial trips and spurts at sea, instead of 
 Tequiring a larger engine, which will really run the greatest part of its life at 
 only 75 per cent, of the full power at which it can run, to do that full power at 
 maximum efficiency. 
 
 To the Engines of Warships the same argument may be applied in a general 
 way, but to them some other considerations are applicable which may modify 
 the decision of the designer. Here, in order that the utmost power may be 
 obtained under the more restricted conditions of weight and space, there 
 may be need of the highest efficiency, in order to get the greatest possible 
 output of power from the boilers. But, judging by modern practice in such 
 ships, the small cylinders, with the necessary low rate of expansion for them 
 at full power, prevails to a greater extent than in even the mercantile marine, 
 because a warship seldom runs at even so small a reduction of speed as the 
 10 per cent. ; with her a reduction of speed of 40 per cent, is common, and 
 that means only about 22 per cent, of the full power is required. 
 
 The Larger the Cylinders are the more costly they must be to manu- 
 facture and maintain ; moreover, the clearance losses and those from con- 
 densation will be greater in the larger than the smaller cylinders. Further, 
 the pipes and connections, the rods, valve gears, columns, shafts, and framing 
 vary with the size of the cylinder, many of these parts being influenced by 
 the maximum rather than the mean pressure or load on them. It is, of course, 
 obvious that for a given power any decrease in cylinder capacity must be 
 accompanied by a decrease in rate of expansion, all other things being the 
 same. But all other things are not the same ; the steam pressure from the 
 boilers and the pressure in the condenser are the same, but the steam efficiency 
 in the smaller set of cylinders may be such that the actual mean pressure 
 at the same normal rate of expansion under working conditions is considerably 
 greater than in the larger ones, due to less wire-drawing, radiation, etc., 
 cylinder condensation, and decrease in back pressure due to larger ports, 
 etc. 
 
 Economy is found, therefore, in Practice with rates of expansion lower 
 than theory indicates, and consequently of late years great increases of boiler 
 pressures have not been accompanied by the corresponding increases in rate 
 of expansion such as was formerly anticipated. Hence, for the full powers 
 .required for trial speed the rate of expansion is expressed by p -5- 15, where 
 
230 MANUAL OF MARINE ENGINEERING. 
 
 f is the absolute pressure of the steam supplied to the engine. Following. 
 this rule — 
 
 195 
 Rate of expansion of triple-stage engines steam 180 lbs. = — = 13. 
 
 „ quadruple „ 210 lbs. = ~l = 15. 
 
 ID 
 
 225 lbs. = ~ = 16. 
 lo 
 
 On service such ships will work with rates of expansion of 16 in the triples 
 and 18 to 20 in quadruples. 
 
 In Naval Service and Express Cross-Channel Steamers the rates of ex- 
 pansion are somewhat less, so that their measure is made by dividing the 
 absolute initial pressure by 18. For the triple-expansion engines of such 
 ships with steam at 200 lbs. gauge, or 215 lbs. absolute — 
 
 215 
 Rate of expansion = -r^- = 12. 
 
 lo 
 
 The Service Speed of Short Distance Express Steamers is about 5 per cent- 
 under the maximum or trial trip speed, and since the mean pressure referred 
 to the L.P. cylinder will vary approximately with the square of the speed,, 
 the mean pressure on service will be (0 - 95) 2 , or - 9 of that on trial. Itt 
 designing the engine, therefore, it should be borne in mind that it is desirable- 
 to have maximum efficiency with a mean pressure 10 per cent, below that 
 of the maximum possible from the cylinders, and as a corollary the cylinders 
 must be large enough to develop 11 per cent, more power than wben on 
 service — that means the "overload" is 11 per cent, in pressure, but it will 
 be 16*7 per cent, of the I.H.P. 
 
 Instead of Speed Margin with such Ships, it would be better to have a 
 power one, for it is seen that with only 5 per cent, of the former there is 
 16'7 of the latter, and, after all, it is more important in a general way to 
 have a reserve of power, which will ensure maintaining the service 
 speed. 
 
 With Cargo Steamers (here is a Disturbing Element not experienced with 
 the warship or passenger steamer — viz., the difference in draft of water 
 with different cargoes, and the possible difference in speed of revolution — 
 in fact, it is equivalent in effect on the engines of a change in size of propeller. 
 The engines can now move at a much higher rate of revolution with a smaller 
 mean pressure of steam, and unless checked by throttling or "linking up." 
 proceed at a rate far in excess of that anticipated by the designer, and pro- 
 vided for by him. Under such conditions the propeller is very liable to 
 sudden and severe racing, in spite of governors and hand gove*nin£. Such 
 engines must, therefore, for safety sake, have in themselves that which goes 
 toward restraint under such conditions. 
 
 The Indicated Horse-power of an engine is the product of the mean pres- 
 sure, the piston area in square inches and feet passed through by the piston 
 in a minute divided by 33.000. It may be said, therefore, that the power 
 depends on the area of the piston on the speed of piston and the pressure on 
 
DIAMETER OF CYLINDERS OF A MARINE ENGINE. 231 
 
 it. With a given engine the two latter may be varied arbitrarily, inasmuch 
 as the pressure may be increased or decreased by the throttle valve at will, 
 or may be altered permanently by changing the cut-off of the distributing 
 valves on the cylinders ; the speed of piston, all other things remaining the 
 same, being dependent on it will vary with the mean pressure. But with 
 the same mean pressure the piston speed may be permanently reduced by 
 increasing the pitch of the screw. When designing an engine the mean 
 pressure aud speed of piston intended must be decided on in order to cal- 
 culate the area of piston. As a matter of fact, it is really the rate of revolu- 
 tion that must be decided on as a basis of calculation for a reciprocating 
 engine, just as it is for a turbine. With a vertical engine under ordinary 
 conditions the length of stroke may be generally of any reasonable amount, 
 so that the real limit, after all, is the practical one of how fast may a piston 
 move with safety and consistent with continuous good working. 
 
 It must not be overlooked that the maximum velocity of movement of a 
 
 piston is ^, or 1*571 times the mean, so that if the mean speed is 1,000 feet 
 A 
 
 per minute, the velocity of the piston at about the middle of the stroke is 
 
 at the rate of 1,571 feet. 
 
 The Diameter of Cylinders of a Marine Engine is ascertained by first deter- 
 mining that of the low pressure from the referred mean pressure and speed 
 of piston decided on, and making the high pressure of such a size as to permit 
 of that mean pressure with the boiler pressure provided. The generality 
 of engines now in use are of the compound type, triple or quadruple. The 
 rate of expansion will be chosen with regard to the conditions under which 
 the engine has mostly to work. In a general way it is true, as already stated, 
 that the smaller the cylinders are the better, both in prime cost and economic 
 working. It has been shown that the consumption of steam in Naval reci- 
 procating engines at full speed per I.H.P. per hour is somewhat higher than 
 at a somewhat reduced one, while, on the other hand, at low speeds, notwith- 
 standing the higher rate of expansion obtaining, the consumption is often 
 even greater than at full speed. This means that with the comparatively 
 small cylinders the economy at an extravagant rate of expansion is better 
 than that in the same cylinder at a high and presumably economic rate. 
 
 In the mercantile marine the lowest rate of steam consumption per horse- 
 power is also often at a higher rate of expansion than that at the maximum 
 power, but not much more so. With the cargo steamer, where economy 
 of fuel is of prime importance, and the engines are run nearly always at full 
 speed, the consumption of steam per I.H.P. is a minimum, or nearly so at 
 that speed. Mr. Parsons found that the steam consumption of the triple- 
 expansion engines of the s.s. "Vespasian" at 70 revolutions was 16*9 lbs. 
 per horse-power hour, at 65 revolutions it was 17 '6 lbs., and from that the 
 rate gradually rose as the revolutions fell, till at 50 revolutions it was as much 
 as 19 - 8. At 65 revolutions the rise was small, as it was only 17 '6 lbs., and, 
 assuming the same rate of slip, the following comparative results are true. 
 The total consumption of water at 70 revolutions was 17,250 lbs. per hour, 
 while at 65 revolutions it was 14,200 lbs. For the same distance, therefore, 
 if at 70 revolutions the consumption is 17,250, while when done at 65 revolu- 
 tions it will be only 15,140 — that is, as 1*14 to 1*0 — or a saving of 14 per cent, 
 in spite of an increase per I.H.P. of 4*14 per cent. 
 
232 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Results op Trials of Various Ships, showing Consumption of Steam 
 and Fuel in Lbs. per H.P.-Hour. 
 
 H.M.S. Hn , 
 Max. I.H P.. , 18,500 I.H.P. 
 Engines Recipros. 
 
 H.M.S. Cn., 
 23.500 I.H.P. 
 Recip. Trpls, 
 
 H.M.S. Tz., 
 
 9,800 I.H.P. 
 Reeip. Trpls. 
 
 U.S A. Dlw., 
 28,500 I.H.P. 
 Recip. Trpls. 
 
 H.M.S. As., H M.S. Hlr 
 24,000 I.H.P. 10,400 I.H.P. 
 Recip. Trpls. Recip. Trpls. 
 
 I.J.N ilk.. 
 
 27.000 S. If. P. 
 Curtis Tribs. 
 
 
 Steam. 
 
 Coal. 
 
 Steam. 
 
 Coal. 
 
 Steam. 
 
 Coal. 
 
 Steam. 
 
 Coal. 
 
 Steam. 
 
 Coal. Steam. 
 
 1 
 
 Coal. 
 
 Steam 
 
 Oil. 
 
 Full power, ' 17-2 
 50 to 75 %. 15-2 
 20 to 10%, 15-1 
 
 1 
 
 1-80 
 1-76 
 1-94 
 
 16-0 
 14-4 
 15-3 
 
 1-99 
 2-04 
 2-00 
 
 13-91 
 15-45 
 16-25 
 
 2-65 
 
 2-28 
 2-06 
 
 • 
 21-0 
 18-7 
 22-0 
 
 ... 
 
 • 
 19-9 
 15-37 
 16-95 
 
 2-03 I .. 
 
 1-85 
 
 1-88 
 
 1-407 
 
 1-49 
 
 1-64 
 
 13-77 
 
 14-76 
 20-57 
 
 1-10 
 
 1-27 
 1-80 
 
 * This consumption was for all purposes. 
 
 Max. S.H.P., 
 
 Engines, 
 
 R.M.S. La., 
 64.600 S.H.P. 
 Direct Turbs. 
 
 U.S.A. Nvd. 
 23,3Q0 S.H.P. 
 Geared Turbs. 
 
 U.S.A. Fda. 
 40,500 S.H.P. 
 Direct Turbs. 
 
 U.S.A. Pkn., 
 11,700 S.H.P. 
 Curtis Turbs. 
 
 U S.A. Esn., 
 17,800 S.H.P. 
 Geared Turbs. 
 
 \ 
 
 R.M.S. Ok., i U.S.A. Clfa. 
 
 6,900 S.H.P. 29,000 S.H.P. 
 
 Mixed R. & T. 1 Electric Turb. 
 
 
 Steam. 
 
 Coal. 
 
 Strain. 
 
 Oil. 
 
 Steam. 
 
 Coal. 
 
 Steam. Oil. Steam. 
 
 Oil. 
 
 Steam. 
 
 Coal. 
 
 Steam. 
 
 Oil. 
 
 Full power, 
 50 to 75 °/ , 
 10 to 20%, 
 
 12-77 
 13-90 
 21-23 
 
 1-43 
 1-56 
 2-52 
 
 15-40 
 16-16 
 22-95 
 
 1-32 
 1-30 
 1-72 
 
 13-40 
 23-77 
 
 1-59 
 
 2-23 
 
 14-49 
 15-50 
 21-49 
 
 1-42 
 1-40 
 1-67 
 
 15-76 
 17-40 
 28-75 
 
 1-23 
 1-22 
 2-03 
 
 14-12 
 14-10 
 15-20 
 
 • • 
 
 11-90 
 11-10 
 14-60 
 
 
 It is to be observed that, whereas a marine engine slows down as the 
 rate of expansion increases, a similar engine on shore driving a mill or dynamo 
 will run at constant speed. This is important, because economy of steam 
 consumption is very much influenced by the rate of flow through the cylinders, 
 just as it is in the turbine, the low consumption of which is in no small measure 
 due to the rapidity with which the steam passes from entrance to the con- 
 denser. In a reciprocating engine, when at 150 revolutions, and of the triple- 
 expansion type, the steam takes 0-8 second to pass through the system, 
 but at 75 revolutions the time is 1-6 seconds. 
 
 Mean pressure is modified in quite a satisfactory way by linking up ; 
 for example, the s.s. " Al." was intended to work at sea with about 1.000 
 H.P. ; on trial trip at full speed 1,232 I.H.P. was developed. The following 
 shows how this reduction was effected by simply notching up to the extent 
 of two turns of the hand wheel, and what were the consequences : — 
 
 TABLE XXX.— Trials of a Typical Cargo Steamer. 
 
 
 Steam. 
 
 Vacuum. 
 
 Revs. 
 
 Mean Pressures. 
 
 Indicated Horse-Power. 
 
 Total. 
 
 High. 
 
 Medium. 
 
 Low. 
 
 High. 
 
 Medium. 
 
 Low. 
 
 Trial Runs, 
 
 full, 
 Sea spocd, 
 
 linked, . 
 
 145 
 145 
 
 25 
 25 5 
 
 64-2 
 581 
 
 50-4 
 452 
 
 24 51 
 
 22-9 
 
 7 91 
 
 7 20 
 
 3857 
 3134 
 
 434 5 
 367 2 
 
 4121 
 339 9 
 
 1 : 232 3 
 1,020 5 
 
 The coal was weighed during these trials, and the consumption was found 
 to be at the rate of 1*51 lbs. per horse-power at full, and 1*55 lbs. at 
 the reduced speed. Now, had smaller cylinders been provided, and the 
 
DIAMETER OF THE L.P. CYLINDER. 233 
 
 revolutions raised from 58 to 64, the rate of consumption would not have 
 exceeded the 1'51 lbs., but probably have been less. 
 
 The Back Pressure in the L.P. Cylinder should be not more than 15 lbs. 
 greater than that in the condenser — that is, with a vacuum of 28 inches it 
 should not exceed 2"5 lbs. absolute. In a well-designed low-pressure cylinder 
 at moderate piston speeds the vacuum in it at exhaust should be 95 per cent, 
 of that in condenser, when that is not more than 28 inches ; 93 per cent, is 
 common experience with good engines when working with a vacuum of 
 27*0 inches at full speed. For purposes of calculation, therefore, an allow- 
 ance of 90 per cent, will be quite on the safe side. With modern air pumps 
 and a good condenser 28 inches can be maintained even in the tropics ; 90 per 
 cent, of this is 25*2 — that is, the back pressure is about 2i lbs. If the ship 
 is likely to see much service in the tropics, 27 inches will be safe, and 90 per 
 ■cent, of it is 24*3 inches, so that the back pressure then will be 2 "85 lbs. With 
 a turbine the back pressure in it is very nearly that in the condenser top, as 
 there is practically no exhaust pipe, and certainly no valve obstructions. 
 For general traders the back pressure should be assumed at 3 lbs., and for 
 those other ships casually visiting the tropics 2A lbs. is sufficient allowance 
 for calculations. 
 
 Example (1). — To ascertain the probable mean pressure referred to the 
 L.P. cylinder of a passenger steamer whose service will be through the tropics, 
 the boiler pressure to be 185 lbs. 
 
 Here the initial absolute pressure is 200 lbs. 
 
 The rate of expansion 200 -5- 15 = 13-33. 
 
 Referring to Table xvii. for this rate of expansion, the ratio — = 0*269. 
 Back pressure is 3. Pi 
 
 Then theoretical mean pressure = (200 X 0-269 — 3) = 50*8 lbs. 
 
 The actual mean pressure will be found by multiplying this amount by 
 the factor given for such engines on (Table xviii., p. 125), say, 0*7. Thus : — 
 
 Probable mean pressure = 50 - 8 x - 7, or 35'56 lbs. 
 
 The Diameter of the L.P. Cylinder may be found* then as follows :— When 
 p m is the referred mean pressure. Let d be the diameter in inches, R the 
 number of revolutions, and S the stroke in feet. 
 
 T , , ■ . d 2 I.H.P. x 33,000 
 
 Then area of piston = «- = ^ x r x 2S • 
 
 flufe-Diameter L.P. cylinder = / I-H.P. X 21,000 = 7 I.H.P. X ,42,q gT 
 
 V pm X S X R y p m X piston speed 
 
 If the stroke is not decided, S X R is half the mean piston speed. 
 
 The volume of steam at cut-off for a rate of expansion E must be = capacity 
 •of L.P. cylinder -•- E. 
 
 The equivalent cut-off in the H.P. cylinder for this will be expressed 
 as a fraction of the stroke = ratio of cylinders -=- E. 
 
 If the cut-ofi is to be early, as it should be, to avoid excessive " drop/ 
 
 • The X.E. Coast Inst. E. aiul S. Rule for this in cargo ships is : — 
 
 ;j HP v -'l-a 
 D — \/ * * ; the I.H.P. is that developed while on voyage, 
 
 b + 4 
 
234 MANUAL OF MARINE ENGINEERING. 
 
 it will be seen that the ratio of cylinders must be smaller than if a later cut- 
 off were admissible or desirable. 
 
 Then, if x be the cut off as a fraction of the stroke, and clearance is 
 neglected 
 
 Then the ratio of L.P. to H.P. cylinder = E X x, 
 or x = cylinder ratio -e- E. 
 
 Taking the cut-off at 0*55 of the stroke, the ratio of L.P. to H.P. will be 
 
 0'55 X rate of expansion, and if that is taken as absolute pressure -=-13 for 
 
 economic engines. 
 
 ^ , -^ . , TT . TTT , ,-. - 55 X absolute working pressure 
 Rule. — Katio of L.P. to H.P. cylinders = — — „ 
 
 absolute working pressure 
 ° r = 23^6 ' 
 
 Example. — To find the sizes of cylinders for a ship whose engines are to 
 develop 3,600 I. H.P. on the conditions named before, with the piston speed 
 at a mean of 700 feet per minute. 
 
 Here the diameter of cylinder . . = \ - '_ _ „ z^-—- = 78 inches. 
 
 ,K -56 x 700 
 
 /3.600 > 42,000 
 
 "V"35-£ 
 
 0'55 X 200 
 Ratio of L.P. to H.P. cylinder . = — = 73. 
 
 /78 2 
 Diameter of H.P. cylinder . .. = */— = 
 
 15 
 
 28-8. 
 
 The Size of Intermediate Medium-Pressure Cylinder to avoid drop should 
 be larger than it would be if determined by such practical considerations 
 as low initial loads to avoid shock, variation in temperature, etc. The true 
 mean between the L.P. and H.P. cylinder would be as follows : — Where d 
 is the diameter of the H.P., and D that of the L.P. cylinder. 
 
 / jo TY2 
 
 Diameter of M.P. cvlinder 
 
 =V" 
 
 This, however, is too large, and not in accordance with practice ; the 
 following rule may be followed, therefore, with triples : — 
 
 Rule. — Diameter of medium-pressure cylinder = a/ — , 
 
 or, Simply „ „ = {d + D) x 0*45. 
 
 For the engine in the above Example, 
 
 Diameter medium cylinder =a/ — =48*1. 
 
 /28 
 
 For such diameters of cylinder in an ordinary passenger steamer the stroke 
 would be 48 inches. Then the engines would have cylinders 28|, 48. and 
 78 inches diameter, and 4 feet stroke, the revolutions are then 700 -=- 8, or 
 87*5 per minute. 
 
ARRANGEMENTS OF CYLINDERS. 235 
 
 The Arrangement of the Cylinders of a marine engine is now a much more 
 important, as also a more interesting, problem than was the case formerly/ 
 when there were only two of them. To-day, with the multiplicity of cylinders 
 and cranks that are commonly found, especially in oil internat combustion 
 engines, the arrangement is governed by circumstances quite beyond the 
 ken of the older engine builder, who was never troubled with problems 
 involving the nice balancing of the moving parts so as to avoid vibration, 
 or to study the economy of steam piping involved in determining the sequence 
 of cylinders so as to get the best flow of steam from boiler to condenser with 
 the least expenditure on balance weights. 
 
 Now the designer has to face such questions, as well as all that is involved 
 in the division of cylinders to bring them within reasonable limits of size 
 when the power to be developed is large. Large cylinders can be made as 
 well now as formerly, but modern engineers prefer to have two of moderate 
 size in the place of the one of 130 or even 140 inches diameter fitted by their 
 predecessors, even though the highest pressure in them is only half that 
 in these older and larger ones. Moreover, the solid steel pistons of to-day 
 are lighter and certainly safer than the old hollow cast-iron ones, and in the 
 vertical engines they are so very much easier to examine, overhaul, or remove 
 than those of the horizontal engine of old days. Nevertheless, the tendency 
 is to split the L.P. member of a compound system into two or more portions, 
 for thereby a triple-stage engine may have four cranks, and be balanced on 
 the Schlick system, and each L.P. cylinder can have a larger ratio of port 
 and steam passage section, whereby the back pressure in it is less than that 
 obtaining in the single large cylinder, and thereby benefit by a high vacuum 
 in the condenser. 
 
 The Size of the Steam Ports in a Cylinder do not vary in practice with 
 the square of the diameter, but at a somewhat less rate, inasmuch as the 
 transverse measurement may be in proportion to the diameter, while the 
 longitudinal or axial breadth does not increase so rapidly, therefore the 
 larger is the cylinder the smaller is the ratio of maximum port to piston 
 area ; consequently with the same speed of piston the flow of steam is of 
 necessity higher through ports and passages. If, therefore, the number 
 of cylinders be multiplied, so as to keep their sizes comparatively smaller, 
 the higher will be the steam efficiency. The mechanical efficiency, however, 
 will probably be considerably less, on account of the multiplicity of working 
 parts, glands, etc. On the whole, however, within reasonable limits and 
 with modern workmanship and material, the general efficiency of the engine 
 with the larger number of cylinders is higher than any older one with the 
 less. 
 
 There are other influences at work to-day, which incline the marine engine 
 builder to more subdivision of cylinders. The oil engine, with its very 
 high initial pressures and temperatures, requires not only a very strong and 
 simple cylinder, but one of quite a limited diameter. Since the single- 
 acting two-cycle engine is likely to be the type for competition with the 
 turbine and reciprocator on shipboard, it is more than ever sure to have 
 cylinders of limited diameter. The Diesel oil engine (fig. 49), using heavy 
 oil of sorts, and without special ignitors, is looked to as the one specially 
 suitable among internal combustion engines for marine purposes. This 
 engine, with its initial pressure of 550 to 650 lbs. per square inch, and liable 
 
No. 1. 
 
 H.P. 
 
 No. 3. 
 
 No. 2. 
 
 LP. 
 
 M.P. 
 
 H.P. 
 
 
 No. 4. 
 
 H.P. 
 
 No. 5. 
 
 Fig. 73. —Various Arrangements of the Cylinders of Triple- Expansion Engines. 
 
ARRANGEMENTS OF CYLINDERS. 237 
 
 to a temperature exceedingly high, requires to be quite moderate in size and 
 strong in construction. 
 
 Even the steam engine of to-day is subject to quite high pressures, and 
 when the steam is superheated it may have a temperature of 600° F., with 
 a pressure of 230 lbs. per square inch, which requires similar care and 
 limitations as the oil engine, although in a lesser degree, however, to be 
 observed in the design of the H.P. cylinder. Under these circumstances 
 lubrication of the internal working parts is at all times uncertain, and never 
 to be depended on as being positive, hence the smaller the cylinder, the less 
 liability to serious derangement. 
 
 The Arrangements of Cylinders possible for a Triple- and Quadruple-Stage 
 Engine may be studied by referring to their diagrams on figs. 73 to 74. The 
 early triple-expansion engine of Dr. Kirk had always three cranks, with the 
 natural sequence of cylinders (v. fig. 76). The very earliest triple-stage 
 engine, however, had only two cranks, with the H.P. and M.P. cylinders in 
 tandem over one and the low pressure over the other, as in fig. 73. No. 1. 
 
 Inasmuch as the older compound engines were made with the L.P. rods 
 and valve gear much heavier than those of the high pressure, it was found 
 better, when tripling such engines, to fit the third or new H.P. cylinder over 
 the low pressure (fig. 73, No. 2) ; the engine so treated was not only safer 
 but better balanced and easier to start than when the new H.P. was placed 
 over the original H.P. cylinder. 
 
 When the larger two-stage compound engines were tripled two new H.P1 
 cylinders were fitted one over each of the old ones, as shown in No. 3, then 
 the load was evenly divided on the cranks and the combination very quick 
 in handling. Some of the old and large two-stagers were tripled by the 
 addition of a complete new H.P. engine before the old ones, and coupled to 
 them in the best way circumstances permitted. 
 
 The old three-crank two-stage compound engines were tripled by adding 
 a new H.P. cylinder tandem to each low pressure, and some new engines of 
 large size have been made on this plan. 
 
 No. 1 design of fig. 73a is one rather for treating an existing expansive 
 engine having two equal-sized cylinders than to copy for a new engine. If,. 
 however, large power were wanted in small floor space, but with unlimited 
 height, it is quite a suitable arrangement. The same may be said for No. 2. 
 on the same page when larger engines still are wanted on a limited floor 
 space. In this case, however, there is a simplicity in design and limitation 
 in dimensions that are attractive. In fig. 73a, Ex. 2, there are three cylinders- 
 in line of equal diameter ; above them tandem-wise are three other cylinders- 
 in line and of equal diameter. Of these the first is the high pressure of the 
 system, the next two are the medium ones, and the lower three are the L.P. 
 cylinders. The ratio of M.P. to H.P. is thus 2, and as that of the L.P. to the 
 H.P. is 6. that of L.P. to M.P. is 3. As a concrete example, the L.P. cylinders, 
 may be taken as each 60 inches diameter, the H.P. and two M.P. cylinders 
 are therefore = 4x3x 60 2 = 42'4 inches diameter. 
 
 Diagram 3 (fig. 73a) shows the four-crank three-stage compound 
 arrangement of cylinders, with the two low pressure of equal weight of moving 
 parts at the ends, as is the common practice. Diagram 4 shows a variation, 
 with the two L.P. cylinders next one another, with a valve box common 
 to the two and their cranks at right angles, so that the flow of steam isi 
 
238 
 
 MANUAL OF MARINE ENGINEERING. 
 
 M.P. 
 
 M.P. 
 
 H.P. 
 
 No. 1. 
 
 No. 2. 
 
 LiR 
 
 VIP 
 
 HP 
 
 LP 
 
 -^^ 
 
 =#r^*; 
 
 No. 3. 
 Four cranks, L. P. rods reduced. 
 
 MP 
 
 L,P. 
 
 L<P 
 
 Y 
 
 s 
 
 ^^w 
 
 H\R 
 
 1 
 
 Four cranks, all rods same size. 
 Fig. 73a —Various Arrangements of the Cylinders of Triple- Expansion Engines. 
 
ARRANGEMENTS OF CYLINDERS 
 
 M.P. 
 
 H.P. 
 
 No. 1. 
 
 No 2. 
 
 lSt\M.P 2ndiM.fi 
 
 No. 3. 
 Four Sets of Valve-Gear. 
 
 LP 
 
 2ndW.P 
 
 lst\M.P 
 
 H\R 
 
 T^gS^ l | / J i ^^\i^, ^ 7D ,^\ 
 
 3 if 
 
 m 
 
 Q&m 
 
 zSx 
 
 /Vo. 4 
 Two Sets of Valve-Gear. 
 
 Fig. 74.— Various Arrangements of the Cylinders of Quadruple-Expansion 
 Two-, Three-, and Four- Crank Engines. 
 
240 
 
 MANUAL OF MARINE ENGINEERING. 
 
 practically continuous through quite short exhaust pipes, and the flat L.P. 
 valves are fitted with a balancing frame between them. 
 
 The Arrangement of Cylinders of the Quadruple Engine were originally 
 as shown on diagram 1, fig. 74 and fig. 75, in order that a compact engine 
 taking little floor space should have an advantage over the three-crank triple. 
 Moreover, the four-stage compound system was capable of a good arrange- 
 ment of valves when in tandem, as may be seen in fig. 75. Another tandem 
 arrangement is shown in diagram 2 suitable for large engines and three- 
 cranks. As in the case of the triple-stage six-cylinder engine, there may be 
 
 Fig. 75. — Cylinders of a Two-Crank Quadruple- Expansion Engine. 
 
 with the quadruple-stage only two sizes of cylinder, for here the first of the 
 lower three may be the second M.P. cylinder, while the second two ol the 
 upper three are the first M.P. cylinders. Then, if the ratio of L.P. to H.P. 
 is 8, then the upper ones will be just half the diameter of the lower ; the 
 ratio of second M.P. to first M.P. will be 2, and of L.P. to second M.P. cylinder 
 also 2. Thus, if the lower cylinders are each 60 inches diameter, the upper 
 ones will be 30 inches each ; quite a simple arrangement. 
 
DIAMETER OF L.P. CYLINDER. 241 
 
 Designs 3 and 4 show the cylinders as arranged for four cranks, the latter 
 being the natural sequence of cylinder, while with the former the smaller 
 cylinders are outside and the three last in sequence, so that the rockin^ 
 action is reduced by the arm of the couple being a minimum. 
 
 Figs. 69 and 70 are examples of multi-cylinder engines operating on 
 one crank, the axis of each engine being virtually in the same transverse 
 plane as the others. Considerable fore and aft space is saved by such 
 a design, and although very much out of balance statically, such engines 
 will work satisfactorily, and are quite suitable and useful for tug boats and 
 river craft working in smooth water, where space and prime cost are of great 
 importance. They are largely employed on the American rivers and harbours, 
 where the tonnage question does not affect the designer. 
 
 Fig. 283 is another example of compactness, whereby four cylinders are 
 caused to operate on two cranks without being in tandem. In this case, 
 however, the engines are balanced both statically and against inertia forces. 
 
 The Ratio of Cylinders in Practice depends somewhat on the service of 
 the ship. The cargo boat, with triple three-crank engines and a working 
 pressure of 185 lbs. the ratio of L.P. to H.P. cylinder is about 6-5 * while 
 with a higher boiler pressure up to 200 lbs. it is 7-0. With express steamers 
 and such pressures the ratios will be from 5-5 to 6-0 ; in Naval ships the ratio 
 is no more than 6-3 when the working pressure is over 200 lbs., and 7-0 when 
 250 lbs. Quadruple engines are limited to service in the mercantile marine 
 and with them the cylinder ratios are from 8 to 9, the latter being the rule 
 with boiler pressures up to 230 lbs. per square inch. In small high-speed 
 craft, as torpedo boats, destroyers, and other similar ships, the ratio with 
 180 to 200 was only 4='5 to 5-0. • 
 
 To determine the Ratio of the L.P. to the H.P. cylinder in any compound 
 system, the following simple rule holds good : — When p is the absolute pres- 
 sure of the steam at H.P. valve-box, ratio of cylinders = -p ~ K, where in 
 the mercantile marine K is 27 for cargo steamers, and 34 for express steamers. 
 In the naval service K is 35 for cruisers and battleships, and 42 for scouts 
 and destroyers. The economy of these vessels at low speeds is quite good, and 
 even at high are not so very extravagant considering the low steam efficiency. 
 
 Having determined the Diameter of the L.P. Cylinder necessary for the 
 power required from the engine, and deduced from the conditions and cir- 
 cumstances of the particular case the diameter of the H.P. and M.P. cylinders, 
 the designer can proceed with the dimensions and arrangements of all that 
 pertains to them. It is, however, of prime importance that the pipe through 
 which the steam is supplied to the engine from the generators, and that 
 through which it passes away from it to the condenser are of adequate size, 
 and as short as circumstances will permit, for otherwise there may be a drop of 
 pressure more than is desirable from the boiler to the valve chest, and, what 
 is worse still, a more serious drop from the L.P. cylinder to the condenser. If 
 the ratio of L.P. to H.P. cylinder is 7-0, the loss of a pound of pressure at the 
 L.P. cylinder will require an increase of 7 lbs. mean pressure at the H.P. 
 cylinder to make compensation. Further, if the referred mean pressure can 
 be raised from 33 to 34 lbs. by a good condenser and adequate exhaust pipe, 
 the gain in mean j)ressure is 3 per cent., and the gain in power somewhat more 
 if the engine is quickened by it as it would be. In all reciprocating engines the 
 
 * X.E. Coast Inst. E. and S. have adapted 7-5 in their standard specification for 180 lbs. 
 
 16 
 
'242 
 
 MANUAL OF MARINE ENGINEERING. 
 
 flow of steam to and from them is intermittent more or less ; the higher the 
 rate of revolution the smaller is the variation due to it. With slow-running 
 
 60 
 
 a 
 
 '35 
 
 e 
 
 c3 
 
 & 
 X 
 
 "3< 
 
 a 
 
 eS 
 
 05 
 Si 
 
 H 
 
 eg 
 
 -8 
 
 s 
 
 o 
 
 I— 
 
 _ 
 
 engines the H.P. valve-box may be with advantage of such a size a* to act 
 as a receiver, and so prevent the wire-drawing action at entry ot steam being 
 
MAIN STEAM PIPE. 243 
 
 excessive ; but there can never be the steady flow that is possible with a turbine, 
 whereby there is no check to the stream with its consequent pulsation that 
 must detract from the efficiency of the reciprocator. Then there must be 
 likewise adequate means for conveying the steam from cylinder to cylinder 
 with least loss in the process. There must be, of course, some drop in pres- 
 sure at every stage, as without it there could be no flow of steam, but the drop 
 should be small, and only a means to an end, and not a cause of loss beyond 
 the actual needs. If the flow is practically continuous, the sizes of pipes 
 and passages may be quite moderate compared with what they must be 
 in a slow-running engine with intermittent demand. This must be remem- 
 bered, that at the top of the stroke the acceleration of piston velocity is more 
 rapid than at or near the bottom, and that with an early cut-off the velocity 
 is less than the mean, so that calculations based on mean piston velocitv 
 mil give ample areas of cut-off, etc. 
 
 Since with slide valves at both steam entry and exhaust the orifice is an 
 «xpanding and contracting one, while the section of the channels or passage? 
 is constant, and considering that clearance space is detrimental to economy, 
 the latter need not be of the same area as the ports ; or, what perhaps is the 
 safer dictum, the ports should always have a larger area than the passage 
 sections (v. fig. 77). Also the passages should be made as short as possible. 
 and as free from turns and corners where eddie3 may set up as the general 
 design permits ; especially should this be followed with the L.P. cylinder 
 where the loss from clearance is irrecoverable. 
 
 Drop Valves have been used successfully with engines of the marine type 
 on shore ; these, however, are not required to reverse, and do not as a 
 rule work a 24-hour day or a 7-day week ; moreover, a stoppage from any 
 cause is not likely to be fatal at any time. If such valves could be trusted 
 •on board ship, and arranged for reversal, much of the steam and thermal 
 losses of the marine engine would be avoided, and besides, the pretty 
 indicator diagrams produced be without the drawbacks, for which, from the 
 marine engineer's point of view, they are inadequate compensation. Judging 
 by what is occurring in the motor-car world, the drop valve is not the desider- 
 atum of the oil engine builder, for one by one they are reverting to slide valves. 
 
 Main Steam Pipe. — The main steam pipe, which supplies a cylinder with 
 steam, should be of such a size that the mean velocity of flow through it 
 does not exceed 8,000 feet per minute. When this is not exceeded, the loss 
 of pressure between the boiler and the valve-chest is very slight indeed. 
 If, however, the valve-chest is large, the cut-off in the cylinder is before 
 half -stroke and the rate of revolution moderate, the area of transverse section 
 of this pipe may be smaller than given by this rule, for as stated the 
 piston speed is below the mean velocity at the early part of the stroke, and 
 the space in the steam-chest acts as a reservoir for steam, so as to keep up 
 a steady supply during admission. If the space is not less than one-half 
 the volume swept through by the piston at cut-off, the velocity of steam 
 in the pipe may be assumed to be 9,000 for engines of 150 N.H.P. to 250 N.H.P., 
 and 10,000 for those above that power ; for smaller engines, owing to the 
 comparatively larger resistances of small pipes, it is not advisable to take 
 a higher speed than 8,000. On the other hand, if the cut-off is later than 
 half stroke, and the valve-box small, the assumed velocity should be at 
 least 10 per cent, less than that given above. 
 
244 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Fig. 77.— H. P. Cylinder (Naval). 
 
LOW-PRESSURE CYLINDER. 
 
 245 
 
 Fig. 77n.— L.P. Cylinder (Naval) with Triple Steam Parts. 
 
246 MANUAL OF MARINE ENGINEERING, 
 
 Taking 8,100 feet as the mean velocity, S the mean speed of piston in 
 feet per minute, and D the diameter of the cylinder, then, 
 
 «'.' - r ■ ' ■ /D' X S D ,_ 
 
 Diameter of mam steam pipe = W ft ,, , =-^. v». 
 
 Example. — To find the diameter of the main steam pipe to a cylindei 
 4o inches diameter and 5 feet stroke, the revolutions at full speed to be 60 "per 
 minute. 
 
 Here S = 2 X 5 X 60 = 600, and D = 45 inches. Therefore, 
 
 4.5 
 
 Diameter of main steam pipe = — s/600 = 12*25 ins. 
 
 When the main steam pipe is abnormally long, as is the case with large 
 ships with long boiler rooms, the divisor should be 85 to 87. 
 
 Area through Stop and Throttle Valves. — Although the loss of pressure 
 at the valve-box is often attributed to want of sectional area in the main 
 steam pipe, it is more frequently due to contracted area past these valves. 
 The friction through a number of small openings is considerably more than 
 through one of an area equal to the collective areas of those openings, especially 
 if the sum of the perimeters of the latter largely exceed that of the single 
 opening, and the " loss of head " will be large if due allowance is not made. 
 For this reason there should be always an excess of area around valves and 
 other obstructions to the free passage of steam, and the passages leading to 
 and from them should be as easy as possible, so as to avoid violent changes 
 both of direction and velocity of flow. 
 
 Steam Ports and Passages. — Since, in most engines, the steam has to exhaust 
 through the same ports and passages by which it was admitted, their size 
 must be governed by the proper flow of emission rather than of admission. 
 The area of section of steam ports should be such that the mean velocity 
 of flow at exhaust should not exceed 6,000 feet per minute. The ports should 
 be somewhat larger than the section of the passages, especially when certain 
 kinds of valves are used, which will be dealt with later on ; as a rule, they 
 have nearly the same area as the sectional area of the passages. To avoid 
 excessive clearance, however, the capacity of the passages should be as small 
 as possible consistent with free flow of steam, and as this depends greatly 
 on their sectional area, so that the reduction in capacity can only be attained 
 by making them as short as possible. Not only will the evils arising from 
 clearance be avoided, but the loss through resistance be very materially 
 lessened by shortening the distance between the valve face and cylinder,, 
 and their cooling effect on the incoming steam temperature. 
 
 Area of steam ports and of section through passages 
 
 Area of piston X speed of piston 
 67000 - 
 
 _ (Diameter of cylinder)' 2 x speed of piston 
 ' 7,686 
 
 Opening of Port to Steam. — It is advisable so to design the valve, etc.,. 
 that the opening for admission of steam to the cylinder is sutticient to avoid 
 
EXHAUST PASSAGES AND PIPES. 
 
 247 
 
 any serious loss by " wire-drawing ; " but in actual practice, unless special 
 gearing is designed so as to give a quick motion to the valve at the instant 
 of cut-off, there is very considerable loss of pressure shown on the indicator- 
 diagram ; and, what is worse still, from deficient opening, the loss is generally 
 not limited to the period near to cut-off, but during the whole time of admission. 
 The ordinary valve gears do not give that quick motion, either at opening or 
 at cut-off, which is such a desideratum. Separate expansion valves and special 
 valve-gearings admit of such a motion, and consequently the opening to 
 steam with them may be smaller than when cut-off is effected by the ordinary 
 slide-valve and link-motion ; they are, however, the cause of more loss than 
 gain, and are now never used. 
 
 Hence, when only common valves and gear are to be used, the area for 
 opening to steam when at its greatest should be such that the mean velocity 
 of flow does not exceed 10,000 feet per minute. In actual practice the amount 
 of opening is often much less than that given by the above rules, but it always 
 results in loss of pressure in the cylinder throughout, and excessive " wire- 
 drawing " previous to cut-off. and, therefore, should not be less but 
 greater. 
 
 Exhaust Passages and Pipes. — The area of section of exhaust passages 
 should be such that the mean velocity of steam does not exceed 6,000 feet 
 per minute, and if the distance from the cylinder to the condenser is com- 
 paratively great, a much larger area is advisable. There should not be a 
 greater difference than 1J lbs. between the pressure in the cylinder and that 
 in the condenser when exhausting, even with a high vacuum. 
 
 
 
 TABLE 
 
 XXXI. 
 
 
 
 Piston Speed. 
 
 Diameter 
 
 Diameter* 
 
 Area 
 
 Opening 
 
 Area * 
 
 Feet per 
 
 Main Steam 
 
 Exhaust 
 
 Main Steam 
 
 to Steam 
 
 Exhaust 
 
 Minute. 
 
 • 
 
 -r D. 
 
 -r-D. 
 
 + A. 
 
 -r A. 
 
 -i- A. 
 
 •200 
 
 0158 
 
 182 
 
 025 
 
 020 
 
 0333 
 
 250 
 
 0-177 
 
 204 
 
 0313 
 
 0-025 
 
 0417 
 
 300 
 
 0194 
 
 0-223 
 
 0-0375 
 
 0-030 
 
 0500 
 
 350 
 
 0-209 
 
 241 
 
 0437 
 
 0-035 
 
 0583 
 
 400 
 
 0-224 
 
 0-258 
 
 0500 
 
 Q-040 
 
 0667 
 
 450 
 
 237 
 
 0274 
 
 00562 
 
 045 
 
 0750 
 
 500 
 
 0-250 
 
 288 
 
 0625 
 
 050 
 
 0833 
 
 550 
 
 262 
 
 0-302 
 
 0-0687 
 
 055 
 
 00917 
 
 600 
 
 0-274 
 
 0316 
 
 00750 
 
 0-060 
 
 1000 
 
 650 
 
 0285 
 
 329 
 
 00812 
 
 065 
 
 01083 
 
 700 
 
 296 
 
 0-341 
 
 0875 
 
 0-070 
 
 01167 
 
 750 
 
 306 
 
 353 
 
 0937 
 
 0-075 
 
 01250 
 
 800 
 
 0-316 
 
 0-365 
 
 o-iooo 
 
 0-080 
 
 01333 
 
 850 
 
 326 
 
 376 
 
 0-1062 
 
 085 
 
 01417 
 
 900 
 
 0-335 
 
 387 
 
 01125 
 
 0-090 
 
 0-1500 
 
 950 
 
 0-344 
 
 397 
 
 01187 
 
 095 
 
 1583 
 
 1000 
 
 353 
 
 0-400 
 
 01250 
 
 o-ioo 
 
 01667 
 
 The exhaust passages from the high-pressure cyliuder of a compound 
 engine to the next cylinder should be such that the flow of steam does not 
 exceed 4,500 feet per minute, in order that the difference between the pressure 
 
 * N.E. Coast Tnst, E 
 3.900 feet, and from L.P. 
 
 and S. recommend for exhaust from H.P. cylinder 3.600 feet, from M.P. cylinder 
 cylinder 4.500 feet, and difference iu pressure 1} lbs. at working sea speeds. 
 
'248 
 
 MANUAL OF MARINE ENGINEERING. 
 
 during admission in the medium-pressure cylinder and exhaust in the high- 
 pressure cylinder may not be excessive, and also from the medium-pressure 
 to the low-pressure cylinder the nominal rate of exhaust should not exceed 
 4,500 feet. The pressure in the receivers is not sensibly constant, as it is in 
 the condenser, being subject to sudden fluctuation when the high-pressure 
 valve opens to exhaust and the medium-pressure valve opens to lead, and 
 so on. 
 
 Table xxxi. gives the relation between the various passages, etc., 
 and the piston in accordance with the foregoing rules, and is based on the 
 assumption of a mean velocity of flow of 8,000 feet per minute for the main 
 steam pipe, 10,000 for opening to steam, and 6,000 for exhaust ; A is the 
 area of piston, D its diameter, and ratio of D to length not more than 60. 
 
 In naval and other ships, the engines of which are only occasionally, and 
 for short periods, run at full speed, so that high efficiency at those times is 
 not of first consideration, the exhaust-pipe to the condenser may be of such 
 a size when short that the flow is 7,300 feet in large engines and 6,500 feet 
 in smaller ones ; the exhaust-pipes from cylinder to cylinder may be corre- 
 spondingly reduced, as may also the ports, etc. 
 
 TABLE XXXII. — Weight of Steam in Lbs. at 100 Lbs. Pressure 
 Delivered per Minute through Pipes with a Drop of 
 
 1 Lb. only. 
 
 t 
 
 
 
 
 By E. 
 
 C. Sickles. 
 
 
 
 
 
 Diameter of Bore. 
 
 
 
 Length of Pipe. 
 
 
 i 
 
 
 
 
 
 
 
 
 
 26 Ft. 
 
 50 Ft. 
 
 76 Ft. 
 
 100 Ft. 
 
 125 Ft. 
 
 150 Ft. 
 
 its n. 
 
 1 inch, 
 
 4-40 
 
 310 
 
 2-54 
 
 2-20 
 
 1-96 
 
 1-79 
 
 1-60 
 
 1J .. 
 
 
 
 9-TG 
 
 6-90 
 
 5 63 
 
 4-88 
 
 4-36 
 
 3*93 
 
 3-6I) 
 
 n » 
 
 
 
 13 04 
 
 9-29 
 
 7 69 
 
 6-52 
 
 5-83 
 
 5-33 
 
 4-95 
 
 2 „ 
 
 
 
 30 54 
 
 21-60 
 
 17 60 
 
 15-20 
 
 13-60 
 
 12 50 
 
 11 -5s 
 
 2i „ 
 
 
 
 50-80 
 
 35-90 
 
 29-30 
 
 25-40 
 
 22-70 
 
 20-80 
 
 19-2,i 
 
 3 „ 
 
 
 
 ... 
 
 65-17 
 
 ... 
 
 46 00 
 
 • > • 
 
 .* . 
 
 35-00 
 
 Si „ 
 
 
 
 • •• 
 
 98-30 
 
 ... 
 
 69-50 
 
 • •• 
 
 
 ., 
 
 52-70 
 
 4 „ 
 
 
 
 ... 
 
 1381 
 
 *•• 
 
 97 60 
 
 • •• 
 
 
 , , 
 
 73-80 
 
 44 .. 
 
 
 
 ... 
 
 187 9 
 
 • •• 
 
 132-9 
 
 • •• 
 
 
 # , 
 
 102 
 
 5 „ 
 
 
 
 ... 
 
 255-6 
 
 • • • 
 
 180-7 
 
 • •• 
 
 
 ,, 
 
 136-8 
 
 6 „ 
 
 
 
 ... 
 
 419-4 
 
 • •• 
 
 296 5 
 
 • •• 
 
 
 .. 
 
 229 1 
 
 7 „ 
 
 
 
 ■ • • 
 
 618-8 
 
 • •• 
 
 437 
 
 • •• 
 
 
 .. 
 
 330-0 
 
 8 „ 
 
 
 
 ... 
 
 890-0 
 
 • •• 
 
 629 
 
 • •• 
 
 
 , , 
 
 472-0 
 
 9 „ 
 
 
 
 • •• 
 
 1206 
 
 • •• 
 
 5530 
 
 • •• 
 
 
 , , 
 
 644 
 
 10 „ 
 
 
 
 ... 
 
 1592 
 
 • • • 
 
 1126 
 
 ... 
 
 
 , . 
 
 851-0 
 
 11 „ 
 
 
 
 ... 
 
 2046 
 
 ... 
 
 1447 
 
 • •• 
 
 
 . , 
 
 1093 
 
 12 „ 
 
 ... 
 
 2575 
 
 ... 
 
 1887 
 
 ... 
 
 
 
 1428 
 
 Cylinder Liner. — In order that a suitable material may be supplied to 
 resist the rubbing action of the piston without wearing away, and one that 
 shall be capable of taking and retaining a polished surface, so as to minimise 
 the friction of the piston, and which, when worn, may be easily and cheaply 
 renewed, an inner bush or false barrel is fitted, usually called the cylinder 
 
CYLINDER LINER. "249 
 
 liner. This liner should be made of a hard, close-grained metal having con- 
 siderable strength, but not so hard as to resist the action of a cutting tool or 
 file, and capable of taking and keeping a polish when rubbed by the piston- 
 rings lubricated with soft water ; it should also be such that the expansion 
 caused by heat is practically the same as the cast iron of which the cylinder 
 itself is made. It is usual to make these liners of cast iron, strengthened, 
 closed, and hardened by mixing with it certain kinds of pig iron, or by the 
 addition of a small quantity of steel {vide Chap. xxx.). The Admiralty 
 prefer the liners to be made of steel, hammered or rolled to a proper size 
 for machining ; and some engineers use cast steel. Although the steel gives 
 good results, it can be equalled by the specially-made cast iron, so far as 
 good wearing is concerned, but, of course, steel exceeds cast iron in tensile 
 strength ; this latter quality was necessary to a higher degree for the hori- 
 zontal engine than for the vertical engine ; the Admiralty were justified in 
 going to the expense of the steel, however, as it enable them to fit much 
 lighter liners than would be admissible if made of cast iron. In the merchant 
 service, with the vertical engine the cast-iron liner does exceedingly well, 
 and it is not likely to be superseded by steel, even if this material can be 
 manufactured much cheaper than at present. Liners are usually made with 
 an inside flange at the bottom end (tig. 78), which tits into a recess in the 
 cylinder end, and is secured there by screw-bolts. The upper end is turned 
 for a few inches, so as to fit tightly into the cylinder shell at that part. The 
 joint at the cylinder bottom is made with red-lead paint, while leakage 
 between the liner and the cylinder shell is prevented at the other end by 
 stuffing a few rounds of asbestos rope or Tuck's packing into a recess formed 
 for that purpose, and preventing it from coming out by securing a flat wrought- 
 iron ring to the liner so as to cover the packing. Sometimes in lieu of a 
 stuffing-box, the outer edge of the liner and the edge of the turned part of 
 the cylinder shell are champhered so as to form a groove ; into this groove 
 a turn of Tuck's packing or asbestos rope is pressed with a ring as before. 
 Some engineers, preferring to rely on metallic contact, turn a slight recess 
 instead of champhering the edge of the liner, and caulk into it a ring of soft 
 copper. The Admiralty method of making a steam-tight joint at the outer 
 end of the liner is by means of a flat copper ring covering the joint and secured 
 to the liner, as also to the cylinder by means of iron rings and screw bolts. 
 The copper ring is deeply grooved between these iron rings so as to permit 
 of a slight movement of the liner endways with respect to the cylinder {vide 
 fig. 78a). The liners are sometimes secured without a flange at the bottom, 
 by screwing studs through the cylinder shell and liner, and making the ends 
 steam-tight as before. 
 
 The space between the liner and shell should not be less than 1 inch, and 
 may be filled with steam so as to prevent condensation in the cylinder, and 
 heat the steam during expansion. If the cylinder has to be jacketed, this 
 is really a better plan of doing it than by casting the cylinder and inner 
 cylinder together, as was very generally done formerly. Independently of 
 the advantage derived from the harder metal of which the liner may be 
 made, compared with that which is suitable for so intricate a casting as a 
 cylinder, there is another very great advantage to the manufacturer. Since 
 it is a necessity that the inside walls of the cylinders be sound and free from 
 sponginess, as well as blowholes, a casting may be condemned for a defect 
 
250 
 
 MANUAL OF MARINE ENGINEERING. 
 
 which in no way detracts from its strength or usefulness, excepting that it does 
 not admit of the piston working on it steam-tight. If a liner is to be fitted, 
 a little sponginess, or even a blowhole in the cylinder casting is of no con- 
 sequence, and therefore the extra cost of fitting a liner does not, as a rule, 
 exceed the reasonable premium which would be allowed for assuring good 
 and sound castings ; and this is especially so in the case of large cylinders. 
 
 False Faces. — For the same reason that liners are fitted to the cylinders, 
 the cylinder face, on which the valve slides, needs a false face. This is usually 
 made of hard, close-grained cast iron, of the same quality as the liner, and 
 secured to the cylinder (fig. 78) by brass screws having cheese heads sunk 
 
 Fig. 78. — Section through Cylinder. Fig. 78a.— Admiralty Method of Fitting Liners. 
 
 in a recess, so as to be considerably below the surface. Care should be taken 
 to lock these screws, so that they cannot slack back ; the simplest way of 
 doing this is to cut a slight nick in the side of the recess, and caulk or drift 
 the metal of the screwhead into it, after the screw is tightened in place. 
 
 False faces were sometimes made of hard gun-metal or phosphor-bronze 
 in the engines of warships. The superior strength of these metals over 
 cast iron admits of the face being much thinner, but besides being much 
 more expensive, there is great risk of damage to the cylinder itself, owing to 
 the greater expansion of the bronzes by heat : they never permit of such a 
 
DOUBLE-PORTED VALVES. 251 
 
 good working surface as does cast iron ; and even if danger from this is slight, 
 some difficulty has been experienced in keeping the joint between the two 
 metals steam-tight. 
 
 By connecting the recesses for the screwheads with grooves cut in the 
 face, the rubbing surfaces are well lubricated, and a considerable amount of 
 relief given to the valve itself by the reduction of the effective pressure on it, 
 caused by the steam flowing through these grooves, etc. 
 
 The corners of the ports, both in the false face and cylinder face, should 
 be well rounded, as the casting is very apt to crack at them if they are sharp. 
 
 The Width of the Steam Ports, in the direction parallel to the cylindei 
 bottom, is usually 0*6 to 0*8 of the diameter, but engines of longer stroke 
 than usual require a larger proportion than this to obtain the necessary port 
 area without having excessive length (measured in direction parallel to the 
 axis). It is obvious that, at the cylinder bore, the width of port cannot 
 exceed the diameter, and must really be somewhat less ; in actual practice 
 it seldom exceeds 0'8 of the diameter ; but at the cylinder face it may, and 
 sometimes does, exceed the diameter, the length being such that the area of 
 section in the passage is uniform throughout. 
 
 Piston Valves. — If the very broad cylinder face is bent into the form of 
 a cylinder, there will be the same area of orifice, while the space occupied 
 in direction of the width of the port is less than one-third of that required 
 for the flat face. The valve for such a face must be cylindrical, or composed 
 of two circular discs or pistons having the same depth of edge as there 
 would be of bearing surface at each end of the ordinary slide valve. 
 Such a valve (fig. 138) is called a piston valve, and besides possessing the 
 advantage of occupying little space, has the more valuable one of being free 
 from lateral pressure, requiring no balancing or relief, and moving with the 
 least resistance of any slide valve. For these reasons, the piston valve is 
 an exceedingly good form when high pressures of steam are used and for very 
 large engines. It is a very general thing for engineers to fit a piston valve to 
 the high-pressure cylinder of compound engines of all sizes and types ; many 
 makers also fit the medium-pressure cylinder with piston valves, and a few 
 fit them to all three cylinders, especially of large engines, and of engines 
 running at high revolutions, as in torpedo boat destroyers. The earliest 
 compound engines of Woolf, made in the early years of the nineteenth century, 
 had piston valves. 
 
 " Drop " Valves, with some form of Corliss valve gear, may take the place 
 of piston valves in larger engines, at least if superheated steam comes into 
 more general use. They are extensively used on the large slow-speed marine 
 type of engine on shore, and in the slow revolution paddle steamer in 
 America (v. fig. 30, p. 73). 
 
 Double-Ported Valves. — Although there is of necessity only one opening 
 of the steam passage into the cylinder, there may be two or more openings 
 through the cylinder face into the steam passage. The combined area of these 
 openings need not materially exceed that of the section of passage, but it 
 is better to make it so ; as each will be open to steam by the same amount 
 as the single port, if the valve has tbe same travel, lap, etc., the total opening 
 is in this case double that of the single port for a double-ported face, and 
 treble for a treble-ported face. When the face is treble-ported, the valve is 
 generally arranged so as to admit steam through all three ports, but to exbaust 
 
252 MANUAL OF MARINE ENGINEERING. 
 
 through two only, as there is seldom any difficulty in getting full opening to 
 exhaust. 
 
 Steam Jackets. — It is not necessary here to enter into the question of the 
 economy of steam-jacketing. If the economy is doubted, it is, at least, 
 .certain that it admits of the cylinders being gently warmed before starting, 
 without moving the working parts. It was customary at one time to form a 
 jacket around the cylinder by casting it with two thicknesses of metal ; but 
 this was often inconvenient, and always risky to the moulder. Now, as a 
 loose liner is fitted, the jacket is automatically provided. When the engines 
 are of short stroke, the ends present almost as large a surface to the steam 
 as do the walls, and should be steam-heated if the jacketing is to be effective. 
 For strength of structure, too, all large cylinders should have hollow bottoms 
 and covers, as stiffening them with ordinary webs is not always sufficient, 
 .and is in some cases of iron castings even a source of danger. If the pressure 
 is on the same side as the webs, they are safe and do add to the strength 
 of structure ; if on the opposite side, then they are in tension, and their outer 
 edge liable to extreme tension ; so that if there be a nick or other such defect 
 from which to start a crack, or if subject to a sudden application of the stress, 
 the outer edge is apt to crack, which will develop and spread into the main 
 body of metal, finally causing serious damage. This arises from the fact 
 of cast iron possessing so low a power of resisting stress in tension, compared 
 with its power against compression. Care should be taken to thoroughly 
 drain the steam jackets, and to this end no webs should so be placed as to 
 stop the flow of water to the drain-cocks. Steam, when supplied to the 
 jackets of the low-pressure cylinder, should not much exceed the pressure 
 that is in the receiver ; for this purpose, a reducing valve is fitted between 
 the boiler supply pipes and the jacket. 
 
 Boring Holes. — The diameter of the boring hole depends generally on the 
 size of the boring bar employed, and should not be less than one-fifth the 
 diameter of the cylinder. When there is a single piston-rod, the stuffing-box 
 is formed in the cover of the boring hole. When there are two or more 
 piston-rods, the boring hole should be large enough to admit a man ; and, 
 therefore, not less than 14 inches diameter, and, when possible, 16 inches 
 •diameter. The doors are sometimes made with the flange fitting into a 
 recess inside the cylinder, so that the piston-rod, having a butt end, may be 
 drawn from the cylinder with the piston ; when this is required, the boring 
 hole must be of sufficient diameter to admit of the piston-rod end drawing 
 through it. 
 
 Auxiliary Valves. — To render engines handy, so that they may be started 
 from any position of the cranks, it is necessary to arrange the gear so that 
 the main valves may be operated by hand, or else to fit smaller valves, which 
 may be readily worked by the engineer ; these are called auxiliary valves, 
 and should have a port area equal to - 002 the area of the piston. It is 
 usual to fit such valves to the low-pressure cylinder of compound engines. 
 Pass cocks, by which steam can be admitted to the valve boxes of inter 
 mediate cylinders, should be fitted, and by their means the use of the L.P. 
 auxiliary valve is often avoided. 
 
 These auxiliary valves are usually only flat plates, without even an exhaust 
 cavity ; but for very large engines they should be piston valves, or some 
 form of balanced valve. 
 
COLUMN FACINGS AND FEET. . " 253 
 
 Escape or Relief Valves. — These are simply spring-loaded safety valves, 
 to allow of the escape of water caused by priming or condensation when the 
 piston drives it to one end of the cylinder. They are fitted to each end of 
 the high-pressure cylinders of all fast-running marine engines. It is suffi- 
 cient in vertical engines if there is only an escape valve at the bottom of the 
 medium-pressure and low-pressure cylinders. The diameter of these valves 
 should be one-fifteenth the diameter of the L.P. cylinder of a compound 
 engine. In the Navy it is usual for all large cylinders to have a fair of escape 
 valves at each end. 
 
 Drain-Cocks. — These should be placed wherever any water is likely to 
 accumulate in the cylinder and casings, and their size 0*023 the diameter of 
 cylinder + h mcn - They should be connected to a pipe leading into the 
 condenser bottom ; for if led to the bilge the engine-room is filled with steam 
 when open, and the receiver and low-pressure cylinder will seldom drain — in 
 fact, during the greater part of the stroke, instead of letting water out, they 
 let air into the low-pressure cylinder and spoil the vacuum. Care should be 
 taken when the drain-pipe is led to the condenser that the water, etc., does 
 not impinge on the tubes, or even on the condenser sides, so as to do serious 
 damage, and a non-return valve fitted to prevent water getting back. 
 
 Receiver Space. — The space between the valve of the high-pressure 
 cylinder and that of the medium-pressure cylinder, and that between the 
 valves of the medium-pressure and the low-pressure cylinders, should be 
 from 0'6 to l'l times the capacity of the exhausting cylinder, when the 
 cranks are set at an angle of 120°. When the cranks are opposite or nearly 
 so, this space may be very much reduced. The pressure in the medium- 
 pressure receiver should never exceed 0*7 the boiler pressure, and is gener- 
 ally much lower than this. It is usual to fit a safety valve to the low-pressure 
 receiver, loaded by weight or spring to a pressure of 20 to 30 lbs. per square 
 inch ; otherwise, owing to the large flat sides between the two cylinders, and 
 in the valve-box when a flat valve is employed, great risk of explosion would 
 be run. This safety valve is usually of the same size and design as the 
 cylinder escape valves. The receivers of three-crank compound engines need 
 not be so large as the above, as the cranks are usually at angles of 120° ; 
 in the case of triple-compound engines with the medium-pressure leading 
 the high-pressure, a smaller receiver will do. 
 
 Column Facings and Feet. — It was very usual at one time to form merely 
 facings for the jointing of the cylinder to the frames and columns ; but as 
 this necessitated the use of studs, or else driven bolts with the heads inside 
 the cylinders, it is now abandoned, distinct projections or feet being cast to 
 the cylinder bottom, having flanges corresponding to those on the columns 
 or frames, so that they may be connected by driven bolts, which are always 
 accessible. The only objections to this method are, that it is more expensive 
 to mould, and a certain amount of risk is run of getting the casting sound 
 and strong where the feet meet the main casting. The former should be 
 disregarded in considering so important a part as the cylinder, and the latter 
 is always avoided by a good moulder. 
 
 Great care should be exercised in designing these column feet, for through 
 them the load due to the steam on the cover (which is greater than that 
 on the piston) is transmitted ; and as the load is a recurrent one and always 
 applied suddenly, verv ample section of metal should be provided to sustain 
 
254 MANUAL OF MARINE ENGINEERING. 
 
 it. The area of section through these feet should be such that the stress 
 does not exceed 600 lbs. per square inch — that is. the area in square inches 
 is not less than that given by dividing the maximum load on the cover in 
 pounds by 600. The webs from the flanges of the feet should be well spread 
 over the cylinder bottom and towards the sides, so as to distribute the strain. 
 
 Holding-down Bolts. — The bolts connecting the cylinder to the columns 
 or frames should be such that the stress on them does not exceed 4,000 lbs. 
 per square inch, taking the section at the bottom of the thread, and when 
 there is a large number of comparatively small size it shoidd not exceed 
 3,000 lbs. per square inch (v. Chap, xxviii.). 
 
 Horizontal Cylinders. — In addition to the facings or feet for connecting 
 to frames, additional feet were necessary for the cylinders of horizontal 
 engines to rest on and be secured to the engine bed. These feet, too, had 
 webs so arranged as to distribute the strain caused by the reaction from the 
 weight of the cylinder, pistons, etc. The front part of the cylinder was 
 rigidly bolted down, while the back end, especially of long cylinders, was 
 held down only, and free to move horizontally when expanded by the heat. 
 But since cast iron will expand only one-tenth of an inch in 8 feet, by an 
 increase of 180° Fahr. of temperature, there was seldom need to make any 
 special provision, beyond boring the holes for the back bolts rather larger, or 
 making them slightly oval in the cylinder feet. 
 
 Oscillating Cylinders. — The chief peculiarity of these cylinders is the 
 method of supporting by trunnions, which also serve as steam and exhaust- 
 pipes (vide fig. 27). Half the load on the piston is taken on each trunnion; 
 and .since they are of such ample diameter, it is sufficient to assume that the 
 metal is subject only to shearing stresses, and therefore the area of section 
 should be such that the stress does not exceed 950 lbs. per square inch. The 
 diameter of the trunnions is governed by the size of the exhaust-pipe, since 
 the steam must exhaust through one of them, and it is usual and convenient 
 to make them both of the same size. In the case of a compound oscillating 
 engine, the trunnions of both cylinders should be of the same size, which 
 will depend on the size of exhaust of the low-pressure cylinder. The trunnions 
 of the high-pressure cylinder, being so much larger than is necessary to accom- 
 modate the steam-pipe, allow of a space between the outer or working part 
 and the inner part or stuffing-box, which, if left open to the air, is well venti- 
 lated, and so reduces the heat on the bearing due to steam of high temperature. 
 
 The length of the trunnion journal or bearing should be such that the 
 pressure per square inch on the area, made by the multiple of its diameter 
 and length, does not exceed 350 lbs. ; generally it is from one-third to one- 
 half of the diameter. 
 
 The trunnions have interposed between them and the cylinder body a 
 belt, which conveys the steam to and from the valve-boxes. This belt should 
 be very strong and well ribbed to the body of the cylinder, immediately 
 above and below the trunnions, and when the cylinder is fitted with a liner, 
 it is better to form the outer shell in the shape of a beer barrel, so that the 
 belt projects inside and not outside, as it would be were there no liner ; the 
 strain from the trunnions is then at once taken by the cylinder sides without 
 the intervention of webs and ribs. 
 
 The cylinder valve faces of oscillating engines should be set so that the edge 
 next the steam entrance should be the nearest point to the cylinder — that is, 
 
CYLINDER COVER STUDS AND BOLTS. 4 255 
 
 the plane of the cylinder face touches a cylinder whose axis coincides with 
 the axis of the cylinder-bore at this edge. When this is so, the lead of the 
 steamway into the valve-box is short and easy, and the opening into the 
 exhaust-belt on the side opposite is large, without causing the valve-spindle 
 centre to be unnecessarily far out from the cylinder. 
 
 Cylinder Covers. — Like the cylinder end or bottom, the cover has to be 
 strong enough to take the full steam pressure, but as a rule it has no load 
 to distribute to any other part. The same remarks as to webs, etc., equally 
 apply to the covers, and all above 24 inches diameter for high-pressure 
 cylinders, and 40 inches diameter for low-pressure cylinders, should be made 
 hollow with two thicknesses of metal. Those of vertical engines are better 
 made in that way for all sizes, inasmuch as it is necessary to fill in the spaces 
 between the webs, when they are so made, to prevent the lodgment of water, 
 etc., and it is usual to add a false cover, either polished or cast with a pattern 
 to give a good appearance. This can always be accomplished by casting the 
 covers hollow. The cylinder covers of naval engines and for large express 
 steamers are made of cast steel, and formed with ribs having bull-nosed 
 flanges to strengthen them. In small naval engines, such as those of destroyers, 
 the cover?, especially of the medium-pressure and low-pressure cylinders, are 
 often made of manganese or other strong bronze, which has an elastic limit 
 higher than that of ordinary steel, and permits of them being cast much 
 thinner and lighter. Covers, when of steel or bronze, are sometimes cor- 
 rugated as well as coned to get the necessary stiffness without the complica- 
 tions of ribs. The depth of the cylinder cover at the middle should be about 
 one-quarter of the diameter of the piston for pressures of 80 lbs. and upwards ; 
 that of the low-pressure cylinder cover of a compound engine should be 
 at least - 15 its diameter. Since, however, the size of the piston-rod is the 
 best measure of the pressure on the cover, it is better so to design the cover 
 that its depth at the middle is not less than T45 times the diameter of the 
 piston-rod. The depth of the cover at the edge depends on the steam port ; 
 a recess being formed for the steam-way, and the inside of the cover otherwise 
 being parallel to the piston. 
 
 It is the custom with some engineers to end the cylinder a little beyond 
 the extreme travel of the piston, the steam port-opening being then in the 
 same plane with the cylinder flange ; the cover has a large recess in it, 
 and its flange so extended as to enclose the port-opening (fig. 79). The 
 advantage of this method is the decreased length and weight of cylinder, and 
 being able to secure the cover in way of the steam port direct to the 
 main casting, instead of to the comparatively weak bridge of metal across the 
 port. On the other hand, however, the cover occupies considerably more 
 room, and not being of circular form at the flange, cannot be turned so easily. 
 For larger engines this plan is a very good one, and may be adopted with 
 advantage, but for small ones and those of moderate size it is not always 
 so convenient as the older one. 
 
 Cylinder Cover Studs and Bolts should be made of the best steel, and of 
 such a size that the stress on them does not exceed 5,000 lbs. per square inch 
 of section at the bottom of the thread, as they are subject to sudden and 
 intermittent loads and severe and sudden shocks when priming occurs, also 
 to considerable wear and tear from the frequent removals of the covers for 
 examination of the pistons. In large engines it is usual to fit the cylinder 
 
256 MANUAL OF MARINE ENGINEERING. 
 
 covers with manholes for purposes of examination ; as for such engines larger 
 studs may be fitted, a higher stress is permissible if desired, so that the section 
 may be such that with the maximum pressure there is a stress of 5,500 lbs. 
 per square inch. From one or two causes the resistance to pressure on the 
 cover may not be evenly distributed over the whole of the studs, so that a 
 good nominal margin of safety should be allowed in such a very important 
 part ; this is especially so when a large number of small studs are fitted ; 
 in this case, when of less diameter than l£ inch, the allowance should not 
 exceed 4,500 lbs. per square inch ; under f inch no more than 3,500 lbs. 
 
 Cylinder Flanges. — The width of the cylinder flange need not exceed 
 three times the diameter of the bolts or studs, but if the former are fitted 
 this allowance is not sufficient to give space for the heads. Studs are 
 now nearly always fitted to marine cylinders in great measure for this 
 reason. 
 
 Clearance of Piston. — If both cylinder end, piston, and cover were accu- 
 rately turned, as is very desirable and required by the Admiralty, and the 
 brasses did not wear, a very small amount of space would suffice for clearance 
 between the piston and cylinder end ; but as it is usual to leave these parts as 
 they come from the foundry, and the bearings, however well made, may wear 
 in course of time, it is necessary to make due allowance for this. Small 
 engines up to 50 N.H.P. require an allowance of f inch at each end for rough- 
 ness of castings, and T g- inch for each working joint — that is, for any part 
 between the piston and the shaft journals where wear can take place ; engines 
 from 50 N.H.P. to 100 N.H.P., ~v mcn au d ;;% mcn for each working part ; 
 engines from 100 N.H.P. to 200 N.H.P., \ inch and ^ inch for each working 
 part ; engines over 200 N.H.P., f inch and \ inch for each working part. 
 Fast running engines should have a larger allowance, or be turned as above 
 recommended. This allowance is usually called the stroke clearance. 
 
 For example, take the case of a vertical direct-acting engine of 120 
 N.H.P. ; the parts which wear so as to bring the piston nearer to the bottom 
 are three — viz., the shaft journals, crank -pin brasses, and piston-rod 
 gudgeon brasses, so that the clearance at top wiJl be | inch, and the clearance 
 at bottom § inch + 3 X ^, or § inch in all. Here the total clearance in the 
 cylinder is 1£ inch. 
 
 Valve-Box Covers. — The covers are usually formed of a flat plate stiffened 
 by ribs or webs. So long as the stiffening webs are on the same side as the 
 pressure this form is satisfactory, especially when the covers are not very 
 large ; but when they are large and it is inconvenient to web them on that 
 side they should be made hollow, with the back rounded like a hog-back 
 girder. These covers in Naval ships and express steamers are now very 
 generally made of cast steel. 
 
 The studs and bolts with which these doors are secured should be arranged 
 in accordance with the rules laid down for those of cylinder covers. 
 
 Small Doors and Covers. — As it is essential to examine from time to time 
 the internal working parts, and as this examination is more for the sake of 
 seeing that everything is in good order, rather than in the expectation of 
 having to execute repairs, and generally there is little time at the disposal of 
 the engineer for these purposes, small doors should be fitted to the large 
 heavy doors, secured with only a few studs, so as to be quickly taken off and 
 refitted. These may for this reason, with advantage, be of cast steel or of 
 
LAGGING AND CLOTHING OF CYLINDERS. 
 
 257 
 
 steel plate pressed to shape. There should be, when possible, a doorway in 
 the bottom and cover, and to the valve casing of the low-pressure cylinder, 
 large enough to admit a man. To the valve-boxes of all cylinders there 
 should be peep-holes, through which to ascertain the leads and cut-off of the 
 valves, and to press the flat valves to the cylinder faces should they have 
 become blown off. 
 
 Lagging and Clothing of Cylinders. — All hot surfaces, from which loss of 
 heat may arise by radiation, should be covered with a non-conducting sub- 
 stance. Felt was formerly employed and well suited for this purpose, being 
 enclosed in polished teak or mahogany lagging, secured with brass bands 
 wherever in view in the engine-room, and by pine lagging or canvas when not 
 in view. Sheet-steel is, however, now used as being more enduring than 
 wood, and when carefully fitted and highly "barfed " looks as well as polished 
 wood, and it lasts much longer. 
 
 Cement of kinds, silicate, cotton, and asbestos are specified by engineers for 
 the cylinder covering, but silicate is very objectionable, on the ground that 
 the dust coming from it through the lagging, owing to the vibration, etc., is 
 
 Fig. 79. — Shortened Cylinder with Port in Cover. 
 
 very apt to get into the bearings and guides, and cause serious trouble. 
 Asbestos fibre and magnesia may, however, be used with advantage, 
 especially on the high-pressure cylinders ; a cement made of this fibre, etc., 
 has been found very successful ; a cement, consisting largely of carbonate of 
 magnesia, has also proved an excellent covering for these and other hot 
 surfaces. It is very necessary to cover up all hot surfaces when using steam 
 of such high temperatures as are now common, and considerable gain is 
 obtained by putting muffles on the cylinder covers and all metallic-exposed 
 surfaces. The value of the different non-conducting materials are given in 
 Chap. xxvi. 
 
 The following rules are for the scantlings of the cylinder and its 
 
 connectioiis : — 
 
 D is the diameter of the cylinder in inches. 
 
 p the load on the safety-valves in lbs. per square inch. 
 
 p x the absolute pressure of steam at H.P. valve-box. 
 
 f a constant multiplier = thickness of barrel + '25 inch. 
 
 17 
 
258 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Thickness of metal of cylinder barrel or liner, not to be less than 
 p x D -J- 3000 when of cast iron.* 
 
 Thickness of cylinder barrel = 
 „ liner 
 
 D 
 
 6000 
 = 0-8 x j 
 
 (p + 50) + 0-2. 
 
 For purposes of calculation p may be taken as follows : — 
 H.P. cylinder p =» boiler pressure, or p x -15 
 
 M.P. 
 
 M.P. 
 
 M.P. 2 
 
 LP. 
 
 LP. 
 
 >> 
 
 >> 
 
 »> 
 
 triple p = 0-6 
 quadruple p = 07 
 „ p = 0-45 
 „ P = 0-22 
 triple p =025 
 
 boiler pressure 
 
 it 
 
 Thickness of liner when of steel 
 
 metal of steam ports 
 
 ,, valve-box sides 
 
 ,, covers 
 
 cylinder bottom 
 
 5> 
 
 covers 
 
 cylinder flange 
 
 „ cover flange 
 
 »» 
 
 i« 
 
 valve-box flange = 1 "0 
 door flange 
 face over ports 
 
 false face 
 
 = 065 x /. 
 
 = 06 x /. 
 
 = 0-65 x /. 
 
 = 0-7 x /. 
 
 = 1*1 x /, if single thickness. 
 
 = 0*65 x /, if double ,, 
 
 = l'O x /, if single 
 
 = 06 x /, if double 
 
 */• 
 x/. 
 
 x/. 
 
 x/. 
 
 x/. 
 
 x /, when there is a false face. 
 
 x /, when cast iron. 
 
 = 14 
 = 13 
 
 >> 
 •i 
 
 0-9 
 = 1-2 
 = 10 
 = 08 
 
 For torpedo boats and gunboats, destroyers, and other such ships where 
 extreme lightness is a necessity, and full power is only developed at intervals 
 and for a short time, the scantling may be reduced by 25 per cent. 
 
 Pitch of Studs or bolts in cylinder-cover or valve-box door in inches 
 
 should not exceed 
 
 V 
 
 t x 130 
 
 , t being the thickness of the cover or door 
 
 flange in sixteenths of an inch, p, the pressure per square inch in pounds 
 on it. 
 
 Flat Surfaces. — All flat surfaces of cast iron should be stiffened by webs, 
 
 < 2 x 50 
 V 
 
 These 
 
 or stays of some form, whose pitch should not exceed \J 
 
 webs should be of the same thickness as the flat surface, and their depth at 
 least 2-5 times the thickness. 
 
 The L.P. Cylinder Body or Barrel should be stiffened by external flanges or 
 webs, at about 12 times the thickness of metal apart ; these webs should 
 be 1-5 X / thick, and stand at least 0*75 X / beyond the surface of the 
 cylinder. Some engineers, however, prefer to do without these stiffening 
 webs, and make the cylinder somewhat thicker instead. The low-pressure 
 cylinder, however, differs from the others in size, and in being exposed to 
 external pressure in excess of the internal ; therefore it should have webs, 
 especially when of large diameter. 
 
 •When made of exceedingly good material, at least twice melted, the thickness may 
 be 8 of that given by the above rules. 
 
STUFFING-BOXES AND GLANDS. 
 
 259 
 
 Stuffing '-Boxes and Glands. — For obvious reasons, it is useless giving 
 any definite rules for the sizes of these, as they will differ from different 
 circumstances, and many of the parts do not vary when others are varied. 
 
 Fig. 80. - Combination Metallic Packing. 
 
 T 
 
 ? 
 
 
 "7% 
 YS////, 
 
 -*\- 
 
 C3 I 
 
 i 
 
 i 
 
 Fig. SOre 
 
 Fig. 806. 
 
 The following Table of sizes gives such as are found in good practice, and 
 will be of more use than any abstract rules : — 
 
260 MANUAL OF MARINE ENGINEERING. 
 
 STUFFING-BOXES, etc. 
 TABLE XXXIII.— Stuffing-Boxes for Elastic Packing (Fig. 80a). 
 
 A 
 
 B 
 
 c 
 
 50 lbs. 
 
 D 2 
 
 100 lbs. 
 
 I>, 
 
 150 lbs. 
 
 E 
 
 F 
 
 
 
 to 
 
 
 60 
 
 eft 
 
 ti 
 
 
 
 n 
 
 o 
 
 5 
 
 o 
 
 M 
 
 no 
 
 a 
 
 s . 
 
 .— ia 
 .at- 
 
 .5 
 
 "8 
 
 5 ei ^ 
 
 c -6 * 
 
 
 ^ 
 
 
 "o + 
 
 
 
 
 
 06 d -i_ 
 
 t- o 
 
 u 
 
 •i rt 
 
 *.< 
 
 3"* 
 
 ~« 
 
 »-•< 
 
 o 9 ?^ 
 
 55S** 
 
 5 - 
 
 6 
 4a 
 
 o a 
 
 
 °x 
 
 °x 
 
 
 l<n-« X 
 
 2 *; 
 
 ft) 
 
 a 
 
 2 
 
 So 
 
 5® 
 
 .eo 
 
 4) " 
 
 5* 
 
 a 
 
 2*5 
 
 2 « o 
 
 5 S 
 
 5 
 
 i£ 
 
 a 
 
 q 
 
 o 
 
 o 
 
 o 
 
 03 
 
 a 
 
 Ins. 
 
 Ins. 
 
 Ins. 
 
 Ins. 
 
 Ins. 
 
 Ins. 
 
 Ins. 
 
 Ins. 
 
 Ins. 
 
 i 
 
 § 
 
 H 
 
 1§ 
 
 If 
 
 If 
 
 1 
 
 3 
 
 2 of f i 
 
 i 
 
 § 
 
 11 
 
 14 
 
 If 
 
 2 
 
 1 
 
 s 
 T3 
 
 2 ., | 
 
 1 
 
 A 
 
 n 
 
 If 
 
 2 
 
 2i 
 
 1 
 
 i 
 
 9 i 
 
 H 
 
 A 
 
 2 
 
 U 
 
 24 
 
 2| 
 
 I 
 
 i 
 
 2 „ 4 ' 
 
 l| 
 
 T 
 
 2| 
 
 11 
 
 2i 
 
 24 
 
 I 
 
 I 
 
 2 „ 4 ' 
 
 18 
 
 A 
 
 2i 
 
 2 
 
 2f 
 
 2| 
 
 1 
 
 i 
 
 2., § 
 
 li 
 
 4 
 
 24 
 
 24 
 
 24 
 
 21 
 
 1 
 
 i 
 
 2 ,, • . 
 
 if 
 
 1 
 
 7 
 
 2| 
 
 2i 
 
 n 
 
 3i 
 
 H 
 
 i 
 
 2 „ | 
 
 2 
 
 • 
 
 34 
 
 24 
 
 34 
 
 34 
 
 U 
 
 I 
 
 2 „ | 
 
 2i 
 
 A 
 
 31 
 
 2| 
 
 34 
 
 3J 
 
 H 
 
 5 
 
 2 „ I 
 
 2i 
 
 § 
 
 3| 
 
 3 
 
 3| 
 
 4i 
 
 14 
 
 5 
 
 | 2 ., | 
 
 2f 
 
 § 
 
 4 
 
 3i 
 
 4 
 
 4| 
 
 if 
 
 A 
 
 2 ., 1 
 1 3 „ 2 
 
 3 
 
 H 
 
 *S 
 
 3i 
 
 4£ 
 
 5 
 
 IS 
 
 
 12., 1 
 |3., I 
 
 3i 
 
 li 
 
 4| 
 
 31 
 
 4£ 
 
 5| 
 
 IS 
 
 1 
 
 II:: 1 ! i 
 
 »i 
 
 i 
 
 5 
 
 4 
 
 5 
 
 5| 
 
 2 
 
 § 
 
 12 „ H 
 3 ■„ 1 
 
 3| 
 
 I 
 
 5i 
 
 4* 
 
 5£ 
 
 6 
 
 24 
 
 1 
 
 1 2 „ U 
 |3 „ 1 
 
 4 
 
 13 
 
 Iff 
 
 5| 
 
 4* 
 
 54 
 
 6! 
 
 2J 
 
 1 
 
 12..H 
 (3 ., 1 
 
 H 
 
 i 
 
 6i 
 
 4| 
 
 6 
 
 7 
 
 21 
 
 7 
 
 J3 „ li 
 
 5 
 
 1 • 
 Tff 
 
 6| 
 
 5 
 
 64 
 
 71 
 
 2§ 
 
 7 
 1» 
 
 /2 „ 11 
 
 13., H 
 
 54 
 
 1 
 
 74 
 
 5| 
 
 7 
 
 84 
 
 21 
 
 A 
 
 4„ 14 
 
 6 
 
 H 
 
 8i 
 
 6i 
 
 71 
 
 9i 
 
 3 
 
 7 
 Iff 
 
 4 „ )4 
 
 6J 
 
 IS 
 
 8f 
 
 6| 
 
 8i 
 
 9| 
 
 H 
 
 4 
 
 4 „ 14 
 
 7 
 
 n 
 
 9i 
 
 7 
 
 8| 
 
 104 
 
 34 
 
 4 
 
 4 „ li 
 
 74 
 
 n 
 
 10 
 
 74 
 
 91 
 
 in 
 
 3f 
 
 4 
 
 4 ., li 
 
 8 
 
 if 
 
 10| 
 
 8 
 
 10 
 
 12 
 
 4 
 
 9 
 
 T8 
 
 4 ,. U : 
 
 84 
 
 13 
 
 Hi 
 
 84 
 
 104 
 
 12g 
 
 4i 
 
 9 
 
 4 „ 11 
 
 9 
 
 i} 
 
 12 
 
 81 
 
 n 
 
 13| 
 
 41 
 
 s 
 
 18 
 
 4 ., l! 
 
 94 
 
 14 
 
 12* 
 
 91 
 
 n§ 
 
 144 
 
 4| 
 
 1 
 
 4 „ 1| 
 
 10 
 
 n 
 
 13 
 
 92 
 
 124 
 
 14f 
 
 45 
 
 i 
 
 4 „ 11 
 
 Iii the case of the cylinder, it is usual to make the stufiing-boxes, for 
 uniformity's sake, for the low-pressure and medium -pressure piston-rods of 
 the same depth as that of the high-pressure rod. The stuffing-boxes of the 
 valve-spindles, too, are usually exceptionally deep, on account of their 
 
STUFFING-BOXES. 261 
 
 liability to leak, and the trouble of packing them. The packing of the 
 stuffing-boxes when in the cylinder covers of vertical engines is very liable 
 to give trouble with steam of high pressure from the want of moisture ; the 
 lubricant affects only the top layers of packing, and keeps them soft, 
 while the bottom ones get hard and charred. Metallic packings (v. fig. 80) 
 are the best for use with steam of high-pressure, and although some patent 
 vegetable, packings work very well, these latter are gradually being superseded 
 by the former. The metallic packings are generally arranged in a series of hoops 
 of triangular section, the pressure on the rod being caused either by a second 
 set of hoops outside the first, causing a wedging action on the gland being 
 pressed home, or else by an arrangement of springs or spring clips. The 
 newer and better forms are now giving great satisfaction, and have taken the 
 place of vegetable and asbestos in the high-pressure cylinder, and the medium- 
 pressure cylinder also. Many engineers fit metallic packing in the low-pres- 
 sure cylinder glands, others strongly object to do so, much preferring good 
 vegetable packing on account of the excessive moisture in this cylinder. 
 
 Metallic packing is, however, that now always used in marine engines 
 of 50 N.H.P. and upwards for piston-rods, and generally for all rods exposed 
 to steam above 3 inches diameter. 
 
262 MANUAL OF MARINE ENGINEERING. 
 
 CHAPTER XL 
 
 THE PISTON — PISTON-ROD — CONNECTING ROD. 
 
 The Piston is essentially only a disc, strong and stiff enough structurally to 
 withstand the pressure of the steam on it, and fitting steam-tight in the 
 cylinder. The piston in this simple form is seen in the Richard's Indicator, 
 and is often so fitted to small engines. 
 
 In the early days of steam-engine construction, when there existed no 
 machine capable of truly boring out a cylinder, the bore was not perfectly 
 cylindrical, nor the sides very smooth, and, consequently, unless some form 
 of elastic packing was interposed between the piston and the cylinder sides, 
 it could not work steam-tight. It was customary to form the piston with 
 a recess on the rim, into which rope or junk was coiled, just as it was usual 
 to do with all pumps. This packing could not be examined or renewed 
 without drawing the piston from the cylinder, a tedious operation at all 
 times ; to remedy this the recess was made without a flange at the top or 
 side of the piston next the cylinder cover, but a false flange or loose ring 
 was bolted to the piston so as to retain the junk packing in place, and admit 
 of its being removed or added to without removing the piston from the 
 cylinder. This ring was called the " junk-ring," and retains that name, 
 although junk is no longer used to pack pistons. After a few weeks' work 
 the cylinder was rubbed smooth and fairly true, when the piston would work 
 steam-tight with very little friction, and with moist steam of low-pressure and 
 temperature the packing lasted a considerable time. 
 
 A solid piston — that is, one without packing — is really the best for good 
 working, so long as it remains steam-tight ; but as there is always some 
 slight amount of wear, especially when the cylinder is fresh from the boring 
 mill, and leakage past the piston is most serious, more particularly when the 
 engine is standing still, it is necessary to have some means of adjustment, 
 whereby the piston is maintained a steam-tight fit in the cylinder. 
 
 In lieu of the vegetable packing, which is not admissible with steam of 
 high pressure, engineers now fit metallic rings, called " packing-rings," in 
 various forms, which are pressed outwards against the side of the cylinder 
 by their own elasticity or by springs. These rings are maintained in position 
 steam-tight by the junk-ring as of old. 
 
 When these metallic rings are once in place so as to fit closely to the 
 cylinder sides, there is no need of further lateral pressure until by wear the 
 piston becomes slack, and steam permitted to pass it. However, nearly all 
 existing pistons are automatic in this respect, and the consequence is that 
 the packing-rings press so tightly on the cylinder sides, that the loss by 
 friction seriously impairs fche efficiency of the engine ; and it is only when 
 the ring or cylinder is considerably rubbed away, that the piston works with 
 
PISTON SPRINGS. 263 
 
 ease. From these causes many really good pistons have been condemned 
 after having been made to cause serious damage. 
 
 Perhaps the first remove from the primitive piston is to be found in the 
 form usually fitted in locomotives, and generally known as Ramsbottom's. 
 
 Ramsbottom's Rings (fig. 81). — The late Mr. Ramsbottom, of the L. & 
 N.W.R. Co., was the first to pack pistons by one or more narrow metal rings, 
 turned somewhat larger in external diameter than that of the cylinder bore, 
 and which, after being cut across so as to be capable of being compressed 
 to suit the bore of the cylinder, are fitted into recesses turned in the piston 
 edge. The rings fit easily into these recesses, and as they are so placed 
 that no two of the joints are in a line, the piston is practically steam-tight, 
 and works very well in locomotives and other quick-working engines of 
 small size : but for large engines, and engines undergoing the same vicissi- 
 tudes as those on shipboard, there is an objection to this form of piston. It 
 will be seen that the rings cannot be examined or removed without drawing 
 the piston, and that there is no means of preventing steam from passing 
 where the spring is cut across, besides which the rubbing surface is very small, 
 and the spring is always exertion; its maximum effort. The first of these 
 objections is overcome (fig. 81a/ by fitting a junk-ring, having cast with it 
 a spigot or ring, which goes down into the recess around the piston for the 
 packing-ring, and made steam-tight ; into grooves turned in the outer surface 
 of this spigot the Ramsbottom rings are fitted. 
 
 For small engines these rings are made of steel ; for such engines as may 
 be standing unused for many days, engineers prefer to fit hard bronze rings. 
 For larger engines, where the section may be three-quarters of an inch square 
 and upwards, the rings are better of tough and hard cast iron, which works 
 very well indeed. 
 
 Common Piston-Rings (figs. 82, 85a, 86, and 87) consist only of a single 
 hoop made of very tough, close-grained, cast iron, made on the same principle 
 as the Ramsbottom rings, but fitted steam-tight between the piston flange 
 and the junk-ring, and free to move laterally. This packing ring is usually 
 turned to a diameter about 1 per cent, in excess of that of the cylinder, and 
 either cut across diagonally, or formed so that one end has a tongue fitting 
 into a recess in the other (fig. 87), a brass cover-piece being fitted behind the 
 gap, so as to prevent steam leaking into the space behind the rings. The 
 ring is then fitted to the piston flange steam-tight by scraping both surfaces ; 
 the ring is raised by interposing very thin pieces of paper between it and the 
 flange, and the junk-ring is then fitted steam-tight to the piston and packing- 
 ring by scraping, etc. Some makers of pistons pro/ess to turn the piston 
 and rings so accurately as to require no scraping, and with the perfection of 
 the modern lathe and the high-class tool steel now used this is possible with 
 care ; other engineers prefer to grind the rings tight after coming from the 
 lathe. In whatever w T ay the object is attained is of small moment compared 
 with the necessity of having the ring perfectly steam-tight between the flange 
 and junk-ring. 
 
 Piston Springs. — When the piston is of comparatively small diameter, the 
 elasticity of the packing-ring itself is sufficient to keep it steam-tight against 
 the cylinder sides for a very considerable time after it is fitted ; and even 
 larger rings may be made of sufficient strength to do this, but they would 
 then be open to the same objection as raised against the Ramsbottom rings. 
 
264 
 
 MANUAL OF MARINE ENGINEERING. 
 
 The old method of pressing the ring out by means of dished springs or coach- 
 springs, as shown in fig. 85a, is now seldom used in new engines ; the objec- 
 tions to it are the uneven and unknown pressure exerted, and the reaction 
 of the piston itself, from the fact of the springs pressing on it. It was a very 
 difficult thing to set every spring so that the pressure on the ring was uniform ; 
 
 Fig. 81. 
 
 Fig. 82. 
 
 Fis. 85a. 
 
 Figs. 81 to 87. — Various Forms of Piston Packings. 
 
 and the range of action of this form of spring is very limited, so that, although 
 the ring might be very tight when first fitted, after a few days running it 
 might be passing steam. The surface taking the pressure too was small, 
 and the springs were apt to bed themselves into the ring, and in doing so 
 
VARIOUS FORMS OF PISTON PACKINGS. 265 
 
 wear through their curved ends. These defects were partially remedied by 
 adding to each one or more subsidiary springs on the principle of coach- 
 springs, but that only tended to aggravate the other evil spoken of — viz., 
 the reaction of the piston itself. 
 
 When a piston is moving through its course, and guided therein by tbe 
 rod at one end and the tail rod (or back guides in case of a horizontal engine) 
 at the other, it should be quite free laterally from the packing-ring, which 
 may follow its course freely. When the bore of the cylinder is quite true, 
 and its axis coincides with the line of motion of the piston centre, it is of no 
 consequence if the springs do bear on the piston ; but if the cylinder wears 
 somewhat out of truth in either direction, it is important that the spring-ring 
 ■shall follow the sides of the cylinder freely ; it cannot do this if the springs 
 react from the piston body. 
 
 Cameron's Patent. — Fig. 86 shows a piston-ring pressed out with a corru- 
 gated ribbon of steel ; the lateral pressure here is obtained by the resistance 
 of the spring to being bent into a circle, and by the pressure exerted by the 
 corrugations when the ends of the spring are pressed apart. This spring 
 exerts an almost uniform lateral pressure on the packing-ring without touching 
 the body of the piston, and by making the packing-ring comparatively thin, 
 it will adapt itself to the shape of the cylinder when worn. The pressure 
 on the ring can also be easily and nicely adjusted by packing pieces between 
 the ends of the spring. One great advantage of this spring is that it can be 
 fitted to any piston without condemning any of the parts beyond the springs. 
 
 Mather and Piatt's Patent. — It was found that metallic packing-rings not 
 •only wore sideways, but also on the edges, so as to become slack between the 
 flange and junk-ring ; a very slight amount of play with heavy rings 
 causes a very large degree of slackness from the continual concussion on 
 change of motion at every stroke. To obviate this the ring was formed with 
 inside flanges, as shown in fig. 83, and split into two, a spiral hoop, having 
 three or four turns, being coiled inside the rings, whose action is to press the 
 packing-rings outward against the cylinder sides, and up and down against 
 the flange and junk-rings. This form has been generally very successful, and 
 pistons so fitted have worked very well indeed ; but there is the objection 
 that no adjustment of the spring is possible, and it is always exerting its 
 maximum effort. The chief part of the elasticity of the spring, however, is 
 exerted in pressing the rings against the flange and junk-ring, and the friction 
 so caused helps to prevent undue pressure, on the cylinder side, so that in 
 practice it is not found that there is excessive side pressure when first fitted, 
 nor lack of it when the cylinder is worn. These springs are made of steel, or 
 very strong cast iron, cut out of a ring of either metal. They are also some- 
 times cast to the form required. 
 
 Buckley's Patent consists of two rings, of section as shown in fig. 84 ; a 
 spiral coil of steel wire is bent into a circle, and inserted between the two 
 packing-rings. Pressure is exerted in the same way as in Cameron's spring, 
 and tends to press the packing-rings both outwards and against the flange 
 and junk-ring. This form of piston is still often used at the present time, 
 and when properly adjusted works very well. The spring, however, is a very 
 stiff one, and requires but little end displacement to exert a very considerable 
 pressure on the sides of the cylinder. 
 
 Qualter and Hall's Patent. — This piston has two packing-rings of tri- 
 
266 
 
 MANUAL OF MARINE ENGINEERING. 
 
 angular section, with a third ring inside and between them, as shown in fig. 
 85. so that on this inner ring being pressed outwards, it exerts a wedge action 
 on the two packing-rings so as to force them against the flange and junk-ring. 
 Coach-springs are employed to press the ring outwards, which are each held 
 in a brass frame having a tapered piece on the back, which fits into a recess 
 in the piston and against a tapered cotter ; this cotter can be pressed down 
 by a set screw in the junk-ring, and any required pressure is imparted to the 
 ring through the spring in this way. This piston has, therefore, the advan- 
 tage of being capable of adjustment without removing the junk-ring, and the 
 adjustment can be made to a nicety with very little trouble, and also, in 
 case of the engines not being required for some weeks, the pressure on the 
 springs may be relieved until required again. 
 
 Fig. 88. — Restrained Packing-rings (Admiralty Plan). 
 
 Rowan's Patent and MacLaine's Patent each consists of two strong rings 
 of square section, or of U section, pressed outwards by springs at the cross 
 cut, and held against the flange and junk-ring by wave springs fitted in a 
 groove between them. This arrangement has been found to work well. 
 
 Restrained Packings. — In the early days of piston-making designers were 
 chiefly concerned in providing means for keeping the packing-rings on tbe 
 cylinder walls ; as pressures of steam increased they still experienced the 
 same anxiety lest there should be loss by leakage. To-day we have less 
 anxiety on that score, but a more serious, how to avoid the loss due to the 
 rings being pressed unduly hard on the cylinder by the steam now used 
 
FORGED- AND CAST-STEEL PISTONS. 
 
 267 
 
 getting behind them. We cannot prevent the steam from getting behind 
 the rings, so it is necessary to restrain them ; this was first done by securing 
 the ends or locking them together by the same device that set them. With 
 the very high pressures now employed even that is not sufficient, so it is the 
 practice in the Navy to fit solid restraining-rings, as shown in fig. 88, to keep 
 the cut rings in place when fitted, as in right-hand illustration of same. 
 
 Body of the Piston. — Pistons of small size are usually made of a single 
 thickness of metal without, of course, any stiffening webs or ribs (v. fig. 89). 
 
 Fig. 89.— Forged -steel Pistons. 
 
 Pistons of very considerable size have been made of cast steel in this way ; 
 they are in the form of a cone, and shaped to suit the cylinder end (v. fig. 90) ; 
 by that means they have the requisite degree of stiffness; originally they 
 were made in this form to save weight, being for fast-running horizontal 
 engines, but now they are used extensively in all fast-running engines. Cast- 
 steel pistons are frequently used in the mercantile marine in engines of all 
 sizes ; chiefly, however, when in those running at high speeds, and always in 
 those of large size. They may be used with advantage in most engines, and 
 are not much more costlv than hollow cast-iron ones, especially when of large 
 
 Fig. 90. — Cast-steel Pistons. 
 
 size. The pistons for very light, fast-running machinery are usually made of 
 forged steel, and are very thin at the flange and near it. Solid cast-iron 
 coned pistons can be used with advantage. 
 
 Pistons of ordinary marine engines above 20 inches diameter for H.P., 
 and 40 inches diam. for L.P. cylinders, are usually made of cast-iron cellular 
 — that is, with two thicknesses of metal stiffened or connected by ribs and 
 webs, and either by the thickness of metal or by the depth of body made 
 strong enough structurally to safely withstand, not only the load due to 
 
268 
 
 MANUAL OF MARINE ENGINEERING. 
 
 steam pressure exerted on it and transmitted to the rod, but also the shocks 
 to which it is liable when priming occurs. 
 
 The piston body also must be so designed that it may be safely cast, for 
 in the early days of large pistons it was not at all an uncommon thing for a 
 piston to crack in cooling, or do so mysteriously afterwards. For this reason 
 any rules must of necessity be empirical which set out the thickness of metal 
 of the different parts of the body ; but care must always be exercised by the 
 designer that no one part is too small for the stresses to which it is subject. 
 For example, there must be sufficient metal in the immediate neighbourhood 
 of the piston-rod boss to resist the tendency to force out the centre by shearing 
 the metal. Again, the piston may be taken as consisting of a number of sectors, 
 and by considering one of such small sectors loaded with the pressure on its area 
 at the centre of gravity of its figure, the bending moment at any section may be 
 found, and the thickness of metal tried whether it be sufficient for the purpose. 
 
 For the section of an ordinary piston having a single rod, the following 
 table gives the multipliers for obtaining the thickness of metal and sizes of 
 the different parts. 
 
 Details of Construction of the Ordinary Piston. — Let D be the diameter of 
 the piston in inches, p the maximum effective pressure per square inch on it, 
 x a constant multiplier, found as follows : — 
 
 D ._ 
 
 X = 50 ( v V + !) ' 
 
 For high-pressure cylinders p may be taken at half the absolute boiler 
 pressure ; for medium-pressure cylinders a quarter that pressure ; and for 
 low-pressure cylinders half the pressure, divided by ratio of low-pressure to 
 high-pressure cylinders. For quadruple engines the value of p in the first 
 medium-pressure is 03 the press., and in tb* second is 016 the absolute pressure. 
 
 ^Hollow Cast-Iron Pistons. 
 The thickness of front of piston near the boss 
 
 >> 5 » >» 5 j rim 
 
 ,, back ,, 
 
 ,, boss around the rod 
 
 ,, flange inside packing-ring 
 
 „ at edge 
 
 ,, packing-ring 
 
 ,, junk-ring at edge . 
 
 „ ,, inside packing-ring 
 
 ,, ,, at bolt holes . 
 
 „ metal around piston edge 
 
 The breadth of packing-ring . 
 ,, depth of piston at centre 
 ,, lap of junk-ring on the piston . 
 „ space between piston body and packing-rin 
 „ diameter of junk-ring bolts 
 „ pitch „ „ 
 
 „ number of webs in the piston . 
 
 ., thickness ,, .... 
 
 . =0-2 
 
 X X. 
 
 . = 0-17 
 
 X x. 
 
 . = 0-18 
 
 X X. 
 
 . = 0-3 
 
 X x. 
 
 . = 0-23 
 
 X x. 
 
 . = 0-25 
 
 X x. 
 
 . = 0-15 
 
 X x. 
 
 . = 0-23 
 
 X x. 
 
 . = 0-21 
 
 X x. 
 
 . = 0-35 
 
 X x. 
 
 . = 0-25 
 
 X x. 
 
 . = 0-63 
 
 X x. 
 
 . = 1-4 
 
 X x. 
 
 . = 0-45 
 
 X x. 
 
 g=0-3 
 
 X x. 
 
 . = o-i 
 
 X x + 0-25 in 
 
 . = 10 diametets. 
 
 I) + 20 
 12 * 
 
 . = 0-18 
 
 X x. 
 
solid packings. 269 
 
 Solid Pistons. 
 
 Cast Steel. Cast Iron 
 
 Thickness near boss . . . . = - 26 x x = 052 X x. 
 „ rim . . . . = 0-15 x x = 022 x x. 
 
 Forged-Steel Pistons. 
 
 Thickness near boss . . . . . . . = 02 x x. 
 
 „ rim = 0*1 x x. 
 
 When made of exceptionally good metal, at least twice melted, the thick- 
 nesses of cast-iron pistons may be as much as 20 per cent, less than given by 
 the rules ; but, on the other hand, if made of other than really good metal, 
 they should be thicker. The piston should be made of good metal always,, 
 and for fast-running engines it is better made of steel. The packing-ring 
 was sometimes made thicker in the part opposite the cut than given above, 
 in order to have sufficient elasticity of itself to press steam-tight against the 
 cylinder ; but it is better to let the springs perform their function wholly, 
 and leave the ring to act only as the packing. 
 
 Junk-ring Bolts. — When screw-bolts are used to hold the junk-ring in 
 place, they are either screwed into a brass nut let into a recess in the side 
 of the piston (v. fig. 85a), or else screwed into a brass plug, which has been 
 screwed tightly into the piston. The former plan is most general, and ha& 
 the advantage that if the bolt thread is torn away the nut can be easily 
 replaced, and owing to the length of body the bolts cannot slack themselves 
 back. Some engineers, however, prefer to screw the bolts directly into the 
 cast iron, making the tapped hole as deep as possible ; and although it may 
 be supposed the bolts would set fast by rust, practice has shown that such 
 does not take place, nor do the cast-iron threads wear quickly away. 
 
 Studs are often used instead of screw-bolts (v. fig. 88), but although. to> 
 some extent, possessing advantages over the latter, they are not so con- 
 venient, and have all to be withdrawn when any refitting of the packing 
 and junk-rings is necessary. 
 
 Safety-rings and Lock Bolts. — The vibration of the junk-ring has a tend- 
 ency to slack back the bolts, and although it is a rare occurrence to find 
 such a thing happen in a vertical engine, very serious accidents have fre- 
 quently been caused by the junk-ring bolts getting loose in a horizontal 
 cylinder. To prevent the possibility of a casualty the piston bolts of all 
 engines should have a light wrought-iron ring secured to the junk-ring by 
 studs, having square bodies and nuts secured with split-pins ; this ring 
 (r. fig. 88) fits close to the heads of the bolts, and prevents them then from 
 turning. When studs are used their bodies should be square or heart-shaped 
 with a projecting side, the holes in the junk-ring corresponding in size and 
 form to them, so that, when on, it prevents them from unscrewing ; the nuts- 
 may be prevented from slackening by a ring, or each stud may have a split- 
 pin through its end. 
 
 There are some other methods, but none of them are either so efficient or 
 so inexpensive as the above. 
 
 Solid Packings. — In order that the weight of the piston of a horizontal 
 engine may be taken by the broad packing-ring, instead of by the compara- 
 tively narrow flange and junk-ring, it was customary and advisable to fit a. 
 
270 
 
 MANUAL OF MARINE ENGINEERING. 
 
 cast-iron packing between the body of the piston and the packing-ring for 
 about one-third of the circumference in lieu of springs. The pistons of 
 ■diagonal and oscillating cylinders are also better if fitted in this way, and 
 
 gjCg 
 
 
 i 
 
 
 
 
 
 i 
 
 J. 
 
 1 
 
 i 
 
 i 
 
 L 
 
 o 
 
 Fig. 93. 
 Piston-rod Ends and Guide Blocks, &c. 
 
 heavy pistons of vertical engines may be also dealt with in this way to provide 
 for the heavy rolling of the ship. 
 
PISTON-ROD. 271 
 
 Piston-rod. — It is usual to have only one rod to each piston of a direct- 
 acting engine, but some manufacturers, to suit a particular style of crosshead 
 and connecting-rod, fit two. The single rod is preferable from practical 
 considerations, even for large engines, because it requires very considerable 
 care on the part of the' workman to bore the two holes in the piston, cylinder 
 bottom, and crosshead so exactly that the rods will fit into their place without 
 adjustment ; the friction of the two stuffing-boxes will be very considerably 
 more than that of the one larger one ; the cost of labour will also be nearly 
 double that for the single rod, and there are two stuffing-boxes, which require 
 packing, and two glands demanding attention, instead of one. 
 
 Return connecting-rod and large steeple engines of necessity required 
 two rods to each piston, and Messrs. Humphreys fitted four rods in the case 
 of the very large pistons of H.M.S. " Monarch," the better to distribute the 
 load over the piston face, and to admit of a better form of crosshead. 
 
 Diameter of Piston-rod. — Since the piston-rod is secured in the piston, 
 and usually well guided at the other end, so that it cannot bend without 
 meeting with considerable resistance, it may be treated as a strut or column, 
 secured at both ends ; but when the outer end fits into a crosshead, which 
 would offer little or no resistance to bending, as in a paddle-wheel diagonal 
 engine, then the rod must be treated as a column loose at one end and secured 
 at the other. 
 
 From Mr. Hodgkinson's experiments, and Mr. L. Gordon's investigations 
 the following are the formulae for computing the strength of columns : — 
 
 / S 
 
 (1) For a column fixed at both ends, P = — ~~w- 
 
 / S 
 
 (2) For a column loose at both ends, P = '- — „. 
 
 2 / S 
 
 (3) For a column fixed at one end only, P = '■ — p. 
 
 2+5^ 
 
 P is the load, I the length of the column in inches, d the diameter in inches. 
 a for solid wrought iron and mild steel fI / OT , and / 36,000 lbs. per square 
 inch, S being the area of section of the rod in square inches. Taking this 
 value of / in the above formula?, P is the breaking load ; since it is usual to 
 have a factor of safety for all important parts of a marine engine, of at 
 least 6, the value of / for this reason should not exceed 6,000 lbs. for 
 mere strength ; but as the piston-rod is liable to great shock, is always 
 working with alternating stresses, and always receives its load suddenly, 
 and must be rigid and without quiver, 3,000 lbs. should be taken as the 
 value of / to calculate the diameter. These formula?, however, are too com- 
 plicated for general use, but the size of a piston-rod may be checked by them 
 easily after having been calculated by an empirical formula. 
 
 Since, however, an approximate value may be safely taken for the relation 
 
272 
 
 MANUAL OF MARINE ENGINEERING. 
 
 between / and d of ordinary marine engines, the formulae may be reduced to 
 a verv workable form. Hence, since 
 
 P = — -xp; 
 
 and / S = -r- X /. 
 
 D being the diameter of the piston, and p the effective pressure on it in 
 pounds per square inch. 
 
 p.D 2 = /d 2 -l-f JL 
 
 Let 
 
 / 
 
 -7 be represented by r, 
 
 Then 
 
 taking / = 3,025. 
 
 = D V? 
 
 (1+ar 2 ) 
 
 D 
 
 Rule : — Diameter of piston-rod = rr vp (1 -f- « r 2 ) 
 
 00 
 
 (1) 
 
 >/p 
 
 The following are the values of r : — 
 
 . r = 25. t- \/p 
 
 . r = 30 
 . r = 35 -5- 
 . r = 40 4- Jp 
 
 . r = 50 -5- -Jp_ 
 „ „ medium ,, and compound, r = 37 -s- Vp 
 
 The following rules, however, will give results sufficiently accurate for 
 
 all practical purposes : — 
 
 _ , ~. . . , , diameter of cylinder 
 Rule : — Diameter of piston-rod = 
 
 Short stroke direct-acting engines. 
 Long 
 Very long 
 
 Oscillating engines of short stroke 
 
 long „ 
 
 F 
 
 J¥. 
 
 The following are the values of F 
 
 Naval engines, direct-acting, . 
 
 ,, return connecting-rod, 2 rods, 
 
 Mercantile ordinary stroke, direct-acting, . 
 
 long 
 medium 
 
 oscillating. 
 
 F = 50 
 F = 70 
 F = 45 
 F = 42 
 
 F = 40 
 
 Note. — Long and very long, as compared with the stroke usual for the 
 power of engine or size of cylinder. 
 
 D is the diameter of the low-pressure and d that of the high-pressure 
 cylinder of a compound system ; d v rf.>, etc., the diameters of the intermediate 
 cylinders ; R is the ratio of low-pressure to high-pr ssure cylinder volumes ; 
 /•j the ratio of volumes of the first intermediate and high-pressure cylinders ; 
 r 2 that of the second intermediate and high-pressure cylinders ; if the stroke 
 be the same in each cylinder 
 
 D 2 dj dj 
 
 d 2 ' Tl ~ d-' % ~ d 2 ' 
 
PISTON-ROD ENDS. 273 
 
 The maximum pressure repeatedly applied on the pistons when working 
 full speed are approximately as given by the following formulas : — Where 
 p„ is the absolute pressure at or near the engines, and may be taken as the 
 load on safety-valve plus 15 lbs. for purposes of calculation of sizes of piston- 
 rods, connecting-rods, columns, etc., and their bolts and fittings. 
 
 Values of p or maximum effective working pressure»in 
 
 High-pressure cylinder of a compound engine, . .' . p a X 0*75 
 
 Low-pressure ,, ,, ,, p a + (R X 1*2) 
 
 High-pressure ,, triple-compound engine, . . p a X - 63 
 
 Medium-pressure ,, ,, ,. . . p a -*- (r x X 1*5) 
 
 Low-pressure ,, ,, ,, . . p a -i- (R X 1 - 5) 
 
 High-pressure ,, quadruple-compound engine, . p a X 0*55 
 
 1st medium-pressure cylinder of a quadruple-compound engine, p a -f- (r : X 1*65) 
 
 2nd ,, ,. ,, ,, „ p a -f- (r. 2 X 1 '55) 
 
 Low-pressure „ „ „ „ p a -r- (R x 1-65) 
 
 The piston-rods in a compound engine are all of one size, except in the 
 case of the triple, with two low-pressure cylinders, whose rods may be and 
 usually are smaller than those of the high-pressure and medium-pressure 
 cylinders. It is at all times desirable that the piston-rod shall move through 
 the stuffing-box without vibration, but especially is this so when metallic 
 packing is used. It is, therefore, a good thing to have the piston-rods larger, 
 rather than smaller, than given by the above rules. 
 
 Piston-rod Ends. — It is absolutely necessary that the rod should fit perfectly 
 steam-tight into the piston, and also be of such a taper as not to " draw " 
 in the least when subject to shock. If the rod end were made cylindrical or 
 1 parallel," as it is technically called, and fitted in to satisfy the above con- 
 ditions, it would be very tedious and difficult to get it out again. For this 
 reason principally it is usual to turn the part fitting steam-tight into the 
 piston " taper " or conical. If the taper is very slight the rod can be easily 
 made a tight fit, but unless formed with a shoulder at the end of the taper, 
 it would in time become so tightly held by the piston as to withstand all 
 attempts at withdrawal ; moreover, there would be at all times a great 
 danger of splitting the piston by the wedging action. If the taper extends 
 the whole depth of the piston, it should be at the rate of f inch to the foot ; 
 that is, the diameter of the rod at back is less than that at the front by one- 
 sixteenth of the length of taper. Even with so liberal an allowance as this, 
 great difficulty is often experienced in withdrawing the rod after a few months' 
 work ; for this reason, and to obtain a larger screwed end, some engineers 
 do not extend the taper the full depth of the piston. The most convenient, 
 and at the same time reliable, practice is to turn the piston-rod end with a 
 shoulder of -3- inch for small engines, and | inch for large ones, make the 
 taper 3 inches to the foot (fig. 94) until the section of the rod is three-fourtls 
 of that of the body, then turn the remaining part parallel ; the rod should 
 then fit into the piston so as to leave £ inch between it and the shoulder 
 for large pistons, and y T inch when small. The shoulder prevents the rod 
 from splitting the piston, and allows of the rod being turned true after long 
 wear without encroaching on the taper. 
 
 18 
 
274 
 
 MANUAL OF MARINE ENGINEERING. 
 
 It was usual to prolong the piston-rods of vertical engines to admit of 
 the " tail" end passing through a stuffing-box in the cylinder-cover, and so 
 help to guide the piston, and prevent its unduly wearing the cylinder. Since 
 gravity prevents moisture getting to the packing of this stuffing-box, and the 
 lubricant applied externally soon gets carried through to the cylinder, some 
 trouble is experienced in keeping it steam-tight ; the rolling of the ship also 
 causes the piston to exert pressure sideways on the gland and packing, and 
 further aggravates the evil : in other words, it is an unsatisfactory guide. 
 For these reasons it is preferable to simply fit a brass or white metal bush 
 in the cover for the " tail " end to work in, and case it with a dome or sheath 
 fitted steam-tight and true on the cover ; a couple of spiral grooves in the 
 side of the bush will admit and release the steam. But, on the whole, it 
 is very doubtful if these tail-rods are efficacious, and certainly they cannot 
 be so beneficial to the good working of the cylinder as a piston with broad 
 bearing surfaces. It should be noted that when the packing-ring is pressed 
 out by springs acting independently of the body of the piston, it is advisable 
 to form the piston with greater depth of flange and junk-ring. 
 
 The piston is secured to the rod by a nut, and the size of the rod at the 
 
 Fig. 94. — Piston-rod Crosshead. 
 
 nut should be such that the stress on the section at the bottom of the thread 
 does not exceed 7,000 lbs. The depth of this nut need not exceed the diameter 
 which would be determined by allowing this stress. To avoid the large cavity 
 which is necessary in the cylinder-cover for the piston-rod nut, some engine 
 builders recess it into the piston ; this recess does not materially affect the 
 strength of the piston, and the plan may be followed with advantage. Al- 
 though piston-rod nuts seldom work loose, and those of vertical engines are 
 less liable to this than are others, still as a measure of safety in all cases a 
 taper-split pin should be fitted to the rod behind the nut, and in the case of 
 large engines it is usual to fit a " lock" plate to the nut itself, or to adopt 
 some other means of preventing it from moving at all when at work. 
 
 Cast-steel Piston-rod Crosshead. — Fig. 95 represents the modern form of 
 crosshead made of cast steel and designed in such a way as a casting permits 
 of. It is, of course, much lighter and cheaper than that of forged material, 
 shown in fig. 94, and is especially suited to fast-running engines which require 
 a large bearing surface. These crossheads can now be cast quite sound and 
 of excellent material so as to be practically as good as a steel forging. 
 
PISTON-ROD GUIDES. 
 
 275 
 
 The stresses on the various sections are, as a rule, light, as the governin° 
 requirements of surface naturally causes it to be much larger than would be 
 the case for mere strength. The gudgeons can be cast solid, or lightened 
 out as shown. The example shown in fig. 95 has white metal fitted to both 
 the head-going and stern-going sides, and is without a loose slipper. Such 
 however, can always be fitted to this type of crosshead if desired. 
 
 95. — Cast-steel Crosshead. 
 
 When cast steel is objected to as material for the gudgeons, a composite 
 design has been adopted whereby the crosshead is made of hard-wrought 
 steel and fitted to the cast-steel slipper, etc., and secured by studs and nuts. 
 
 Piston-rod Guides. — The pressure on the piston is transmitted through 
 the piston-rod to the connecting-rod, and the reaction of the latter rod acts 
 in the direction of its length ; consequently, when the connecting-rod is not 
 in line with the piston-rod, the force of its reaction can be resolved into two 
 component forces, one in the direction of the piston-rod, and the other per- 
 pendicular to it. This latter force is usually called the " thrust of the 
 connecting-rod," and unless specially prevented, would tend to bend the 
 piston-rod. To prevent such an occurrence, and to preserve the piston-rod 
 in its true course, a guide is provided, and the piston-rod end fitted with 
 blocks or slippers to work in it. This thrust varies from at the end of the 
 stroke to its maximum point, which is towards the point when the crank is 
 
 B 
 
 Fig. 96. 
 
 rit a right angle to the centre line through the cylinder, and depending on 
 the cut-off point; when steam is cut off past half-stroke, then, neglecting 
 inertia effects, this is exactly the point of maximum thrust. To determine 
 the magnitude of the thrust when P is the total effective load on the piston, 
 S the stroke, and L the length of connecting-rod, represented by A B, fig. 96 ; 
 completing the parallelogram by the dotted lines A E, BE, A C, then the 
 
276 MANUAL OF MARINE ENGINEERING. 
 
 reaction, R, of the connecting-rod, represented in direction and magnitude 
 by A B, will be resolved into two forces, P and Q, represented by the lines 
 E B, A E, both in direction and magnitude. P must equal the load on the 
 piston, and Q the thrust on the guides, may be obtained by measuring B C 
 on a graphically constructed diagram, or by geometry calculated as- 
 follows : — 
 
 A C 2 = A B 2 - B C 2 = L 2 
 
 -(!)' 
 
 Now, 
 
 
 
 
 P 
 
 :Q: 
 
 :AC:B 
 
 Therefore, 
 
 
 
 
 
 
 S 
 
 Q = 
 
 p 
 
 BC 
 
 p 
 
 
 
 2 
 
 
 * AC " 
 
 
 
 
 -- /S\2 
 
 -p s 
 
 ^/4L 2 - S 2 
 
 V " V2/ 
 Or, by Trigonometry, 
 
 Q = R sine B A C, 
 
 p 
 
 P = R cosine B A C, or R = — : ~ . _, = P secant BAG. 
 
 cosine B A C 
 
 Therefore, 
 
 sineBAC ptanBAC 
 
 cos B A C 
 
 The angle B A C is found by knowing its sine to be half the stroke- 
 -r length of connecting-rod. 
 
 Example. — To find the thrust taken on the piston-rod guide of an engine 
 whose piston load is 100,000 lbs.; the length of stroke is 60 inches, and the 
 connecting-rod is 120 inches long. 
 
 Thrust = 100,000 x 6 ° = 25,819 lbs. 
 
 ^4 x (120) 2 - 60 2 
 
 Surface of Guide-block. — The area of the guide-block, or slipper-surface 
 on which the thrust is taken, when going ahead should be sufficiently large 
 to prevent the maximum pressure exceeding 100 lbs. per square inch. When 
 the surfaces are kept well lubricated this allowance may be exceeded, but 
 the reduction in surface should be effected by making shallow grooves and 
 recesses in the face of the slipper, in which *he lubricant can lodge and impart 
 itself to the guide as it is carried along. A good method of carrying this 
 into effect is to provide a surface calculated on the allowance of 100 lbs. per 
 square inch, and by cross planing so as to leave shallow recesses about -^ inch 
 deep, reduce the actual surface which touches the guide to about f of the 
 original area ; there will then be strips across the slipper 1-^ inches wide, 
 with depressions between them f inch wide, filled with grease. 
 
 When, however, the piston speeds exceed 300 feet per minute the area 
 of guide-block should be increased to ensure continuous safe running, and 
 follow this rule : — 
 
 Working pressure per square inch = 1,760 -=- vS + 100. 
 
 For example, an engine running at a piston speed of 900 feet per minute 
 
PISTON-ROD CROSSHEADS AND GUDGEONS. 277 
 
 should have guide-blocks such that the maximum pressure per square incb 
 
 does not exceed Jm + & , or o5 lbs. 
 
 Guide-blocks designed on these lines for ahead motion have always ample, 
 provision for astern-going. Then in any engine 
 
 r> , r, t A M 1 V. T X J (S + 100) 
 
 itw/e : — Gross area of guide- block shoe = 7ft fT~ sc l' ms ' 
 
 Cast iron, hard and close-grained, is really good material for the guide 
 plates ; its surface, after a few days' work, becomes exceedingly hard and 
 highly polished, and offers very little resistance to the slipper or guide-block. 
 So long as this hard skin remains intact, no trouble will be experienced, but 
 if abrasion takes place from heating or other cause, it rarely works well after 
 and should b? at once planed afresh, or, better still, ground smooth. 
 
 The slippers or facing plates fitted to the piston-rod or crosshead were 
 sometimes made of bronze ; but bronze never gets that smooth hard skin so 
 essential to good and efficient working, and when once the surface is grooved 
 and scratched, it will wear away very rapidly. Fine grain cast iron is, after 
 all, the best metal for this purpose, if care is taken at the first working of the 
 engine to run for a few hours at easy speed, so as to rub down and polish 
 the surfaces ; after this is once thoroughly done, cast-iron surfaces will 
 continue to work well with very slight attention. White metal is, however, 
 now generally used for the facing of slippers, and works very well, and for 
 high speed is reliable for good and safe working. The best way of using 
 white metal for this purpose is to fit strips of this material into grooves planed 
 across a cast-iron slipper, and leave them standing from ^ to \ inch above 
 the cast iron. The strips should be about 2 inches wide, and the space 
 between them from § to \\ inches, into which the lubricant can collect and 
 lodge, as before described. A slipper fitted in this way is shown in fig. 92. 
 It is cheaper, however, and more general to cast the white metal practically 
 over the whole surface of the slipper, where it is retained in place by " tinning " 
 and underlying the edges of the recess provided for it. It is usual to cut 
 •oil- ways in the face of the guide, which distribute the oil across it, and metal 
 combs secured to the slipper dip into the oil receiver at the end of the guide, 
 and smear the face on the return stroke. 
 
 The guide plates are sometimes planed across, instead of the slippers, for 
 the same purpose of retaining the lubricant, or have circular grooved rings 
 dotted about their surface. 
 
 Piston-rod Crossheads and Gudgeons. — When there are two piston-rods, 
 as in the case of the return connecting-rod engine, they are united to a 
 common crosshead, having a turned journal in the middle for the connecting- 
 rod to work on ; or else a bearing is fitted to the crosshead, in which a gudgeon 
 on the connecting-rod works, ^he former (fig. 97) is the better and usual 
 plan under ordinary circumstances. This crosshead is of wrought-iron or 
 ■steel, and made of a form suitable to the circumstances, and arranged to 
 work in guides. The diameter at the middle must, of course, be sufficient 
 to withstand the bending action, and generally from this cause ample surface 
 is provided for good working ; but in any case the area, calculated by multi- 
 plying the diameter of the journal by its length, should be such that the 
 
278 
 
 MANl'AL OF MARINE ENGINEERING. 
 
 pressure does not exceed 1,200 ibs. per square inch, taking the maximum 
 load on the piston as the total pressure on it. 
 
 Let L be the distance of the centres of the piston-rods in inches, and P 
 the maximum load on the piston in pounds, then for strength 
 
 Diameter of crosshead should not be less than 
 For good wearing, / being the length of the journal. 
 
 Diameter of crosshead should not be less than 
 
 V P X L 
 18 
 
 P 
 
 1.200 x r 
 
 With fast-running engines the following rule may be followed where P 
 is the maximum load on the piston, d is the diameter and / the total length 
 of gudgeon or crosshead arms in inches, and R the revolutions per minute. 
 
 Then for any engine 
 
 Diameter of crosshead = 
 
 P X * /R + 100 
 12,500 
 
 P x ^R + 100 
 12,500 x / " 
 
 Fig. 97. — Crosshead and Guide-block for double Piston-rods. 
 
 Of course, the maximum load on the crosshead is really the reaction of 
 the connecting-rod, but, to avoid any complication of the calculation, it is 
 sufficient to take the load on the piston. 
 
 When the diameter of the crosshead journal is calculated by the first 
 rule, the length is usually made equal to it. 
 
 Direct-acting engines have sometimes a gudgeon secured to the connecting- 
 rod (fig. 98), which works in a bearing in the piston-rod end (figs. 91 and 
 92) ; or have the gudgeon at the piston-rod end (figs. 93, 94. and 95). and 
 connecting-rod (figs. 99 and 100) swung on it by brasses, etc., on either side. 
 By the latter plan larger bearing surfaces are obtainable, and the brasses, 
 being on the outside of the rods, are much easier watched and adjusted ; 
 on the other hand, there are two sets of bolts, brasses, etc., to lubricate 
 and keep in order, and there is the liability by careless adjustment to put the 
 
PISTON-ROD CROSSHEADS AND GUDGEONS. 
 
 279 
 
 whole of the load on one side only. In the main, however, this plan is a 
 preferable one for large and heavy rods, and it is one which admits of the 
 piston-rod being fitted into its end (figs. 94 and 95) instead of forged with 
 it. This latter advantage is well worthy of consideration, for it is highly 
 important that the piston-rod shall be quite free from flaws and reeds on its 
 
 I 
 — — i — 
 
 I 
 
 
 ^ 
 
 Fig. 98. 
 
 Fig. 99. — Connecting-rods. 
 
 surface, otherwise the packing soon gets damaged, and it is then found im- 
 possible to keep the glands from leaking. A steel rod rarely has them, but 
 if the rod simply fits into the crosshead or rod end, this latter may with 
 advantage be made of cast steel, while the rod may be of forged steel of any 
 desired quality, easily returned or ground in the lathe and cheaply replaced. 
 Smaller engines are better with the gudgeon shrunk into the jaws of the 
 
280 .MANUAL OF MARINE ENGINEERING. 
 
 connecting-rod, and working in brasses fitted into a recess in the piston-rod 
 end (fig. 91), and secured by a wrought cap and two bolts. 
 
 The diameter of the gudgeon = T125 X diameter of piston-rod. 
 ., length „ ,, = T4 X ,, ,, 
 
 The area obtained by multiplying these is exactly double the area of the 
 piston-rod section, and so if the maximum stress per square inch on the rod 
 does not exceed 2,400 lbs., the above rule holds good ; if this allowance of 
 stress is exceeded in calculating the rod, then the length of the gudgeon should 
 be increased until the pressure on the section, as calculated by multiplying 
 length by diameter is in accordance with the following : — 
 
 12 500 
 Rule. — Pressure per square inch on gudgeon = _^!_. — — lbs. for revo 
 
 lutions R per minute. v R + 100 
 
 When the gudgeon is fixed in the piston-rod (fig. 93), or formed with the 
 guide ends as in fig. 94, the length of each end should be not less than - 75 the 
 diameter of the piston-rod. 
 
 The brasses when fitted into the piston-rod should be square-bottomed 
 and of hard gun-metal, with good oil-ways, as owing to the motion being 
 through a comparatively small angle only, the lubricant is not so easily 
 spread. Some white metal does not work well on gudgeons and crossheads ; 
 it has a tendency to abrade, and to wear the journal oval. This is generally 
 observed whenever white metal is used on a bearing subject only to a small 
 angular motion. White metal, however, can be used with mild steel gudgeons, 
 especially in fast-running engines with advantage. 
 
 Gudgeons, when fitted to the connecting-rods, should be made of hard 
 iteel or mild steel case-hardened ; great care is required, if of the latter, 
 that they are carefully ground true after hardening. 
 
 The bolts securing the brasses should be of the best tempered mild steel ; 
 they should be of such a size that the stress per square inch of section at 
 the bottom of the thread does not exceed 6,500. It is true that where extreme 
 lightness of machinery is a sine qud non, these stresses have been exceeded 
 by some engineers ; but even under these circumstances it is unwise to exceed 
 them by more than 10 per cent., as the saving of weight effected is exceedingly 
 small, especially when compared with the risk run (v. Table lxxix.). 
 
 When the bolts are less than 2 inches in diameter, owing to the uncertainty 
 of the depth of thread, etc., the allowance should be not more than 6,000 lbs. 
 
 Rule. — Diameter, of piston-rod bolts at the bottom of thread 
 
 -»/ 
 
 r 
 
 D is the diameter of piston, p, the effective pressure on it per square inch. 
 
 Where there are two bolts of steel/ = 13,000 lbs. ; under 2", 12,000. 
 four „ J = 24,500 lbs. ; under 2", 22,000. 
 
 The cap is of the same width as the piston-rod end — viz., 1*15 X diameter 
 of the rod, and its thickness equal to the diameter of the bolts at the bottom 
 of the thread. 
 
 Connecting-rods. — The length of the connecting-rod measured from the 
 centre of the gudgeon to the centre of the crank-pin should, if possible, be not 
 
CONNECTING-RODS 281 
 
 less than twice the stroke. Quick-running vertical engines have almost 
 invariably the connecting-rods twice their s'troke in length ; other vertical 
 engines from twice to three times, but generally two and a quarter times is 
 not exceeded. 
 
 A connecting-rod may be viewed as a strut loose at both ends, the 
 
 f S 
 formula for which, as given on p. 271, is R = yr- . 
 
 1 + 4«-jT 
 
 a 1 
 The value of R will be found by multiplying the load on the piston by the 
 secant of the angle of obliquity of the connecting-rod (vide p. 276)- Or by 
 geometry 
 
 B.l. 2L , 
 
 J4: L 2 - S* 
 
 P being the load on the piston, S the stroke, and L the length of the 
 
 connecting-rod as before. 
 
 Simplifying the above formula by assuming a value r for the ratio of 
 
 <ir d 2 
 I to d, and substituting — - for S ; and taking the value of /at 3000 lbs. 
 
 Diameter of connecting-rod in the middle = — — ^= — . 
 
 48*5 
 
 Example. — To find the diameter at the middle of the connecting-rod of 
 
 an engine, 60 inches stroke, whose length is 120 inches, and the load on the 
 
 piston 100,000 lbs., r being taken at 15. 
 
 100,000 x 2 x 1 20 inQQ . Q „ 
 R = — , = 103,358 lbs. 
 
 ^/4 x 120 2 - 602 
 
 J 103,358 (l + ^~) 
 
 Diameter at middle = — — ttt-= = 7-6 inches. 
 
 48-5 
 
 The following are the values of r in practice : — 
 
 Naval engines — Direct-acting • . r = 9 to 11. 
 Mercantile engines, ,, ordinary r — 12. 
 
 ,,, ,, long stroke r = 13 to 16. 
 
 Taking 10 as the average value of r for naval engines, and 13 for mercantile; 
 
 then, 
 
 For a naval engine, 
 
 Diameter of connecting-rod at middle = . / ~ ft( ,», 
 
 For mercantile engines, 
 
 R 
 
 Diameter of connecting-rod at middle = A / ttttttv- 
 
 /: 
 
 V 1! 
 
 The sizes given by these rules, although large enough for strength, are 
 somewhat smaller than found in actual practice in the mercantile marine 
 generally. 
 
282 MANUAL OF MARINE ENGINEERING. 
 
 The following empirical formula will be found a very useful one, and the 
 results given by it agree very closely with good modern practice : — 
 
 Diameter of connecting-rod at middle = — • 
 
 L is the length of the rod in inches, and 
 
 K = 0-03 x ^Effective load on the piston in lbs. 
 
 Example. — To find the diameter of the connecting-rod, 100 inches long, 
 for an engine having a load of 55,000 lbs. 
 
 K = 0-03 x ^55,000 = 70, 
 
 Diameter =-^- — = 6*6 inches. 
 
 * 
 
 The diameter of the connecting-rod at the ends may be 0*875 of its 
 diameter in the middle. The tapering of rods, or making them barrel- 
 shaped, is usual in the case of those having single brasses at both ends ; 
 then the diameter of the crank-pin end is 0"925 of the diameter at middle. 
 Direct-acting engines have usually the connecting-rods tapering from the 
 gudgeon end to the middle, and then parallel, or nearly so, to the crank-pin 
 end. 
 
 It is, however, simpler and sufficiently accurate to follow, in the case of 
 connecting-rods, a similar rule to that laid down on p. 272 for piston-rods, 
 viz. : — 
 
 -r.. , , ,. , Diameter of cylinder / — 
 Diameter ot connecting-rod = = — sjp. 
 
 Here p is as calculated by the formulae on p. 273. F may be taken at 55 for 
 the crosshead end of fast-running light engines, and 52 of mercantile engines 
 of ordinary type. The diameter of the rods at the middle may be got by 
 dividing by F — L, where L is the length of connecting-rod in feet. 
 
 Connecting-rod Bolts. — The diameter of the bolts may be calculated by 
 allowing the same stress per square inch as that given for piston-rod bolts. 
 It is usual now, from practical considerations, to make the bolts of both 
 piston and connecting-rod of the same size ; the bolts therefore should be 
 calculated from the load on the connecting-rod. In order that the whole of 
 the stretch shall not come on one section, as at the bottom of the last thread 
 of an ordinary bolt, it is better to turn part of the body of connecting- and 
 piston-rod bolts to the same diameter as at the bottom of the thread, leaving 
 it a little larger than the diameter over the thread close to the head, and in 
 way of any joint — that is, the bolt is made with a phis thread, and bearing 
 collars where required. 
 
 Connecting-rod Brasses. — The crank-pin brasses are more severely tried 
 than any others about an engine, and, therefore, should be most carefully 
 designed, and made of the very best material. Some engineers make the 
 brasses to form the end of the rod (fig. 99), and retained to it by bolts and a 
 steel cap ; others prefer that they shall only act as bushes or liners to the 
 connecting-rod, sometimes fitting them into a square or octagonal recess in 
 the rod end,' and held in place by a flat cap and bolts, just as is generally 
 done to piston-rod ends; .but more generally they are fitted in duplicate 
 
CONNECTING-ROD BRASSES. 283 
 
 halves, as shown in fig. 98. The former plan is an expensive one when they 
 are ol large size and made of brass, on account of the great weight required, 
 and consequently are also costly to renew when worn, besides which they 
 are very liable to get out of shape when heated, and to crack through the 
 crowns. The latter plan avoids the use of so much brass, gives a good solid 
 bed to the brasses, and leaves the bolts free of all stress except tension. When 
 rods are made in this way it is usual to forge the head of the rod solid, and 
 turn it and the cap at the same operation ; the hole for the brasses is bored 
 or slotted out (the latter when the hole is 9 inches and upwards in diameter) 
 roughly ; the head is then slotted through or parted in the lathe so as to cut 
 off the cap, the space left by the tool being equal to twice the difference in 
 thickness of the brass at the crown and sides ; the cap is then bolted close 
 to the rod, and the hole bored out to the diameter of the brasses measured 
 across the rod. The brasses are kept from turning by a brass distance piece 
 secured between the cap and rod and projecting between the brasses, and in 
 the case of large brasses a short feather is fitted close to each flange in the 
 crown. All brasses have a tendency to close on the pin or journal after 
 having been hot, because the inner surface becomes warm first, and the 
 metal in expanding tends to straighten the curved part ; this is resisted by 
 the other part of the brass and the bed in which it is fitted, and in conse- 
 quence this inner ourface gets compressed permanently, so that on cooling 
 down it contracts, and tries then to give the brass more curvature, and so 
 presses hard on the journal. It is for this reason that some bearings will 
 never work cool but always a trifle warm ; this slight amount of heat causes 
 the brass to expand so as truly to fit the journal. It is now not at all un- 
 common to make these fittings of good cast or malleable cast iron when lined 
 with white metal ; indeed, cast iron carries white metal better than brass 
 does, and when of really good mixture is quite as strong as gun-metal, and 
 stronger than the common brass so often used with white metal. 
 
 Fig. 100 is an example of a modern connecting-rod with the jaws made 
 as part of a cone, instead of as in fig. 99, thereby providing a better connec- 
 tion between each jaw and the body of the rod, as well as permitting of its- 
 being machined much more cheaply. This rod is fitted with cast-iron bearings- 
 at each end, lined with white metal. This form of rod and " brasses " is now 
 very generally used in the mercantile marine and Navy for engines of all- 
 sizes. 
 
 White metal is better than bronze for the rubbing surface of the crank- 
 pin " brasses," it is important, therefore, that the white metal shall project 
 beyond the " brass," so that it alone shall bear on the pin. For this purpose 
 strips of white metal should be fitted into grooves planed in the " brass " 
 and be well hammered, so as to thoroughly fill the spaces, after which it 
 should be smoothly bored and fitted to the pin. Brasses which have not 
 been originally designed for white metal may be fitted in this way, or by 
 boring some shallow holes, whose diameter at the bottom is more than at the 
 surface, casting into them buttons of white metal, which, after hammering 
 down, are bored out so that the white metal stands out beyond the original 
 wearing surface. 
 
 A very good plan, but somewhat more expensive and not more efficient 
 than the one above described, is to run the white metal into recesses cast 
 with the brass, hammer it well in place, bore out, and then plane out the brass- 
 
284 
 
 -MANUAL OF MARINE ENGINEERING. 
 
 intervening between the white metal patches, leaving only slight ridges 
 surrounding the latter, which prevent it from being spread out. The most 
 
 Fig. 100.— Connecting-rod. Modern Form. 
 
 general plan, however, is to have a recess in the " brass" or shell, which is 
 now often of cast iron instead of bronze, tin the surface and run the metal 
 
GUDGEON END OF ROD. 
 
 285 
 
 in place while the shell is warm. When cold bore out as usual, trim the 
 edges, etc., and cut grooves for the oil to circulate in the surface of the metal 
 (v. fig. 100). 
 
 Caps of Connecting-rod Brasses. — The width of the connecting-rod end 
 should be such as to efficiently support the brasses ; its thickness (in direc- 
 tion of the length) = 0"6 x diameter at middle of rod. 
 
 The thickness of cap at middle = 0*8 X diameter of body of bolts -j- 
 0*1 X pitch of bolts. 
 
 Thickness of cap at ends = diameter of bolts at bottom of thread. 
 
 For ease of manufacture caps are generally made straight and of thickness 
 given by the first rule. 
 
 Gudgeon End of Rod. — The jaws of every connecting-rod are subject to 
 forces which tend to open them on the down stroke and close them on the 
 up. When fitted with a gudgeon-pin secured in the eyes, as in fig. 98, the 
 jaws are subject to bending moments at the various sections, which become 
 maxima at a point just below the gudgeon and also at an angle of about 40° 
 to the vertical through the centre of curvature of the inner portion. When 
 the rod is fitted with "brasses" working on a crosshead, as in fig. 99, the 
 maximum bending moment is at an angle of about 40° to the vertical or axis 
 of the rod, and is greater than when with a gudgeon secured to the rod ends. 
 When the diameter of the gudgeon or crosshead ends is 
 
 then 
 
 Diameter of cylinder #— 
 ~I5~ " X V? ' 
 
 Let P be the maximum re-current load on the piston. 
 
 L the width from centre to centre of gudgeon-rings or of crosshead 
 brasses when so fitted. 
 
 F 
 
 = IT 
 
 V 3,000' 
 
 then 
 
 Diameter of gudgeon ring . 
 Thickness ,, ,, 
 
 Width of jaw . 
 Thickness of jaw in line at 40 c 
 
 . = diameter of gudgeon + D25 X F. 
 . = 0-85 x F. 
 
 . = 1*35 X F, or 1-1 X diameter of rod. 
 . = '48 F X L when with gudgeon. 
 . = *52 F X L when with brasses. 
 
 Fig. 100a. — Connecting-rod End (Michel). 
 
286 MANUAL OF MARINE ENGINEERING. 
 
 CHAPTER XII. 
 
 SHAFTING — CRANKS AND CRANK-SHAFTS, ETC. 
 
 The Shafting of a Modern Marine Engine is made wholly of mild steel, 
 produced from a Siemens furnace, and generally having a tensile ultimate 
 strength of about 30 tons. In section it is, of course, circular, and in the 
 mercantile marine generally solid ; in Naval ships it is always made hollow 
 by boring out the centre so as to remove the core of doubtful metal and 
 provide the strongest shaft with the least weight, inasmuch as, with a large 
 reduction in material, there is only a small reduction in strength by this 
 treatment. For example, if a shaft have a borehole in diameter half that 
 of the outside, the reduction in weight is 25 per cent., while the strength is 
 diminished by 12J per cent. only. The cost of hollow shafts is, of course, 
 greatly in excess of that of solid ones ; notwithstanding, it is found worth 
 while to fit them to many of the high-speed express steamers crossing channels 
 and entering shallow harbours. 
 
 In Practice every Shaft in a Ship is subject to torsion, and, therefore, to 
 the shearing stresses caused by torque. It has also to resist the additional 
 shearing stress due to its own weight, but this addition is, as a rule, not great 
 by comparison. But the weight causes a bending moment to act on every 
 part of the shaft, which, if the bearings are wide apart, will cause a stress 
 and deflection which must be taken account of. 
 
 Throughout the Line of Shafting from the motor to the propeller is trans- 
 mitted to the latter the torque generated in the former, and when the screw 
 is heavy and racing badly in a sea-way, there mav be, and often is, a reverse 
 action by the inertia stored in the latter transferring back a torque to the 
 motor. 
 
 The Propeller-shaft of every Ship is subject to a more severe load from 
 the overhung heavy instrument setting up severe bending moments, due to 
 gravity in smooth water, and in rough water to the inertia effects and the 
 uneven action of the blades on the water. When a ship is pitching in a heavy 
 sea the stern drops with great rapidity, so that the vertical velocity of a heavy 
 screw is serious ; this velocity is checked with sufficient abruptness to put 
 a very heavy bending moment and shearing force on the shaft end, where 
 it ceases to be supported by the stern bush, as is equally the case when the 
 wave motion causes a rapid rise of the stern. The side throw or lurch at the 
 stern from wave action sets up horizontal inertia forces, also severe and 
 trying, although not so intense as the vertical ones. The differential pres- 
 sures on opposite blades of the screw also produce bending moments with 
 the corresponding shearing forces. Hence the stern shaft of a cargo steamer, 
 which is sometimes deeply laden and sometimes light, and always somewhat 
 lively at the stern, is tried very severely, and should always be of ample size ; 
 
TUNNEL SHAFTING. 
 
 287 
 
 in fact, larger than that of the finer-lined express steamer of the same power, 
 but with a screw lighter, because of smaller diameter and often of bronze. 
 Crank-shafts of reciprocators are subject to complex loads at their various 
 parts, and always have bending moments of kinds as well as torque to resist. 
 The shafts of turbines also have to bear considerable bending moments, 
 due to the weight of the rotor, but in their case both it and the torque are. 
 constant and uniform ; there is no variation in torque through the revolution 
 as with the reciprocator. It does, however, have to resist inertia forces of 
 a kind when the ship is in a sea-way, and when the turbine is well aft, as in 
 fig. 40, the stresses due to them will be heavy in bad weather. 
 
 The Crank-shaft of the ordinary Oil Engine is subject to severe bending 
 moments, and to shocks which produce heavy shearing forces. It is claimed 
 for the Diesel engine that, owing to the very great amount of compression, 
 there is little or no shock at explosion, and the diagram (fig. 101) would 
 seem to prove this. The combined torque diagram from an oil engine on the 
 
 IN0ICATOR DIAGRAM OF DIESEL 
 2 STROKE CYCLE ENGINE-, 
 
 Fig. 101. —Diesel Engine Diagram. 
 
 four-cycle system, however,' shows a very large ratio of maximum to mean 
 torque as well as the big bending moment. 
 
 The Tunnel or Intermediate Shafting of a ship is subject theoretically 
 to torque only ; it is, of course, in compression axially from the screw shaft to 
 the thrust block, but this affects the torque so little as to be negligible (about 
 1 \ per cent.). If the tunnel bearings are close together, and the rate of revolu- 
 tion is not high, then torsion alone is the governing factor in determining its 
 size. But, since the rate of revolution has increased so largely, especially 
 where turbines are employed, it is necessary now to cousider the bending 
 moments that may or do come on these shafts. However close, in reason, 
 the bearings may be to one another, the weight of the shaft must cause a 
 certain amount of sag ; if the centre of gravity of the shaft does not coincide 
 with its axis of rotation, it may then whirl — that is, the C.G. will tend to 
 move in a circle, whose centre is on the axis of rotation. At low speeds 
 the whirling action is comparatively small, and not sufficient to further 
 bend the shaft, but at higher speeds the centrifugal force may be so great 
 as to be harmful. 
 
288 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Ill order that a shaft may revolve freely without any danger of whirling 
 due to the " sag " or deflection caused by its own weight, the bearings should 
 never be further apart than the distance given by the following rule, where 
 d is the diameter in feet and z is a multiplier depending on the rate of revo- 
 lution : — 
 
 Rule. — Pitch of plummer blocks should not exceed z Jd. 
 
 When the revolutions do not exceed 200 per minute, z = 40. 
 
 5 »» 
 
 » >5 
 
 5 >♦ 
 
 » J> 
 
 > >J 
 
 > )J 
 J >> 
 J J> 
 
 99 
 
 250 
 
 > ? 
 
 2 = 36. 
 
 »> 
 
 300 
 
 j> 
 
 2 = 315. 
 
 >} 
 
 400 
 
 j» 
 
 2 = 29-0. 
 
 >> 
 
 500 
 
 5J 
 
 2 = 25-5. 
 
 >» 
 
 600 
 
 JJ 
 
 2 = 23-5. 
 
 »> 
 
 700 
 
 »; 
 
 2 = 21-5. 
 
 >> 
 
 800 
 
 >f 
 
 z = 20-0. 
 
 >> 
 
 900 
 
 5> 
 
 2 = 19-0. 
 
 >> 
 
 1,000 
 
 JJ 
 
 2 = 18-0. 
 
 Considering that in a screw ship the lengths of tunnel shafts near the 
 stern will be subject to considerable inertia forces, the pitch of bearings should 
 be less than the limit given above for still water, and about 60 per cent, should 
 not be exceeded. For example, a shaft 12 inches diameter at 200 revolutions 
 should be supported every 24 feet, while at 400 revolutions it should be 
 17*4 feet, and never exceed in smooth water 29 feet. 
 
 Alternating Stresses. — If a shaft having a heavy body, such as a screw or 
 paddle-wheel, at its overhung end is revolving under the action of torque, 
 the material from the weight to the supporting bearing is subject to tension 
 and compression, and shearing due to that weight and to shearing at every 
 section throughout its length due to the torque. But the bending moment 
 due to the overhung weight sets up tension on all particles above the neutral 
 axis and compression on all below it, and inasmuch as the shaft is turning, 
 so that what was at the top at one moment is at the bottom the next, hence 
 the particles, from being subject to tearing apart when above, are quickly in 
 a state of being crushed together below. The magnitude of the forces ranges 
 from steadily to the maximum at top and bottom, so that the load and 
 change of load is gradual ; nevertheless, seeing that at 120 revolutions the 
 change from plus to minus is no less than four times in a second, while with 
 turbine-driven shafts there is often as many as 24 alternations per second, 
 the endurance of the material is severely tried. Such stresses as these which 
 it has to resist are called Alternating. 
 
 Piston-rods, Connecting-rods, Valve-rods, and Columns are all subject 
 to these reversals of load and stress, but in their case, the application and 
 release are sudden and made with more or less shock, but modified somewhat 
 by the action of inertia forces and the effect of cushioning. 
 
 The Arm of a Lever vibrating through an angle when transmitting energy 
 is subject to alternating stresses also, more or less abruptly applied and 
 released. 
 
SAFE WORKING STRESS. 289 
 
 But the Arm of a Crank-shaft, which is a lever moving always angularly 
 in one direction, is subject to no such changes of stress. On the one side 
 on which the force is applied the molecules are subject to tension only, and 
 thos-e on the opposite side to compression only ; but in the case of an engine 
 crank the Load and stress, as measured by the turning moment, are not so 
 abrupt in their application, etc., and they also are modified by inertia forces 
 and cushioning. With the reversal of the engine into " astern gear " these 
 stresses are, of course, reversed. 
 
 The Bolts of Connecting-rods, Main Bearings, Cylinder Feet, and 
 Cylinder Covers are all subject to a load of one kind, but it is applied 
 and released, then ceases for a moment, to be again applied more or 
 less suddenly. Such loads and the stresses caused by them are called 
 Intermittent. 
 
 Iron, Steel, and all Materials suffer more or less severely from these inter- 
 mittent and alternating stresses, but much more so from the latter, as the 
 effect on their structures is to gradually destroy them. The higher the stress 
 and the greater the number of alternations or intermissions per minute, the 
 quicker is the strength or virtue of the material annihilated. Of steels some 
 varieties have a longer life than others, and of the bronzes it is even more 
 marked. It does not follow of necessity that because a steel or bronze has 
 a very high elastic limit that it can be used with greater longevity in a rapidly 
 active structure than one whose elastic limit is less. Professor Arnold has 
 shown that bar steel, by rolling and drawing cold, might have its yield 
 point raised by as much as 8 to 12 tons, while the endurance was reduced 
 by 38 per cent. ; in the case of aluminium copper alloys the endurance 
 of one was more than double that of others not differing largely from it in 
 composition. 
 
 The Influence of Numbers of Reversals of Stress and their magnitudes 
 has been demonstrated also by Prof. Arnold and others, and the results of 
 experiments on steel and iron is startling. 
 
 Wohler's experiments showed that the iron used for axles then (1870) 
 required 56,430 repetitions with a plus and minus load of 15*3 tons per square 
 inch to produce fracture, while with 8 - 6 tons each it required over 19 millions. 
 Similar experiments with Krupp's steel showed that with plus and minus loads 
 of 20 - l tons per square inch 55,100 repetitions produced fracture at 15 - 3 tons; 
 it was done with a little over 3 millions, while some bars with 143 tons took 
 over 45 millions. It is clear, then, that since some definite number of alter- 
 nations will break down the tenacity of a metal, and that the greater the 
 stress the fewer they will be, that for long life a low working stress is 
 necessary. 
 
 Safe Working Stress on the material must, therefore, vary with the con- 
 ditions under which it performs its duty. In a general way the highest 
 stress allowable under any circumstances is one-half that of the elastic limit, 
 and for margins of safety to cover inaccuracies and small hidden defects 
 the highest working stress should not exceed 40 per cent, of the elastic limit. 
 If the material is subject to intermittent stress, 90 per cent, of this maximum 
 or 36 per cent, of the elastic limit should be observed in the design. If, 
 however, the stresses are alternating, 66 of the maximum or 26*4 per cent, 
 of the elastic limit should not be exceeded. 
 
 The elastic limit of mild steel, as found in any workshop, is about 
 
 19 
 
290 MANUAL OF MARINE ENGINEERING. 
 
 30,000 lbs. The working stresses on it with moderate speed of revolution 
 should be 12,000, 10,800, and 7,920 lbs. per square inch in tension. 
 
 For high rate of revolution even these allowances are too great, and 
 over 500 revolutions per minute the reduction in stress should be quite 
 10 per cent., so that for screw shafts of fast-running engines the stress 
 allowed in calculating the sizes should not exceed 7,000 lbs. for continuous 
 working. 
 
 For work where the alternations or intermissions are effected with con- 
 siderable shock the working stress should be at least 10 per cent, less than 
 when effected gently. 
 
 It should be noted, however, that certain qualities of steel and bronze 
 stand shock and alternation of stress much better than others. 
 
 Twisting Moment. — If a force is acting on a shaft so as to turn it, or tend 
 to turn it, round on its axis, it is called a twisting force, and the effort of this 
 force is measured by multiplying it by its distance from the axis, and called 
 the twisting moment or torque. Suppose P is the thrust along the connecting- 
 rod when at right angles to the crank, and L is the distance of the centre 
 of the crank-pin from the centre of the shaft, P X Lis the twisting moment 
 on the shaft. 
 
 When one force is acting on the shaft as above described, the second 
 force, which completes the couple, is the reaction of the bearing, which is 
 equal to P, but acts in the opposite direction. If the force P and the reaction 
 R act in a plane perpendicular to the axis of the shaft, they will cause no 
 bending action on the shaft, but there will be a force R tending to shear 
 the shaft across. But in actual practice it is almost impossible that P and 
 R shall act in such a plane, and they usually act in planes parallel to one 
 another, and perpendicular to the axis ; hence, the shaft is also subject 
 to a bending action. But if a shaft is turned by means of two equal forces 
 acting in opposite directions, one on either side of the shaft and equidistant 
 from the axis and in the same plane, then the shaft is balanced, these forces 
 will cause no pressure on the bearings, and it is subject, therefore, to twisting 
 strains only. If one shaft is coupled to another shaft, from which it is to 
 transmit power by two coupling bolts equidistant from the centre, it will 
 only receive a twisting strain. Such is the state of the shafting from the 
 crank-shaft to the propeller-shaft of a screw steamship. 
 
 Resistance to Twisting. — Let T be the twisting moment on a shaft in 
 inch pounds, d the diameter of the shaft in inches, and / the stress per 
 square inch on the transverse section of the shaft. Then (Rankine, Applied 
 Mechanics, p. 355) for solid shafts of diameter d, 
 
 T = '^-f = 0-1964/d 3 ; or d = M x 51. 
 
 For hollow shafts with a bore diameter d lt 
 T = 0-1964/(— ^j-^V 
 Example. — To find the diameter of a shaft subject to twisting only, the 
 
DIAMETEK OF A SHAFT SUBJECT TO TORSION. 291 
 
 force being 100,000 lbs. acting at a distance of 24 inches, stress to be 8,000 lbs. 
 T = 100,000 x 24 = 2,400,000 inch-lbs. 
 
 d- 3 A40 
 V 8 
 
 400,000 _ . .... , 
 
 x 5-1 = 11 *5 inches. 
 
 8000 
 
 Diameter of a Shaft subject to Torsion. — If a constant force P were applied 
 to the crank-pin tangentially to its path, then the work done per revolution 
 
 will be P x ■— l — ; L being the length of the crank in inches ; then if R be 
 
 the number of revolutions per minute, 
 
 2 it L 
 Work done per minute = P x '* x R. - - (1) 
 
 But this work is equal to I.H.P. x 33,000 ; and the twisting moment is 
 P x L constantly. Then 
 
 (PxL)x^xR = I.H.P. x 33,000, 
 
 and 
 
 „ T I.H.P. x 33,000 x 12 
 P x L = ^ • 
 
 2 ir x R 
 
 That is, 
 
 (2) 
 
 .3) 
 
 Diameter of shaft = a/-^^ X 42-84. - ... (4) 
 
 But as shafts must be strong enough to resist the maximum twisting stress, 
 it is necessary always to base calculations on it instead of on the mean twisting 
 moment. The factor 42 "84 must, therefore, be multiplied by the ratio of 
 maximum to mean moment, as given in Table xxxv. 
 
 Professor Rankine directs (Rules and Tables, p. 250), in order to find the 
 greatest twisting moment from the mean : if a shaft is driven by a single 
 engine, multiply by 1*6 ; if by a pair of engines with cranks at right angles, 
 by ri ; if by three engines with cranks at angles of one-third of a revolution, 
 by 1-05. 
 
 These values are, however, very much lower than usually obtained in 
 modern practice if the effect of inertia forces are neglected. 
 
 For a three-crank engine, cranks at 120°, cutting off at half to two-thirds 
 stroke, multiply by 1*15. 
 
 For a two- and four-crank engine having cranks at right angles, cutting 
 off steam at half to three-quarter stroke, multiply by 1*3. 
 
 For a single-cylinder engine cutting off steam at half stroke, by 2"0. 
 
290 MANUAL OF MARINE ENGINEERING. 
 
 30,000 lbs. The working stresses on it with moderate speed of revolution 
 should be 12,000, 10,800, and 7,920 lbs. per square inch in tension. 
 
 For high rate of revolution even these allowances are too great, and 
 over 500 revolutions per minute the reduction in stress should be quite 
 10 per cent., so that for screw shafts of fast-running engines the stress 
 allowed in calculating the sizes should not exceed 7,000 lbs. for continuous 
 working. 
 
 For work where the alternations or intermissions are effected with con- 
 siderable shock the working stress should be at least 10 per cent, less than 
 when effected gently. 
 
 It should be noted, however, that certain qualities of steel and bronze 
 stand shock and alternation of stress much better than others. 
 
 Twisting Moment. — If a force is acting on a shaft so as to turn it, or tend 
 to turn it, round on its axis, it is called a twisting force, and the effort of this 
 force is measured by multiplying it by its distance from the axis, and called 
 the twisting moment or torque. Suppose P is the thrust along the connecting- 
 rod when at right angles to the crank, and L is the distance of the centre 
 of the crank-pin from the centre of the shaft, P X Lis the twisting moment 
 on the shaft. 
 
 When one force is acting on the shaft as above described, the second 
 force, which completes the couple, is the reaction of the bearing, which is 
 equal to P, but acts in the opposite direction. If the force P and the reaction 
 
 R. ant in a. r»la.np v\prneru]if>iilsir t.n fhp mris ni thp shaft t.ViPV will pause no 
 
 The reference on page '291 to Table WW. should be 
 Table XXXIV., page 298. 
 
 transmit power Dy two coupling uvivo cijumiciuau UU u. ^„~ , 
 
 only receive a twisting strain. Such is the state of the shafting from the 
 crank-shaft to the propeller-shaft of a screw steamship. 
 
 Resistance to Twisting. — Let T be the twisting moment on a shaft in 
 inch pounds, d the diameter of the shaft in inches, and / the stress per 
 square inch on the transverse section of the shaft. Then (Rankine, Applied 
 Mechanics, p. 355) for solid shafts of diameter d, 
 
 - d 3 It 
 
 T = —/ = 0-1964/d3; ovd = ^ x 5-x. 
 
 For hollow shafts with a bore diameter d u 
 
 T = 0-1964/(-- d -^ 4 ). 
 Example. — To find the diameter of a shaft subject to twisting only, the 
 
DIAMETER OF A SHAFT SUBJECT TO TORSION. 2\)l 
 
 iorce being 100,000 lbs. acting at a distance of 24 inches, stress to be 8,000 lbs. 
 T = 100,000 x 24 = 2,400,000 inch-lbs. 
 
 'J 
 
 2,400,000 _ . .... , 
 
 x 5-1 = 11 '5 inches. 
 
 8000 
 
 Diameter of a Shaft sutject to Torsion. — If a constant force P were applied 
 
 to the crank-pin tangentially to its path, then the work done per revolution 
 
 2 t L 
 will be P x — — — ; L being the length of the crank in inches ; then if R be 
 
 the number of revolutions per minute, 
 
 2tL 
 
 Work done per minute = P x x R. - - (1) 
 
 But this work is equal to I.H.P. x 33,000; and the twisting moment is 
 P x L constantly. Then 
 
 (PxL)x^xR = I.H.P. x 33,000, 
 
 and 
 
 n T I.H.P. x 33,000 x 12 
 
 P x L ^ 
 
 J T X R 
 
 That is, 
 
 T H P 
 
 Mean twisting moment = -^-^ — ■ x 63,000. - - - - (2) 
 
 And as before 
 
 d _ 3 /I.H.P. x 63,000 x 5-1 
 
 R xf 
 
 I.H.P. 321,300 
 
 "V""R~ x ' / 
 
 If/ be taken at 7,500 for mild steel 
 
 (3) 
 
 Diameter of shaft = */ •■■' ~ x 42'84 
 
 V R-X 42-84. - - . . ( 4 ) 
 
 But as shafts must be strong enough to resist the maximum twisting stress, 
 it is necessary always to base calculations on it instead of on the mean twisting 
 moment. The factor 42 "84 must, therefore, be multiplied by the ratio of 
 maximum to mean moment, as given in Table xxxv. 
 
 Professor Rankine directs (Rules and Tables, p. 250), in order to find the 
 greatest twisting moment from the mean : if a shaft is driven by a single 
 engine, multiply by 1 -6 ; if by a pair of engines with cranks at right angles, 
 by 1*1 ; if by three engines with cranks at angles of one-third of a revolution, 
 by 1-05. 
 
 These values are, however, very much lower than usually obtained in 
 modern practice if the effect of inertia forces are neglected. 
 
 For a three-crank engine, cranks at 120°, cutting off at half to two-thirds 
 stroke, multiply by 1*15. 
 
 For a two- and four-crank engine having cranks at right angles, cutting 
 off steam at half to three-quarter stroke, multiply by 1/3. 
 
 For a single-cylinder engine cutting off steam at half stroke, by 2*0. 
 
292 MANUAL OF MARINE ENGINEERING. 
 
 The following rule holds good for the ordinary engines, as found in general 
 use iu the merchant service : — 
 
 /r u p 
 
 Diameter of the tunnel shafts = ?/ - * ' : X F. • • (5) 
 
 Two-stage compound engines, cranks at right angles — 
 
 F = 8 Jp, 
 
 where p is the absolute actual pressure, or boiler pressure + 15 lbs. 
 Triple-compound, three-cranks at 120 c — 
 
 F = 4Wp. 
 
 Quadruple-compound engines, two cranks, right angles, F = 5*5 Jv. 
 „ „ four cranks, F = 5-C Jp. 
 
 Expansive engines, cranks at 90°, and the rate of expansion 5, F = 9 - 5 J p. 
 Turbines, F = 54. 
 Single-crank compound engines, pressure 80 lbs., F = 10 s '« 
 
 The shafts of torpedo-boats, destroyers, and fast craft which are run at 
 full speed only occasionally, and for short periods, may be designed by taking 
 F at about a half of the above values. 
 
 The Torsional Stiffness of a Shaft is of more importance to the marine 
 engineer to-day than it has ever been before, inasmuch as it is employed as 
 the measure of the torque of the turbine, and from it the horse-power trans- 
 mitted is measured (v. p. 150). 
 
 The torsion meter employed on shipboard to obtain the measure of the 
 power of a turbine is really in essence only an instrument for indicating the 
 angle of twist of a definite portion of one of the shafts that previous to fitting 
 in the ship was tested by torque to ascertain the angular displacements 
 corresponding to the magnitude of each application. The following formula 
 is a means for computing what that angular displacement should be in degrees, 
 with a shaft whose external diameter is d and the diameter of bore d x ; the 
 twisting moment or torque is T l in inch-pounds, and C is the modulus of 
 stiffness, which for solid shafts of best mild steel is 11,750,000, and for hollow 
 shafts 12,150,000. is the angle in degrees; L is the length of the shaft 
 under observation in inches, R the revolutions per minute. 
 
 («) 
 
 584 X T x XL 
 
 C X./ 1 
 T,xL 
 
 for solid shafts. 
 
 20, 120 d* 
 
 584 x T, X L 
 
 ' CX («**-(*!*) 
 
 (I>) 
 
 - for hollow shafts. 
 T lX L 
 
 - &A J 
 
 21.147 {d*-d,*) 
 
BENDING MOMENT. 293 
 
 nT , , , ,, S.H.P. x 33,000 , , „ 
 (c) I he torque on a shaft = — = — - foot-lbs. 
 
 = ^5^ X 63,000 inch-lbs. 
 
 17 *-- / x rp ^X 20,120 d* S.H.P. . Q/vvn 
 Prom equation (o) Tj = ' = —= — x 63,000. 
 
 (b) T l = dx21Ml( d —=^) = ^^ X 63,000. 
 Then S.H.P. = f x *L f or so lid shafts, 
 
 K rf 4 — (7 * 
 
 and S.H.P. = f- X nQ i for hollow shafte. 
 
 Li Ji'VO 
 
 Example. — Find the value of 6 for a shaft 9 inches diameter and 6 inches 
 "bore when transmitting 5,000 S.H.P. at 350 revolutions per minute. The 
 length under observation is 40 inches. 
 
 5 000 
 Here the torque = -^ X 63,000, or 900,000 inch-lbs. 
 
 u< *■ ,m /, 900.000 x 40 ft _ oq , , 
 
 Equation (6), fl = 21,147 (6,581 - 1,296) = ° -323 of a de ^- 
 
 Bending Moment — If a force is acting on a shaft tending to bend 
 it only, its effort is called the bending moment, and is measured by- 
 multiplying the force by the distance at which it acts from the support 
 of the shaft. 
 
 If the shaft is overhung like a cantilever, and a force P is applied at a 
 •distance L from the point of support, 
 
 The bending moment = PxL- - -(1) 
 
 If supported on two bearings, whose distance apart is L, and a force P is 
 ■applied at a point midway between these two bearings, 
 
 P X L 
 
 The bending moment = — j — • " - - (2) 
 
 If the bearings are long — that is, exceeding the diameter of the shaft in 
 length, and are also strong and rigid, so that the shaft is held by them suffi- 
 ciently to prevent flexure taking place in the bearing, 
 
 P X L 
 
 The bending moment = — « — -. - - - (3) 
 
 o 
 
294 MANUAL OF MARINE ENGINEERING. 
 
 Since, however, few shafts are so secured as to comply with these conditions 
 exactly, any shaft supported in strong hearings not less than 1 diameter long, 
 and whose distance apart does not exceed 10 diameters, and which has to 
 work freely in its bearings, may be treated as partly complying with these 
 conditions, and 
 
 The bending moment = — - — . - - (4) 
 
 When the bearings are not rigid enough to prevent flexure of the shaft 
 within them, L must be measured from the centres of the cap bolts, so that 
 where each cap is held down by a pair of bolts L is measured to the centres 
 of bearings. If the caps and bearings are strong and rigid enough to resist 
 any tendency to bend by the action of the shaft, L may be measured from 
 the edge of the bearing or cap. If the bearing is fitted with brasses, which 
 project beyond the cap and bed so much as to receive little or no support 
 from them, L must still be measured from the edge of cap. In a few words, 
 the distance must be measured from what would be the actual points of sup- 
 port if it is bent by severe pressure. 
 
 Resistance to Bending. — The strength of a circular section shaft to resist 
 bending is only half of that to resist twisting. If M is the bending moment 
 in inch pounds, and d the diameter of the shaft in inches, 
 
 M = ^x/;andrf = *fi&~ W -2 
 
 f is a factor which depends on the material of which it is composed, and 
 the value may be based on the fact that it is in tension and compression. 
 The only shafts in a marine engine which are subject to bending only are 
 some weigh-shafts having double-ended levers, similar to the side levers of 
 paddle-wheel engines, and their diameter is determined from other considera- 
 tions than that of mere strength ; but with them, as with the crossheads of 
 return connecting-rod engines, care should always be taken that the size suit- 
 able for good working in the bearings should be sufficient for strength. 
 
 Equivalent Twisting Moment. — When a shaft is subject to both twisting 
 and bending simultaneously, the combined stress on any section of it may be 
 measured by calculating what is called the equivalent twisting moment — that 
 is, the two stresses are so combined as to be treated as a twisting stress only 
 of the same magnitude, and the size of shaft calculated accordingly. Pro- 
 fessor Rankine gave the following solution of the combined action of the two 
 stresses (vide Rankine, Rules and Tables, p. 227) : — Let T be the twisting 
 moment on a shaft when M is the bending moment on a section, then taking 
 T T as the equivalent twisting moment, 
 
 Tj = M + JM? + T 2 . 
 
 Example. — To find the diameter of a section of a shaft at which the bend- 
 ing moment is 40,000 inch-pounds, when the twisting moment is 250,000 inch- 
 pounds. The shaft of steel / = 7500 lbs. 
 
 Here 
 
 T x = 40,000 + X A0,000 2 + 250,000 2 
 
 = 40,000 + 10,000 Vi* 2 + 25* 
 = 293,170 inch-pounds. 
 
 i . s/T 5 .! _ 3/ 293,170 x 5- 1 = 
 V/ V T.50U 
 
CURVE OF TWISTING MOMENTS. 
 
 295 
 
 Crank Shafts. — These shafts are subject always to twisting, bending, and 
 shearing stresses ; the latter are so small compared with the former that 
 they are usually neglected directly, but allowed for indirectly by means of 
 the factor f, as already stated. 
 
 The two principal stresses vary throughout the revolution, and the maxi- 
 mum equivalent twisting moment can only be obtained accurately by a series 
 of calculations of bending and twisting moments taken at fixed intervals, 
 and from them construct a curve of strains. 
 
 Curve of Twisting Moments. — The twisting moment at any position of the 
 crank is equal to the pressure on the piston multiplied by the distance inter- 
 cepted by a line through the connecting-rod on a line at right angles to centre 
 line through centre of cylinder. 
 
 Fig. 102. 
 
 Let AB (fig. 102) be the centre line of the engine through the cylinder and 
 shaft centres, A C the position of the crank, B C the connecting-rod, and A D 
 a line at right angles to A B. Produce B C to cut the line A D, and drop 
 from A a line A E perpendicular to B C. P is the load on the piston, and R 
 is the thrust on the connecting-rod. It will easily be proved that the angle 
 D A E is equal to the angle A B D, called for convenience a. Then 
 
 P = R cos a ; and AE = AD cos a. 
 The twisting moment = RxAE = RxAD cos a = P x A D. 
 
 Let the twisting moment be calculated at equal intervals of say 10° of 
 angular movement of the crank, so that in the whole revolution there will be 
 36 observations, or 18 in the half revolution. Draw a line AB (fig. 103), and 
 divide it into 18 equal parts, Aa 15 a x a 2 , &c; erect at these points perpen- 
 diculars, and cut off parts a 1 b v a 2 b 2 , &c, to represent the value of the twisting 
 moments at each corresponding position of the crank to a suitable scale. 
 Through the points b x b 9 , &c, draw a curve, which represents the curve of 
 strain on the shaft during the forward movement of the piston ; by producing 
 A B, and going through a similar operation for the second half of the revolu- 
 tion, the curve of strain during the backward movement of the piston can be 
 obtained. 
 
296 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Divide the area enclosed between this curve and the line A B by the 
 length of A B, and the quotient is the mean twisting moment, and repre- 
 sented by A M in fig. 103, so that the rectangle A M N B is equal in area to 
 the figure A B C. 
 
 The value of A M may be calculated by taking a mean of the values oi 
 Ojfej, a 2 6o. etc. 
 
 When there are two engines — that is, two pistons operating on one shaft 
 — the combined twisting moment is found by drawing the curve of twisting 
 
 Fig. 103. — Curve of Twisting Moments. 
 
 moments of each crank separately, transposing that of one on that of the 
 other in a position corresponding to the relative position of the cranks. In 
 fig. 103a A C B is the curve of strain on one crank, and A 1 C 1 B l that on the 
 other, which is at an angle with it of degrees represented by AA X . The 
 combined twisting moment at any period a is represented by a d, which is 
 equal to a b 4- a c, and the dotted curve CdC lt etc., represents the curve of 
 combined twisting moments. 
 
 The maximum twisting moment will be at the point where the curve is 
 
 A A. ;a B B 
 
 Fig. 103a. — Curve of Combined Twisting Moments. 
 
 highest, and the ordinate may be measured and its value found by referring 
 to the scale to which the curve is drawn. The mean twisting moment may be 
 found by measuring the area included between the dotted curve and the base 
 line and terminal ordinates, and dividing by the base line, or by taking a 
 mean value from the ordinates as before. 
 
 If there are three engines, a similar operation will indicate the maximum 
 twisting moment. 
 
CURVE OF TWISTING MOMENTS. 
 
 297 
 
 There is another, and perhaps a better, method of showing the curve of 
 torque whereby the magnitude of the twisting moment at any angle through- 
 out the revolution is found by taking the length of the radial line intercepted 
 between the crank-pin circle and the curves, which are constructed by using 
 that circle as a reference instead of a base line. Fig. 104 is a good example 
 of this method, as prepared by Professor Jamieson, for a triple-compound 
 three-crank engine working under ordinary conditions as given below, and 
 will be found instructive. 
 
 Fig. 104. — Crank Effort Diagram of a Triple- 'Expansion Engine. 
 
 Ratio of expansion, 10-4. 
 Length of connecting-rod = 9 feet. Stroke = 4-5 feet. 
 Mean efficiency of steam, or ratio of area of work in cylinder to full theoretical diagram 
 
 55 per cent. 
 
 Cylinder's diameter, . 
 Area, ..... 
 Ratio, ..... 
 Mean pressures, lbs. per sq. in., . 
 Range of temperatures, Fah., 
 
 HP. I.P. L.P. 
 
 . 28" 46" 77" 
 . 615-8 1661-9 4656-6 
 1 2-80 7-56 
 67-6 28-2 9-7 
 64-3° 74-9° 80-6° 
 
 Steam, 164 lbs ; Vacuum, 26£ ins. ; Receivers, 52 and 5 lbs. ; 
 Revolutions, 62i per minute. 
 
 Cut-off H.P. =- 33A ins. 
 
 „ LP- = "„ 
 ,, L.P. — „ 
 
 I.H.P. = 710 L.P. 
 „ = 799 I.P. 
 „ = 764 H.P. 
 
 
 2,273 total I.H.P. 
 
 
 ssss 
 
298 
 
 MANUAL OF MARINE ENGINEERING. 
 
 The pressures at the different points may be taken from actual indicator 
 diagrams, or by constructing steam diagrams from the conditions under 
 which the engine is to work. 
 
 The bending moment on a section of the shaft will vary exactly with 
 the pressure on the crank-pin, and to find the maximum equivalent twisting 
 moment on a section, it is only necessary to construct a secondary curve from 
 
 the formula Tj = M + -/M 2 -f- T 2 between the point of maximum twisting 
 and that at which the pressure on the piston is greatest. 
 
 When steam is not cut off in the cylinder before 0"4 of the stroke, the 
 maximum load on the piston may be used to calculate the bending moment, 
 which is to be combined with the maximum twisting moment to find the 
 maximum equivalent twisting moment. Only when steam is cut off earlier 
 than this does the point of maximum equivalent twisting moment differ 
 much from the point of maximum twisting. 
 
 Momentum of Moving Parts. — In making these calculations it has been 
 assumed that the moving parts, such as the piston and rods, have no effect 
 on the force exerted on the shaft ; but this is never strictly true, for since 
 these parts are of considerable weight, a part of the energy of the steam is- 
 absorbed at the commencement of the stroke in overcoming their inertia, and 
 consequently the load on the crank-pin is less then than is represented on 
 the curves. Again, towards the end of the stroke the momentum, or energy 
 thus stored in these moving parts, is given out on the crank-pin, and causes 
 larger loads on it than shown by the curve. The further consideration of 
 the effect of the inertia and momentum of the moving parts will be found 
 in Chap. xxix. It is sufficient to say here that the general tendency is to 
 modify the stresses on the crank shaft, as also those of the connecting-rods, 
 piston-rods, etc., so that if any of these are strong enough to withstand the 
 stresses due to external forces, they are sufficiently strong for the engines 
 when moving. Indeed, it is only at very high speeds that momentum need 
 be taken seriously into account. 
 
 The following Table (xxxiv.) gives the relation between the maximum 
 and mean twisting moments of engines working under various conditions, 
 the momentum of the moving parts being neglected : — 
 
 TABLE 
 
 XXXIV. 
 
 
 Description of Engine. 
 
 Steam cut-off at 
 
 Max. Twist. 
 Mean Twist. 
 
 
 0-2 
 
 2-625 
 
 ii 
 
 
 
 0-4 
 
 2 125 
 
 ii 
 
 
 
 06 
 
 1-835 
 
 
 
 
 08 
 
 1-698 ! 
 
 Two-cylinder, cranks at 90*, - 
 
 
 . 
 
 o-i 
 
 1-872 
 
 n »» • ■ 
 
 
 . 
 
 02 
 
 1-616 
 
 i» i» 
 
 
 . 
 
 3 
 
 1-415 
 
 »» >• 
 
 
 • 
 
 4 
 
 1-298 
 
 n i» • ■ 
 
 
 . 
 
 5 
 
 1-256 
 
 ii »» 
 
 
 - 
 
 6 
 
 1-270 
 
 ii ii 
 
 
 - 
 
 07 
 
 1-329 
 
 ii ii 
 
 
 - 
 
 0-8 
 
 1-357 
 
 Three-cylinder compound, three ci 
 
 •anks at 120°, 
 
 H.P. 5, L.P. 6G 
 
 1-40 
 
 „ triple, „ 
 
 06 
 
 115 
 
 Four-cylinder quadruple, four cranks at 90°, 
 
 6 
 
 1-26 
 
OVERHUNG CRANK. 
 
 299 
 
 Overhung Crank. — The simplest form of crank is that known as the over- 
 hung crank, such as is usually fitted in mill engines, but hardly found now 
 in marine engines. The shaft projects beyond the bearing, and has keyed to 
 its end a lever or disc, in which is secured the crank-pin. 
 
 The pin is subject to bending and shearing forces, due to the thrust on 
 the connecting-rod. The maximum bending moment on the part of the pin 
 close to the crank is found by multiplying the greatest thrust of connecting- 
 rod by the distance to the centre of the connecting-rod. 
 
 If R is the thrust of the connecting-rod, and I the length of the pin, then 
 
 , . R x l 
 * Bending moment on crank-pm = 
 
 2 
 
 and diameter of pin 
 
 R x I 10-2 
 ■ x 
 
 y^rr; 
 
 5-1. 
 
 Example. — To find the diameter of the crank-pin whose length is 14 inches 
 1 
 
 Fig. 105. — Cranks of Paddle-wheel Engine. 
 
 and the thrust of connecting-rod is 125,000 lbs., /being of steel and taken at 
 9,000 lbs. 
 
 Diameter 
 
 -v/- 
 
 ,000 X u 
 
 9,000 
 
 X 51 = 9*97 inches. 
 
 The crank-arm (fig. 105) is to be treated as a lever, so that if a is the 
 thickness in direction parallel to the shaft axis, and b its breadth at a section 
 x inches from the crank-pin centre, then 
 
 Bending moment M at that section = R x x, 
 
 a x 62 M 
 6 -/' 
 
 6M 
 
 and 
 
 or 
 
 a = 
 
 62 x/* 
 
 If a crank-arm were constructed so that b varied as v x (as given by the 
 above rule), it would be of such a form as to be inconvenient of manufacture, 
 and consequently it is customary in practice to find the maximum value of 
 
 * For other conditions as to size of pins, vide p. 309. 
 
300 MANUAL OF MARINE ENGINEERING. 
 
 b, and draw tangent lines to the curves at the points ; these lines are gener- 
 ally, for the same reason, tangential to the boss of the crank-arm at the shaft. 
 The bending moment decreases as the distance from the crank-pin 
 •decreases, while the shearing stress is the same throughout the crank-arm ; 
 •consequently this latter stress is large compared with the bending stress 
 close to the crank-pin, and so it is not sufficient to provide there only for 
 bending stresses. The section at this point should be such that, in addition 
 to what is given by the calculation from the bending moment, there is an 
 extra square inch for every 8,000 lbs. of thrust, on the connecting-rod. 
 Moreover, the crank-arm is subject to twisting from the action of the pin ; 
 strictly speaking, therefore, it should be calculated from the formula 
 
 Tj = M + VM 2 + T*. 
 
 The length of the boss h, into which the shaft is fitted, is from 0"75 to 
 1 "0 of the diameter of the shaft, and its thickness e must be calculated from 
 the twisting stress R X L. 
 
 The crank turns the shaft (fig. 105) by exerting a force S on the key, 
 whose centre of effort is on the circumference, and therefore at a distance 
 of half the diameter from the axis of the shaft, so that 
 
 SxB=RxL; orS = 2Rx|. 
 
 If the crank is loose, the area of the section of the key parallel to the 
 shaft must therefore not be less than S + 10,000 lbs. And the load on 
 the section of the crank-boss opposite the key is 
 
 2L 
 
 Q -S- R=R( 2 ^ - l) 
 
 The stress on *,he section of the boss crossways is T, so that 
 
 T — "t-? = R x L ; or T = 2 R _ L -. 
 2 D + e 
 
 The stress on this section should not exceed 9,000 lbs. 
 
 To avoid a complicated expression, it is convenient to assume a relation 
 
 between h and e, and to substitute the value of e thus found in the above 
 
 expression. The value of — in practice varies from 2, when there is not 
 
 much space for the crank, to 3, when there is ample room. 
 
 Example. — To find the section of the boss of a wrought-steel crank 8 inches 
 long ; the pressure on the crank-pin is 54,000 lbs., the diameter of the shaft 
 
 10 inches, and — assumed at 2 - 2. Stroke of piston 60 inches. 
 
 Here assume 
 
 8 
 e = — = 3 - 67 inches. 
 
 30 
 T = 2 x 54,000 iq-tT3^7 = 237,015 lbs. 
 
 Area = 237,015 -v- 9,000 = 26 '33 square inches. 
 
 And since // = 8 inches, e = ^— = 3 3 inches. 
 
 o 
 
PADDLE-SHAFTS. 301 
 
 The cranks of marine engines are always of steel, wrought or cast, and 
 generally of the same materials of which the shaft is made, so that the length 
 and thickness of boss may bear a constant relation to the diameter of the shaft. 
 
 When h = D, then e = 032 D. 
 „ h = 0-9 D, „ e = 0-34 D. 
 „ h = 0-8 D, „ e - 0-35 D. 
 „ h = 07 D. .. e = 0-36 D. 
 
 The crank-eye or boss into which the pin is fitted should bear the same 
 relation to the pin that the boss does to the shaft. 
 
 Cranks are always shrunk on to both shaft and pin, and when this opera- 
 tion is carefully and well done, a key to the latter is almost unnecessary, and 
 some engineers have latterly omitted to fit one to even very large pins ; some 
 engineers simply drill a hole half into the shaft and half into the crank, and 
 drive into it a steel pin so as to answer the purpose of a key. 
 
 The diameter of the shaft end on to which the crank is fitted should be 
 1*1 X diameter of the journal. Overhung cranks are never fitted now to 
 screw engines, as they often proved to be very unsatisfactory, from the fact 
 of the whole of the pressure coming on one bearing, and the whole of the 
 bending and twisting stresses being taken by one crank and journal. 
 
 Paddle-shafts. — The cranks of a paddle-wheel engine (fig. 105) are still 
 often overhung, and in the case of double engines, the arm to which the pin 
 is secured is the one fitted to the intermediate shaft ; the pin fits loosely into> 
 an eye on a crank or disc secured to the paddle-shaft, and so drives this latter 
 shaft. The effect of this arrangement is to give a very equable strain to the 
 paddle-shaft, for the pressure of the pin is always at right angles to the crank 
 on the paddle-shaft ; and in smooth water the power of each engine is very 
 nearly equally divided between the two wheels, and the bending action on 
 the paddle-shait never exceeds half that due to its own cylinder, for when 
 near the dead points the bending moment is at its maximum, and is whollv 
 taken on the crank-arm to which the pin is secured. For these reasons the 
 shaft to which the arm having the crank-pin secured is fitted must be stronger 
 than the outer shafts, especially when the ship is intended to work in rough 
 water, as it is liable then to have to transmit the whole twisting force of one 
 engine, and always takes, during certain periods of the revolution, the whole 
 bending force from that engine. Hence, if T be the maximum twisting 
 moment from one piston of a double paddle-wheel engine, and M the maxi- 
 mum bending moment from that piston, the 
 
 Maximum equivalent twisting moment on the intermediate shaft 
 
 = M + JW + T 2 , 
 
 And maximum equivalent twisting moment on the paddle-shaft 
 
 4+vW^ 
 
 Exception may be taken to the latter, since at times when one wheel is out 
 of water the whole of the twisting force of both engines is transmitted 
 through the shaft of the wheel which is deeply immersed ; but when the 
 maximum combined effect of twisting is on this one shaft, the bending 
 
 M 
 
 moment on the crank- journal is probably less than — -, and is that due to the 
 
302 MANUAL OF MARINE ENGINEERING. 
 
 force found by dividing the maximum twisting moment by the length of the 
 
 T x J2 
 crank, which is approximately T ; the distance at which this force acts 
 
 is measured from the face of the crank-arm to the edge of the casting, into 
 which the journal brass is fitted. 
 
 Example. — To find the diameter of intermediate and paddle-shafts of 
 a double paddle-wheel engine, having cylinders 80 inches diameter and 
 60 inches stroke, using steam of 45 lbs. pressure absolute, cutting off at 
 0-6 the stroke ; the distance between the bearing beds being 50 inches. 
 
 Maximum effective pressure in the cylinder will be about 40 lbs. per 
 square inch. Hence 
 
 Load on piston - - = 5026 x 40, or 201,040 lbs. 
 
 Maximum twisting moment = 201,040 x 30 = 6,031,200 inch-lbs. 
 
 Maximum bending moment = - — - — -. — = 2,513,000 inch-lbs. 
 
 (1) Maximum equivalent twisting moment on intermediate shaft 
 
 = 2,513,000 + ^2,513,0002 + 6,031,200 2 = 9,056,000 inch-lbs. 
 Diameter of shaft 
 
 (2) Maximum equivalent twisting moment on paddle-shaft 
 
 _ 2,513,000 = J(*fim*f + 6>03 ^ _ 7i417>800 inoh . lb . 
 
 Diameter of shaft 
 
 ,eyM£^ + 5 . 1 = 16 . 8inohes . 
 
 8000 
 
 The outer part of a paddle-wheel shaft is subject to twisting and bending 
 from the reaction of the water on the floats, and from bending due to the 
 weight of the wheel itself. The pressure on the float can be found by divid- 
 ing the twisting moment by the distance to the centre of the pressure of the 
 float from the shaft axis in inches. It is practically sufficiently accurate to 
 measure to the centre of fixed floats, and to gudgeons of feathering floats. 
 
 For example, the reaction of the water on the floats of the engine in the 
 preceding example, whose radius to float centres is 140 inches, will be found 
 
 .p _ , 6,031,200 .oAoniu 
 
 Pressure on floats = , ' — , or 43,080 lbs. 
 
 140 
 
 The twisting moment on the shaft is the same at the outer bearing as at 
 the inner, and is 6,031,200 inch-lbs. The weight of the wheel is 20 tons, or 
 44,800 lbs., and the distance of its centre from the bearing is 40 inches. 
 
 * Maximum bending moment 
 
 = (44,800 + 43,080) x 40 inches = 3,515,200 inch-lbs. 
 
 * In smooth water the bending force is really the resultant of the weight and reaction 
 on the floats, and may be taken = J weight 2 + reaction 2 . 
 
CRANK-SHAFT OF SCREW ENGINES. 303 
 
 Maximum equivalent twisting moment 
 
 = 3,515,200 + V3,515,200 2 + 6,031, 200 2 = 10,515,200. 
 Diameter of shaft 
 
 . lySflgS „ 5 -l , !M8 inches. 
 
 The outer end of a paddle shaft is subject to alternating stresses due to 
 the weight of wheel, and the inner part to intermittent ones (v. Table lxxxiii.). 
 
 The crank-shafts of paddle engines are now made in the same way as those 
 of screw engines. Sometimes they are made in one piece from a " solid " 
 forging, and sometimes in one piece " built up " (v. fig. 19). Those in very 
 large engines have a separate shaft for each cylinder, coupled as in screw 
 engines. 
 
 Crank-shaft of Screw Engines. — In case of 'the forward crank of a double 
 or treble engine, and the crank of a single engine having two arms, there is 
 the action of one engine only on it. On the forward journal and crank-arm 
 there is a twisting action sufficient to overcome the friction, and to drive the 
 eccentrics if fixed in this part, and half of the whole bending moment due to 
 the thrust on the crank-pin. On the aftward journal, the other half of the 
 bending moment, and the whole of the twisting moment, except the small 
 portion required as above ; this portion is at certain periods of the revolu- 
 tions so small, that in calculations for the journals it may be neglected. 
 
 Then equivalent twisting moment on aftward journal 
 
 "♦y®^ 
 
 2 
 
 M 
 
 Strain on forward journal = — . 
 
 In multiple-crank engines the aftward crank has not only to resist the 
 action of its own piston, but also to transmit the twisting strains of the 
 forward engines. There will be strains from its own piston, which may be 
 calculated in the same way as those on the forward cranks, and to these must 
 be added the twisting strain of the forward engine. 
 
 Let T 2 be the maximum twisting moment on the after engine from its own 
 piston, and M 2 the corresponding bending moment, T x the twisting moment on 
 the forward engines at the same period. 
 
 Then on the forward journal of the after crank, the twisting moment is T,. 
 M 9 v 
 
 and the bending strain -5—, so that — 
 
 Equivalent twisting moment on forward journal of after crank 
 
 On the after journal of the aftward crank, the twisting moment is 
 T 2 + T p and the bending moment -3, so that — 
 
 Equivalent twisting moment on after journal of aftward crank 
 
 f ♦ Ar) * v. 
 
304 MANUAL OF MARINE ENGINEERING. 
 
 The bending moment on the after-arm of the aftward crank will be found 
 by calculating the maximum force on the crank-pin tending to twist the shaft. 
 
 Let T„ be the maximum combined twisting moment, as found by the 
 methods indicated before, L the length of the crank or half-stroke of piston. 
 Then the maximum twisting force at the crank-pin is T„ ■•- L. 
 
 The maximum bending moment at any section of the after crank-arm of 
 the aftward crank, whose distance from the centre of the crank-pin is x 
 
 inches, is—-? x x. 
 
 The maximum bending moment on a section of the forward arm of the 
 
 same crank is — l x x. 
 
 Example. — To find the sizes of the parts of a crank-shaft of a double 
 expansive engine of 1000 I.H.P., the length of stroke is 40 inches, the cut- 
 off 0-6, and the stroke and the cranks at right angles. Revolutions 60 per 
 
 minute. Mean twisting moment of one engine = -— x 63,000. Since the 
 
 cut-off is 0-6, the ratio of maximum to mean twisting moment is 1-835 
 (Table xxxiv. ) ; therefore 
 
 Maximum twisting moment of one engine 
 
 = 1-835 x — x 63,000 = 963,375 inch-lbs. 
 60 
 
 Mean twisting moment of both engines 
 
 = l ™ x 63,000. 
 60 
 
 Ratio of maximum to mean twisting moments is 1-27 (Table xxxiv.); 
 therefore 
 
 Maximum twisting moment of both engines 
 
 = 1-27 x 1229 x 63,000 = 1,333,500 inch-lbs. 
 60 
 
 Maximum turning force on forward pin 
 
 = 963 > 375 ^ 48,168 lbs. 
 20 
 
 Maximum turning force on aftward pin 
 
 = 1 > 333 > 5QQ = 66,675 lbs 
 20 
 
 Assuming the distance between the bearings on which the brasses are 
 bedded to be 30 inches, 
 
 The maximum bending moment on each of the two forward journals 
 
 = 48 ' 168 u X 3 ° = 180,630 inch-lbs. 
 
 8 
 
 That on the two journals of aftward crank 
 _ 66,675 x 30 = 250)0()() . nch lba 
 8 
 
TWISTING MOMENT. 305 
 
 Then diameter of foremost journal 
 
 80 ' 0Q x 10-2 = 6-13 inches. 
 
 \r- 
 
 The maximum equivalent twisting moment on after journal of forward 
 crank 
 
 = 180,630 + 7l80,630 2 + 963,375 2 = 1,160,630 inch-lbs. 
 Diameter of journal 
 
 /l,160,630 
 
 8000 
 
 x 5-1 = 9*04 inches. 
 
 The maximum equivalent twisting moment on fore journal of aftward 
 crank 
 
 - 250,000 + 7250,0002 + 963,375 2 = 1,245,000 inch-lbs. 
 Diameter of journal 
 
 3 /1,245,000 _ . n __ . , 
 V 80Q0 - 5-1 = 9-25 inches. 
 
 Maximum equivalent twisting moment on aftermost journal 
 
 = 250,000 + s/250,000 2 + 1,333,500* = 1,606,700 inch-lbs. 
 Diameter of journal 
 
 3 A,606,700 
 
 8000 
 
 x 5*1 = 10*1 inches. 
 
 The aftermost crank-arm will be 1 1 inches across the face ; to find its 
 thickness 18 inches from the pin. 
 
 Bending moment at that section = 66,675 x 18 = 1,000,000 inch-lbs. 
 
 Thickness = , 19 '" or L n = 7*44 inches. 
 11* x oOOO 
 
 In actual practice the crank-shaft would not be made with the four 
 journals all of different diameter, but some engineers make the shafts partly 
 in accordance with theory, by arranging the two forward journals of the same 
 diameter, and the two aftward journals of the same diameter ; that is, for 
 the example given above, the journals of the forward crank would be each 
 9 04 inches diameter, and those of the aftward one 10*1 inches diameter. 
 
 Example. — To find the dimensions of the crank-shaft of a single engine, 
 whose cylinder is 30 inches diameter and stroke 50 inches, the steam used is 
 65 lbs. per square inch absolute pressure, and the cut-off at 0*3 the stroke. 
 The distance between foundation facings for shaft brasses is 40 inches. The 
 connecting-rod is 100 inches long. Back pressure and loss at piston are 
 5 lbs. 
 
 The maximum pressure on the piston is 60 x 706 = 42,360 lbs. 
 
 The maximum twisting moment occurs just at the cut-off in this case, and 
 is 42,360 x 24, or 1,016,640 inch-lbs. 
 
 The bending moment on each journal at that period — — tt or 
 
 211,800 inch-lbs. 
 
 20 
 
306 MANUAL OF MARINE ENGINEERING. 
 
 The bending moment on the after arm, at a distance of 22 inches from the 
 centre of crank-pin, is 42,360 x 22, or 931,920 inch-lbs. 
 
 Diameter of fore journal 
 
 */ 
 
 211,800 .... 
 
 x 10-2 = 6-46 inches. 
 
 8000 
 Maximum equivalent twisting moment on after journal 
 
 - 211,800 + N /211,800 2 + 1,016,640 2 - 1,250,200 inch-lbs. 
 Diameter of after journal 
 
 V 
 
 f/lZgjF. «M =9-27 inches. 
 
 If the crank-arm is 10 inches wide at the face, then thickness of crank- 
 
 ™ . ^ , • 6 x 931,920 _ . . 
 arm at 22 inches from pin = -y™ 8000 ~ ' incnes - 
 
 The bending moment at the centre of the pin of a solid or rigidly built 
 op crank-shaft is n — . 
 
 Note. — A solid shaft, or one whose continuity of strength is unbroken from end to 
 end, is treated, so far as bending stresses are concerned, as a girder or beam secured at 
 its points of support ; or as a continuous girder supported at several points when there 
 are more than two journals. Hence, the bending moment in the middle between two 
 
 "R. v T Ft x L 
 
 journals is — - — ; and at the points ofsx^pport also — ^ — , since change of flexure takes 
 
 T 
 
 place at a distance — from the supports. Marine crank-shafts whose arms are at least 
 
 0*7 of the diameter thick, and whose bearings thoroughly support the shaft close to the 
 crank -arms, are really subject to little or no bending action at the journals. 
 
 At and near the end of the stroke the crank-arms are subject to a bending 
 moment applied very suddenly when the steam enters the cylinder, on opening 
 to lead after only slight compression. The force should be taken at twice 
 the load on the piston (2 P), and if L x is the length of the pin -f- the thickness 
 of the crank-arm as close to the shaft, then 
 
 „ ,, , 2PxL x PxL, 
 
 Bending moment = - L =s — - — -. 
 
 4 Ji 
 
 Crank-shafts are subject to intermittent loads, and therefore to inter- 
 mittent stresses (v. Table lxxxiiL). 
 
 And the bending moment on the section of each arm caused bv this force 
 
 PxL 
 
 is — -r — -, and acts at right angles to the bending force, due to the force on 
 
 the crank tending to twist the shaft. Hence, 
 
 . JPXL,) 3 P x Lj 
 
BUILT CRANK-SHAFTS. 
 
 307 
 
 In the previous example L x may be supposed to be 23 inches, and P is 
 42,360 lbs., a is 7 lbs. ; then 
 
 b = 
 
 3x 42,360 x 23 
 2 x 1- X 8,000 
 
 = 3 - 6 inches, 
 
 so that the forward crank-Arm must not be less than this thickness at 
 any part. 
 
 Crank-shafts for mercantile screw engines, when above 10 ins. diameter, are 
 generally made in duplicate pieces, so that in case of damage to one only a 
 part of the shaft is condemned, and a spare piece can be easily carried on 
 foreign voyages. And also by this plan there is less labour in replacing the 
 damaged part, than if the whole shaft is moved. 
 
 Built Crank-shafts. — Shafts above 10 inches diameter are better built up 
 than in one forging, and they can then be made of steel at much less cost 
 than a solid one ; indeed, many engineers now make crank-shafts of all 
 sizes on this principle. The crank-arms are usually of the same thickness 
 at the pin as at the shaft, and equal to 0*7 to - 8 of the diameter of shaft 
 
 I 
 
 
 — 
 
 
 j 
 
 
 
 • — 
 
 
 
 i 
 
 1 
 
 
 
 1 
 
 i 
 
 -r~- 
 
 i 
 i 
 
 j 
 
 
 
 
 
 
 
 
 r 
 
 
 
 "\ 
 
 Fig. 106.— Built up Crank-shaft. 
 
 journals ; the end view, as in fig. 106, shows the usual shape for large 
 cranks — smaller ones are often straight on the sides. Great care is 
 required properly to construct such a shaft so as to be perfectly true 
 when finished, and to have the arms shrunk on sufficiently tight without 
 leaving the metal around the pins, and shaft-ends in such a state of tension 
 as to be dangerous. 
 
 The thickness of the metal around the shaft, etc., can be calculated, as 
 before stated for the overhung crank. 
 
 The crank-arms are sometimes forged with the shaft-ends, and the pins 
 shrunk into eyes in the arms. This method has advantages, but it is very 
 unsightly, and misses one of the chief merits of a built crank-shaft. Another 
 arrangement is to make the crank-pin and arms in one piece, generally a 
 steel casting, and shrunk it on the shaft-ends or shanks ; and sometimes 
 it is secured to the latter by flanges, bolts, etc. 
 
 There are a number of other patented forms of crank-shafts, some having 
 the crank-arms of cast steel, and some of forged steel and iron, so arranged 
 a? to couple the shafts at the cranks instead of between them. 
 
308 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Fig. 107 shows a piece of crank-shaft as generally made for fast-running 
 engines, and indicate the modifications made when further weight is to be 
 saved, as in naval and other high-speed ships. 
 
 Couplings. — It is usual now to have the coupling forged with the shaft 
 instead of keyed on as formerly. The tube shafts of twin-screw engines, 
 however, generally have one coupling keyed on. As a rule, the only stress 
 to which a coupling is subject is due to twisting ; hence, if t be the 
 thickness of the flange, and r the distance of any part of it from the centre 
 of the shaft which is subject to a twisting moment T : the section of metal 
 resisting the force is 2 n r t ; and if / be the stress per square inch on this 
 section, acting at the distance r . then 
 
 T 
 
 T = 2 cr r t f x r = 2 t r- . f . t, that is, thickness of flange = .-: =-> 
 
 J J ° 2 t r-f 
 
 If r is the radius of the shaft subject to twisting only, so that -q-/is 
 equal to T. Then 
 
 Thickness of flange = 
 
 irr' 6 f 
 
 3*i*f, or T 
 
 Ci 
 
 
 / Section, ( 
 
 
 "7 
 
 P®®»r 
 
 I 
 
 < 
 
 
 — 
 
 
 
 
 • 
 
 
 r 
 
 Fig. 107.— Naval Crank-shaft. 
 
 From practical considerations the thickness of the flange should not be 
 less than the diameter of the coupling bolts, and since the strength of a 
 coupling is somewhat impaired by the holes drilled for the bolts, it should 
 be about '27 the diameter of the shaft subject to twisting only. 
 
 Coupling Bolts. — When shafts are close coupled, and the bolts are a good 
 fit in the holes, they are subject to a shearing force only, caused by the 
 torque on the shaft ; hence, if d be the diameter of the bolts, whose 
 number is n, K is the distance from centre of bolts to centre of shaft, T the 
 twisting moment, and D the diameter of the shaft subject to twisting only. 
 Then 
 
 T = n—r—/ x K ; or d 
 
 V o- 
 
 T 
 
 7854/. K.r* 
 
 but 
 
 Hence, 
 
 tD 3 
 
 16 
 
 Diameter of bolts = -^ . / ^. 
 
 2 \ n \ K. 
 
SURFACE OF CRANK-PINS AND SHAFT-JOURNALS. 309 
 
 If K is always taken at 0-8 x D. Then 
 
 Diameter of bolts = — . / zr— . 
 
 '2 V 0-8 x n 
 
 Then when there are 5 bolts, 
 
 Diameter of bolts — -~ */ = j—z = t • 
 
 2 v x - o 4 
 
 If 3 bolts, diameter of bolts = diameter of shaft 
 
 „ 5 
 
 ., 6 
 „ 7 
 „ 8 
 ., 9 
 „10 
 
 »i 
 >> 
 
 
 »> 
 
 »t 
 
 »» 
 
 >i 
 
 it 
 
 i» 
 
 of shj 
 
 ift -=- 3-10. 
 
 »» 
 
 3 58. 
 
 >> 
 
 4 00. 
 
 ii 
 
 4 38. 
 
 »> 
 
 4 73. 
 
 >» 
 
 5 06 
 
 >» 
 
 5-37. 
 
 11 
 
 5 67. 
 
 The number of bolts in a coupling depends sometimes on circumstances, 
 but usually there should be two, and an allowance of one more for each 
 2 inches of diameter of shaft, and the above proportions are based on this 
 allowance ; but when it is necessary to have the couplings as small as possible 
 the number may be increased, and with the consequent decrease in diameter, 
 the centres of bolts may be nearer to the centre of shaft. 
 
 The couplings of a two-crank engine, whose shaft is in duplicate halves 
 at right angles, should have four or a multiple of four bolts ; and for those 
 of a three-crank engine, whose shaft is in three duplicate pieces, the number 
 of bolts must be a multiple of three. 
 
 With shafts of mild steel, the bolts should be of a harder kind ; indeed, 
 as they are subject to shearing stresses only, they may be made with 
 advantage of a steel 40 to 45 tons tensile strength. 
 
 Surface of Crank-pins and Shaft-journals. — Measuring, as in the case of 
 gudgeons and crossheads, the effective bearing surface as the diameter multi- 
 plied by the length of the bearing, the bearing surface of crank-pins was 
 such that the pressure per sq. in. did not exceed 500 lbs. ; this, however, is 
 "hardly sufficient for the pins of high-speed engines, which should be such 
 
 , 6,500 
 ■that the pressure per square inch does not exceed ,- _ , _ = lbs. — that is, 
 r r n v n + 100 
 
 if d is the diameter of the crank-pin, and I its length, L the maximum 
 recurrent load on piston, R the revolutions per minute, then 
 
 . . T 6,500 WlTflOO 
 (ZXJ = L "-v/R + 100 ; ° r ~^500 . 
 
 "When the brass is recessed, so that it bears only in parts on the shaft, the 
 actual bearing surface should not be exposed to more than 600 lbs. pressure 
 per square inch. 
 
 The pins of paddle-wheel engines, owing to the comparatively slow speed 
 of shaft, may be designed, if necessary, to take a pressure of 800 to 900 lbs. 
 per square inch. 
 
 The main bearings in which the crank-shaft runs should be such that the 
 pressure never exceeds 600 lbs. per sq. in. in paddle engines, and in screw 
 •engines it should not exceed 400 lbs. The main bearings of screw engines, 
 
310 MANUAL OF MARINE ENGINEER TNG. 
 
 when room admits, should be such that the nominal pressure in lbs. per 
 
 ■ , , 4 > 300 
 
 sq. in. does not exceed -, p ■ /, measuring the whole of the bearing. 
 
 If, then, d be the diameter of the journals and / the length of each — 
 
 dxl - 2 X 4,300 * 
 
 The length of the crank-pin is from 1 to 1*5 of the diameter, and that of 
 each journal from 1*2 to 1*5 the diameter of the journal. Vertical engines 
 have usually sufficient space for a crank-pin 1*25 the diameter, and each 
 journal 1*5 the diameter. The foremost journal of a compound engine is 
 often made much shorter than the others, to allow the eccentric sheaves to- 
 be nearer the centre, so as to come in line with the valve-rod. 
 
 Owing to the comparatively small pressure on the crank-pins and journals 
 of three-crank compound, and triple- and quadruple-compound engines, the 
 intermediate journals may be generally somewhat shorter ; but the first and 
 last journals should not be materially less than was usual in a two-crank 
 engine. 
 
 Drivers. — In order to avoid any of the thrust of the propeller coming on 
 the crank-shaft and its bearings, the coupling-bolts connecting it to the 
 thrust-shaft are sometimes made without heads, so that they are free to move 
 in and out of the holes in the coupling flange of one of these shafts while held 
 firmly in that of the other. When this is so they should be of larger diameter 
 than the ordinary coupling-bolts, and the part projecting from the face 
 of the flange into which they are secured, should be larger still, so as to form 
 a shoulder, against which they may be tightened up, and give the necessary 
 strength to resist bending. The faces of the flanges when thus loosely coupled 
 should be from j to J inch apart. It was found necessary, generally, to pro- 
 vide means for lubricating these drivers, especially in heavily armoured and 
 high-powered warships with single screws. Drivers are still necessary in all 
 ships which are liable to change shape from varied loading, as in cargo steamers 
 in heavy seas, and in long steamers of light build. As a rule, however, it 
 is better to connect the crank and thrust-shafts in the ordinary way with 
 bolts, etc., and to leave such clearance in the bearings and brasses that the 
 crank-shaft may have some longitudinal " play." 
 
 Taper Bolts are often used in lieu of the ordinary parallel ones, especially 
 for the shaft next the propeller-shaft, to facilitate their withdrawal. Taper 
 bolts can be used with advantage when the flange is by necessity small, for 
 the screwed end is much smaller than the part at the junction of the two- 
 shafts subject to shearing. These bolts also are necessarily a tighter fit in 
 the hole, since the tightening of the nuts draws them farther into it. 
 
 Cross-keys are sometimes fitted to couplings. Half the key is bedded 
 into a recess in the face of each flange, and so it takes the shearing stress 
 from the bolts. 
 
 Since with a number of bolts or drivers it is possible, from wear or bad 
 workmanship, that the load is taken only by a part of them, it is usual to- 
 provide an excess of strength. This provision can be conveniently effected 
 by proportioning them to the diameter of the crank-shaft as if it were subject 
 to twisting only. 
 
PROPELLER-SHAFTS. 311 
 
 Propeller-Shafts,* sometimes called "screw" shafts, and sometimes "tail- 
 end " shafts. The propeller shaft is subject always to the torque of the 
 engine, and to bending due to the weight of the propeller. In rough weather, 
 when the ship is pitching, the strains are increased and often become 
 very severe ; for when the screw is partially immersed, in addition to the 
 twisting moment by the reaction of the water acting on one side only, there 
 is a bending stress as on a paddle-shaft ; besides which the momentum of 
 the screw when pitching also adds severe bending stresses, all of which are 
 moreover alternating. 
 
 In still water the bending moment on the shaft is the weight of the screw 
 multiplied by the distance of its centre from the stern bush. To provide 
 for the strains in rough weather, the bending moment should be taken at 
 twice this value ; for ships which may cross the Atlantic at any time in 
 ballast trim, even this is insufficient to produce a shaft sufficiently large to 
 last a satisfactory length of time, inasmuch as the material of a shaft working 
 under such conditions is subject to alternating stresses of considerable intensity, 
 which tend to degrade it and render it unfit to resist such stresses as come 
 on it. The diameter of the screw is also an important factor in determining 
 the size of screw shafts of ships subject to rough weather. 
 
 Hence, if T is the maximum twisting moment on the crank-shaft. \V the 
 weight of the propeller in pounds, and L the distance of its centre from the 
 stern bush, 
 
 Maximum bending moment = 2 W x L ; 
 and 
 
 Maximum equivalent twisting moment T\ 
 
 = 2 W x L + n/(2W x Lp 7~T2; 
 
 and as before, 
 
 /t 
 
 Diameter of screw-shaft = I ' — ~ x 5'1. 
 
 V / 
 
 Example. — To find the diameter of the screw-shaft for an engine whose 
 maximum twisting moment is 1,333,500 inch-lbs. The weight of the screw 
 is 6000 lbs., and the distance of its centre from stern bush is 20 inches. 
 
 The max. bending moment 
 
 = 2 x 6000 x 20 = 210,000 inch-lbs. 
 The max. equivalent twisting moment 
 
 = 240,000 + n/240,000 2 + 1,333,500 2 = 1,594,000 inch-lbs. 
 Diameter of shaft 
 
 =y- 
 
 1,594,000 
 
 x 5*1 = 11*2 inches. 
 
 6000 
 
 It is such a very serious matter when the screw-shaft breaks, that it 
 should always be of ample size, and for ships" in the Atlantic trade it should 
 be specially strong. It was usual to make it the same diameter as the crank- 
 shaft, but in some ships even this is not sufficient, and it is now not at ail 
 an unusual thing to make them 20 per cent, stronger than the crank-shaft. 
 
 Where the screw is fitted in a " banjo " frame for lifting above water when 
 the ship is under sail, the shaft is, of course, nearly wholly free from bending 
 stresses. 
 
 * The outboard shafts of naval ships and large express steamers are generally hollow, and for stilfncis 
 are of large diameter, with the hole as much as 0-7 x the external diameter. 
 
312 MANUAL OF MARINE ENGINEERING. 
 
 Outer Bearing. — It was formerly customary to provide an outer bearing 
 on or in the rudder-post, for the extreme end of the screw-shaft to rest upon ; 
 but since the rudder-post practically gives no support sideways, and a very 
 precarious one in any direction, the practice is now obsolete. Also, it was 
 found that when ships so fitted touched the ground with the heel, the screw- 
 shaft was often bent, and sometimes dangerously so. The strongest argu- 
 ment in favour of this outer bearing is, that it prevents the loss of the screw 
 when the shaft is broken ; but if the shaft is broken, and the ship has to 
 depend on the sails, it is better perhaps to be without the screw ; and if the 
 shaft is broken diagonally, as is often the case, and the screw is caused to 
 revolve from the motion of the ship, there is great risk of splitting the stern 
 tube. If there is no outer bearing, and the fracture is well within the tube, 
 the screw will not be lost, but will go back until the shaft-end butts against 
 the rudder-post, and revolves then without danger. If the shaft breaks close 
 to the propeller, and there is an outer bearing, the danger of damage to the 
 rudder-post and rudder is very great indeed, from the wrenching of the bearing 
 on the propeller falling out. 
 
 Screw-shaft End. — The shaft end fitted into the screw boss should be 
 turned to a taper of f inch to the foot ; if the taper is less than this, as was 
 sometimes the case, extreme difficulty is experienced in getting the screw off. 
 The screw should be secured by a key extending the whole length of the 
 boss, and driven into place after the screw is thoroughly well driven on. The 
 screw is retained in place by a nut, the screw-thread of which is the reverse 
 hand of that of the screw itself. A tail key through the shaft end was pre- 
 ferred by some engineers as a means of retaining the screw in place ; but 
 although it is a very safe plan, it is not so convenient as the nut. When a 
 nut is employed a safety key or pin is fitted in rear of it (v. fig. 173), or else 
 a set-screw or other simple means of locking it is used. These nuts are now. 
 often of cast steel or bronze, and made so as to cover the end of the shaft. 
 When the propeller is of bronze a " cap " nut of the same material is 
 necessary. Such nuts are secured by set-screws in the end " out of centre," 
 or set-screws in the boss-fitting in recesses in the side of the nut (v. fig. 174). 
 
 Screw-shafts are encased with bronze from the propeller to the inner end 
 of the stern tube in H.M. Navy ; in merchant ships this is now frequently 
 the case, although objected to partly on account of the expense, and partly 
 because it prevents examination of the shaft and the detection of flaws, which 
 may extend unobserved until rupture takes place. Bronze casings, on the 
 bearing parts, are used not so much to protect the shaft from corrosion, as 
 to provide a better wearing surface when running on lignum vita?, and to admit 
 of wear without weakening the shaft thereby. Lloyd's Register now require 
 all casings to be the full length of the stern tube. 
 
 When working in sandy water lignum vita? wears very quickly and grinds 
 away the bronze casing. Ships which are often exposed to this are better 
 without the bronze casing and the lignum vita?, and should be fitted in lieu 
 with Fenton's white metal, and the shaft either without casing or cased 
 with a hard steel liner, which may be renewed when worn. 
 
 The Stern Bush should be of such length that the pressure per square 
 
 inch (measured as stated for bearings) does not exceed — . lbs. This 
 
 v & v/R 4- 100 
 
 bush has to sustain the weight of the propeller and that of a considerable 
 
THE STERN BUSH. 313 
 
 portion ol its shaft ; in a seaway it has also added to its load that due to 
 inertia. The above allowance, however, for smooth-water conditions provides 
 a large enough bush for sea-going ones. If d is the diameter of the shaft 
 over the bronze casing, d 1 = - 9 d that of the shaft under it ; I the length 
 of the bearing part of the bush, all in inches, and W the weight of the propeller 
 in pounds, and the length of shaft, whose weight is borne by the bush, as 
 15 diameters, or 15 d t = 13*5 d. 
 
 Weight of shaft = ^ x 15 d x x 0-28 = 3*3 tf x 3 , or 2-4 dK 
 
 Then total weight on bush = W + 2*4 r/ 3 pounds. 
 
 Then lx dx . 5Q ° = W + 2*4 d 3 , 
 s/R + 100 
 
 and I = {-j + 2*4 d*) X >-^— • 
 
 In a general way this total weight on the bush = 1*3 W for solid shafts, 
 and 1*225 W for hollow ones. 
 
 TV, i a *i. t. -Wv'E+100 
 
 Then, length of bush = ^ 1 , 
 
 ° ¥ X d 
 
 where for solid shafts F = 385, and for hollow shafts F = 408. 
 
 Example (1). — What should be the length of the bearing strips of a stern 
 bush for a ship whose screw weighs 25,000 lbs., revolves at 100 per minute, 
 and the shaft is 17 inches over the liner. 
 
 T . /25,000 . -, , 1Q .\ v'200 „ nQ . , 
 Length = f — ^ (- 2*4 X 289 ) -^r- = 60-a inches. 
 
 Example (2). — For a turbine-driven ship the screw is 6,600 lbs. in weight, 
 the shaft 12 inches over the liner, revolutions are 500. 
 
 T . 6,600 x s'600 
 
 Length = 409 x 1 J~ = 33 mches - 
 
 The stern bush in practice is of a length equal to three to four times the 
 diameter of bore. The stern-shaft should be supported on a bearing in the 
 tunnel when possible ; when this is either not possible or inconvenient, it 
 should rest on a bush in the stern tube just abaft the stuffing-box. 
 
 In the mercantile marine the screw-shaft, when partly cased with bronze, 
 and running on bushes fitted with lignum vita?, the brass casings should 
 extend from the screw boss to an inch or two beyond the inner end of bush, 
 and should do so also where the shaft passes through the stuffing-box. and 
 inner bush when there is one. Of late years there has been a tendency to 
 follow the Admiralty method of casing in the whole length covered by the 
 stern tube. In order to make the process as easy and inexpensive as possible, 
 the stern tube is made very short, so that the casing is very little longer 
 than the sum of the lengths of the two liners as formerly fitted. 
 
 The Thickness of Brass Casing at bearing parts is 0-3 inch + 0*035 x 
 diameter. In order to easily withdraw the shaft, the diameter of the inner 
 casing should be £ inch larger than the outer. 
 
 * Lloyd's Register requires it to be at least 4 diameters in length. 
 
314 MANUAL OP MARINE ENGINEERING. 
 
 Stern Tube. — In the Navy the stern tube is always of bronze, and is within 
 another steel tube secured to the framing of the ship. The lignum vita? 
 strips are fitted into the tube, either in separate grooves for each strip, or the 
 strips fit in side by side with a brass strip at the top secured by screws to 
 the tube which keys them so as to form a bush. A similar but shorter set 
 of strips is fitted next the stuffing-box. The brass stern tube fits into 
 the stern-post or spectacle piece accurately and tightly, and is secured to 
 the bulkhead by a flange, etc.* 
 
 In the mercantile marine the stern tube is nearly always of cast iron, 
 whoso thickness is 0*5 inch + # 08 the diameter. The stern bush of bronze 
 fits accurately into the tube, and the stuffing-box neck ring and gland are 
 lined with brass. There are various ways of fitting the stern tube in place. 
 
 The common plan adopted by most engineers is to turn the outer end so 
 as to fit accurately and tightly into the hole bored in the stern forgings or 
 castings, and secure it by a nut screwed on its end. The inner end has a 
 flange, and next it a projecting rim, which is turned to fit in the hole bored 
 in the bulkhead ; the flange is bolted to the bulkhead after a liner is fitted 
 between them (v. fig. 108). 
 
 The tube was sometimes bolted to the stern-post by two lugs cast with it, 
 one above and one below it. 
 
 Another plan of fixing the stern tube is to fit ifcs outer end into a recess 
 bored in the stern-post, and secure it by bolts to the bulkhead as before 
 described, and by two strong draw-bolts passing through the flange to a 
 partial bulkhead two or three frame spaces nearer the stern. In this case 
 the stern bush is partly in the tube and partly in the stern-post. 
 
 Stern Bushes. — When made of white metal they should be of a thick- 
 ness = 05 inch -(- 0"03 x diameter. Those fitted with lignum vitae are of 
 bronze, and formed with a flange, which is secured to the stern tube by screws, 
 which prevent it from turning or coming out. Stern bushes are often made 
 of strong cast iron, with white metal strips driven in, or with it run in as in 
 a main bearing. Such bushes have flanges like the bronze ones. The lignum 
 vita? is sometimes fitted in strips, as in stern tubes, and sometimes into 
 square holes, the bush being cast as a skeleton to hold the wooden blocks. 
 This latter plan is very convenient for small ships, but not a good one for 
 large ones, as the wood by the continued concussion gets impressed on the 
 cast-iron tube. Lignum vitse wears best in end grain, especially when it is 
 of inferior quality ; when cut from a good tree of large size it wears equally 
 well either way. 
 
 Lignum vita? and white metal strips should be from f-inch to f-inch 
 thick, and about three -to four times their thicknesses in breadth; they 
 must be bevelled so as to leave free watercourses between them. The brass 
 behind the strips should be 0'04 X diameter in thickness, and the metal 
 ridges between the wooden strips of the same thickness. Sometimes the 
 bronze bush is dispensed with, and the strips fitted into the cast-iron tube 
 as in the brass tube, but this is not good practice. 
 
 A pipe is fitted leading from the top of the stern tube to the bulkhead, 
 through which the water may run from the tube so as to cause a fresh supply 
 to enter from the sea, and thereby prevent heating (v. fig. 108). 
 
 * It is usual, however, to fit a separate bush crntaining the lignum vitae strips in all 
 but small ships for convenience of repair. 
 
f .V-'- ' ' ' * * ssss s 
 
 
 315- 
 
 5.^25351 
 
 e 
 
 o 
 
 5 
 
 - 
 
 u 
 
 «2 
 
 ■a 
 a 
 
 *- 
 a 
 a, 
 
 u 
 
 4> 
 
 6 
 
 e 
 
 s 
 E* 
 
 c 
 
 O 
 CQ 
 
 CO 
 
 o 
 
 OK 
 
•316 MANUAL OF MARINE ENGINEERING. 
 
 From time to time attempts have been made to exclude sea water from 
 the stern tube, and lubricate the shaft bearings with oil or soapy water. 
 Ordinary stuffing-box, gland, etc., have been fitted to the tube end, and the 
 lubricant fed through a pipe from the upper deck, so as to give a " head " 
 superior to that of the sea water, and so prevent any leakage inward. No 
 great degree of success has followed this plan, as no provision is made for the 
 wear of the bush. But if the bush did not wear the violent action of the 
 screw in a seaway would inevitably cause leakage. This difficulty has been 
 got over by Mr. Cedervall, who prevents water from entering by the means 
 shown in fig. 108, which have been generally successful in practice. 
 
 Thrust-shaft. — Although the crank-shaft was sometimes made with a 
 ■collar or collars on it, to take the thrust, it is not good practice, especially 
 for large engines. The crank-shaft should be required to take only such 
 loads and motions as are due to the direct action of the pistons, and be free 
 to move around in its bearings without end pressure ; and since any longi- 
 tudinal displacement of the crank-shaft tends to throw abnormal strains 
 •on the working parts, it is better to remove all causes of such a derangement. 
 To this end the thrust collars should be on one of the intermediate shafts, 
 and for convenience on that one next the crank-shaft. If possible the thrust 
 bearing should bean the engine-room, and it is for this purpose chiefly that 
 the collars were sometimes on the crank-shaft. 
 
 Thrust. — To find the thrust along the shaft of a screw engine, it is neces- 
 sary to know the speed of the ship and the effective horse-power. The 
 •effective horse-power is in this case the power actually employed in pro- 
 ducing thrust, and, of course, its relation to the indicated horse-power depends 
 •on the combined efficiency of the engines and propeller. This ratio may be 
 taken as 0*77 with the best high-speed engines, and - 68 with the ordinary 
 merchant steamers' machinery. With turbines it is probably 08. If P be 
 the pressure in pounds exerted by the propeller against the thrust bearing, 
 .and S the speed of the ship in feet per minute, then 
 
 Work done in moving the ship = P x S, 
 and therefore Effective H.P. = PxSt 33,000, 
 
 P X S 
 
 or I.H.P. x/ = 
 
 33,000' 
 
 +X. T, T TT T, 33,000 - 
 
 then P = I.H.P. x -~ — k/. 
 
 Now, if K be the speed in knots, 
 Then 
 
 „ v 6080 
 
 p = i.h.p. x 326 K X £ . 
 
 J? is called the mean normal thrust. 
 
 Thrust may be ascertained fairly accurately from the formula — 
 
 i t> D* x VA x V 2 * 
 
 Thrust in pounds P = 5 — X tr. 
 
THRUST. 317 
 
 D is the diameter in feet, A the acting surface in square feet, V is = pitch X 
 revolutions per second, G = 0*4 to 0"5, and P r is the pitch ratio. 
 
 Example. — To find the thrust on the shafting of an engine whose I.H.P: 
 is 2,000, and the speed of 'the ship 12 knots, the efficiency 0*66 
 
 P = 2,000 X 326 * °' 66 = 30,166 lbs. 
 
 Now, it will be seen that P varies with the I.H.P., and inversely as the 
 speed, so that the thrust of a particular screw may vary very considerably ; 
 for if from some cause the speed is decreased, without a corresponding decrease 
 in the power, the thrust must of necessity increase. This actually occurs 
 in practice, and must be provided for always. The times when the actual 
 thrust exceeds the normal thrust are when the engine first moves, and its 
 power is employed in overcoming the inertia of the ship, also when the ship 
 is towing, and when driving against a head wind or sea. It is also to be 
 noted that the speed of ship means speed through the water ; for it is on this- 
 account that so little strain comes on the moorings of a ship whose engines 
 are working at full speed when in a dock or confined piece of water. In this 
 case it is only at first starting that any great tension is thrown on the moor- 
 ings, for as soon as the water is set into motion so as to flow past the ship in 
 a steady stream, the power is absorbed in facing the stream and really pro- 
 pelling the ship through the water. 
 
 Another cause of variation in the thrust is the variation in the twisting 
 moment, which is, as before shown, very great in certain classes of engines. 
 
 The surface exposed to thrust may, however, be calculated from the mean 
 normal thrust, and allowance made for all emergencies. This surface should 
 be such that the pressure per square inch from the mean normal thrust does 
 not exceed 70 lbs. ; and for tugboats or ships especially exposed to severe 
 weather, or service analogous to either of these, it should not exceed 50 lbs. 
 Ordinary merchant ships have usually such surface that the nominal thrust 
 does not exceed 60 lbs., and naval ships 45 lbs., as they may have to tow 
 or do analogous work at full power. 
 
 Friction Loss at the Thrust Block is not so great as generally supposed 
 by those who devise and patent special means for its reduction. As a matter 
 of fact, with carefully turned steel shafts with multiple collars running 
 against good white metal well lubricated the loss is never more than 1 '5 per 
 cent, of the I.H.P. transmitted by the shaft, and generally is as follows : — 
 
 The ordinary merchant steamer, . . . 04 per cent. 
 
 Express steamers and large naval ships. . . 0'5 ,, 
 
 High-revolution steamers and naval ships. . 0'65 ,, 
 
 Turbine steamers of large size, . . . 0"70 ,, 
 
 ,, high revolution, . . 1*0 to 1 "3 per cent. 
 
 The following rule holds good for all : — 
 
 Loss per cent. = 0*05 vrevs. per min. 
 
 ♦Pressure per Sq. In. on a Thrust Bearing should not exceed 2,700 + VRrf+100, 
 R being the revs, per min and d the diameter of thrust-shaft in inches. 
 
 Diameter ot Thrust Collars. — Let P be the mean normal thrust, d the 
 
 * With Michel bearings the pressure may be as much as 17,500 + y/Hd + 100. 
 
318 MANUAL OF MARINE ENGINEERING. 
 
 diameter of the shaft, and D the diameter of the thrust collars, whose number 
 is n, and the pressure 60 lbs. Then 
 
 P _ 60 (li^! - I*) n = 47 (B 2 - d 1 ) 
 
 and 
 
 n 
 
 °= y 
 
 2 + 
 
 47 n 
 The thickness of each collar for mere strength 
 
 »s» 
 
 = --=- (rrd x 1000) = 
 
 n v ' 3142 dn 
 
 In practice the thickness of each collar = 0*4 (D — d). 
 
 (1) Space between the collars, if rings are of solid brass = 0*4 (D — d). 
 
 (2) Space between the collars, if rings are of cast iron faced with bronze or 
 white metal = 0*75 (D — d). 
 
 (3) Space between the collars, if rings are of hollow bronze for water to 
 circulate through = D — d. 
 
 The number of collars depends very much on the size of the engine and 
 the prejudice of the designer. If there are many collars, they are of necessity 
 somewhat small, and although the chances are in favour of the majority of 
 them acting efficiently, provision must be made for the contingency of the 
 whole thrust coming on only one of them, and the larger the number of 
 collars, the less able is each one separately to resist the whole thrust. The 
 chief objection to a few collars is, that they are of necessity of comparatively 
 large diameter, and have, therefore, a greater resistance to turning owing to 
 the longer radius, besides a higher speed of rubbing surface ; there is also 
 the consideration of cost of forging against large collars. On the other 
 hand, when there are a few large collars a better design of thrust-block is 
 possible, and the rings can be made adjustable without removal. 
 
 The number of collars should vary with the size of the shaft, and a very 
 good rule is, that there should be one collar for shafts up to 5 inches diameter, 
 and then an additional collar for every 1*8 inches of diameter beyond this 
 when the collars are large, and the engines slow running ; fast-running naval 
 engines require more than this to get sufficient surface, say an additional 
 collar for each additional inch and a quarter. That is, 
 
 A 5 
 
 Number of collars for naval and express = 1 -f- 
 
 mercantile engines = 1 -f- 
 
 1-25 
 d-o 
 
 1-8 
 
 Michel's Thrust-block, shown in fig. 109, is based on the principle followed 
 by this engineer in all rubbing surfaces, whereby perfect lubrication is obtained 
 and freedom from metallic contact of surfaces ensured. This arrangement 
 requires one shaft collar only ; on each side in the block is a horse-shoe-shaped 
 inverted collar, each containing six segments of suitable 'metal — bronze or 
 white metal — loosely held and pivoted out of centre so as to tip slightly but 
 sufficiently to admit the passage of oil freely as the shaft revolves and main- 
 tain oil films between the surfaces, which are sheared continuously. The 
 inventor claims that this resistance to shear is constant and about J lb. per 
 
THRUST-BLOCKS. 
 
 319 
 
 square inch, whatever the thrust pressure may be ; the heat generated is, 
 therefore, 
 
 B.T.U. = number of square inches X speed -fr- 2 X 778. 
 
 Fig. 109. — Self-contained Michel Marine Thrust Bearing. 
 
 In practice the coefficient of friction is only 0-0015, and a pressure of 500 lbs 
 per square inch is not excessive under ordinary working conditions. With highly 
 viscous oil us much as 5 tons per square inch has been maintained satisfactorily. 
 Fig. 110 is the form of block generally used. Here the horse-shoes fit 
 
 L_ 
 
 Fig. 110— Thrust-block with Adjustable Collars. 
 
 over two screwed bars, one on either side of the block; nuts are fitted to 
 these bars, so that each collar may be adjusted by its own nuts, or the whole 
 of them by the nuts at the end. A simpler and cheaper method is to turn 
 
320 MANUAL OF MARINE ENGINEERING. 
 
 these side bars so that there is a collar between each pair of shoes instead 
 of a pair of nuts — that is, the bars have a set of collars corresponding to those 
 of the shaft — and take the thrust of the shoes as before. 
 
 Both these plans are most successful in practice, in great measure 
 due to the fact that the collars are open and exposed at the top, so 
 as to be easily lubricated and cooled by the air, and to their running 
 in oil. or in a mixture of oil and soapy water contained in the trough 
 below them. 
 
 It is most important that a bearing be placed close to the thrust, 
 so that the shaft cannot vibrate and cause uneven pressure over the 
 surface of the collars. The function of the thrust bearing? is to take 
 only end pressure. This is particularly the case when designed with 
 horse-shoe rings. 
 
 The Length of the Bearings of Tunnel Shafting will depend on the size of 
 the shaft and their distance apart. If d is the diameter and L the distance 
 in inches, the weight of the shaft which each bearing has to support is 
 
 W = jXLx 0-282, or 0-22 rf 2 xL 
 
 The bearing surface is taken at I X d square inches, and the pressure 
 should be about 60 lbs. per square inch in smooth water when the rate of 
 revolution does not exceed 100 per minute ; taking 
 
 * 
 
 Pressure per square inch = 800 -s- vrevs. + 100. 
 
 800 
 Then 0-22 d* x L = d X I X 
 
 </R + 100' 
 
 , „ . , , , . rfxLx Jr + loo . , 
 
 and Length of bearing =-. inches. 
 
 Example. — What should be the length of tunnel shaft bearings in a turbine 
 ship having shafts 6 inches diameter running in bearings 10 feet apart at 
 800 revolutions per minute ? 
 
 t -T, 1 1 • 6 X 120 X x/800 + 100 _ ... . . 
 Length of bearing = ■ = 0*93 inches. 
 
 Diameter of Shafts, Rules for, in Practice. — As almost every ship is classed 
 with one or other of the shipping registers, the shafting must in no case be 
 smaller than provided for in the rules of the corporation with which she is 
 classed, and from whom a machinery certificate is necessary. All British 
 ships carrying more than 12 passengers must have a certificate from the' 
 Board of Trade, whose surveyor has to report that, inter alia, the shafting is 
 fit and proper, etc. ; in his case it means that the diameters are not less than 
 given by the Board's Rules. 
 
 The Admiralty now generally specify the sizes of the various shafts, 
 which are almost invariably hollow steel. 
 
BOARD OF TRADE RULES. 
 
 321 
 
 Board of Trade Rules for Shafts. 
 
 For compound condensing engines with two or more cylinders, when the 
 cranks are not overhung : — 
 
 ~T5* 
 
 s 
 
 3 / C x P x PS 
 
 p = 
 
 / X S« 
 
 D«\ 
 
 8" ( n D»\ 
 
 Where S = diameter of shaft in inches. 
 
 d 2 = square of diameter of high-pressure cylinder in inches or sum 
 of squares of diameters when there are two or more high 
 pressure cylinders. 
 D 2 = square of diameter of low-pressure cylinder in inches or sum 
 of squares of diameters when there are two or more low- 
 pressure cylinders. 
 
 P = absolute pressure in lbs. per square inch — that is, boiler 
 pressure plus 15 lbs. 
 
 C = length of crank in inches. 
 
 f = constant from following table. 
 
 Note — Intermediate pressure cylinders do not appear in the formulae. 
 
 For ordinary condensing engines with one, two, or more cylinders, when 
 the cranks are not overhung : — 
 
 S 
 
 -V E 
 
 x P x D 8 
 
 P = 
 
 3 x / x S 3 
 C x D 2 ' 
 
 3 x/ 
 
 Where D 2 = square of diameter of cylinder in inches ; or sum of squares 
 of diameters where there are two or more cylinders. Other symbols as above 
 
 TABLE XXXV. — Board of Trade Factors for Shafts. 
 
 For Two Cranks- 
 Angle 
 between Cranks. 
 
 For Crank and Thrust Shafts. 
 
 For Tunnel Shaft. 
 / 
 
 For Propeller Shaft. # 
 / 
 
 90° 
 100° 
 110° 
 120° 
 130° 
 140° 
 150° 
 160° 
 170° 
 180° 
 
 1047 
 
 966 
 
 904 
 For paddle engines of ""' 
 
 ordinary type multiply „ . _ 
 
 constantin this column 7fi8 
 
 suitable for angle of ^..^ 
 
 cranks by 1*4. ij.i 
 
 743 
 
 740 
 
 1221 
 1128 
 1055 
 997 
 953 
 919 
 894 
 877 
 867 
 864 
 
 890 
 821 
 768 
 727 
 694 
 670 
 651 
 638 
 631 
 629 
 
 For Three Cranks. 
 120° 
 
 1110 
 
 1295 
 
 943 
 
 Note. — When there is only one crank the constants applicable are those in 
 
 the Table opposite 180°. 
 
 * The portion of the propeller shaft which is forward of the stern gland, and all 
 the thrust shaft, with the exception of the part enclosed in the thrust bearing, may be 
 the same diameter as the intermediate tunnel shafting. 
 
 21 
 
322 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Lloyd's Rules with regard to Shafting. 
 
 The diameters of crank and straight shafts are to be not less than those 
 given by the following formulae : — 
 
 TABLE XXXVI.— Diameters of Shafts— Lloyd's Rules. 
 
 Description of Engine. 
 
 Compound — Two cranks 
 
 at right angles, 
 Triple — Three cranks at ' 
 
 equal angles, 
 Quadruple — Two cranks 
 
 at right angles, 
 Quadruple— Three cranks, 
 
 Do. — Four cranks, 
 
 Diameter of Intermediate Shaft in 
 
 Inches. 
 
 
 (•04 A+ 006 D+ -02 S) x 3 JF. 
 (038 A + 009 B + -002 D + 0165 S) x 
 
 (034 A + -011 B + '004 C + "0014 D + 
 
 (•028 A + -014 B + 006 C + "0017 D + 
 (•033 A + -01 B + -004 C + "0013 D + 
 
 •016 S) x 
 
 •015 S) x 
 
 •0155 S) x 
 
 VpT 
 
 Where A is diameter of H.P. cylinder in inches. 
 B „ first LP. „ 
 
 C „ second LP. „ 
 
 D „ L.P. „ 
 
 S is stroke of pistons in inches. 
 P is boiler pressure above atmosphere in lbs. per square inch. 
 
 ), where P is the diameter of the propeller 
 
 The diameter of the crank-shaft and the thrust-shaft between the collars to 
 be at least f £ of that of the intermediate shafts. The diameter of thrust- 
 shaft may be tapered off at each end to the same size as that of the inter- 
 mediate shafts. 
 
 The diameter of the screw-shaft to be equal to the diameter of intermediate 
 
 shaft multiplied by ( - 63 -\ — ^ 
 
 .and T that of the intermediate shaft, both in inches. In no case, however, 
 must the diameter of the screw-shaft be less than TOT T. 
 
 Note. — This size of screw-shaft is intended to apply when continuous brass liners 
 are fitted the whole length of stern tube. If no liners are used, or if two separate ones, 
 then the shaft must be ? J of that given by the Rule. 
 
 Lloyd's Register Rules for Oil Engine Shafts. 
 
 The diameter of the shafts of oil engines for driving screw propellers may 
 be determined in the following manner when made of ordinary mild steel and 
 the maximum pressure in cylinders does not exceed 500 lbs. per square inch : — 
 
 (1) For petrol or paraffin engines for smooth-water service, diameter of 
 crank-shajt in inches = C ^D 2 X S. D is the diameter, and S the stroke 
 in inches. C = 0-34 ; for engines with six cylinders C = 0-36 ; if for open 
 sea work add 0-02 to each. 
 
 (2) Single-acting Diesel type engines : — 
 
 Diameter of crank shaft = N^D 2 X (A S + B L). 
 Here L is the distance apart of bearings (inner edges) ; the values of (A S+B L) 
 are as in this table. 
 
BRITISH CORPORATION RULES. 
 
 323 
 
 
 TABLE XXXVII.— Crank Coefficients. 
 
 
 Four-cycle Single-acting 
 Engine. 
 
 Two-cycle Single-acting 
 Engine. 
 
 Value of Coefficient, 
 AS + BL. 
 
 a 
 b 
 c 
 d 
 
 4 or 6 cylinders. 
 
 8 
 10 to 12 „ 
 16 
 
 2 to 3 cylinders. 
 
 4 
 5 to 6 
 
 8 »> 
 
 •089 S + -056 L 
 •099 S + -054 L 
 •111 S + -052 L 
 •131 S + -050 L 
 
 For auxiliary engines values 5 per cent. less. 
 
 In solid forged shafts, breadth of webs 1-33 diameter, and thickness 
 0-56 diameter of shaft or of equivalent strength. 
 (3) When no flywheel is fitted — 
 
 Diameter of intermediate shafts = coefficient X ^D 2 x S. 
 The value of this coefficient is as follows : — 
 
 TABLE XXXVIIa. 
 
 Four-cycle Single-acting. 
 
 Two-cycle Single-acting. 
 
 Value of 
 Coefficient. 
 
 4 cylinders. 
 
 6, 8, 10, or 12 cycles. 
 
 16 cycles. 
 
 2 cylinders. 
 
 3, 4, 5, or 6 cylinders. 
 
 8 cylinders. 
 
 0-456 
 0-436 . 
 0-466 
 
 N.B.— When the 
 stroke is not less than 
 12 nor more than 16 
 diameter of cylinder, 
 then 
 
 Diameter of internal shaft = coefficient (735 D + "273 S). 
 (4) When flywheels are fitted, then coefficient : — 
 
 TABLE XXXVIIb. 
 
 Four-cylinder 
 
 Single-acting 
 
 Engine. 
 
 Two-cycle 
 
 Single-acting 
 
 Engine. 
 
 Value 
 
 of 
 
 Coefficient. 
 
 Four-cylinder 
 
 Single.acting 
 
 Engine. 
 
 Two-cycle 
 
 Single-acting 
 
 Engine. 
 
 Value 
 >i 
 
 Coefficient. 
 
 4 cylinders. 
 6 " „ 
 8 „ 
 
 2 cylinders. 
 
 3 „ 
 4 
 
 0-405 
 0-400 
 0-409 
 
 10 cylinders. 
 12 „ 
 16 
 
 5 cylinders. 
 
 6 „ 
 
 8 „ 
 
 0-420 
 0-427 
 0-461 
 
 (5) Diameter of screw shaft 
 
 = t(o- 
 
 63 + 
 
 0-03 P 
 
 )• 
 
 -), and never less than 
 
 107T; T being diameter of intermediate shaft and D that of propeller in inches. 
 The Board of Trade Rules for Shafting of Oil Engines is the same as the 
 above rules of Lloyd's Register. 
 
 The British Corporation Rules for Shafts. 
 
 Diameter of Shafting. — The minimum diameters of crank, thrust, pro- 
 peller, and intermediate shafts may be found from the following formulae, 
 except where the ratio of length of stroke to distance between main bearings 
 is unusual, when they will receive special consideration : — 
 
 »-y 
 
 P X L 2 x S 
 
 C 
 
 X B. 
 
 Where D 
 P 
 S 
 L 
 B 
 B 
 B 
 
 diameter of shaft. 
 
 absolute pressure — i.e., boiler pressure + 15 lbs. 
 
 stroke of engine, in inches. 
 
 diameter of low-pressure cylinder, in inches. 
 
 1 "0 for crank- and thrust-shafts. 
 
 - 95 for intermediate shafts. 
 
 for propeller shafts to be taken from the following Table 
 
324 MANUAL OF MARINE ENGINEERING. 
 
 TABLE XXXVIII. — British Corporation Factors for Shafts. 
 
 Coefficient of 
 
 i 
 Ratio of Diameter of Propeller to Diameter of Crank-shaft. 
 
 Displacement of 
 
 
 
 
 Vessel at four-fifths 
 
 
 
 
 
 
 
 Moulded Depth. 
 
 13 
 
 14 
 
 15 
 
 16 
 
 17 
 
 18 
 
 •6 
 
 10 
 
 101 
 
 102 
 
 103 
 
 1-04 
 
 105 
 
 •62 
 
 101 
 
 102 
 
 103 
 
 1 
 
 04 
 
 105 
 
 1 
 
 06 
 
 •64 
 
 102 
 
 103 
 
 1-04 
 
 1 
 
 05 
 
 1-06 
 
 1 
 
 07 
 
 •66 
 
 103 
 
 104 
 
 105 
 
 1 
 
 06 
 
 107 
 
 1 
 
 08 
 
 •68 
 
 104 
 
 105 
 
 1-06 
 
 1 
 
 07 
 
 1-08 
 
 1 
 
 09 
 
 •70 
 
 1-05 
 
 106 
 
 107 
 
 1 
 
 08 
 
 1 09 
 
 1 
 
 10 
 
 •72 
 
 1 06 
 
 1-07 
 
 1-08 
 
 1 
 
 09 
 
 110 
 
 1 
 
 11 
 
 •74 
 
 107 
 
 1-08 
 
 1-09 
 
 1 
 
 10 
 
 1-11 
 
 1 
 
 12 
 
 •76 
 
 1-08 
 
 1-09 
 
 1-10 
 
 1 
 
 11 
 
 1-12 
 
 1 
 
 13 
 
 •78 
 
 109 
 
 1-10 
 
 111 
 
 1 
 
 12 
 
 113 
 
 1 
 
 14 
 
 •80 
 
 110 
 
 111 
 
 112 
 
 113 
 
 114 
 
 115 
 
 L2 
 
 The value of the divisor C in the formula depends on the ratio t™, 
 
 where L = diameter of low-pressure cylinder and H of high-pressure 
 cylinder, in inches : — 
 
 
 Two Cranks at 90°, Compound 
 
 
 
 L 2 
 
 or Quadruple, also Three 
 
 Three Cranks at 120°, 
 
 Four Cranks at 90°, 
 
 BP 
 
 Cranks at 120°, Quadruple 
 Expansion. 
 
 Triple Expansion. 
 
 Quadruple Expansion. 
 
 Ratio 3 
 
 9,910 
 
 
 
 
 i H 
 
 10,160 
 
 ••• 
 
 .. 
 
 
 
 , H 
 
 10,410 
 
 ••• 
 
 ,. 
 
 
 
 , 3§ 
 
 10,660 
 
 • •• 
 
 •.. 
 
 
 
 , 3£ 
 
 10,910 
 
 •• . 
 
 •• 
 
 
 
 , 3f 
 
 11,160 
 
 ••• 
 
 • • 
 
 
 
 . H 
 
 11,410 
 
 •• • 
 
 a# 
 
 
 
 , H 
 
 11,660 
 
 ••• 
 
 ,. 
 
 
 
 , 4 
 
 11,910 
 
 ••> 
 
 • • 
 
 
 
 , H 
 
 12,160 
 
 • a • 
 
 .. 
 
 
 
 , 4£ 
 
 12,410 
 
 • •• 
 
 .. 
 
 
 
 . 4g 
 
 12,660 
 
 ... 
 
 .. 
 
 
 
 , *i 
 
 12,910 
 
 13,650 
 
 .. 
 
 . 
 
 
 , H 
 
 13,375 
 
 14,160 
 
 .. 
 
 
 
 , 5 
 
 13,840 
 
 14,670 
 
 •• 
 
 
 
 , 5i 
 
 14,305 
 
 15,180 
 
 .. 
 
 
 
 , 5* 
 
 14,770 
 
 15,690 
 
 .. 
 
 
 
 , 5f 
 
 15,235 
 
 16,200 
 
 ,, 
 
 
 
 , 6 
 
 15,700 
 
 16,710 
 
 .. 
 
 i 
 
 
 . 6i 
 
 16,630 
 
 17,730 
 
 ,, 
 
 
 
 , 7 
 
 17,560 
 
 18,630 
 
 ,. 
 
 
 
 . 7* 
 
 18,410 
 
 19,530 
 
 
 
 
 > 8 
 
 19,260 
 
 20,430 
 
 22,'66( 
 
 
 , H 
 
 20, 1 10 
 
 21,330 
 
 23,660 
 
 
 , 9 
 
 20,960 
 
 22,200 
 
 24,660 
 
 
 , 9* 
 
 21,750 
 
 23,070 
 
 25,660 
 
 
 , 10 
 
 22,540 
 
 23,940 
 
 26,580 
 
 
 , ioi 
 
 23,330 
 
 24,810 
 
 27,500 
 
 
 , 11 
 
 24,120 
 
 25,660 
 
 28,420 
 
 
 , Hi 
 
 24,900 
 
 26,500 
 
 29,340 
 
 
 , 12 
 
 25,680 
 
 27,340 
 
 30,260 
 
shafts for screw engines. 325 
 
 Rules of the Bureau Veritas. 
 
 Shafts for Screw Steamers. 
 
 (a) Crank Shafts. — § 7. When the crank of a screw engine is not over- 
 hung, the diameter of the shaft shall be determined by one of the following 
 formula? : — 
 
 For non-compound condensing engines : « 
 
 , 3 / n P L D2 ... 
 
 <*- V— C - - - (A) 
 
 For double, triple, and quadruple-expansion engines : 
 
 • d= y PLKD-f 0-lnD'X . (B) 
 
 For shafts having a single overhung crank, the form under the radical 
 sign is to be multiplied by 
 
 8 + 7* 2 + 1. 
 
 For two-cylinder single-crank tandem engines the formula will therefore 
 
 be 
 
 = V pL(D» + 0-1D«)(« + J«» + l) . (C) 
 
 In those formulae : 
 
 d = diameter of the after shaft bearing in inches. 
 % = number of high-pressure cylinders. 
 
 D 1 = diameter of each high-pressure cylinder in inches. If there are 
 several high-pressure cylinders the diameters of which are not 
 the same, n x D x 2 represents the sum of the squares of theii 
 respective diameters. 
 n = number of low-pressure cylinders. 
 
 D = diameter of each low-pressure cylinder in inches. If there are 
 several low-pressure cylinders the diameters of which are not the 
 same, n D 2 represents the sum of the squares of their respective 
 diameters. 
 
 N. B. — For triple or quadruple-expansion engines the intermediate cylinders 
 do not come into account in the formula?. 
 
 L = length of stroke in inches, common to all pistons. 
 
 P = boiler pressure above atmosphere in pounds per square inch. 
 
 s = - (see fig. 110a). In order to determine a, B is supposed to be situated 
 half-way the length of the bearing, unless the latter be longer 
 than lh times the diameter ; in this case B C may be considered 
 as being equal to f of the diameter. 
 C = a constant, the values of which are given below for certain cases. 
 The values given apply to navigation in a sea-way ; for smooth 
 water (tugs excepted) the constants may be increased by 
 30 per cent. 
 If it is above 15 inches, it should be increased by an amount to be 
 determined by the Administration ; for built-up shafts, however, this latter 
 increase will not be required. 
 
326 
 
 MANUAL OF MARINE ENGINEERING. 
 
 For hollow shafts the diameter must be increased by 
 1 per cent, if the diameter of the hole is 0*4 of the outside diameter. 
 
 » 
 »» 
 
 >> 
 
 10 
 
 0-5 
 0-6 
 0-7 
 
 eF 
 
 l_ 
 
 '0"'-'<0i 
 
 -•!•—£- 
 
 
 >> 
 
 
 
 
 
 >» 
 
 
 
 
 
 Q] 
 
 ider 
 
 0-4 
 
 of 
 
 the 
 
 
 
 increase 
 
 will 
 
 be 
 
 V 
 
 — K 
 
 Fig. 110a. 
 
 If the hole is 
 outside diameter, no 
 required. 
 
 The Administration mav allow a 
 reduction on the diameter in certain 
 special cases, for instance in well-balanced 
 engines with light moving parts, or for 
 very superior workmanship, etc. On the 
 other hand, the Administration may re- 
 quire an augmentation for engines which 
 differ much from the average propor- 
 tions found in practice, thus, for instance, 
 for engines having a comparatively small stroke ; for compound engines, the 
 low-pressure cylinder of which has a very large size compared with the high- 
 pressure cylinder, etc. 
 
 Values of Constant in Formula (A) is 6,230 ; in (B) for double-expansion 
 two-cranks 90° it is 3,400 ; for triple three-cranks 120°, 3,900 ; for quadruples 
 and triples four-cranks 90°, 4,000, and for special crank setting for minimum 
 torsion add 100. 
 
 Other Cases. 
 
 (6) Propeller, Tunnel, and Thrust Shafts. — The diameter of the crank-shaft 
 
 must be ( 1*7 -j — 15 J per cent, in excess of the diameter of the crank-shaft 
 
 calculated from one of the formulae (A), (B), or (C) ; D being the diameter 
 of the propeller in inches, and d the diameter of the crank-shaft in inches. 
 It is recommended to fit the propeller shaft in such a way that it cannot 
 move endways, if for some reason or other it has been uncoupled from the 
 rest of the shafting. Liners fitted on propeller shafts to be tapered off at 
 ends. 
 
 For tunnel-shafts a reduction of 6 per cent, on the diameter of the crank- 
 shaft will be allowed. 
 
 The diameter of thrust shaft at the bottom of the collars, both between 
 and immediately beyond these latter, to be equal to that of the crank- 
 shaft, and tapered off at each end to the smaller diameter of the body 
 of the shaft. The thrust of the screw propeller must be taken up by 
 an efficient thrust block, so as to prevent any fore and aft strain on 
 the crank-shaft. 
 
 Shafts for Paddle Steamers. 
 
 § 8. In side wheel steamers having double, triple, or quadruple-expansion 
 engines with an intermediate shaft, each end of which carries an overhung 
 crank-pin fitting loosely into an eye of the paddle-shaft crank, the bearing of 
 
HORSE-POWER TRANSMISSIBLE THROUGH SHAFTS. 
 
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328 
 
 MANUAL OF MARINE ENGINEERING. 
 
 the latter (see A of the following sketch) must have its diameter calculated 
 from the formula : 
 
 3 / P L {n~Df + 0-1 n D 2 ) (s + Js 2 + 1) 
 
 (D) 
 
 where the letters have the same meaning as before (§ 7), except that a, for 
 
 determining s = -, is to be measured as shown in the sketch below, the point 
 r 
 
 B being the middle of the bearing. 
 
 Fig. 1106. 
 
 For two-cylinder compound receiver engines with two cranks at 90°. 
 
 C = 13,000 for navigation in smooth water. 
 7,100 for coasting vessels. 
 5,700 for sea-going vessels. 
 
 For triple-expansion engines with three cylinders and cranks at 120°. 
 
 C = 14,900 for navigation in smooth water. 
 8,150 for coasting vessels. 
 6,540 for sea-going vessels. 
 
 The diameter of the outer bearing of the paddle shaft and of the inter- 
 mediate shaft to be submitted to the Administration or the Surveyors for 
 approval. The same applies to other cases not dealt with in this paragraph. 
 
 Turbine Shafting (Bureau Veritas) 
 
 Diameter of tunnel-shafts 
 
 XT p 
 
 70x-^=(^ 
 
 R 
 
 H.P. is the shaft horse-power, R the corresponding revolutions per minute. 
 The rotor shaft diameter = 1 "05 X dt at least. 
 
 diameter of screw in inches 
 
 Diameter of screw-shaft = dt + 
 
 160 
 
SHAFTS OF OIL ENGINES. 
 
 329 
 
 Shafts of Oil Engines {Bureau Veritas). 
 
 Diameter of crank shaft = C \/D 2 x S, 
 
 where D is the diameter of cylinder, and S the stroke of piston in inches 
 
 S 8 
 
 C is a factor which varies inversely as =., and is 0*493 when ^ is 2*0. and O560 
 
 when it is only - 85. — is frequently T5. Then C = 0-520. 
 
 Steel Shafts (Bureau Veritas). 
 
 § 9. Shafts made of steel must have a tensile strength of 26 to 30J tons, 
 and test pieces cut from the forging must satisfactorily withstand the pre- 
 scribed tests. When it is desired to make use of steel of softer or harder 
 quality, the corresponding alteration in the diameter must be submitted 
 to the Administration (of the Bureau Veritas) for approval. 
 
 Summary. — The following is a summary of certain parts of this Chapter 
 which will be useful for reference : — 
 
 Rule 1. — Diameter of shaft 
 
 V ihr 
 
 V revoiutio: 
 
 oiutions 
 
 X F. 
 
 The various values of F can be ascertained from the following table, where 
 p is the absolute initial pressure, or that at which the safety valves are loaded 
 -+ 15 lbs., when there is no reducing valve : — 
 
 TABLE XXXIX.— Values of Factor F in Formula for Shafts 
 
 d 
 
 -^ 
 
 H.P. 
 
 R 
 
 x F. 
 
 
 
 
 F for Crank 
 
 F for Tunnel 
 
 
 
 
 Shaft. 
 
 Shaft. 
 
 Single-crank expansive 
 
 1 cyl. 
 
 Cut-off, 0.2 stroke, 
 
 170 + 3Vp 
 
 135 + 3\ p 
 
 )J M 
 
 1 „ 
 
 „ 0-G „ 
 
 113+ 3Vp 
 
 85+3 v'p 
 
 „ compound 
 
 2 „ 
 
 „ 0-G „ 
 
 100 + 3Vp 
 
 75 + 3 V / 
 
 Two-crank 90° expansive 
 
 2 „ 
 
 „ 0-2 „ 
 
 86 + 3Vp 
 
 70 + 3\> 
 
 )» if 
 
 2 
 
 „ 0-6 „ 
 
 68 + 3 Vp 
 
 55 + 3\p 
 
 „ compound 
 
 2 „ 
 
 „ 0-6 „ 
 
 63 + 3 Vp 
 
 50 + 3Vp 
 
 „ quadruple 
 
 4 » 
 
 „ 0-6 „ 
 
 58 + 3Vp 
 
 45 + 3Vp 
 
 Three-crank 120 c compound 
 
 3 „ 
 
 „ 0-5 „ 
 
 51 + 3Vp 
 
 40 + 3Vp 
 
 „ triple 
 
 3 „ 
 
 „ 0-6 „ 
 
 41 + 3Vp 
 
 30 + 3 v'p 
 
 Four-crank 90° triple 
 
 4 „ 
 
 „ 0-6 „ 
 
 47 + 3Vp 
 
 35 + 3 Vp 
 
 „ quadruple 
 
 4 „ 
 
 . „ 0-6 „ 
 
 55 + 3 Vp 
 
 38 + 3Vp 
 
 Three-crank triple-compound naval 
 
 engines (hollow) . , 
 
 72 
 
 63 
 
 Four-crank „ 
 
 M 
 
 n ( « ) • 
 
 75 
 
 61 
 
 Turbines taking S.H.P., . 
 
 • 
 
 . . • • 
 
 • ■ 
 
 56 
 
330 MANUAL OF MARINE ENGINEERING. 
 
 The Diameter of the Propeller Shaft of a screw ship can be obtained with 
 a close approximation to the correct size by adding to the diameter of the 
 tunnel-shaft an allowance based on the diameter of the screw and the con- 
 ditions of working. If d be the diameter of the screw-shaft at the outer 
 bush, d x the diameter of the tunnel-shaft or that to resist torque only, both 
 in inches, and D the diameter of the screw in feet. Then 
 
 d = (d x -f- x D) inches. 
 
 The value of x for a single screw ship working in smooth water only is - 06. 
 
 For ocean-going ships with single screws, . . . x = - 075. 
 
 For twin-screw ships with outer brackets, . . . x = - 09. 
 
 For Example. — An ocean-going single-screw ship has a propeller 18 feet 
 diameter, and the tunnel-shafts are 13 inches diameter, what should be that 
 of the propeller shaft. 
 
 Diameter of propeller-shaft = 13 -f 0-075 X 18 = 14-35 inches. 
 
 Example. — A twin-screw steamer having stern brackets and a screw 
 12 feet diameter, has tunnel-shafts 11^ inches diameter, what size should be 
 the outer propeller-shaft ? 
 
 Diameter shaft = 1125 + 0-09 x 12 = 12-33. 
 
 The crank- and screw-shafts of ships which are run at full speed only 
 occasionally, and for short periods, may be calculated by taking F at half the 
 above values. 
 
 Example. — To find the diameter of the crank-shaft for a three-crank 
 triple-compound engine having cylinders 30 inches, 45 inches, and 75 inches 
 diameter, and a stroke of 50 inches. The load on safety valve is 165 lbs., 
 the revolutions 90, and the I.H.P. 3,300. 
 
 (1) By above Rough Eule. 
 
 Diameter of crank-shaft = ^/^° X 41 +3 N /l80 = 14-39 ins. 
 
 (2) By Board of Trade Rule. 
 
 / 25 x 180 x 75 2 
 Diameter of crank-shaft = 8 / lUQ U + 75*\ = 14-12 ins. 
 
 (3) By Lloyd's Rules. 
 Diameter of crank-shaft 
 
 = \* (-038 x 30 + -009 x 45 + -002 x 75 + -0165 x 50) Vl65 = 14-5 ins. 
 
 (4) By British Corporation Rule. 
 
 /Tu 
 
 Diameter of crank-shaft = \ / — — ■,„' nn „" — = 14-3 ins. 
 
 [80 x 752 x 50 
 17,300 
 
 (5) By the Rules of the Bureau Veritas. 
 
 ' , , .. 3 / 165 x 50 (30 2 + 0-1 x 752) , A .,. 
 Diameter of shaft = \/ vJOO ™ im ' 
 
DETAILS OF CRANK-SHAFTS. 
 
 331 
 
 TABLE XXXIXa. — Shafts foe Paddle Engines. 
 
 Description of Engine. 
 
 Value of F 
 Intermediate 
 Shaft Journal. 
 
 Value of F 
 
 Paddle-shaft 
 
 Inner Journal. 
 
 Value of F 
 
 Paddle-shaft 
 
 Outer Journal. 
 
 Single-crank single-cylinder, 
 
 Two - crank two - cylinder ; cranks 1 
 virtually at right angles, and con- > 
 nected by link, - - - - ) 
 
 Two-crank two-cylinder, with inter- j 
 mediate shaft ; cranks at right > 
 angles, ) 
 
 Two-crank two-cylinder ; solid crank- ) 
 shaft ; cranks at right angles, - ) 
 
 ••• 
 
 58 
 
 • •• 
 
 80 
 53 
 
 50 
 55 
 
 100 
 65 
 
 65 
 
 65 
 
 1 
 
 For paddle steamers working only in smooth water the above values of F 
 may be reduced 20 per cent. 
 
 The above values of F were for iron shafts, but it is better not to 
 reduce them when employing mild steel, unless the forgings are from ingots 
 of the highest quality. 
 
 Details of Crank-shafts. 
 
 Crank-arms if forged solid with the shaft — 
 
 *Breadth - - »- - = 1*1 x diameter of shaft. 
 Thickness - - - = 0'75 
 Diameter of coupling - = 2*0 
 Thickness „ - - = 027 
 Number of coupling bolts = 2 
 
 x 
 
 X 
 X 
 
 )) 
 >» 
 
 Diameter 
 
 diameter of shaft in inches 
 4« + 19 
 
 = diameter of shaft -?- 
 
 10 
 
 Diameter of crank-pins - = 1 to 11 x diameter of shaft. 
 Length „ - = 1 to H the diameter of shaft. 
 
 Length of journals - - = 1 to 1£ „ 
 
 The following convenient rules give sizes closely approximating to those found 
 in practice, and may be used to obtain the diameter of the shafts, preliminary 
 to making a more elaborate calculation. They are, therefore, very useful in 
 the initial stages of a marine engine design to enable the designer to get on 
 with the work ; but being purely empirical they should be used with some 
 caution : — 
 
 d is the diameter of the cylinder in inches. 
 H MP 
 
 D ,, ,, L.P. ,, 
 
 S the stroke also in inches. 
 
 F a factor, which for the crank-shaft of ordinary compound engines 
 with cranks at 90° is 12 ; and for the tunnel-shafts is 13 ; for 
 three-crank triple-compound engines F is 14 for the crank shaft 
 and 15-2 for the tunnel-shafts ; four cranks, 13 and 14'2. 
 
 * Breadth* x thickness = 0-9 X 4*. 
 
332 MANUAL OF MARINE ENGINEERING. 
 
 (1) Ordinary compound engines 
 
 Diameter of shaft = 
 
 d + D+S 
 
 {2) Triple-compound three-crank engines 
 
 d+^+D+S 
 
 Diameter of shaft = 
 Taking the same example as before 
 
 F 
 
 Diameter of crank-shaft •= r-j = 14*3 inches. 
 
 14 
 
 Rule for Determining quickly the Diameter of Crank-shaft. — Taking D 
 as the diameter of each L.P. cylinder of a compound system, and S the stroke 
 both in inches, and p the initial pressure absolute. Then the diameter of 
 the crank-shaft may be found rapidly and with close approximation to correct- 
 ness by the following : — 
 
 TABLE XXXIX6. — Crank-shafts of Screw Engines. 
 
 Description of Engine. 
 
 Diameter. 
 Crank-shaft. 
 
 Two cranks at 90° compound, two cylinders, .... 
 Three cranks at 120° compound, three cylinders, . . 
 Three cranks at 120° triple-compound, three cylinders, . . 
 Four cranks at 90° triple-compound, four cylinders, . . . 
 Four cranks at 90° quadruple-compound, four cylinders, . 
 
 1-6 D + S r 
 no vp 
 
 2-85 D H- S ,- 
 150 N P 
 
 2 D + S i- 
 
 185 ^ 
 
 3 D + S ,- 
 
 205 Vp 
 2.5D + S V - 
 225 7 
 
 The tunnel -shafts should be 0*95 X diameter of crank- shaft. 
 
 Board of Trade Rule for Shafts of Turbine Ships is as follows : — 
 
 Diameter 
 
 3 /I.H.P. .__ _. 
 = ^/ -jj— X 40-2 x C, 
 
 where R is the number of revolutions per minute. 
 
 For the intermediate or tunnel-shafts, . . - . C = T500. 
 
 ,, propeller-shafts. . . . . . C = 1*667. 
 
 Hollow Shafts, now so extensively used in special ships, are especially 
 suited for the propeller-shafts of multiple-screw ships, inasmuch as they are 
 less liable to sag under their own weight, and, therefore, freer from the ten- 
 dency to whirl than solid ones would be under the same circumstances. 
 
DETAILS OF CRANK-SHAFTS. 333 
 
 If the diameter of the hollow shaft is d, and that of its bore d v then 
 
 — A — L * 
 The relation of the weights of these shafts will be -™ — "TT 
 In practice d x is about a half of d ; assuming this to be so, 
 
 j 2 
 
 Ratio of weight of solid to hollow = ■ * ,j . 
 
 d, = Vd* (1 - 0-0625) = 0-979 d. 
 
 Substituting this value of d s , 
 
 958 
 Ratio of weight of solid to hollow = ==» or 1-278. 
 
 That is, the solid shaft of the same strength is nearly 28 per cent, heavier. 
 
 The Outboard Shafts of large naval ships with twin screws and ordinary 
 brackets are of large diameter (as much as 28 inches), with a bore-hole 0-7 of 
 the external diameter. Under these circumstances — 
 
 (a) d s = ^/ i '- = 0-9W. 
 
 , , . , . , (0-91rf) 2 -828 
 
 (6) The ratio of weights = ^ _ ^ t = -^ Q - 
 
 That is, the solid shaft is 1-6 times the weight of the hollow one ; but in salt 
 water the difference is greater. Taking the specific gravity of steel as 7-86 
 and salt water as 1-03, then — 
 
 (c) Hollow shaft = d\\ - 049) X 7-86 - l-08eP = 2-98. 
 
 Solid shaft = (0-91cZ) 2 X (7-86 - 1-03) = 5-66. 
 The ratio is now 1-9. 
 
334 MANUAL OF MARINE ENGINEERING. 
 
 CHAPTER XIII. 
 
 FOUNDATIONS, BED-PLATES, COLUMNS, GUIDES, AND FRAMING. 
 
 For the good working of an engine it is essential that the fixed parts, such as 
 bed-plates, framing, etc., shall not only be strong enough to resist the strains 
 to which they are subject, but rigid and stiff enough to prevent any tendency 
 to rack or change of form which would throw abnormal stresses on the working 
 parts. With this object in view it is usual to construct such parts of cast iron 
 or cast steel ; and from their form and general construction thus enables the 
 engineer to manufacture a cheap structure having the necessary qualities. 
 But since ordinary cast-iron has not a high tensile strength, and is compara- 
 tively unsuited to withstand sudden shocks, Structures made of it cannot 
 be so light as if made of wrought-iron or steel ; so that when extreme lightness 
 of machinery is aimed at, the framing is usually made of wrought steel, and 
 rigidity given to it by cross bracing, etc. This latter system is, of course, 
 an expensive one in most engines, and only adopted when economy of weight 
 is of more importance than economy of manufacture. Although cast-iron 
 framing and bed-plates are undoubtedly cheaper and better for engines 
 generally, and continue to be used for all sizes in the mercantile marine, a 
 system of construction with wrought iron or steel for very large engines may 
 be adopted with advantage, and taking into account cost of patterns and 
 risk in casting may not be less economical. 
 
 Steel manufacturers can now, however, produce large and moderately 
 complicated castings in steel, and the foundations and frames of express 
 engines of large sizes are being made wholly of that material. Increased 
 experience in the making of steel castings has no doubt given greater faith in 
 their soundness as well as permitted of a lower price, and with improvement 
 will be more freely used for columns and foundations (fig. 32). At present 
 they can only be used to advantage in large engines owing to the considerable 
 thickness still demanded by the steel moulders. 
 
 Bed-plates and Foundations. — Vertical engines are usually built on a 
 structure called by these names, but are sometimes known as the sole-plate. 
 It contains the main bearings for the crank-shaft, and on it are the facings 
 for the columns, condenser, etc., and it used sometimes to contain the water- 
 ways leading from the condenser to the pumps. The foundation generally 
 contains only the main bearings, and has facings for the front columns only 
 when it is bolted to feet on the front of the condenser, so that with the latter 
 it forms the engine base. The condenser is in this case lower down, and the 
 weight and cost of half the sole-plate is saved. The part of the foundation 
 fitting to the feet on the condenser front should be of good depth, the flanges 
 strong, notched one into the other, and strongly bolted at top and bottom. 
 The transverse parts of the bed-plate, into which the main bearing brasses 
 are fitted, are sometimes formed like inverted bowstring girders, and are 
 
MAIN BEARINGS. 335 
 
 unsupported by the bed built in the ship, but span the space between the fore 
 and aft bearers. This is convenient sometimes, especially when the shaft 
 must be low down in ships having a good rise of floor, and also for very small 
 engines ; but it must be of strong form, and as it depends for strength only 
 on the connection to the longitudinal parts of the foundation, it is a most 
 convenient form when steel castings are used, as also for very light engines ; 
 additional support may be, and often is, given in the case of large engines by 
 bolting to the ship at the middle (fig. 114). When this particular style is 
 adopted great care should be exercised in designing the foundation, so that 
 the transverse portions have a good extended connection to the longitudinal 
 ones, especially in the direction of the column bases. If the longitudinal 
 parts are flat and straight on the bottom, so as to be in the same plane as 
 the rest of the foundation, they may be formed with flanges and bolted to the 
 steel seating in the ship, and from it receive support and strength. 
 
 Main Bearings. — The bearings in which the shaft journals run should 
 approximate, as far as possible, to a hole through a solid support. If it were 
 practicable a hole with a bush of suitable metal in it would form the best 
 possible bearing for a shaft ; but since the bearing, however well designed 
 and made, is liable, in course of time, to wear somewhat, it becomes a neces- 
 sity that there shall be some means of adjusting the brasses, so as to prevent 
 the shaft from having too much movement when they are worn.* In the case 
 of the vertical engine, the weight of the shaft, and the pressure from the 
 piston, act very nearly in the same direction, so that the wear is only verti- 
 cally above and below the shaft ; consequently the adjustment is necessary 
 only in a vertical direction. The greatest loads on the bearings, however, 
 are during the first half of the stroke, and consequently the position of mean 
 pressure on the journals is not exactly vertical ; this is also somewhat modi- 
 fied on the upstroke by the tendency of the shaft to roll on the surface of the 
 brasses, and on the downstroke it is aggravated from the same cause. In 
 fitting the brasses for a vertical engine, this should be borne in mind, and 
 every allowance made for taking the wear due to these causes. It is of the 
 utmost importance for the good working and endurance of a crank-shaft 
 that the bearings be rigid in themselves, and that the framework containing 
 them shall be rigid enough to sustain them perfectly in line one with another. 
 Crank-shafts are more severely tried by the giving or springing of the bearings 
 than any other cause, and they were more often broken from want of rigidity 
 in the bed-plates and seatings, than from the normal strains from the pistons,! 
 so that a shaft may be of ample size to bear the twisting and bending moments 
 if properly supported in its bearings, and yet give way after a few weeks' 
 work because it is in a light foundation in a weak ship. 
 
 The brasses were usually formed as shown in fig. Ill, and carefully bedded 
 into the recesses provided for them in the foundation. At one time it was 
 usual to design them with projecting facings, called chipping strips, to avoid 
 the labour of chipping and filing the whole of the surface ; this was, however, 
 found to be highly objectionable as engines increased in size ; with the in- 
 crease of boiler pressure, and consequent increased percussive action due 
 to the high initial loads, such an effect was produced on these strips and 
 the cast-iron surface on which they were borne, that engineers have gradually 
 abandoned the practice ; the planing machines also have rendered such a 
 device unnecessary, as it is practically as cheap to fit brasses now to bear 
 
 • With forced lubrication and the bearings shut off from dust, etc., tiicre is in practice no wear. 
 
336 
 
 MANUAL OF MARINE ENGINEERING. 
 
 over the whole surface as to do so only on strips. The square bottom brass 
 is objectionable on two grounds ; one being that it is impossible to remove 
 it in most engines without lifting the shaft, and the other that when it becomes 
 hot it is invariably distorted, from the variation in thickness of metal, with 
 the ultimate result that it is broken through the middle longitudinally. 
 
 Fig. 111. — Crank-shaft Bearing. 
 
 The first of these evils is avoided by making the bottom brass round and 
 of even thickness, so that it can be got out when relieved of the weight of 
 
 Fig. 112. — Improved Form of Crank-shaft Bearing. 
 
 the shaft, by being moved around until it is on the top of the journal. The 
 second evil is also partly avoided by making it of an even thickness ; but 
 this form of brass is often found cracked, and is liable to heat from its want 
 of stiffness. Both these brasses, when first heated by. abnormal friction, 
 
MA.IX-BEARIXG BOLTS. 337 
 
 tend to expand along the inner surface, which is in contact with the shaft ; 
 this would open the brass, and make the bore of larger diameter, if not resisted 
 by the cooler part near the cast iron, and by the bed-plate itself. If the 
 brass has become hot quickly and excessively, the resistance to expansion 
 produces permanent set on the layers of metal near the journal, consequently 
 on cooling this contracts, the brass closes and tends to grip the shaft ; it will 
 then set up sufficient friction to become again hot, and expand sufficiently 
 to ease itself from the shaft, when so long as that temperature is maintained 
 the shaft runs easily in the bearing. This is the reason why some bearings 
 always are a trifle warm, and will not work cool. A continuance of heating 
 and cooling will set up a mechanical action at the structure of the brass, 
 which must end in rupturing it, just as a piece of sheet metal is broken by 
 continually bending backwards and forwards about a certain line. 
 
 This pernicious action of the brass can be prevented by securing it to the 
 bed-plate, along its two longitudinal edges, as shown in fig. 112, by an H- 
 shaped strip, which holds both top and bottom brasses, so that they cannot 
 move in their beds. This method is a very simple one, and has been found 
 most successful in engines of all sizes. 
 
 It is also essential that the bearing to be efficient should be rigid through- 
 out its whole length ; this is not the case when the brasses have long over- 
 hanging ends, which afford little or no support to the shaft. To this end 
 it is better, when possible, to extend the bed for the brasses, so as to support 
 them over the whole of their length, as shown in fig. 112. 
 
 Caps or Keeps for Main Bearings are very generally made of wrought iron, 
 but as stiffness is as necessary as strength, cast steel, as in figs. 113 and 114, or 
 even cast iron, as in fig. 112, may be used with advantage in their construction. 
 A wrought-steel cap, which may be amply strong, is often far from stiff enough, 
 while a cast-iron cap, which is stiff enough for good working, is generally 
 amply strong. 
 
 Let d be the diameter of the main-bearing bolts (when there are only two 
 to each cap), t the thickness of the cap, and b its breadth, I is the pitch of the 
 bolts, all in inches ; / a factor, which for wrought iron and forged steel is 1, 
 for cast steel 1 % and for cast iron 2. 
 
 Thickness of cap - - = d . / - x f. 
 
 Thickness of brass at middle = — . / — . 
 
 3 V b 
 
 Main-bearing Bolts. — Each cap is usually held by two bolts, but very 
 
 large bearings have four bolts, two on each side, so as to avoid large bolts 
 
 and heavy nuts, and to distribute the load over the cap. When everything 
 
 is in good order and properly adjusted, the load from the piston should be 
 
 equally divided between the bolts ; but since, from a very slight difference in 
 
 setting of the nuts, the load may come on three, and sometimes even on two 
 
 bolts only, due allowance must be made for this. To meet this it should be 
 
 assumed that each bolt is capable of sustaining one-third the load on the 
 
 piston. If P is the maximum load on the piston in lbs., 
 
 P 
 Area of each bolt at bottom of thread = .,-,• 
 
 <V 
 
 22 
 
338 
 
 MANUAL OF MARINE ENGINEERING. 
 
 For mild steel/ = 5,000 lbs. for small and 7,000 lbs. for large bolts (r. Table 
 lxxxiii.). /— - 
 
 Also, Diameter of main-bearing bolt = diameter of cylinder X ^J JL,. 
 
 p is the maximum pressure per square inch, and is, as stated on p. 273, and 
 may be taken generally at 0'75 X absolute boiler pressure for the high-pres- 
 sure cylinder of a triple engine, to equal the load on the low-pressure piston. 
 
 Fig. 113.— Solid Steel Main Bearing Frame. 
 
 Fig. 114. — Cast-steel Main Bearing Frame for Naval Engines. 
 
 Brasses, so called because formerly always made of brass. — They should 
 be made of a metal which itself will withstand wear without wearing the 
 shaft journals, and whose surface is such that the shaft runs on it with 
 minimum amount of friction. The metal must also possess sufficient strength 
 in itself or its shell not to fracture under the percussive loads of the piston, 
 and be free from brittleness, so as not to crack when quickly cooled. Good 
 gun-metal or bronze possessed all the qualities essential for brasses when the 
 journal was. of wrought iron ; but the soft steel as used for shafts will not 
 run satisfactorily on bronze bearings. There are, however, other metals 
 which have certain of these qualities in a higher degree without having them 
 
BRASSES. 339 
 
 all. Cast iron is harder than ordinary bearing bronze, and when once worn 
 to a smooth surface gives equally good results ; but it is liable to fracture 
 from continued shocks and to crack when cooled suddenly. White metals 
 offer least resistance, or produce least friction, but most of them are too 
 soft to be used alone. Of the patent bronzes there are few which are suitable 
 for heavy bearings, and none of them have so far been shown to be much 
 superior to good gun-metal. 
 
 White metal carried in a bronze or cast-iron shell is beyond doubt the 
 best bearing for all qualities of steel journals, and if kept well lubricated 
 there is practically no wear on it or the journals. 
 
 When a bearing is of ample size, properly designed and constructed, well 
 lubricated and looked after, it may be of almost any of the white metals. If, 
 however, the bearing surface is limited there is a considerable difference in 
 the behaviour of the different metals ; but if badly designed and constructed 
 even the best metal will give trouble ; moreover, if not properly looked after 
 by the engineer, the best metal and the most careful design are of no avail. 
 
 Certain of the white metals have given the best results as a bearing surface, 
 and there is every reason for this, inasmuch as they do not cause abrasion 
 of the shaft, and if their own surface is injured it will not, as a rule, form 
 into tine, hard grit, which grinds both the surfaces, as all the hard bronzes do 
 more or less. When white metal is used it is highly important that the 
 shaft shall bear wholly on it, and not partly on it and partly on the metal 
 containing it. and also that efficient courses for the distribution of the lubricant 
 are provided. For heavy loads the tin compounds are the best. 
 
 There are three common methods of fitting the white metal into a setting 
 of other metal — (1) By casting it into oblong recesses ; (2) by casting it into 
 a large number of small circular recesses ; and (3) by driving strips into 
 longitudinal grooves, in the same way as the lignum vita? in a stern bush. 
 The last plan is, on the whole, the most satisfactory with large bearings, for 
 the strips are well secured, and extend over the whole length of the bearing, 
 leaving several oil courses longitudinally, and the shaft bears on the white 
 metal only; this also possesses the advantage that a damaged strip may be 
 taken out, and a new one re-fitted with ease, and without heating the brass 
 and running the risk of distorting it. Cast iron is now very often used as 
 a setting for the white metal, and answers the purpose very well indeed, 
 being much harder than brass, and thereby affording the softer metal a better 
 support. When cast iron is used the shell is generally made thicker than 
 when of brass ; and sometimes advantage is taken of this to cast the shell 
 hollow, so as to admit of water circulating through it. 
 
 The thickness of " brasses " in the crown depends principally on the 
 diameter of journal. 
 
 When of bronze .... =0*10 X diam. of journal + 0*10 m. 
 
 ,, cast iron .... =0*15 x ,, +0-15 in. 
 
 When fitted with white metal thickness = O20 X diameter -f (W5 in. 
 
 When fitted with white metal strips as follows : — 
 
 Thickness of strips . . . = O04 X diam. of journal -f- £ in. 
 
 Breadth „ . . . = 0*16 ,, + \ ,, 
 
 Space between strips . . . = thickness of strip. 
 
 Thickness of metal beyond . . = - 065xdiam. of journal if bronze. 
 
 „ 0-12 X „ „ iron. 
 
310 MANUAL OF MARINE ENGINEERING. 
 
 Columns. — The columns which support the cylinders of a vertical engine 
 are subject to alternate tensile and compressive stresses from the steam 
 pressure acting on the ends and cover of the cylinder ; to a steady compressive 
 stress from the weight of the cylinders, etc. ; and to cross-breaking stresses 
 when the ship is rolling and pitching. As a rule, columns, if designed from 
 considerations of mere strength only, would not be stiff enough for good 
 working. The same reasons, therefore, which decide the using of cast iron 
 for foundations strongly influence most engineers to choose this metal for 
 columns. The front columns are often made, however, of wrought iron or 
 steel, turned smooth (vide fig. 114) ; and all the columns for exceedingly 
 light engines are made of wrought steel, well braced together to prevent 
 vibration (vide fig. 34). Since cast iron is so superior to wrought iron 
 or steel for resisting compression, and so inferior to either for resisting 
 tension, a good composite column is formed by fitting a steel tie-bar or bars 
 through a hollow cast-iron column, the latter supporting the cylinder while 
 the former holds it down. The chief objection, however to both this com- 
 posite column and those of wrought steel for large engines is the difficulty of 
 getting good attachment to the cylinder ; and since it must be outside the 
 cylinder the columns are necessarily far apart, and away from the centre 
 line of compound engines. To avoid this difficulty wrought-steel columns 
 are formed with a flange at each end like a shaft-coupling ; but even then 
 the load on the cylinder is very much concentrated. Columns when of 
 cast iron or of cast steel (v. fig. 32) are made of various shapes, and no rule 
 can be laid down in favour of any particular form. 
 
 The columns should be so arranged as to support the cylinders and resist 
 the reaction on the foundation. Some engineers, in thoroughly effecting the 
 latter, completely hide from view the working parts, and make all the bearings, 
 etc., very inaccessible (v. fig. 38). They should be so placed at the cylinder 
 bottoms that the piston-rod axis is within the lines drawn through the 
 extreme points of back and front columns ; and when there are only two 
 columns to each cylinder, the front ones are better to be spread out some- 
 what, so as to act as struts to resist any tendency to motion of the cylinders 
 when the ship is rolling and pitching, and thereby leave the working parts 
 more open to view. When there are guides on both back and front columns, 
 or when the front columns only have the guides, then this is not possible. 
 
 Back columns are generallv of different form from the front ones, to suit 
 the guides and bearings for the pump weigh-shaft of the levers when so 
 fitted. 
 
 Some engineers have utilised the back columns as exhaust pipes to the 
 condenser, but this is not good practice, inasmuch as the heat of the steam 
 causes them to expand, and when the guides for the piston-rods are on them 
 the heat is conducted to them with prejudicial results ; this latter difficulty, 
 however, was sometimes overcome by placing the guides on the front columns. 
 It is also bad practice to expose any important part of the engine structure 
 subject to heavy strain, to unnecessary wear, such as wasting of the inner 
 surfaces of the casting by corrosion. 
 
 Guide-plates. — In order to have a sound and hard surface for the piston- 
 rod slides or shoes to work on, the guide-plates should be separate from but 
 secured to the columns ; when so fitted they also admit of adjustment, and 
 may be cast hollow, so as to permit of a flow of cooling water through them 
 
FRAMING. 341 
 
 (v. fig. 32). This is especially needful for large quick-running engines, where 
 the speed of piston is very high, and any casual want of lubrication would 
 soon cause most serious damage. Cast iron when once worn smooth pro- 
 vides a splendid surface for a slide, but if by any mischance this surface 
 suffers a little abrasion, it is most difficult to get in good working order 
 again, and cannot be relied on to work well until replaned or ground 
 smooth. 
 
 The face of the guide plates should have good oil courses cut on it, so 
 that the lubricant is well distributed, and they should be cut deep enough 
 to retain it and also prevent them being choked with the soapy deposit from 
 some kinds of oil. The piston-rod slide should always be provided with 
 a comb, which will carry the lubricant from the drip-boxes, and spread it 
 over the face of the guide at every stroke ; the plate may with advantage 
 have recesses in the upper part to catch the oil so spread, where it is 
 retained and stored to trickle down at leisure. 
 
 Framing. — Horizontal engines required a different arrangement of bed- 
 plate and framing from that of the vertical type. Trunk and return con- 
 necting-rod engines had no sole-plate proper, as the cylinders were connected 
 to the condenser casting (fig. 31) by A frames, which contain the main bearings ; 
 the trunk engine required no guides, and those for the crossheads of the 
 return connecting-rod engine were on each side of the condenser. These 
 frames had to be sufficiently strong to take the whole load from the pistons, 
 and stiff enough to be rigid under them, otherwise the crank-shaft was 
 liable to distortion. The usual form approximated to the letter A, the two 
 feet being connected to the cylinder front, one at top and one at bottom, 
 in line with the brackets on which the cylinder sat, and by which it 
 was secured to the seatings in the ship. Projecting feet were cast on the 
 cylinder front to meet those of the frames, so that the connection was 
 made with driven bolts. 
 
 Usually there were only three frames to a two-cylinder engine, and four 
 frames to a three-cylinder engine, the middle ones being very much stronger 
 than the other two, as they were required to take part of the strain of both 
 engines. 
 
 Horizontal direct-acting engines have a sole-plate, which connects the 
 cylinders to the condense casting, and contains the guides for the piston- 
 rods, and the brackets for the main-bearings ; these latter are usually stayed 
 to the cylinder tops by tie-bars through cast-iron struts or by steel tie-rods 
 only. 
 
 The framing for the older diagonal paddle-wheel engines was made some- 
 what on the same principle as that for the horizontal screw-engine, with 
 modifications (v. fig. 19) to suit the altered conditions. The main part of these 
 frames must extend from the cylinder to shaft and down again, so as to form 
 a support for the latter, and having guides for the piston-rod crossheads. 
 Intermediate supports or columns connect this main frame to the foundation. 
 It is now usual to make these frames of steel bars (vide fig. 20), and for some 
 very light draught steamers of large power frames made of steel sections 
 and plates have been found considerably lighter than the ordinary frames, 
 and can be designed to add materially to the stiffness and strength of the 
 ship in the neighbourhood of the machinery (v. fig. 22). A modified form of 
 this kind of frame is imperative, when exceedingly light draught is a necessity, 
 
342 MANUAL OF MARINE ENGINEERING. 
 
 as the hull is so light, to comply with the requirements, as to be unable by 
 itself to stand the racking strains from the engines. 
 
 Even when weight is of secondary consideration, the wrought-steel frame 
 is preferable to the cast iron, and when the cost of patterns is taken into 
 account, it is no more expensive. The bearings for the shaft, and the guide- 
 plates for crossheads, are, of course, of cast iron or cast steel fitted to the 
 wrought-steel work direct. 
 
 Side stiffness and stability are obtained by connecting the four frames 
 by wrought-iron tie-bars through the top (v. fig. 22), and by the main beam 
 before the shaft. The main-bearings for the shaft of a diagonal engine are 
 sometimes so set that the shaft can be lifted vertically ; but a better plan is 
 to set them at a slight angle, so that their centre line is in the direction of 
 the resultant of the iveight of shaft, etc., and mean pressure on the journals 
 due to the thrust on the connecting-rod. 
 
 Entablature of Oscillating and Steeple Engines. — This is usually of cast 
 iron, but may with advantage be made of wrought-steel plates and angles, 
 or, better still, of cast steel, as the strains on it from the overhung cranks 
 are severe and concentrated owing to its being supported at so few points. 
 It was no uncommon thing formerly to find it broken and patched after 
 only a few months' use, and not many work without a certain amount of 
 give-and-take, which must tend to produce rupture in course of time. 
 
 It is usually supported (fig. 27) by four columns to each crank, and addi- 
 tional stiffness and stability are imparted by diagonal cross braces to the 
 foundation at each end. It is seldom possible to place the supporting 
 columns in line with the main girders of the entablature, but when possible 
 this should always be done, so as to avoid the canting action which is caused 
 by the centre of support not being in the same plane with the centre of force 
 on the journals. When this is not possible, the sides of the entablature 
 connecting the main girders or rockers should be of extra stiffness, and well 
 connected to them by spreading out webs or fillets. Special advantage 
 should also be taken of the main beams worked into the ship, to form a 
 powerful tie to the entablature girders, and to prevent their tendency to 
 canting. This can be done generally by multiplying the number of the bolts, 
 and fitting cast-iron filling pieces in lieu of hardwood ones only. The bottoms 
 of the cross-bearers should, when possible, be tied together by bars parallel 
 with the crank-shaft to keep them from " giving " or twisting. In the case 
 of steeple engines (fig. 28) the thrust of the connecting-rod on the guides 
 caused great stress on the entablature and framing, and necessitated diagonal 
 ties between it and the foundation. The horizontal force in an oscillating 
 engine also puts cross or horizontal strains ou the entablature which the 
 ordinary columns are not fitted to resist, hence the diagonal cross frame 
 which Penn generally fitted. 
 
 To resist, as far as possible, the tendency to spring, the supporting columns 
 should be of extra size, with strong and broad Manges. 
 
 When there are four supporting columns of wrought steel, their diameter 
 in the body should be - 7 the diameter of the piston-rod. and at each end 
 0-55. 
 
 The collars or shoulders on which the entablature is supported, should be 
 equal in diameter to that of the piston-rod, and 02 the diameter of the 
 piston-rod in thickness. 
 
ENTABLATURE OF OSCILLATING AND STEEPLE ENGINES. 343 
 
 If the columns happen to come in line, or nearly so, with the centre of the 
 shaft journal, they may be 10 per cent, less in diameter than given above. 
 The breadth of the rockers should be not less than the diameter of the shaft 
 journals, and the depth at the centre should be calculated as for a box-girder, 
 subject to sudden loads applied at the middle of its length, which is measured 
 from column to column. 
 
 Roughly speaking, the depth of the rocker under the bearing brass should 
 not be less than four times the diameter of the piston-rod for engines of 
 ordinary dimensions. 
 
 The thickness of metal of sides of rockers 
 
 = 0-4 \/diameter of piston-rod. 
 
 Thickness of top and bottom = 0*6 ^diameter of piston-rod. 
 
 The bottom brasses of the main bearings should be round, so that the 
 recesses for them may tend to strengthen the rockers rather than weaken 
 them, as would be the case if square-bottomed. 
 
344 MANUAL OF MARINE ENGINEERING. 
 
 CHAPTEK XIV. 
 
 .THE CONDENSER. 
 
 The function of the condenser is to cool the steam on leaving the engine, so 
 as to reduce its pressure to a minimum ; in doing this the steam is converted 
 back into water, 'ine very early engines could only work by the aid of 
 condensation, as the steam with which they were supplied was generally 
 of a little lower pressure than the atmosphere ; it is, in fact, owing to this 
 that the steam-engine owes its birth, for steam was preferred by the early 
 mechanicians because it was so readily changed from vapour to liquid, and 
 so produced that vacuum which Nature was supposed to abhor, and to fill 
 which she would perform the work of horses. The proper relation of the 
 condenser to the engine is better understood by following the early history 
 of the steam-engine from the day when the cooling water was admitted to 
 the cylinder containing steam, and then allowed to run freely away from the 
 bottom on the descent of the piston, to the time when Watt, having per- 
 ceived the waste of energy in always forcing the piston up against the atmo- 
 spheric pressure, and in admitting the hot steam into the cold cylinder, made 
 the engine double acting, and effected the condensation in a separate chamber. 
 The jet or spray of water continued long after Watt's time as the means of 
 cooling the steam, and gave in later days the distinguishing name to the 
 condenser, which is now nearly entirely superseded by a more perfect 
 apparatus. 
 
 ' The Common or Jet Condenser, now really uncommon on shipboard, con- 
 sists essentially of an air-tight chamber, into which the steam flows from the 
 cylinder after having been exhausted of its available energy ; at its entry 
 the steam is intercepted by a spray of water, caused by the inrush from the 
 sea through small holes or narrow slits in a pipe placed across the steam- 
 way. If the spray is fine, like a shower of rain, it mixes mechanically with 
 the steam, as well as cools it by surface contact ; should the pipe have slits, 
 so as to cause the water to flow in thin broad streams like ribbons, the cooling 
 is effected principally by surface contact. The result in either case is the 
 turning of the steam into water, which falls to the bottom, and is pumped away 
 by the air-pump. It might be supposed that the mere turning of the steam 
 into water, thereby causing it to occupy far less space, will produce the 
 vacuum in the condenser ; in practice it does, but to so slight ah extent 
 and of such an evanescent nature, that unless some other means were at 
 hand, this condenser would be nearly useless. Water readily absorbs air 
 when freely exposed to the atmosphere, and gives it up again on being heated, 
 or relieved of atmospheric pressure. Fresh feed-water contains air, which 
 becomes mechanically mixed with the steam in the boiler, and passes with it 
 through its various passages, until it enters the condenser, when it parts com- 
 pany with it, and remains as cooled rarified air after the steam is converted 
 
AMOUNT OF INJECTION WATER. 345 
 
 to water. The cooling water also contains air, and readily gives most of it 
 up on becoming heated by the exhaust steam, especially under the diminished 
 pressure of the condenser. After a few strokes of the engine a sufficient 
 amount of air will be accumulated to raise the pressure to that of the atmo- 
 sphere, and although the condenser may be kept quite cool there will be 
 no vacuum, but the reverse. The pump, therefore, which exhausts the 
 condenser draws away the air as well as the water, but since the latter could 
 run away by gravity, it is only the former which is of necessity pumped ; 
 hence this pump is rightly named the air-pump. 
 
 The shape of a jet condenser is immaterial so long as the inlet for the 
 steam is high enough to prevent the water running back into the cylinder, 
 and the bottom so shaped that the water will all drain into the air-pump 
 bottom. It was generally formed to suit the ship and the working parts of 
 the engine, and was often a part of the engine framing ; the back columns 
 of vertical engines were often utilised for the purpose, and did extremely 
 well, except that occasionally rapid corrosion so weakened them as to become 
 dangerous. The frames of the horizontal engines were arranged by :;ome 
 engineers to do duty for condenser, until the Admiralty finally forbade the 
 practice. 
 
 The capacity of the jet condenser should not be less than one-fourth that 
 of the cylinder or cylinders exhausting into it, and need not be more than 
 one-half of it, unless the engine is a very quick running one ; one-third the 
 capacity is generally, however, sufficient. The objection to a large condenser, 
 beyond its cost and weight, is that a longer time is necessary to form a 
 good vacuum in it ; and the objection to a small condenser is its liability 
 to flooding and overflowing to the cylinders, unless the engineer is most 
 attentive. 
 
 The amount of injection water depends on the weight of steam to be con- 
 densed and its temperature ; the exact quantity of water per pound of steam 
 depends on the temperatures of the steam, of the cooling water, and of the 
 " hot-well," or receptacle into which the air-pump delivers the products of 
 the condenser. As the supply of water to the boilers (called the feed-water) 
 is taken from the hot-well, and it is an obvious advantage for it to be as 
 warm as possible, the cooling water used is only such as sufficient to produce 
 & good vacuum. With the jet condenser a vacuum of 24 inches was con- 
 sidered fairly good, and 25 inches as much as was possible with most con- 
 densers ; the temperature corresponding to 24 inches vacuum, or 3 lbs. pressure 
 absolute, is 140°. In actual practice the temperature in the hot-well varied 
 from 110° to 120°, and occasionally as much as 130° was maintained by a 
 careful engineer. To find the quantity of injection water per lb. of steam 
 to be condensed : 
 
 Let T x be the temperature of the steam whose latent heat is L ; T the 
 temperature of the cooling water, whose quantity in lbs. is Q ; T 2 the tem- 
 perature after condensation, or that of the hot-well. 
 
 The total heat of the steam = T, -f L. 
 
 The heat absorbed by the cooling water will be (T x -f- L) — T 2 . 
 
 But the heat absorbed by the cooling water is also represented by 
 Q (T 2 — T c ). Hence 
 
 (T 2 + L) - T 2 = Q (T - T ). 
 Or, Q = (T 1 + L)-T 2 ^(T 2 -T ). 
 
346 MANUAL OF MARINE ENGINEERING. 
 
 Now (Tj + L) — T., is equivalent to the total heat of evaporation from 
 T 2 and at T a , and is therefore equal to 966 c + 0-7 X 212° + 0-3 xT ; - T 2 . 
 
 Or, Q(T 2 -T ) = 1,114° + 0-3 x T x - To. 
 
 Therefore 
 
 Q 1,114° + 0-3 x T x -T 9 
 
 y " To - T 
 
 Example. — To find the amount of injection water required for an engine, 
 the steam at exhaust being at a pressure of 10 lbs. absolute ; the temperature 
 of the sea is 60°, and it is required to keep the hot-well at 120°. 
 
 The temperature corresponding to 10 lbs. is 193°. 
 
 n 1.114° +0-3 X 193° -120° ,--«« 
 
 Q= 120° - 60° =17-53 lbs. ^ 
 
 That is, the amount of injection water is 17 '53 times the weight of steam 
 for this particular case. 
 
 The allowance made for the injection water of engines working in the 
 temperate zone was usually 27 to 30 times the weight of steam, and for the 
 tropics 30 to 35 times, the summer temperatures being about 60° and 80° F. 
 
 The Area of injection orifice and size of pipe is governed by the head 
 of water, vacuum, and length of piping, or, in other words, by the equivalent 
 head at the condenser. 
 
 Neglecting the resistance to flow at the orifice, and in the pipes and 
 passages, the velocity at the condenser may be found as follows : — 
 
 Let h be the head of water above the valve on the condenser in feet, p 
 the pressure in the condenser, and h x the equivalent head n for gravity, 
 
 ftj = h + (15 - p) 2-3, 
 
 and velocity in feet per second = V2 g h x = 8-025 Jh x . 
 
 Example. — To find the theoretical velocity of flow into a condenser in 
 which the vacuum is 26 inches, and the orifice 12 feet below the water-line 
 of the ship. Here 
 
 h x = 12 + (15 - 2) 2-3 = 42 feet. 
 Then 
 
 velocity = 8-025 \/42 = 52 feet per second. 
 
 In practice owing to loss of head due to resistance at valves in the pipes, 
 etc., the actual velocity was only about half that given by the above rule. 
 Hence, in designing it was usual to calculate on a velocity of only 25 feet 
 per second for shallow draught steamers, and 30 for deeper ones. 
 
 From these rules, and with such allowances, the following holds good : — 
 
 Area of injection orifice in square inches = number of cubic feet of 
 injection water per minute -3- 10 - 4 to 12-5 according to circumstances, or = 
 weight of injection water in pounds per minute -f- 650 to 780. 
 
 A rough rule sometimes used was — 
 
 Allow one-fifteenth of a square inch for every cubic foot of water con- 
 densed per hour. And another 
 
 Area of injection orifice = area of piston + 250. 
 
 The injection valve was usually a simple slide or sluice valve, which 
 
SURFACE-CONDENSER, 347 
 
 readily opened or shut, and regulated the amount of water. The handle 
 or lever for working the injection valve must be very near the starting gear, 
 so that the water may be shut off as soon as the engine stops. 
 
 Snifting Valve. — It is usual to fit. at the bottom of a jet condenser, a 
 non-return valve, through which the water, etc., may run or be blown out 
 by steam ; it shuts by its own weight, and is held on its seat by the pressure 
 of the atmosphere. This is called the sniiting valve, and it allowed of the 
 condenser being emptied of water and air by steam before starting the engine ; 
 it likewise prevents the pressure in the condenser exceeding that of the 
 atmosphere to such an extent as to be dangerous in case of mishap with the 
 injection water or steam, and on that account is a useful and necessary 
 adjunct to the condensers of turbine engines, as there is no hindrance to 
 the full and direct flow of steam into it should the rotor stick. The valve 
 through which the steam is admitted is called the blow-through valve, and 
 was a simple mushroom valve, raised by means of a lever, and closed by the 
 steam pressure on the lever being released. 
 
 The snifting valve was usually exposed without a casing, so as to be 
 easily inspected or removed in case of being gagged with dirt, etc. ; to prevent 
 the water from being spurted about the engine-room, the valve was formed 
 with a curved rim, which completely covered and overhung the seat like an 
 inverted saucer. 
 
 The internal injection pipe or rose should be placed below the flow of 
 steam, so that the cooling water may pass in a shower twice through the 
 steam. 
 
 Surface-Condenser. — It has been seen that with jet condensation the con- 
 tents of the hot-well consist of a mixture of condensing and condensed water 
 in the proportion of about 30 to 1, so that the water available for feeding the 
 boiler was nearly as salt as sea water. If the cooling water is kept separate 
 from the condensed steam, the latter, which is pure water, may be used as 
 feed- water. The idea was by no means new, since so far back as 1794 Cart- 
 wright took out a patent for an engine, in which the steam was condensed 
 on the cold surfaces of two metal cylinders placed one within the other, and 
 having cold water through the inner one and about the outer. An engine 
 was made on this patent, and is said to have given great satisfaction. 
 Brunei, in 1822, took out a patent for the same object ; his invention was 
 designed more especially for ships, and consisted of groups of small tubes. 
 In 1833, L. Herbert and J. Don patented an arrangement whereby " the 
 air and steam from the eduction passage is drawn by a fan through or among 
 small tubes, so as to be condensed. The tubes may be below the water." 
 In 1835, W. Symington patented a plan " for condensing the steam from the 
 cylinder, and cooling the surplus water from the air-pump, by tubes laid 
 along the keel exposed to water outside a steam vessel." In 1838, J. B. 
 Humphreys took out a patent for " surface- condensation by leading the 
 steam through tubes in a vessel kept cold by a flow of water." The practical 
 success, however, of this mode of condensation is chiefly due to Mr. Samuel 
 Hall, with whose name the surface-condenser is generally and properly 
 associated. He took out a patent in 1831 for a system of surface-condensa- 
 tion, and in his specification claims, among other things, to condense the 
 waste steam from the safety-valves, and to distil fresh water, to make up 
 loss, by an apparatus, in principle similar to the distillers of to-day. One 
 
348 MANUAL OF MARINE ENGINEERING. 
 
 of the first ships fitted w : th Hall's condenser and appurtenances was the 
 *' Siriua," which made the first voyage under steam from England to America 
 in 1838. In 1837 engines, made for the " Wilberforce," were fitted with 
 Hall's condenser, and it is interesting to note that these engines had piston 
 valves, etc., pretty much as now used. Ignorance and prejudice drove 
 Hall's invention off the market for many years, so that it was not till 1860 
 that the surface-condenser came into general use, and then only slowly ; 
 indeed, it was not until mineral oil was used exclusively for internal lubrication 
 that objections to surface-condensation ceased. In Hall's condensers the 
 steam passed through the tubes and the circulating water outside them ; 
 had the cooling water gone through the tubes it is highly probable they would 
 not have " choked with mud and sand " and been condemned, as they were 
 in the case of the " Wilberforce " in 1841. 
 
 Condenser Tubes. — It is essential that the material on the surface of 
 which steam i3 to be condensed should be metallic, thin, and a good 
 conductor of heat, strong enough to resist the pressure of the water on 
 it, amounting to at least 15 lbs. per square inch, and capable of withstanding 
 sudden changes of temperature without fracture and the corrosive action of 
 salt water and distilled water. 
 
 The circular section being best suited to resist both internal and external 
 pressures, tubes were naturally choser as the means of separating the steam 
 from the water, and small tubes admit of a Very large amount of surface in 
 a small space. 
 
 Copper, being highly ductile and one of the best conductors of heat, was 
 at first chosen as the material from which to make the tubes. But it was 
 soon found that the acids derived from the fatty matter from the cylinders 
 dissolved some of the copper, and produced soluble salts of that metal, which 
 were pumped into the boiler with the feed-water, and there caused great 
 injury to the iron surfaces. This, for a time, gave the surface»-condenser 
 an ill repute, as it was found that the value of fuel saved by them was ex- 
 ceeded by that representing wear and tear of the boilers. This was eventually 
 obviated by having the copper tubes coated with tin, and by discontinuing 
 the use of tallow in the cylinders. But, notwithstanding this, the boilers 
 in H.M. Navy continued to show signs of premature decay, such as was 
 not customary with those receiving water from a jet-condenser. It was 
 attributed to the highly corrosive power of redistilled water on the bare 
 surface of the iron, together with the impossibility of keeping a protective 
 6cale on the surfaces when such water was used ; later research has shown also 
 that some gases which entered the boiler with the original water chemically 
 combined with bases, which kept them comparatively innocuous, were freed, 
 and came back mechanically mixed or in solution with the feed-water from 
 the hot- well, capable of highly destructive action on iron surfaces, carbonic 
 acid being a common and very corrosive one. 
 
 Copper tubes were, of course, expensive, and the tinning added to their 
 cost very considerably, thereby rendering the first cost of a surface-con- 
 denser an appreciable addition to that of the engine. In 1860 brass tubes 
 had long been used for boilers, and as this material was very ductile, and 
 its galvanic action on iron feeble, it was tried as a substitute for copper in 
 the manufacture of condenser tubes with at first mixed success, the want 
 of complete success being partly due to want of care in manufacture, and" 
 
CONDENSER TUBES. 340 
 
 partly to prejudice. The partial has since become a complete success, arid all 
 Condenser-tubes are now made of brass, but somewhat richer in copper. 
 
 The Admiralty, and consequently all foreign governments, required brass 
 tubes to be tinned when the condenser was of iron, and some of the large 
 steamship companies also adopted this practice, but, as a rule, now the tubes 
 are untinned. The tinning adds ljd. per pound extra to the cost of the 
 tubes, and amounts to an increase of 16 per cent. ; it is not necessary, but 
 is an additional safeguard against the formation of copper salts on the one 
 side, and to the corrosive action of sea-water on the other. Brass tubes, 
 untinned, after twelve years' constant use, have been found, on being cleaned, 
 to be nearly as good as when new ; on the other hand, brass tubes have 
 pitted badly, and in places been perforated in a few months from causes 
 practically unknown* It was conjectured that galvanic action had set up 
 from iron filings, carried on to the tube surface by the steam or circulating 
 water, causing a separation of the copper and zinc and the dissolving away 
 of the latter. The investigations Prof. Bengough made for the Institute 
 of Metals has thrown much light on the matter, and will eventually clear 
 up what was somewhat of a mystery. As the condensers of warships are 
 now always made of brass or steel, the tubes are untinned, but the Admiralty 
 require the addition of 1 per cent, of tin to the mixture of copper and zinc, 
 as this composition does not usually pit or corrode so readily as the ordinary 
 brass mixtures have so often done. 
 
 The loss from blowing off the boilers to prevent dangerous incrustation 
 when fed from the hot-well of a jet-condenser, amounted to as much as 
 25 per cent., and was seldom less in general practice than 12 per cent. This 
 loss is almost wholly avoided by the use of a surface-condenser, and an addi- 
 tional saving of no mean importance is effected by cessation of necessity 
 to scale and clean out the boilers as was the case when jet-condensers were 
 used. The net saving of fuel by the use of a surface-condenser averages 15 per 
 cent. ; and in the hands of a careful engineer, the economy has been found 
 to extend to even 20 per cent. 
 
 When sea-water is raised to a temperature of 280° F., which corresponds 
 to a pressure of 50 lbs. absolute, or 35 lbs. above the atmosphere, what are 
 called its insoluble salts are wholly precipitated, and these form a hard scale 
 on the hot surfaces. The principal insoluble sait in sea-water (v. Chapter xxx. } 
 is sulphate of lime ; it is called insoluble, because it does not dissolve in 
 water under ordinary circumstances, and consequently when deposited on 
 the surface of the tubes, etc., will not easily redissolve and wash off again. 
 The carbonate of lime, and the salts of soda and magnesia, are comparatively 
 harmless, for although the former is precipitated, it is only in a soft muddy 
 state, and mixed with the brine products of the latter can be blown off, and 
 easily removed from the boiler when in port. It is for this reason that a 
 surface-condenser is an imperative necessity for engines using steam above 
 35 lbs. pressure, or 50 lbs. absolute. 
 
 For the same reason it is advisable to fit a surface-condenser to steamers 
 running in muddy fresh water, as otherwise the boiler soon fills with deposit, 
 which, unless removed and the boiler thoroughly cleaned, will cause serious 
 damage, and be a source of danger and expense, as well as a constant cause 
 of priming. 
 
 It is seen, then, that by the use of a surface-condenser steam of higher 
 
 * Vide Reports of the Corrosion Committee of the Inst, of Metals, 
 
350 MANUAL OF MARINE ENGINEERING. 
 
 pressure than 50 lbs. absolute may be employed with safety ; a considerable 
 saving of fuel and time is effected ; and there is considerably less risk of 
 burning and otherwise damaging the boiler. A better vacuum is also obtained 
 in it, as a rule, than was possible with the jet-condenser. On the other hand, 
 a surface-condenser is heavier, more costly, and occupies more space than 
 the jet-condenser; a second pump for the cooling water is necessary, and 
 although the air-pump need not be so large as for a jet-condenser, it was 
 often made so in case the jet was used or the tubes leaked. ■ It was urged that 
 the wear and tear and the store account are increased somewhat when there 
 is a surface-condenser, and more care and responsibility are laid on the 
 engineers ; but these are mere trifles compared with the benefits, and are 
 only mentioned now as examples of the arguments used by the opponents 
 of Mr. Hall's system in the bitter controversy that raged in the forties of 
 last century, and prevented its general adoption. For sea-going vessels the 
 surface-condenser is indispensable, and now always fitted, notwithstanding 
 its weight and other drawbacks. 
 
 Condenser efficiency was never, as a matter of fact, seriously considered 
 until quite recent times, for the old school of engineers had other distractions, 
 and were quite content with what is now considered to be only a moderate 
 or even a low vacuum, especially were they complacent if with "it was a hot- 
 well temperature of 120° F., without enquiring very closely as to how it was 
 obtained. Moreover, so long as it was accepted as an axiom that any higher 
 vacuum could be of no service to the marine engine, there was no incentive 
 to enquire into, much less to attempt better things. The advent of the 
 turbine as a marine motor, however, together with the demand for the highest 
 speed with some ships at any reasonable cost, has changed all this, and turned 
 the apathy of the past into the activity of the present with regard to the 
 condenser. Moreover, engineers have been impressed with the apprehension 
 that while little remained to be gained at the upper part of the expansion 
 diagram, at the lower the field was large and fallow. Thoughtful minds have 
 turned, therefore, to it, and made a careful study of the theory and practice 
 involved in the working of a surface condenser. To Mr. James Weir, Mr. 
 Morison, Prof. Weighton, and others, we are much indebted for the better 
 knowledge we possess, and although these gentlemen do not always agree 
 they are all equally frank and communicative of what they discover when 
 experimenting, so that to maintain 29 inches vacuum in a modern condenser 
 in the temperate zone is easy. To the turbine such a small back pressure is 
 of the utmost value, whatever it may be to a reciprocator, and its efficiency 
 with it very high. But it must not be taken now for granted that no advantage 
 can accrue to a reciprocating engine by the rediiction in back pressure. No 
 doubt, owing to restriction in the valve and passage sections existing in 
 most of these engines, the high volume steam has to flow at such excessive 
 speeds that the difference in pressure between the cylinder and condenser 
 is necessarily greater than required between a turbine and its condenser, 
 or even what obtains when the back pressure is greater. At 29 inches of 
 vacuum, or a half-pound pressure, a pound of steam has a volume of 640 
 cubic feet, while at 26 inches it is only 172 ; for the same weight of steam 
 the velocity of flow to the condenser will, therefore, have to be nearly four 
 times at the high vacuum if the pressure difference of it is to be the 
 same as at 26 inches. But it is quite unlikely that, with a decrease in the 
 
CONDENSER EFFICIENCY. 
 
 351 
 
 pressure in the condenser, there will be none in the cylinder, whatever its 
 ports and passages be. That there is some decrease is certain, and such 
 decrease will enable a greater power to be developed ; and that the decrease 
 may be considerable may be seen on examining the diagram shown in 
 fig. 1186. Each pound of back pressure means 2\ to 3 per cent, difference 
 in the power developed by a triple- or quadruple-compound engine, and 
 such an increment as this is quite possible in a properly designed engine by 
 
 Fig. 115.— " Weir.Uniflux " Condenser for a Turbine of 11,000 S.H.P. 
 
 merely increasing the vacuum from 26 to 29 inches. Moreover, if marine 
 reciprocating engines are designed with a multiplicity of cylinders that 
 each may be kept of small size, as they are on the lines of the oil engine, so 
 that instead of one or. at most, two L.P. cylinders per engine there are 
 three or four, then the port and passage ratios may be so much larger that 
 *iigh vacuum would be of advantage to such engines. 
 
352 
 
 MANUAL OF MARINE ENGINEERING. 
 
 The effect on economy of consumption is another matter, and has to be 
 considered from quite a different standpoint. More energy will be required 
 by the air-pumps to maintain such higher vacuum ; and as the temperature 
 due to 29 inches is 80°, against the 126-8° of 26 inches, the feed water must 
 leave the condenser so much cooler ; but whether the reheating will be 
 more costly than the gain is rather a matter of experience and trial than of 
 calculation ; for so great a difference probably it is so ; but as 28 inches 
 can be much more cheaply obtained than 29 inches, and the temperature 
 difference is then only 25° against 46-8° as above, the loss, if really anything, 
 cannot be great in these days of feed heaters. 
 
 Mr. Weir attaches no importance to the question of water deposit on the 
 tubes in a surface-condenser, and thinks there is nothing to be gained, while 
 there is something lost, by fitting, collecting, and training plates among the 
 tubes, whereby it is conducted away as soon as formed without further 
 
 Kinetic Air Pump 
 Suction 
 
 -Water Pump Suction. 
 
 Fig. 116. — Cylindrical Condenser on Morison's System 
 
 contact ; whereas Mr. Morison advocates and fits them to all his condensers, 
 and claims to get thereby a distinct benefit. The former authority is, hew- 
 ever, an equally strong advocate for having warm hot-well water, and he 
 should, therefore, not object to any reasonable natural means of attaining 
 that very desirable end. With that end in view, the exposure of the water 
 when once formed to a further cooling action cannot be helpful ; on the 
 contrary, the already condensed water must lose some of its own heat, and 
 add to that of the cooling water, thereby reducing its efficiency. Further, 
 by means of such plates as Mr. Morison and others fit, the steam and air are 
 dispersed over the whole of the cooling surface without scattering the water 
 when condensed. But, on examining the design of these gentlemen's con- 
 densers, one is struck by the necessity for such arrangements in the one and 
 its absence of it in the other. Mr. Weir's condenser is of triangular cross- 
 
SURFACE CONDENSER. 353 
 
 section or heart-shaped, as an approximation to it (fig. 115). Steam enters 
 at the butt end through such an enormous orifice as to enable it to spread 
 at once over the whole breadth and length (which is generally limited) of 
 the top of the condenser, and gravitate to the contracted bottom as it cools 
 and shrinks in volume. In this case the water, as it forms, can, and does, 
 drop on to the inclined sides, on which it courses to the bottom, which is 
 not very far distant. Here there is little chance of any of the water dripping 
 on to the colder lower rows of tubes from the majority of the upper ones. 
 But the shape of the condenser is not economical of space or suitable to 
 resist unaided external pressure. On the other hand, the cylindrical 
 condenser of Mr. Morison (fig. 116) only requires inexpensive guide plates, as 
 does ako that of Mr. Morcom, which shall cause a perfect circulation of the 
 steam among the tubes and the leading away of the water when formed 
 is accomplished by the same means surely with some advantage. In the 
 case of the compact rectangular section condenser forming part of the engine, 
 and economic of room and cost, as in common mercantile practice, the Con- 
 traflo system (fig. 117), with the similar drainage scheme, is also advantageous 
 — in other words, circumstances alter cases. Fig. 117a is a diagrammatic 
 form of a Contraflo, and shows how the actual one, as in Fig. 117, is 
 virtually the same as the wedge-shaped or triangular section one. 
 
 For efficiency, a complete and rapid circulation of the steam over the 
 whole of the cooling surface is essential, as is also the concentrating of the 
 air and water into a natural flow towards the passage to the air-pump ; the 
 simplest means of effecting these two things are probably the best. 
 
 Equally important is the question of cooling-water distribution. Its 
 quantity depends on the difference in temperature between it as it enters 
 and that of the condenser where it leaves it, the weight of steam to be con- 
 densed, and the rapidity of flow. In practice condenser tubes are usually 
 | inch in external diameter in the mercantde marine, and f inch on naval 
 ships. Now, a tube f inch diameter, 18 gauge thick, and 10 feet long, has 
 a surface of 1*95 square feet, and inasmuch as 30 lbs. of steam per square 
 foot per hour may be condensed on it when clean, it will be necessary for a 
 vacuum of 28 inches that about 1,200 lbs. of cooling water should pass through 
 it per hour in winter time in the temperate zone, and 3,600 lbs. in the tropics ; 
 the flow then will be at the rate of 400 feet per minute in the latter case, 
 which is somewhat excessive, and would require considerable power to obtain, 
 as the friction head would be about 3| feet per tube length. On the other 
 hand, in the temperate zone in winter the flow would be too moderate. 
 
 This means that if a good vacuum is to be maintained in the tropics 
 the tubes must not be very long with so small a diameter, while, on the other 
 hand, if a ship is never to be where the cooling water is above 60°, the tubes 
 may be long with advantage. As, however, the rate of flow will vary as the 
 square of the diameter of tube, while its surface is directly as the diameter, 
 a small increase in diameter will admit of considerable increase in length. 
 
 In practice, the cooling water may have sufficient heat imparted to it 
 to raise it to a temperature very little below that of the condenser top before 
 leaving it. As this warming up, however, is usually done in stages — that is, 
 the water passes and repasses through the tubes on its journey from the 
 entering in to the leaving them — three stages is the most common, so that 
 in a general way the increment added to the temperature of the circulating 
 
 23 
 
Fig. 117. — Ordinary Marine Engine with Contmflo Condenser. 
 
 Fig. 117a— Diagram showing the Flow of Steam in the Contraflo System. 
 
SURFACE CONDENSER. 355 
 
 water at each pass should be only one-third of the total increase. To do this 
 the number of tubes in the lower or first group should be much less than 
 that in the second, and less there than in the third, as the efficiency owing 
 to difference in temperature between it and the condenser is decreasing at 
 each stage. This circumstance accentuates the reason for shorter tubes 
 or of larger diameter when the water is warm and a high vacuum desired. 
 
 Surface Condenser Efficiency has, however, become a matter of first-rate 
 importance ever since the introduction of the turbine on shipboard, for 
 it was there discovered how seriously the efficiency of that instrument was 
 augmented when working with the very low back pressures obtainable there 
 at comparatively cheap rates, owing to the unlimited supply of coolingi 
 water obtainable with little sacrifice of power due to the small resistance 
 " head." Mr. Parsons, and those interested in the success of the turbine, 
 turned their attention to improving the means whereby high vacuum could 
 be obtained and maintained ; Mr. Morison, ably assisted by Prof. Weight on, 
 of Durham, followed on with a series of most interesting and instructive 
 experiments, and their investigations have been the means of throwing 
 quite a flood of light on the subject, and clearly demonstrating exactly what 
 takes place in a condenser and its pumps. 
 
 The effect of Air mixed with Steam or water vapour had been noted by 
 Prof. Osborne. Keynolds, and others years before, and it was well known 
 that the presence of air was always a cause of retardation of condensation, 
 and the reduction of amount of water deposited on a cold surface in a fixed 
 time. If steam alone entered a surface-condenser, it would be wholly trans- 
 formed into water, and a vacuum corresponding to the temperature would 
 be maintained in it. The amount of steam condensed per square foot of 
 surface would be very high, so long as the tube remained clean inside and 
 outside and the cooling water supply plentiful with the flow through the 
 tubes rapid. The only pump necessary under these circumstances would be 
 one to withdraw the condensed water as it falls to the bottom. But, as a 
 matter of fact, it is impossible in practice to work with a circuit so com- 
 pletely closed that no air gets into the system when once the water put into 
 the boilers has been deprived of the air it originally held in solution, for 
 every condenser of a marine engine contains more or less air always, and, 
 therefore, an air-pump is necessary to it, in order to maintain any sort of 
 steady vacuum ; and if the vacuum is to be high, the air-pump must be efficient 
 as well as sufficiently large ; for, however large it may be, if it is not efficient 
 no high vacuum at all is possible. 
 
 Air Leaks to the Condenser may have originated at the glands of the 
 main and auxiliary engines when the pressure inside them is less than that 
 of the atmosphere. Much air comes from the feed-water, which, if taken 
 from storage tanks, will contain from 2J to 4 per cent, by volume of air. Even 
 the water of the hot-well is charged with a considerable amount, and this 
 Mr. Weir endeavours to eliminate in his feed-heaters before it can enter the 
 boiler. Auxiliary feed-water is subject to the same action in the feed-heaters. 
 Leakage may, and does, arise at times from faulty or damaged jointings of 
 the condenser and connections ; these leaks, however, should be found 
 out and stopped ; for this purpose a periodical examination is made by 
 some careful engineers in charge of turbines ; those with reciprocators should 
 do the same. The leakage from auxiliary machinery, however, is often a 
 
356 
 
 MANUAL OP MARINE ENGINEERING*. 
 
 most serious cause of loss, and is said to amount to as much as the cost of 
 the whole power developed in them. No doubt the Admiralty are wise to 
 have separate condensers for the auxiliary machinery, and in large merchant 
 steamships the same practice might be followed with considerable advantage ; 
 even if such condensers were simplified so that the water circulation was 
 " natural," and the condensed water allowed to drain into a tank by gravity, 
 it is better to have them rather than to run the risk of spoiling the main 
 condenser efficiency by admitting aerated auxiliary exhaust to it. 
 
 Beveporismg Chei.iber 
 
 Feed Temperature 
 Regulating Chamber 
 
 Air Pump 
 * Suction 
 
 Feed Temperature Regulating Valve 
 
 Fig. 118. — Gontraflo ©ondenser with Feed Temperature Regulatof. 
 
 Air is heavier than Water Vapour, so that, if the exhaust steam contains 
 but a very small quantity, it will accumulate at the bottom of the condenser, 
 and unless removed it practically blankets the cooling surface there, and 
 virtually reduces the size of the condenser. Even when the air-pump draws 
 it away, if only at the same rate as it flows in, there may be a portion of the 
 surface constantly surrounded by air and effecting no condensation. It is- 
 
MORISON S DIAGRAM. 
 
 357 
 
 necessary, therefore, that the air-pump shall be large enough to keep the 
 inside of the condenser free from air lodgment. 
 
 It follows, then, that, in addition to the large efficient pump, means 
 shall be provided that the flow to it will be as direct as possible ; and at the 
 
 
 
 
 
 
 
 
 
 
 
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 4-5 50* 60° 70 SO 90 100 HO 120 130 140 
 
 Temperature in Degrees Fahr. 
 
 fig. 118a. — Morieon's Diagiam. Ai» Saturated with Water Vapou*. 
 Carves showing relation between vacuum, temperature, volume, and ratio of air to vapour. 
 
 same time that every part of the cold surface should be active ; hence there 
 must be no pockets or eddies anywhere, and if the design of condenser doea 
 
358 
 
 MANUAL OF MARINE ENGINEERING. 
 
 not ensure thi« constant and general flow, the vapour must be directed by* 
 guide plates and baffles. Mr. Morison has given great attention to this, 
 and all else that pertains to the surface-condenser, and been good enough to 
 impart his knowledge in the papers he has read at the meetings of the Insti- 
 tution of Naval Architects, and of the N.E. Coast Institution of Engineers 
 and Shipbuilders, the transactions of which may be studied with advantage. 
 The Flow of Vapour and Water in its passage through the condenser 
 from the exhaust to the air-pump suction must be at all times positive, but 
 separate. When water is once formed, it should not be kept in contact 
 with the cooling surface, or allowed even to touch it again, but at once take 
 its passage to the drain. The vapour should be made to pass or repas3 
 over the whole cooling surface, so that no part of it is inactive. In modern 
 condensers the continuous contact is done by repassing largely in order 
 to economise the space taken by the condenser, which makes it virtually 
 the same as a long one of wedge shape, transverse section, like Mr. Weir's 
 (fig. 117), having the butt end next the cylinder, and the drain to the air-pump 
 at the thin end. In this way the air is concentrated before entering the 
 pump, and is moreover cooled as much as is possible, so that the air-pump 
 
 Top 
 
 K 
 
 A 
 
 4~ 
 
 
 -J 
 
 s. 
 
 Mean Press 3-5 Lbs. 
 Scale fjf 
 
 >C 
 
 \ 
 
 Steam, .. 
 
 Vacuum, 
 
 Revolutions, 
 
 176 His. per square inch. 
 28} inches. 
 66 pur minute. 
 
 High-pressure engine, 
 Intermediate-pressure engine, 
 Low-pressure engine 
 
 6S0 I. If. P. 
 6!)7 ,, 
 
 7o5 ,, 
 
 Total, .. 2,035 „ 
 
 Fig. 1186. — L P. Indicator Diagram showing High Vacuum. S.S. " Xigaristan." 
 
 (Richardson and Westgarth.) 
 
 can be kept cool and fit to produce in its chamber a very high degree of 
 vacuum. For this purpose the lowest part of the condenser is often arranged 
 so that the lowest rows of tubes through which the coldest water passes are 
 " drowned, "so that the water which surrounds them is cooled down below 
 the temperature of the bulk of the condensed water. 
 
 Fig. 118a, Morison's diagram showing the relation between vacuum, 
 temperature, volume, and ratio of air to vapour, is a most interesting and 
 instructive one, and may be referred to with advantage when considering 
 condenser problems. 
 
 Cooling Surface. — Mr. Weir has shown that, with efficient air-pumps, 
 as much as 16 lbs. of water can be condensed on a square foot of cooling 
 
COOLING SURFACE. 359 
 
 surface per hour in a vacuum of 26*75 inches when the sea-water is as hot 
 as 80° F., and the hot-well temperature as high as 106°. With sea-water 
 at 54° F., and the vacuum 28 inches, as much as 28 - 6 lbs. were condensed per 
 square foot per hour. He also demonstrates how as much as 35 lbs. can be 
 condensed with the vacuum at 27*3 inches, with the cooling water at the 
 same temperature, but with the hot- well temperature reduced to 101°, or 
 nearly that in the condenser. 
 
 It would seem, then, from these facts that 1 square foot of cooling surface 
 per I.H.P. is sufficient for any ship, so long as the condenser is fairly clean ; 
 and, further, it is evident that h square foot per I.H.P. is ample allowance 
 for such ships as work in temperate zones with cooling water under 60°. 
 Moreover, as most ships on service run with less power than developed on 
 trial, the allowance for trial conditions practically provides a margin for 
 loss of efficiency of surface afterwards. Experience with turbine-driven 
 ships, having a vacuum in their condensers of 28*5 to 29*0 inches, has estab- 
 lished the fact that with cooling water at 50° F. and the hot-well kept at 
 80° to 85° F., as much as 25 lbs. of steam can be condensed per square foot 
 of surface per hour. 
 
 The Allowances of Cooling Surface may be computed in the following 
 way : — Ships trading to all parts of the world, and, therefore, sometimes in 
 the tropics, should have 1 square foot of cooling surface for each 16 lbs. of 
 steam condensed. Turbine steamers working under the same conditions 
 should, to maintain high efficiency, have 1 square foot for every 12 lbs. 
 
 Ships limited to service in temperate zones, or whose service in the 
 tropics is short, or when there high efficiency is not of consequence, 25 lbs. 
 for reciprocators and 20 lbs. for turbines is not too large an allowance. 
 The temperature of the condenser with a vacmim of 29 inches is 80° F. ; 
 it is obvious, therefore, that, with sea-water at this or a somewhat higher 
 temperature, such a vacuum is impossible, whatever be the surface, and 
 even with 28 inches the amount of cooling water required would be so great 
 as to be almost prohibitive in commercial practice. * 
 
 The temperature of the condenser is somewhat higher near the entering 
 in of the steam than at the bottom when the engine is working at full power, 
 but at reduced power the difference is very trifling ; at full speed, therefore, 
 the circulating water may be nearer the temperature at bottom of condenser 
 than at low powers. The water should, therefore, for this and other reasons, 
 always flow in the opposite course to that of the steam — that is, the coldest 
 water should be where the steam is coldest, and where the steam is entering 
 the condenser, and, therefore, at its hottest the heated cooling water may still 
 take up more heat from it, leaving it with heat still in it for abstraction below. 
 
 The Cooling Surface per Horse-power may be as follows : — 
 
 Triple-compound express steamers, 080 sq. ft. home waters, 125 sq. ft. tropics. 
 
 „ „ economic „ 070 ,, „ 1"06 „ ,, 
 
 Quadruple „ „ „ 65 „ „ 95 „ „ 
 
 Turbine-driven ordinary „ 0'65 „ „ 110 „ „ 
 
 ,, express „ 0"80 „ „ 1'25 „ „ 
 
 For ordinary cargo steamers going to all parts of the world 1 square 
 foot of surface per trial trip I.H.P. is sufficient, as the temperature of sea- 
 water does not exceed 85° F., and is not often over 80°. This allowance is 
 sufficient to permit of good vadium being maintained on service conditions 
 
 • At 80° F. the cooling water must be 63 times the weight of steam ; and at 85° no less than 153. 
 
360 MANUAL OF MARINE ENGINEERING. 
 
 with the surface somewhat foul. Destroyers with reciprocating engines on 
 trial in summer time have condensed 28 lbs. of steam per square foot, and 
 the allowance per I.H.P. was only \ square foot, and the vacuum 25 to 26 
 inches at full power. 
 
 Professor Weighton, with a Contrarlo condenser, got 33 lbs. condensed 
 per square foot from a triple-compound engine: and even 40 lbs. have been 
 got by other experimenters. 
 
 Cooling Surface. — Professor Rankine suggested the following as a means 
 of ascertaining the amount of cooling surface : — Let t denote the temperature 
 of a film of liquid at one side of a metal plate, S the extent of cooling surface ; 
 let heat be communicated to the liquid at a temperature t by some such 
 process as the condensation of steam, and let that be abstracted by the 
 flow of a current of air, water, or other fluid, in contact with the metal plate ; 
 the weight of fluid which flows past per second being W, its specific heat 
 C, its initial temperature T T , being lower than t, but higher than T 2 . Then 
 in all the equations t — T^ is to be substituted for T 1 — t, and t — T 2 for 
 
 T 2 — t in the equation — ^ = a \ pp ™ f ; but he also added 
 
 that there are not sufficient data to obtain the value of the constants. 
 
 Another Basis for Calculating Cooling Surface is obtained by assuming 
 that the maximum mean flow of cooling water permissible is 400 feet per 
 minute, and the units of heat to be abstracted being that latent at the different 
 pressures, and the cooling water in the tropics 80° F., and in home waters 
 60°. The diameter of the condenser tubes is taken aa f inch for naval practice, 
 | inch in general mercantile, and 1 inch in exceptional oases. 
 
 The weight of water passing per minute through a f-inch tube at this 
 rate will be 44 lbs., through a f-incb 60 lbs., and through a tube 1 inch in 
 diameter 107 lbs. The temperature of the water at discharge from the 
 condenser will be taken at about 2 per cent, below that of the hot-well water, 
 or that of the condenser, so that for a vacuum of 28 inches it will be 0-98° X 
 90 - 4 c = 88-6°. Under these circumstances, in the tropics the heat abstracted 
 will be 88-6 — 80, or 8-6°. The latent heat with this vacuum will be 1,051° ; 
 
 hence the weight of cooling water per lb. of steam = ' = 122 lbs. : a 
 
 44 
 
 tube f inch in external diameter will condense, therefore, in an hour 60 X ^ 
 
 or 21-6 lbs. 122 
 
 The greatest length of tube through which the water flows should not 
 exceed 400 diameters, and, therefore, if the water passes three times, as is 
 common practice, the f-inch tube should not be in the aggregate more than 
 (f X W")' or 21 feet, and each tube is consequently only 7 feet long, or if 
 only twice through 10*5 feet. 
 
 Now, 21 feet of f-inch tube has a surface of 3-437 square feet. It follows, 
 then, that the cooling surface in a condenser having f-inch diameter tubes 
 for service in the tropics when 28 inches vacuum is required must be at the 
 rate of 1 square foot for each, 21*6 -=- 3-437, or 6-3 lbs. of steam to be condensed. 
 Taking the consumption of steam in a turbine steamer at 12*5 lbs. per 
 S.H.P., then- 
 Cooling surface of tubes f inch in diameter per S.H.P. for tropics 28 inches 
 vacuum is 2 square feet. 
 
COOLING SURFACE. 361 
 
 If sea-water is 60°, there will be then the following, viz. : — 
 Heat abstracted 88-6 — 60, or 28-6°. 
 
 Weight of cooling water = 1,051 ^ 28*6, or 36-7 lbs. per lb. of steam. 
 Water condensed per hour = 60 X 44 -s- 36-7, or 72 lbs. 
 Rate of condensation per square foot per hour = 72 -f- 3*437, or 21 lbs. 
 For home service, therefore, 12-5 -4- 21, or 0*596 square foot per S.H.P., 
 is sufficient for f -inch diameter tubes. 
 
 For the Mercantile Marine with f -inch tubes and a 28 inches vacuum, the 
 following will hold good, viz. : — 
 
 Heat abstracted in the tropics, as before, 8*6 units. 
 
 Water passed per hour, 60 X 60, or 3,600 lbs. 
 
 Weight of cooling water per lb. of steam, 1,051 -f- 8*6 ; or 122 lbs. 
 
 Weight of steam condensed per hour = 3,600 -=- 122, or 29-5 lbs. 
 
 Maximum length of tubes, f X ^j^ , or 25 feet. 
 
 Each tube being 8*33 feet long if three times through, or 12-5 if twice. 
 
 The surface of 25 feet of -f-inch diameter tube is 4*91 square feet. 
 
 Quantity of water condensed per square foot is then 29*5 -=- 4-91, or 6 lbs. 
 
 If, however, the condenser were made the same length as that with f-inch 
 tubes, the surface would be as before 3-437 square feet, and the quantity 
 of water condensed per square foot 29*5 -s- 3*437, or 8*6 lbs. under these 
 circumstances. 
 
 The cooling surface per S.H.P. = 12*5 -*- 8*6, or 1*45 square feet. 
 
 If the condenser tubes are made 1 inch in external diameter, the water 
 passed per hour is 60 X 107, or 6,420 lbs. ; weight of cooling water as before, 
 122 lbs. 
 
 Steam condensed per hour. 6.420 -5- 122, or 52*6. 
 
 Maximum length of tube = 1 X -^°-, or 33*3 feet. 
 
 So that in this case each tube may be 11*0 feet long. 
 
 The surface of 33 feet of 1-inch tube is 8*64 feet. 
 
 52*6 
 Quantity of steam condensed per square foot = ^i, or 6*1 lbs. 
 
 If the cooling water has a temperature of 60°, the number of heat units 
 abstracted by each pound will be as before 88*6 — 60, or 28*6°. 
 
 The weight.of water for eaeh pound of steam, ? , or 36*7 lbs. 
 
 28"6 
 
 Steam condensed = 6,420 -f- 36*7, or 175 lbs. 
 
 Steam condensed per square foot = 175 -f- 8-64, or 20*3 lbs. 
 
 ' From the above it will be seen that if the maximum combined length 
 of 400 diameters of tube is followed as the rule for condensers, that 
 a vacuum of 28 inches should be maintained with a rate of condensation 
 of about 6 lbs. per square foot of surface when in the tropics, and 20 lbs. 
 in the temperate zone. For the steam consumption per horse-power this 
 would indicate that the cooling surface for tropical work must be at least 
 three times that for cool climates when high vacua arc required and necessary 
 as with turbines. It has been, however, pointed out that under tropical 
 -conditions much more than 6 lbs. of steam per square foot of surface can 
 be condensed. It follows, then, that the length of tube must be reduced, 
 and the combined length be inversely as that quantity is to 6. 
 
362 
 
 MANUAL OF MARINE ENGINEERING. 
 
 That is, if Q be the quantity in pounds, then — 
 
 fi tt 
 
 Combined length = 400 X^X vq- 
 
 d being diameter of tube in inches. 
 
 If, however, a vacuum of 26 inches is sufficient, as is the case with 
 reciprocating engines in the -tropics, then the temperature of condenser 
 will be 120°, and the latent heat 1,030°. 
 
 The number of units abstracted per lb. of cooling water will be 120 — 80, 
 or 40. 
 
 Weight of water for each pound of steam, 1,030 4- 40, or 25*75. 
 
 Taking the tubes f inch in diameter — 
 
 The steam condensed = 3,600 -*- 25-75, or 140 lbs. 
 
 140 
 The steam condensed per square foot surface = r^-r = 28 - 5. 
 
 It is evident from the rates observed by Mr. Weir and others that the high 
 
 vacua are obtained in the ordinary ship's condenser ; it follows, then, that 
 
 either the tubes were short or the rate of flow through them much higher 
 
 than 400 feet per minute or both. The difference in temperature has been 
 
 taken here at only 8 - 6° F. for the tropics, which is, of course, very low. But 
 
 if long tubes are fitted to a condenser the rate of flow must be very high to 
 
 maintain high vacua, and the allowance of surface more than shown above. 
 
 If, therefore, 28 inches is to be maintained with T25 square feet of cooling 
 
 surface per S.H.P., the flow of water through the condenser, whose tubes 
 
 400 x 2 
 are 400 diameters in length, at a rate per minute = — =-^ — , or 640 feet, 
 
 which is high for any tube, but especially so for those only f inch in diameter. 
 The tl head " to overcome the mere resistance of 21 feet of such tube will be 
 at a velocity of 640 feet per minute, probably as much as 45 feet, while at 
 400 feet it will be no more than 16 4 feet. 
 
 With tubes § inch in external diameter and 25 feet aggregate length the 
 "head" to overcome resistance will be 15*7 with a velocity of 400 feet, and 
 35-7 feet if the velocity is raised to 600, while with tubes of 1 inch external 
 diameter, and an aggregate length of 33 feet, the resistance at 600 feet per 
 minute flow will be only 14 feet " head." 
 
 TABLE XL. — Effect of Vacuum on Steam Consumption in Lbs. per 
 I.H.P. in a Turbine and Quadruple-expansion Engine 
 (from Prof. Weighton's Observations), and 
 
 in a Triple-expansion Engine (200 I.H.P.) and 1,000 K.W. Turbine. 
 
 (from Sir C. Parsons' Observations.) 
 
 Vacuum, . . Ins. 
 
 20 
 
 22 
 
 24 
 
 26 
 
 27 
 
 28 
 
 29 
 
 Turbine, Lbs. per H.P. 
 Quadruple, ,, 
 
 19-2 
 16-7 
 
 18-1 
 16-1 
 
 16-9 
 15-5 
 
 L5-6 
 
 150 
 
 14-8 
 14-7 
 
 13-9 
 14-5 
 
 130 
 14-3 
 
 Triple-comp. eng. „ 
 Turbine „ 
 
 14-8 
 19-3 
 
 14-35 
 18-1 
 
 14-0.") 
 16-9 
 
 13-90 
 15-6 
 
 13-8 
 14-9 
 
 13-77 
 14-0 
 
 13-76 
 130 
 
CONDENSER TUBES. 
 
 363 
 
 TABLE XLI. — Temperature, Latent Heat, and Volume of 
 Steam of very Low Pressure. 
 
 Pressure. 
 
 
 
 
 Pressure. 
 
 
 
 
 
 
 Tempera- 
 ture F ». 
 
 Latent 
 
 Volume of 
 
 
 
 Tempera- 
 ture F°. 
 
 Latent 
 
 Volume of 
 
 Lbs. 
 
 
 Heat 
 
 F°. 
 
 1 Lb. of 
 Steam. 
 
 Lbs. 
 
 
 Heat 
 F°. 
 
 1 Lb. of 
 Steam. 
 
 Absolute 
 
 Vacuum. 
 
 
 
 
 Absolute 
 
 Vacuum. 
 
 
 
 
 
 
 
 
 Cub. Ft. 
 
 
 
 
 
 Cub. Ft. 
 
 0-3, 
 
 29-4 
 
 67-5 
 
 1,070 
 
 1,067 
 
 1-7, 
 
 26-6 
 
 120-3 
 
 1,029 
 
 200 
 
 0-4, 
 
 29-2 
 
 740 
 
 1,063 
 
 800 
 
 1-8, 
 
 26-4 
 
 122-4 
 
 1,028 
 
 190 
 
 0-5, 
 
 29-0 
 
 80-0 
 
 1,058 
 
 640 
 
 1-9, 
 
 26-2 
 
 124-6 
 
 1,026 
 
 181 
 
 0-6, 
 
 28-8 
 
 85-5 
 
 1,054 
 
 535 
 
 2-0, 
 
 26-0 
 
 126-7 
 
 1,025 
 
 172 
 
 0-7, 
 
 28-6 
 
 90-4 
 
 1,051 
 
 461 
 
 2-1, 
 
 25-8 
 
 128-6 
 
 1,024 
 
 165 
 
 0-8, 
 
 28-4 
 
 94-5 
 
 1,048 
 
 410 
 
 2-2, 
 
 25-6 
 
 130-4 
 
 1,022 
 
 158 
 
 0-9, 
 
 28-2 
 
 98-5 
 
 1,045 
 
 367 
 
 2-3, 
 
 25-4 
 
 132-2 
 
 1,021 
 
 152 
 
 i-o, 
 
 28-0 
 
 102-0 
 
 1,042 
 
 333 
 
 2-4, 
 
 25-2 
 
 1340 
 
 1,020 
 
 146 
 
 11, 
 
 27-8 
 
 105-0 
 
 1,040 
 
 306 
 
 2-5, 
 
 25-0 
 
 135-6 
 
 1,019 
 
 140 
 
 1-2, 
 
 27-6 
 
 108-0 
 
 1,038 
 
 282 
 
 2-6, 
 
 24-8 
 
 136-9 
 
 1,018 
 
 135 
 
 1-3, 
 
 27-4 
 
 1110 
 
 1,036 
 
 260 
 
 2-7, 
 
 24-6 
 
 138-2 
 
 1,017 
 
 130 
 
 1-4, 
 
 27-2 
 
 113-7 
 
 1,034 
 
 240 
 
 2-8, 
 
 24-4 
 
 139-6 
 
 1,016 
 
 125 
 
 1-5, 
 
 270 
 
 1160 
 
 1,033 
 
 225 
 
 2-9, 
 
 24-2 
 
 141-0 
 
 1,015 
 
 121 
 
 1-6, 
 
 26-8 
 
 118-2 
 
 1,031 
 
 212 
 
 3-0, 
 
 240 
 
 142-2 
 
 1,014 
 
 118 
 
 Condenser Tubes. — They are. as a rule, made of brass, solid drawn, and 
 tested both by hydraulic pressure and steam ; the latter test is a very useful 
 one, as faults which escape detection under water pressure are often found 
 out by steam ; these faults are due generally to minute particles of flux or 
 slag in the original ingot, and sometimes the faults are in the form of cracks 
 done in the process of drawing. It is by no means an uncommon thing to 
 find a few tubes in a new condenser leaking through minute holes of various 
 shapes ; these holes soon become enlarged if the tube is not at once stopped 
 or withdrawn. These faults are not confined to the tubes of a few makers, 
 but may be found in those of all makers at some time or other. Tinning is, 
 as a rule, a preventive, as the defective places are in the process covered 
 or filled with that metal, but it is seldom resorted to now. The Admiralty 
 method of adding a small amount of tin to the alloy of copper and zinc has 
 proved a good preventive of corrosion, and Mr. Philip, Admiralty Chemist, has 
 shown by statistics its efficacy in a large number of condensers in H.M. Navy. 
 
 Condenser tubes were usually made of a composition of 68 per cent, of best 
 selected copper, and 32 per cent, of best Silesian spelter* The Admiralty, 
 however, always specify the tubes to be made of 70 per cent, of best selected 
 copper, and to have 1 per cent, of tin in the composition, and test them 
 to a pressure of 300 lbs. per square inch. To prove that the tubes are of the 
 70/30 alloy, a few pounds of them are melted in a closed crucible, and suffi- 
 cient spelter added to bring the mixture to contain 62 per cent, of copper. 
 The metal is then cast into an ingot, which when cold is rolled into a sheet, 
 strips are cut from it and tested, and if satisfactory should have an ultimate 
 tensile strength of 24 tons per square inch. In the mercantile marine the 
 tubes are, as a rule, J inch diameter externally, and 18 L.S.G. thick (0-049 
 inch) ; and 16 L.S.G. (0-065), under some exceptional circumstances. In 
 H.M. Navy, the tubes used to be 18 to 19 L.S.G. thick, tinned on both sides ; 
 
 * The tubes of the mercantile marine are usually 
 cent, of lead has been tried with success. 
 
 70/30 mixture without additions, but latterly 2 per 
 
364 MANUAL OF MARINE ENGINEERING. 
 
 but now the Admiralty do not require the tubes to be tinned. On account 
 of the economy of space and weight that is effected with small tubes all 
 Naval engines are now fitted with condensers having tubes f inch diameter. 
 The smaller the tubes, the larger is the surface which can be got in a 
 ■certain space. Since larger tubes are of necessity somewhat thicker than 
 the smaller ones, a square foot of surface costs more when they are adopted, 
 -and is not so efficient. Patent tubes made from sheet brass 22 B.W.G. 
 thick, and joined at the seams like a tinsmith's joint and soft soldered, have 
 been tried. The advantage claimed for them is the uniformity, whereby as 
 little as 22 B.W.G. is sufficient .thickness, while it would not be safe to use 
 tlrawn tubes so thin. 
 
 The length of the tube depends on the arrangement of the condenser, but 
 when they are not held tightly in the plates, but only packed, their unsup- 
 ported length should not exceed 100 diameters ; when held with tight-fitting 
 ierrules it may be 120 diameters. 
 
 Tube-plates are now always made of brass, either cast or rolled into plates 
 of suitable size ; the latter is preferable, as the rolled brass is very tough and 
 close grained, and as strong as wrought iron. Formerly it was no uncommon 
 thing to make the tube-plates of cast iron from If to 2h inches thick, and 
 while some were converted into a substance resembling plumbago after two 
 or three years' work, others have been found sound and good after twelve 
 years' continuous work. 
 
 Rolled brass tube-plates should be from T3 to 1*5 times the diameter of 
 tubes in thickness, depending on the method of packing. When the packings 
 go completely through the plates the latter, but when only partly through, 
 the former is sufficient. Hence, for f-incb. tubes the plates are usually f to 
 1 inch thick with glands and tape-packings, and 1 to 1£ inches thick with 
 wooden ferrules. In the Navy the tube-plates are generally 1 to f inch 
 thick, the tubes being £ inch diameter and 19 L.S.G. thick, but in the 
 " Destroyers " the plates are only £ inch thick ; in their case, however, the 
 plates are of small diameter and stayed in the middle ; and it may be added 
 that leakage of tubes was no uncommon occurrence in such ships, so that 
 this is as thin as they can safely be employed. 
 
 The tube-plates should be secured to their seatings by brass studs and 
 nuts, or brass screw-bolts ; in fact, there must be no wrought iron of any 
 kind on the sea-water side of a condenser. When the tube-plates are of 
 large area it is advisable to stay them by brass rods, to prevent them from 
 bulging or collapsing. 
 
 Tube Packings. — All attempts to drift or expand the tubes tightly into 
 holes in a brass plate fail, owing to the softness of both plates and tubes ; and 
 if it could be done it would be found impossible to draw the tubes for examina- 
 tion and cleaning without damage. Fig. 119 shows a very simple plan, and 
 one that proved effective under all circumstances, and essential with cast-iron 
 plates. The ferrule is made of soft wood, such as pine or lime tree, very dry 
 and well seasoned ; they were made nearly an eighth of an inch larger in 
 diameter than the hole into which they had to fit, and a good fit on the tube. 
 Before fitting them into place they were squeezed through a die in a press 
 until they could be easily driven into their holes ; soon after being fitted into 
 place they absorb moisture and expand circumferentially at each end, and 
 become exceedingly tight on the tube and in the hole. After twelve years' 
 
TUBE PACKINGS AND FERRULES. 
 
 365 
 
 service they were found quite sound. It is urged against them that they are 
 apt to shrink and drop out when the condenser is not in use, but this is not 
 the case, as the swelled projecting ends form collars to prevent this, and- 
 they do not shrink so much as is generally supposed, unless by unusual heat. 
 This is one of the cheapest forms of tube-packing, and although not used 
 now, was often employed in the mercantile marine of this and other countries. 
 The plan adopted in H.M. Navy, and generally in the mercantile marine, 
 is that shown in fig. 120. Each tube end passes through a stuffing-box fitted 
 
 Fig. 119. 
 
 Fig. 120. 
 
 Fig. 121. 
 Figs. 119-121. — Condenser Tube Packings and Ferrules. 
 
 with a screwed gland, and kept tight by a tape washer, or some soft cordv 
 as packing. This method is somewhat expensive, but it admits of the 
 water being on either side of the tubes ; the packing is not affected by heat, 
 and the condenser may remain unused for a very long time, and be quite 
 tight at the end of it ; for these reasons it was chosen by the Admiralty. 
 It is likewise the plan used by Hall in his early surface-condensers, and on. 
 
366 MANUAL OF MARINE ENGINEERING. 
 
 the whole the most satisfactory one. Modern methods of drilling and tapping 
 the holes, making the ferrules, etc., have very much reduced the cost, and 
 so practically removed the only objection. Fig. 121 shows a modification 
 of gland specially to suit vertical tubes ; the gland has an inside rim, which 
 prevents the tube from slipping. It is now always used in condensers with 
 horizontal tubes, and is especially necessary when they are long to prevent 
 creeping. 
 
 Steam Side and Water Side of Tubes. — This was somewhat of a vexed 
 question, about which there is much to be said on both sides. The naval 
 practice was formerly to circulate the water outside the tubes, so that the 
 condenser shell may be kept cool and prevented from making the engine- 
 room hotter than can be helped. The almost universal practice of the 
 merchant service and Navy now is to circulate the water through the tubes. 
 Independently of the particular reason for the choice of the Admiralty, the 
 balance of argument is in favour of circulating the water through the tubes ; 
 for when this is the case there is (i.) a larger surface of metal exposed to the 
 hot steam ; (ii.) the grease that may be deposited on the tubes is easily 
 removed by working a trifle warm, and using a solution of caustic soda or 
 potash, and if this does not remove it, the deposit being soft does not prevent 
 the tubes from being easily drawn, as is the case when scale from salt-water is 
 deposited on their exterior surface ; (iii.) the scale from sea-water, which must 
 be removed mechanically, can be so done without removing the tubes ; 
 (iv.) a more extended and complete circulation of the cooling water is possible, 
 and that without risk of air accumulation and without special and expensive 
 diaphragms, etc. ; (v.) the condenser is more easily designed, and fits into the 
 general arrangement of most engines, and is smaller, inasmuch as there is no 
 need of an expansion chamber in front of the tubes, as is the case when 
 steam passes through them ; (vi.) when it is necessary to examine the packings, 
 or to plug a defective tube, only a water joint is broken ; and (vii.) the thin 
 tubes are stronger to resist internal than external pressures, especially of brass. 
 On the other hand, when the steam passes through the tubes the bulk of 
 the fatty matter is deposited on the front tube-plate, and prevented from 
 coating the tubes ; the large flat sides of the condenser are subject to the 
 very slight pressure due to the head of water, and so may be made much 
 lighter ; and there is less hot surface exposed in the engine-room. As little 
 or no oil is now used for direct internal lubrication, and such as gets in with 
 the piston and valve- rods is small, and a mineral oil, the argument as regards 
 grease deposit is considerably modified from what it was when animal and 
 vegetable oils were freely used, and even tallow on occasion, in both cylinders 
 and boilers. 
 
 Spacing of Tubes, etc. — The holes for ferrules or glands are usually \ inch 
 larger in diameter than the tubes ; but when absolutely necessary the wood 
 ferrules may be only J^ inch thick. 
 
 The pitch of tubes when packed with wood ferrules is usually J inch more 
 than the diameter of the ferrule hole. For example, the tube beinsf f inch 
 external diameter, the ferrule hole will be 1 inch and the pitch of the tubes 
 1 J inch. In the Navy, with tubes f inch diameter and the tube-plates | inch 
 to 1 inch thick, the pitch of holes is 1] inch, and in "Destroyers" only 
 }f inch. In the mercantile marine, with tubes f inch diameter and the tube- 
 plates 1 inch thick, the pitch is 1| inch; for |-inch tubes, 1^- inch; and 
 
CONSTRUCTION OF SURFACE CONDENSER. 
 
 367 
 
 for 1-inch tubes, 1-^ inch. The tubes are generally arranged zigzag, and 
 the number which may be fitted into a square foot of plate is as in Table xlii. 
 
 
 
 TABLE XLII. 
 
 
 
 Pitch of Tubes. 
 
 No. in a 
 square foot. 
 
 1 
 
 Pitch of Tubes. 
 
 No. in a 
 square foot. 
 
 Pitch of Tubes. 
 
 No. in a 
 square foot. 
 
 Inches. 
 
 
 Inches. 
 
 
 Inches. 
 
 
 31 
 15 
 
 184 
 
 
 ... 
 
 ... 
 
 
 1 
 
 172 
 
 1 : ' 
 
 128 
 
 h 
 
 lib 
 
 1A 
 
 161 
 
 
 ... 
 
 ... 
 
 
 1A 
 
 152 
 
 1A 
 
 122 
 
 1 ■• 
 
 105 
 
 1 3 
 
 144 
 
 
 
 ... 
 
 
 H 
 
 136 
 
 i£ 
 
 lie 
 
 1 » 
 
 99 
 
 The Body of the Condenser. — The surface-condenser was generally in the 
 form of a cylinder or a rectangular parallelepiped, and sometimes a flattened 
 cylinder ; the first and last forms are the best suited when weight is a great 
 consideration, and the second and most convenient when space is of first 
 importance ; the cylindrical form is by far the cheapest, as the patterns are 
 very simple — the body, when not made of steel, copper, or brass sheets, 
 being struck out ; the covers, tube-plates, and corresponding flanges can all 
 be faced in a lathe ; this form also is by far the lightest, for the two reasons, 
 that the circular plate is the form giving the minimum perimeter for a given 
 area, and consequently a minimum barrel, and that the cylindrical form for 
 . either internal or external pressure is the strongest, so that the thickness of 
 metal will be the minimum. A modification of the cylindrical form possesses 
 many of these advantages. 
 
 The waterways, or chambers at each end of the condenser, are sometimes 
 cast with it, and sometimes cast separately ; in the latter case there is an 
 additional joint, but this is mitigated by its also forming the tube-plate joint ; 
 in the former case there is only one joint less at each end through which aic 
 can leak, but the plates are more troublesome to fit, and, except in the case 
 of the cylindrical form, the tube plate flange is difficult to face. 
 
 Great care should be taken in designing a condenser that free outlet is 
 given to any air that may collect near the tubes, and all pockets, where it or 
 dead water could lie, should be avoided, as a few of the tubes may get hot 
 and leak from the above causes. 
 
 The Construction of the Surface Condenser is to-day in all naval ships, in 
 express steamers, and even in some cargo ships of steel sheets or plates from 
 £ to | inch thick in the body, and of cast iron or brass in the waterways and 
 parts exposed to sea-water in naval ships, and of cast iron in the mercantile 
 marine where weight is not of paramount consideration. It is recognised 
 now that there is no corrosive action on the body exposed to steam and 
 condensed water, since the use of sea-water as the supplementary feed has 
 been given up, and brass in the waterways can be protected by zinc slabs, or 
 even by a few steel studs ; cast iron also is easily protected by zinc slabs. 
 The condensers of such craft as destroyers, where the last ounce must be 
 saved, should be of sheet brass, cylindrical in form, and of sufficient thickness 
 to withstand atmospheric pressure and its own weight. With such condensers, 
 
368 
 
 MANUAL OF MARINE ENGINEERING. 
 
 as also with the brass waterways, the mudhole and peephole doors may be 
 of cast iron or steel with advantage, as then they form the protectors against 
 corrosion of the brass from sea-water, and can be easily and cheaply replaced. 
 The shape of the modern condenser is either as an approximation to 
 a triangle, or what is known as heart-shaped in transverse section, as 
 shown in fig. 118. The cylindrical condenser is generally fitted with baffles 
 of some kind, although with the modern entering in so much enlarged, 
 especially as it is for turbines, the need to protect the upper rows of tubes 
 from the pulsating impact of the steam has almost ceased to exist. With 
 the ordinary rectangular or nearly rectangular condenser of the mercantile 
 marine and the circular section exhaust pipe there should still be this screen, 
 as well as means for draining the condensed water when formed without 
 further contact with other tubes. 
 
 TABLE XL III. — Trials of I.J. Battleship "Ibuki" — Curtis Turbines. 
 
 Date. 
 
 Kinds of 
 Trial. 
 
 Aug. 12, 1909, Full power 
 Aug. 7, 1909, 
 July 31, 1909, 
 July 26, 1909, 
 July 24, 1909, 
 
 Baro- 
 meter 
 
 in 
 Inches. 
 
 29 995 
 29 990 
 
 29 900 
 
 30 300 
 30 320 
 
 Temperature (Deg. Fahr.). 
 
 Super Heat 
 
 ifter through 
 
 Regulating 
 
 Valve. 
 
 Star- 
 board. 
 
 21-8 
 
 150 
 
 9-8 
 
 1-3 
 
 31 
 
 Port. 
 
 22-9 
 
 10-3 
 
 3 9 
 
 2 
 
 4-3 
 
 Main Condenser. 
 
 Sea Water F" 
 
 Inlet. 
 
 Out- 
 let. 
 
 74 6 
 
 71-2 
 
 75 
 73 2 
 
 78 2 
 
 977 
 914 
 93-4 
 
 90-8 
 92 3 
 
 Con- 
 densed 
 Water, 
 
 F°. 
 
 90-9 
 S6 
 
 88-5 
 86-7 
 88-7 
 
 Feed- 
 Water 
 
 after 
 through 
 
 Feed 
 Heater. 
 
 13111 
 
 132-75 
 
 151-6 
 
 136-8 
 
 145-2 
 
 Mean 
 
 Vacuum 
 
 in Main 
 
 Condensers 
 
 at 
 
 Top. 
 
 Bottom 
 
 26-33 
 
 2S-42 
 
 26 90 2S25 
 
 27-10 ' 27 90 
 
 27-80; 2S50 
 
 28-03 
 
 28-32 
 
 Date. 
 
 Boiler. 
 
 8.H.P. 
 
 per 
 Sq. Foot 
 
 of 
 Grate. 
 
 Aug. 12, 1909, 
 Aug. 7, 1909. 
 July 31, 1909. 
 July 26, 1909, 
 July 24, 1909, 
 
 16 45 
 
 12-80 
 
 9-72 
 
 9-23 
 
 9 05 
 
 Water Rate of Main 
 Engine per S.H.P. 
 
 Main Condenser. 
 
 Not 
 Corrected 
 
 to 
 Contract 
 Condi- 
 tions. 
 
 Corrected *&£ 
 Contact I Bfcfi* 
 
 tions. (V ™- J per Hour 
 (Per hour. )Cernom.; .„ Lbg- 
 
 15 050 
 15-686 
 
 16 505 
 18-652 
 21074 
 
 13-77 
 14-76 
 15-62 
 17-75 
 20-57 
 
 1 1 -505 
 9 099 
 7 253 
 5-038 
 3-065 
 
 Circulat- 
 ing 
 Sea \\ ater 
 
 per Lb. 
 of Steam 
 
 in Lbs. 
 
 44 -296 
 39-685 
 41113 
 310§6 
 30 946 
 
 Main Engine. 
 
 Bucket 
 
 Speed 
 
 in Feet 
 
 per Sec. 
 
 Steam 
 
 Velocity 
 
 in Feet 
 
 per Sec 
 
 (IstStage.) 
 
 157 3 
 
 147-7 
 
 136-2 
 
 118-4 
 
 95-6 
 
 2,142-5 
 2,330 
 2,379 
 2,650-0 
 2,943 
 
 Efficiency 
 
 of 
 Turbine. 
 Per cent 
 
 52 45 
 48-95 
 46 30 
 40 70 
 3510 
 
 The Stiffening of Flat Surfaces should be by means of ribs cast with it 
 
 when of cast iron or cast brass, and 25 thicknesses apart for cast iron and 
 
 40 for tough brass and sheet steel and brass, which should have angle or 
 
 ("-iron stitt'eners riveted on. Malleable flat surfaces can be stiffened by 
 
QUANTITY OF COOLING WATER. 369 
 
 corrugations formed with it about 35 thicknesses pitch if fairly deep ; if 
 the corrugating is light, the pitch should be much less. 
 
 Quantity Of Cooling Water. — The necessary amount of circulating water 
 may be calculated in the same way as that for injection water (see p. 345). 
 on the principle that the exhaust steam has a certain quantity of heat which 
 is to be expended in raising a mass of sea-water of a certain temperature to 
 nearly that corresponding to that in the condenser. The quantity of sea- 
 water will, therefore, depend on its initial temperature, which in actual 
 practice may vary from 40° in the winter of temperate zones to 80° of the 
 West Indies and other subtropical seas. In the latter case, with a vacuum 
 of 28 inches, a pound of water requires only 20 thermal units to raise it to 
 100°, while 60 are required in the former. From this it is seen that the 
 quantity of circulating water required in the tropics is three times that of 
 the North Atlantic in the spring of the year. 
 
 As before, let T l be temperature of the steam on entering the condenser, 
 and L the latent heat : T the temperature of the circulating water, and Q 
 its quantity ; T the temperature of the water on leaving the condenser, and 
 T 3 the temperature of the hot-well. 
 
 The heat to be absorbed by the cooling water is now (Tj + L) — T 3 ; 
 and this amount of heat must be equal to Q (T 2 — T ). Hence, 
 
 Q = (T, + L) - T 3 ■*■ (T, - T„). 
 Q= 1,114 + 03 T.-T, 
 
 ■*-2 x 
 
 Example. — To find the amount of circulating water required by an engine 
 whose steam exhausts at 8 lbs. pressure absolute, the temperature of the 
 sea being 60°, and (2) the amount required when the temperature of the 
 sea is 75°. The temperature of the hot- well to be 120°, and th'at of the 
 water at the discharge 100°. Vacuum 26 inches. The temperature corre- 
 sponding to 8 lbs. is 183 . 
 
 (1) Q = UH + 0-3X183 -120 = 
 W w 100—60 
 
 That is, the water required is 26-22 times the weight of steam. 
 
 With the jet condenser under similar conditions the quantity was only 
 17-48 times {v. p. 346). 
 
 (2) When the sea is at 75° 
 
 n 1,114 + 0-3 x 183-120 . 
 
 Q = loo— 75 = 41 9o tlmes - 
 
 It is usual to provide pumping power sufficient to supply 30 times the weight 
 of steam for general traders, and as much as 40 times for ships working in 
 subtropical seas. As will be shown in another chapter, if the circulating 
 pump is double-acting, its capacity may be -^ in the former, and -^ in the 
 latter case of the capacity of the low-pressure cylinder. 
 
 Table xliv. shows the least weight of cooling water required in practice 
 per pound of steam entering the condenser at a pressure of 12 lbs. absolute. 
 Modern air pumps can rffaintain a vacuum of 29 inches, but it will be seen 
 
 24 
 
S70 
 
 MANUAL OF MARINE ENGINEERING. 
 
 by this table that the quantity of cooling water required for it at a tem- 
 perature of 70° is enormous, and with water at 80° is impossible. 
 
 TABLE XUV. — Ratio of Cooling Water to Steam Condense/). 
 
 Vacuum, 
 
 . Ins. 
 
 250 
 
 25-5 
 
 26-0 
 
 26-5 
 
 27-0 
 
 27-5 
 
 28-0 
 
 28-5 
 
 29-0 
 
 29-3 
 
 Sea water, 
 
 . 50 c F. 
 
 13-9 14-7 
 
 15-4 
 
 16-4 
 
 18-2 ! 20-1 
 
 22-8 
 
 28-0 
 
 4(1-4 
 
 65-0 
 
 )> 
 
 . 60° F. 
 
 16-0 
 
 17-1 
 
 18-1 
 
 19-5 
 
 22-1 24-S 
 
 29-0 
 
 38-0 
 
 04-3 
 
 ]3H 1 
 
 »> 
 
 . 70° F. 
 
 18-9 
 
 20-5 
 
 21-8 
 
 24-0 
 
 27-9 32-3 
 
 40-0 
 
 66-3 
 
 156 
 
 . . 
 
 j» 
 
 . 80° F. 
 
 231 
 
 25-5 
 
 27-6 
 
 31-0 
 
 37-9 ! 46-3 
 
 63-0 
 
 133 
 
 " 
 
 
 The highest possible vacuum with the temperature of cooling water at 
 80 c is 28"8 ; and the ratio no less than 270. The highest vacuum when the 
 water is 70° will be 29' 1 inches, and the ratio then 220. In the temperate 
 zone with sea-water at 60° a vacuum of 29 - 4 can be maintained by passing 
 280 times the weight of steam condensed. 
 
 In every-day practice, and the condenser not too clean, the quantity 
 of water required may easily be 10 to 20 per cent, more than that given in 
 this table. 
 
 Passage of Circulating Water. — The water must be caused to pass over 
 a sufficient amount of surface to become duly heated, if the minimum quantity 
 is to be used. In practice it should tra\el at least 20 feet lineally through 
 the tubes before leaving the condenser ; if this cannot be arranged, then it 
 must remain longer in contact with the surface. Hence, in small condensers, 
 where the steam is outside the tubes, the water circulates only twice through 
 them at a slow pace ; in larger condensers it may circulate twice through 
 long tubes, or three or four times through shorter tubes at a higher velocity, 
 due to the larger quantity of water. To obtain the best results the velocity 
 of flow should be not less than 200 feet per minute when the sea-water is at 
 60', and 300 feet when at 75°. When the water circulated outside the tubes, 
 it was necessary to fit baffle-plates, to divert the water and prevent it from 
 taking the shortest course from the inlet to the outlet, and also to prevent 
 the hot water from accumulating at the top part and the cold water from 
 remaining in the bottom. Some skill and ingenuity were often required to 
 fully overcome such difficulties. Condensers are, however, seldom or never 
 made now with the water outside the tubes. 
 
 Where reciprocating pumps are employed, the circulating water should 
 be admitted to the condenser direct from the sea, and pumped from it through 
 an outlet above the level of the top row of tubes ; the pressure in the condenser 
 does not then exceed that due to the head of water, and little or no shock 
 is given to it by the varying velocity of the circulating pump, as is the case 
 when the water is forced through. When a centrifugal pump is used, it 
 should deliver into the condenser direct from the sea. 
 
 The Size of the Circulating Pump Pipes may be calculated in the following 
 manner, where C is the consumption of steam per I.H.P. per hour. K the ratio 
 of cooling water to steam condensed. Then : — 
 
 p w f v T TT P 
 
 Total quantity of cooling water in pounds per minute = — — ■ — '— '■• 
 
SIZE OF PIPES. 371 
 
 If the maximum flow of water be taken as an average of 600 in temperate 
 zones, where K does not exceed 64, and 800 in the tropics when K may be 
 130. The weight of a cubic foot of sea-water being taken at 64 lbs. 
 
 Then volume of water ) C X 64 X I.H.P. 
 
 used per minute J ~ 60 X 64 cublc teet 
 
 C x I.H.P. 
 
 60 
 
 cubic feet. 
 
 C x I H P 
 
 The area of pipe section = -^ — ~^tt square feet in temperate zone. 
 
 C x I.H.P. 
 
 30x800 
 
 square feet for tropical work. 
 
 The diameter of the pipe = y/ ^ ^^ and J~ 
 
 C x I.H.P. 
 
 000 x 0-7854 
 
 x /C x I.H.P. x/C x I.H.P. f 
 
 ~ ~l65~ anCi 13V3 teet * 
 
 Or, diameter of circulating) _ JC X I.H.P. 
 water pipes in inches / x ' 
 
 where for European waters x = 13 - 8, and for tropical waters 11*5. 
 
 Example. — A ship intended to cross the equator is to be of 2,000 I.H.P. ; 
 her engines require altogether 16 lbs. of steam per I.H.P. per hour, what 
 size circulating pipes should she have ? 
 
 Diameter = */l6~>T2,000 + 115 or 15-6 inches. 
 
 Size of Inlet and Discharge Pipes * should be such that the flow of water 
 through them when the engines are working at full speed does not exceed 
 700 feet per minute. In temperate zones 7 lbs. per minute of cooling water 
 per I.H.P. is sufficient for a triple-compound engine, and 6 lbs. will do for 
 a quadruple engine or a triple working at a fairly high rate of expansion 
 with 27 "5 inches vacuum, but for the same vacuum in the tropics it will 
 require 12 and 10 - 6 lbs. (v. Table xliv.). Hence 
 
 Area section inlet and discharge = ' ' sq. ins. 
 
 Where C = 32 triple and quadruple high expansion, general, 
 = 30 „ ,, moderate ,, ,, 
 
 = 26 „ ., tropics ,, (if 27*5 vacuum). 
 
 Extra Supply Cock. — To provide for the water lost in waste, leakage, 
 etc., it is usual to fit a small cock, through which some of the circulating 
 water may be passed to the steam side of the tubes. The pipe for this should 
 be about one-third the diameter of the main feed pipe, the velocity of flow 
 being nearly nine times that usually provided for in feed pipes. Now that 
 evaporators are in general use, so that the waste can be made up with fresh 
 water, the necessity for such a fitting scarcely exists, but as an emergency 
 provision it may remain. For reserve water such a cock is convenient. 
 
 * The flow should not exceed the rates given by this rule : — v =350 v diameter pipe. 
 
372 MANUAL OF MARINE ENGINEERING. 
 
 Man-holes and Mud-holes. — A man-hole is necessary for the purpose 
 of admitting men to clean, repair, or tube the condenser, and smaller holes 
 should be provided through which mud, scale, etc., may be scraped out. 
 Peep-holes are sometimes formed in the doors, especially if they are large 
 and heavy, through which the tube ends may be seen and examined. These 
 latter are useful when steam is condensed inside the tubes, to admit the 
 nozzle of a steam jet to wash away deposit. 
 
 Drain Cocks should be fitted so that the condenser may be thoroughly 
 drained when not in use. 
 
 Testing. — The Admiralty require condensers to be tested to 30 lbs. per 
 square inch before being placed in the ship, and many steamship companies 
 require the same test, while others are content to test with lower pressures. 
 To provide for such loads, the flat surfaces must be stiffened in the same 
 way as laid down for cylinders, and, if flat-sided, stiffened as already pre- 
 scribed, and, when necessary, tied together by stay bars, etc. 
 
 Cementing. — It is a good plan to cover the inside of iron condensers, 
 where there is much wash of condensed water, with a good coating of Portland 
 cement, and under the air-pump and in the pump passages it should be at 
 least a quarter of an inch thick. 
 
 Evaporators. — Since the use of steam of 150 lbs. and upwards the extra- 
 supply from the sea has been avoided as much as possible, fresh water being 
 carried in tanks or double bottoms ; now the employment of " evaporators " 
 is doing away with the necessity for this and providing a long-felt want, 
 and permitting of the use of water-tube boilers and modified forms of ordinary 
 marine types. 
 
AIR-PUMPS. 373 
 
 CHAPTER XV. 
 
 PUMPS. 
 
 Air-Pump. — The function of this pump in all condensers is to abstract the 
 water condensed, and the air which was contained in the water when it 
 entered the boiler : and in the case of jet-condensers, it pumps out in addition 
 the water of condensation and the freed air which it contained. Further^ 
 it must withdraw all air which leaks into the condenser. 
 
 It follows, then, that the size of the air-pump must be calculated from 
 these conditions, and allowance made for the efficiency of the pump ; or, 
 what is the same thing, the result thus found must be multiplied by the ratio 
 between what the pump should do theoretically, supposing its action perfect, 
 and what it does actually in practice. 
 
 Ordinary sea-water contains, mechanically mixed with it, one-twentieth 
 of its volume of air, when under the atmospheric pressure. Now, suppose 
 the pressure in a jet condenser to be 2 pounds, and the atmospheric pressure 
 15 pounds, neglecting the effect of temperature, the air on entering the con- 
 denser will be expanded to \ 5 times its original volume ; so that a ctibic foot 
 of sea-water, when it has entered the condenser, is represented by a cubic 
 foot of water, and ^ of a cubic foot of air. 
 
 Now let q be the volume of water condensed per minute, and Q the 
 volume of sea-water required to condense it ; and let T 2 be the temperature 
 of the condenser, and T 1 that of the sea-water : — 
 
 Then (q + Q) will be the volume of water to be pumped from the con- 
 denser per minute, and 
 
 40 (? + Q) X t 2 + 461 ° the <l uantit y of air '* 
 
 If the temperature of the condenser be taken at 120°, and that of sea- 
 water at 60°. the quantity of air will then be "418 (q + Q), so that the total 
 volume to be abstracted will be 
 
 (q + Q) + -418 (q + Q) = 1-418 (q + Q). 
 
 Now. if the average quantity of injection water be taken at 26 times 
 that condensed, q -f- Q will equal 27 q. 
 
 Therefore, volume to be pumped from the condenser per minute 
 
 = 38 q. 
 
 Example. — To find the theoretical capacity of a single-acting pump for 
 
 * Absolute zero point, or point of no heat, is 461° below the zero of Fahrenheit's 
 thermometer. 
 
374 MANUAL OF MARINE ENGINEERING. 
 
 an engine using 3 cubic feet of water per minute, and the number of stroke? 
 being 40. 
 
 Volume to be pumped out = 38 X 3. or 114 cubic feet. 
 
 114 
 
 Therefore, the capacity of the pump = -j^-, or 2*85 cubic feet. 
 
 If the stroke of the pump be taken at 1*5 feet. 
 
 2*85 
 The area of bucket = y^*, or 1*9 square feet. 
 
 Example. — To find the theoretical capacity of a double-acting pump for 
 an engine using 10 cubic feet of water per minute, the number of revolution* 
 being 90, and the stroke of the pump 2*5 feet. 
 
 a Volume to be pumped out = 38 X 10, or 380 cubic feet. 
 
 380 
 
 Area of bucket = K ^ » x-=, or '844 square foot. 
 
 90 X 2 X 2*5 ^ 
 
 Of course, were these examples worked strictly, it would be necessary 
 to find the relation between Q and q in each case, instead of assuming it at 
 26, as has been done. 
 
 The foregoing calculations, etc., are for jet-condensers, and based on the 
 supposition that the air-pump also abstracts the condensing water ; now in 
 a surface-condenser, not only is the air-pump relieved of this latter duty, 
 but, since the feed-water is condensed steam, and has not had time to absorb 
 much air, it would seem that practically its only function is to draw off the 
 condensed water, and should, therefore, be no larger than a feed pump, as 
 no doubt it might be, did the engine work perfectly; but since ordinary 
 water has to be occasionally admitted to make up for waste, and as slight 
 leakage at the glands and joints very frequently exists, it would be necessary 
 and sufficient to make the air-pump about half the size that would be requisite 
 for jet condensation ; but when a surface-condenser was arranged so as to 
 be worked as a jet-condenser, the air-pump had to be large enough to do the 
 work necessitated by this, unless the circulating pump was fitted so as to 
 be worked as an air-pump. A surface-condenser is now seldom arranged 
 as to admit of jet condensation. Too great care cannot be expended on the 
 design of the air-pump for a surface-condenser, as the success of many an 
 engine in the past was marred by a bad vacuum ; and doubtless an inch or 
 two of vacuum tells wonderfully on an engine, especially on a compound 
 engine, where the degree of vacuum at any time can be told almost by varia- 
 tion of its speed. Since, however, the surface-condenser was required to work 
 as a jet-condenser, when so fitted, only in cases of emergency, which seldom 
 happen, and when such a case does occur it is not of importance that the 
 engines shall work at the highest efficiency or maximum speed, it seems 
 better so to design the air-pump as best to suit surface condensation, rather 
 than to make it of the larger size, and sacrifice a large amount of work in 
 driving it during the long period, when jet condensation is not necessary. 
 In short, the pump should be of such a size as to produce maximum efficiency 
 during the longest time, and if for a surface-condensing engine, it should 
 be designed for that particular service. 
 
AIR-PUMPS. 
 
 375 
 
 Air-Pumps were originally simple lift pumps, having a valve in the bucket 
 and a delivery or " head " valve at the top, through which was passed the 
 water of condensation, mingled with which was the condensed water and 
 the air or non-condensable gaseous matter passing to the cylinder from 
 the boiler, or that emanating from the water of condensation being previously 
 mechanically mixed or in solution with it. The modus operandi was simple ; 
 the pump bucket descended to the bottom of the (vertical) chamber, and 
 in so doing expanded such gaseous matter as remained locked between it and 
 the head valve until the pressure was less than that in the condenser, when 
 
 Fig. 122. — Air-pump cf Ordinary Type. 
 
 there flowed into it the gaseous residuum of the condenser and the water 
 until the bucket ceased to move and the valve closed ; the bucket on ascending 
 compressed the gases till their pressure was higher than that of the atmo- 
 sphere, when the head valve opened, and they together with the charge of 
 water were delivered into and separated at the hot-well. 
 
 With the Horizontal Engine the Air-Pumps were usually horizontal and 
 double-acting, so that the bucket no longer had a valve ; moreover, sometimes 
 in lieu of a cylindrical pump chamber and piston or bucket, there was a 
 
376 
 
 MANUAL OF MARINE ENGINEERING. 
 
 plunger passing through a stuffing-box in the diaphragm separating the back 
 from the front chamber of the pump. Although with a well-designed horizontal 
 pump in good working order quite high vacua were sometimes maintained, as a 
 rule such pumps were not nearly so efficient as a vertical pump. They had the 
 same stroke as the engine, and as the valves were quite apart from the pump, 
 they could be of any reasonable number or size. A return to such pumps 
 is now most unlikely, for, as a fact, even horizontal engines had latterly 
 vertical pumps whenever possible. But some makers of vertical engines have 
 preferred to fit pumps worked direct from the piston, so that their stroke 
 was large and vohlme clearance small, but the passages through their 
 buckets is likewise small, and consequently a source of danger with a 
 leaky condenser. 
 
 Fig. 123. — Edwards' Air-pump. Bucket at Bottom of Stroke. 
 
 The Single-acting Vertical Air-pump (tig. 122), having valves in the bucket 
 as well as foot and delivery valves, is by far the most efficient, -and. when 
 possible, was generally chosen. If the rod of such a pump is enlarged or the 
 bucket has a trunk to surround the rod, which is attached to a joint in its 
 centre, it is to a certain extent double-acting, since on the upstroke it causes 
 the chamber in which it works to fill, on the downstroke it displaces, and 
 consequently discharges a volume equal to the volume of the trunk, and on 
 the upstroke discharges the remainder. If the sectional area of the trunk is 
 half that of the bucket, the discharges are equal, and the pump is virtually a 
 double-acting one. 
 
 Edwards' Air-pump is an ingenious form of single-acting vertical pump, 
 with valves only at the top to check the discharged water and air from 
 
AIR-PUMPS. 
 
 377 
 
 returning to the pump on the downstroke. Fig. 123 shows the «eneral 
 design and action of the bucket or piston of this pump, whereby the water 
 and air are caused to enter the exhausted space between the head- valves 
 and the bucket. These pumps are very simple, have the minimum number 
 of valves, which are easily examined and replaced, work with the least 
 possible resistance and wear and tear, and produce a good vacuum. 
 
 The Double-acting Air-pump. — It is not always convenient to have a 
 vertical pump in the horizontal engine, and consequently a horizontal pump 
 was generally employed, and this was almost of necessity double-acting. 
 The bucket in this case works air-tight in a smooth barrel placed beneath the 
 
 Fig 124. — Weir " Dual " Air-pump (direct-driven). 
 
 condenser, and has a set of foot and delivery valves for each end. Some- 
 times, in lieu of a barrel and bucket, a plunger is fitted, passing through a 
 stuffing-box in the diaphragm-plate dividing the condenser bottom. This 
 latter arrangement generally admits of more room for, and a better dis- 
 position of, the valves. Some engineers adopted this form of the horizontal 
 pump for twin-screw vertical engines, and worked it by means of an eccentric 
 formed on one of the crank-arms. 
 
 Weir's Dual Pumps are shown in fig. 124, and in the diagrammatic form 
 fig. 124rt the full working of this ingenious and effective system is clearly seen. 
 There are two pumps — one called the Wet Pump draws water and such "air" 
 
378 
 
 MANUAL OF MARINE ENGINEERING. 
 
 as it can from the condenser in the usual way, and discharges into the hot- 
 well ; the other, which may be looked on as the additional or auxiliary, is 
 
 «2S- 
 
 "SCO 
 
 called the Dry Pump, inasmuch as it draws no water direct from the 
 denser, but the gaseous products only ; it does not, however, deliver 
 
 con- 
 theni 
 
AIR-PUMPS. 379 1 
 
 to the hot-well or other open receptacle, but into the chamber of the other 
 pump between the bucket and the head valves, and then only when the tension 
 there is low for a delivery or non-return valve prevents the contents of the 
 first pump when at high compression flowing into the second. In this way 
 there is maintained a two-stage delivery of the gaseous products, and a& 
 the two pumps are worked from the opposite ends of a rocking lever, the 
 dry pump is delivering only when the piston of the other is falling and ex- 
 panding the contents of its chamber. The result is that the contents of the 
 dry pump are delivered at quite a low pressure, and the whole compressed 
 and discharged to the hot-well by the wet pump. The pressure in the wet 
 pump under these circumstances is never very low, so that any air it gets 
 must be mechanically mixed with the water, which, under the action of 
 gravity, flows into the wet pump. It is obvious that the work of the two 
 pumps is quite different, while there is active co-operation between them ; 
 also that their efficiency will be high, and the air leakage not nearly so serious 
 as those in pumps where the difference in pressure at every point is greater. 
 But the second or dry pump is somewhat misnamed, inasmuch as it is supplied 
 with water with which to lubricate the moving parts as well as seal them 
 against air leaks, and inasmuch as that water is very cold, it serves the more 
 important function of being the means whereby the gases are cooled and 
 the pump kept cool, so that the low vacuum desired in it may be obtained. 
 This water is cooled in the pipe coils shown, by means of sea-water, and is 
 really a small auxiliary condenser, which serves the same purpose as the 
 few drowned tubes in the Morison condenser (fig. 118). 
 
 The following is Mr. Weir's own description of the dual pumps and their 
 action : — 
 
 Fig. 124a shows in a diagrammatic form the arrangement of surface-condenser, 
 " dual " air-pump, and injection water cooler. In all cases the pump A or wet pump 
 is situated below the steam cylinder, as this pump is the only one which works under 
 any considerable load, the dry pump B is driven by the beam and links in the usual manner. 
 One connection C is made to the condenser, but a branch pipe D is led to the dry pump, 
 the connection being made in such a manner that the water will all pass by C : to the wet 
 pump. Both pumps are generally of the three-valve marine type, but in certain cases- 
 the dry pump may be of the suction valveless type. 
 
 The first and most important difference from an ordinary twin pump consists in 
 the separate suction to each pump, then in the dry pump discharging through the return 
 pipe E, through a spring-loaded valve F, into the wet pump at a point below its head 
 valves. The next point concerns the supply of water to the dry pump for water sealing, 
 clearance filling, cooling, and vapour condensing. When starting the pump the filling 
 valve G must be opened for a minute or so to enable the vacuum to draw in a supply 
 from the hot-well of the wet pump. The valve is then closed, and the water passes from 
 the hot-well of the dry pump by the pipe H to the annular cooler, through which a supply 
 of cold sea-water circulates, and after being cooled passes into the suction of the dry 
 pump, then passing through the pump it becomes heated and again passes to the cooler, 
 and so on in a continuous closed circuit, any excess passing over the pipe E to the wet 
 pump. The spring-loaded valve F is adjusted to maintain about 20 inches vacuum 
 in the dry pump hot-well when the condenser is working at 28 inches vacuum, and thi» 
 8 inches difference of pressure is sufficient to cause the water to overcome the cooler 
 friction and pass into the suction, and at the same time never allow any direct air con- 
 nection between the dry suction and discharge. 
 
 Air-pump Indicator Diagrams, shown on fig. 125, explain the action of 
 each kind of pump now used on shipboard for extracting air from a condenser, 
 assuming that the valve resistance is nil. The pressure in the condenser is 
 
380 
 
 MANUAL OF MARINE ENGINEERING. 
 
 represented by L C, that of the external air J, and the head at delivery 
 above that pressure A H. If no water is passing, A H will be nil. 
 
 A J is the equivalent clearance above the bucket, and D K the full chamber. 
 Taking first a simple pump like the Edwards, and suppose the bucket at the 
 top of its stroke A, and the clearance space charged with air of pressure 
 J. On descending this is expanded, as shown by the curve A B D, so that 
 the pressure in the chamber at the end of the stroke would be L D, while 
 that in the condenser is L C. But before reaching the point D, the piston 
 passes the openings in the barrel at M, when the inflow from the condenser 
 takes place from M to C. The bucket then begins to rise, and the com- 
 pression of air takes place on the curve C G, until the pressure is equal to 
 that in the hot-well, when it commences to discharge till the pressure is at H. 
 If, however, the pump has valves in the bucket and foot-valves without 
 clearance under the bucket, then the following cycle is followed. Expansion 
 
 I Z 3 * 5 6 7 e 3 10 II 12 13 /* IS 16 17 IB 13 20 
 Fig. 125. — Indicator Diagram of Air-pumpe (Theoretical). 
 
 takes place above the bucket as before, but now compression below it also 
 takes place until the pressure below is equal to that above the bucket — that 
 is, at the point B, where the expansion curve A B intersects the compression 
 ■curve F B, during the remainder of the stroke the pressure above and below 
 is L C ; the valves close, and the bucket rises, compressing the air above it 
 on the curve C G as before ; but now expansion takes place below, following 
 the curve C E until the pressure is so much below that in the condenser 
 (0 F in this instance) that vapour and air flow into the chamber, and finally 
 fill it at a pressure F, and so on, as before. 
 
 It will be seen that, with the same clearance on top of the bucket, the 
 Edwards pump cannot produce so good a vacuum as the common one with 
 foot-valves. As a matter of fact, however, the Edwards has naturally 
 smaller clearance space, and both kinds of pump have virtually little or no 
 real clearance, as the space is filled with water. The water itself, however, 
 
AIR-PUMPS. 38] 
 
 has some air in solution, and probably such as fills the space is frothy, 
 there will be thus a modicum of gaseous matter to expand, and the cooler the 
 chamber is the less will the expansion be. Hence the supply of water to seal 
 the pump and fill the spaces should be cold and as free from air as possible. 
 
 The Efficiency of Air-pumps. — The high efficiency of the single-acting 
 vertical pump is due to the certainty of its action in taking the water, etc., 
 through the bucket-valves, and to the valves from their position so readily 
 closing when required ; there is also time for the water to drain into the 
 bottom of the pump during its upstroke, and collect there ready for the 
 bucket when it descends. Moreover, as the flow is always in one direction, 
 the velocity of flow is not checked by diversion. The water always lies on 
 the valves so as to render them air-tight, and there is very little clearance 
 space, as a rule, between the foot and bucket valves, and between the bucket 
 and head valves, and what there is contains water, and, therefore, virtually 
 there is no clearance at all for air. 
 
 The low efficiency of the double-acting horizontal pump is caused by 
 the reverse of some of the above conditions, especially by the failure of the 
 valves in closing, and to the large space between the foot and delivery valves, 
 also by leakage at the gland of the rod, and past the bucket, which is only 
 lubricated by the water on the bottom, and in no small degree by the ever- 
 changing direction of flow. The latter defect is proved by stopping one end 
 of the pump, when it is often found (especially in the case of badly designed 
 pumps, etc.) that the vacuum is not very materially altered, although some- 
 times improved. The foot valves are sometimes kept covered with water 
 by allowing the water to pass back again through a pipe from the hot-well 
 to the pump-chamber. 
 
 If an Air-pump has no Foot Valves, or, like the Edwards pump, the 
 clearance real or virtual must be such that there is a vacuum in the pump 
 chamber superior to that in the condenser an appreciable time before the end 
 of the stroke, so that some of the gaseous contents of the condenser may enter 
 it. If, therefore, f e is the pressure to be maintained in the condenser, x the 
 clearance fraction of stroke S, the atmospheric pressure, say, 15 lbs., and the 
 weight of the valves and water " head " and resistance equal to another 
 pound, then the pump must compress on the upstroke to a pressure of 16 lbs. 
 to deliver its contents, and on the closing of the valves the pressure of gases 
 remaining locked in the clearance will be 15 lbs. 
 
 The clearance (virtual) must not be less than r~ S to effect any inflow 
 
 at all, but as this inflow must commence before the end of the stroke, and 
 in the Edwards pump there must be a much lower pressure than in the 
 condenser to carry away considerable quantities of vapour, x must be much 
 
 less than ^~, and should be £tt at most, 
 lo 20 
 
 During the down-stroke expansion of the gases takes place until the 
 
 pressure is less than that below the piston or bucket, and from that point to 
 
 the end of the stroke the pump is being charged from the condenser through 
 
 the bucket valves'. The pressure of the gases in the pump is then p c , which 
 
 on compression must be compressed to 16, hence the ratio of pump chamber 
 
 to its clearance will be 16 -f- p c when delivery takes place, and since this 
 
 must be before the end of the stroke, it follows that x must be less than £| • 
 
382 MANUAL OF MARINE ENGINEERING. 
 
 If, therefore, a pump is to maintain a vacuum of 28 inches at all, its 
 clearance must be less than T Y, and should be ■£$ to -^ s . 
 
 For so high a vacuum as 29 inches it must be even higher, and not less 
 than -£z, and should be, if possible, T V- 
 
 Since, however, the pump is always charged with water, so as to seal the 
 bucket, the valves and other sources of air leak, and generally fills the clear- 
 ance space, the virtual clearance in a properly designed pump is exceedingly 
 small. 
 
 But to form and maintain in the pump chamber a vacuum under 28 
 inches, the water entering it must never exceed 95° F., and for 29 inches it 
 should be under 78° F. If, therefore, it is desirable to obtain the condensed 
 water as warm as possible with such high vacua, there should be two pumps, 
 one dealing with the water, the other with the gases, and a special supply 
 of cold water to fill the clearance and seal the leaks, as done virtually by 
 Mr. Weir. 
 
 If an Air-pump has Foot-valves the conditions obtaining at the lower 
 end of the pump differ from the foregoing, for on the down-stroke the attenu- 
 ated gases remaining in the pump undergo compression at the same time 
 that the fragment of compressed gases above the bucket are expanded, as 
 already described ; when the pressure below is slightly in excess of that 
 above the bucket valves lift and remain open to nearly the bottom of the 
 stroke ; when they close the pump is charged with gases at the pressure 
 to which they were compressed, and the further compression of them takes 
 place until the 16 lbs. pressure is attained, when discharge into the hot-well 
 takes place. On this upstroke, however, expansion of the gases between 
 the bucket and the foot valves takes place, and continues till the pressure is 
 somewhat below that of the condenser, when inflow from it commences, so 
 that at the end of the upstroke the pump chamber is filled with gas at a 
 pressure very nearly equal to pc. 
 
 Diagram fig. 125 shows the cycle of the top and bottom of an air-pump 
 with foot valves, and the plain dash lines indicate what happens when there 
 is no foot valve. The diagram is constructed on the assumption that the 
 virtual clearance at both top and bottom is equal to 5 per cent., or ^ the 
 stroke, and that expansion and compression follows the law pv = constant, 
 which is sufficiently correct for the purpose here. It is also assumed that 
 there is no resistance to inflow from the condenser or through the bucket. 
 
 In the case of the single-acting pump working under these conditions, 
 it will be seen that the highest vacuum will probably be about 27, and by 
 the diagram is shown as 26-5, whereas with the foot valves fitted it is 29 inches 
 in the pump, and 28 in the condenser. It is likewise obvious that, if the 
 virtual clearance can be kept down below 5 per cent., which is really not 
 difficult, even the Edwards or the common pump without a foot valve can 
 maintain high vacua. 
 
 The resistance of the pump in question, due to compression of gases, 
 is equal to a mean pressure of 4-92 lbs. per square inch of bucket on the 
 upstroke. In the downstroke the pressure above the piston is greater 
 than that below it for 45 per cent, of the stroke, and practically nothing 
 for the remaining 55 per cent. The weight of the bucket and the water helps 
 on the downstroke as much as it resists on the up. But the condensed 
 water has to be lifted a few feet, and there is resistance due to friction. For 
 
AIR-PUMPS. 
 
 Wd 
 
 calculation tf power, a mean pressure during the upstroke of 6*5 lbs. will 
 be sufficient, but it must not be overlooked that during about 8 per cent, 
 the resistance is from 15 to 16 lbs. per square inch, and the ratio of maximum 
 to mean resistance is, therefore, as much as 2 - 46, and has to be overcome 
 by motor or steam piston employed to work the pump. Unfortunately, the 
 maximum resistance comes at the termination of the stroke, so that one 
 cannot fit a cylinder with an early cut-off to the steam. The only relief 
 so far provided is by working the two pumps from the opposite ends of a 
 locking lever, or by having three pumps worked from a three-throw crank- 
 shaft having the cranks at 120° apart. The ratio of maximum to mean 
 is then considerably reduced, and the work of a motor comparatively 
 easy. This latter is the system generally followed --in electric generating 
 central stations (fig. 126). On shipboard such an installation might be 
 
 Fig. 12(3.- — Motor-driven Air-pumps (Edwards) 
 
 driven by a low-pressure turbine taking its steam from the centrifugal pump 
 cylinder exhaust, or from the feed-pump when operated by a cylinder, as in 
 the Weir type. 
 
 Size of Air-pump. — The capacity of the air-pump should be calculated 
 from consideration of the conditions under which it is to work, and by the 
 rules given in this and the preceding chapter, and suited to practice by an 
 allowance made for the efficiency of the pump. If the pump is single-acting 
 and well-designed, and is working under favourable conditions, its efficiency 
 may be taken at 06 ; and if the reverse of this 0*4 ; generally its efficiency is 
 about 05, so that the size in practice should be double that given by theo- 
 retical calculation. The efficiency of the double-acting pump varies from 
 0-5 to 0-3, and generally is not more than 0-35 ; its size for good working 
 should be nearly three times that given by theoretical calculations 
 
384 MANUAL OF MARINE ENGINEERING. 
 
 Hence, when the temperature of the sea is 60°, and that of the (jet) con- 
 denser is 120°, Q being the volume of the cooking water, and q the volume of 
 the condensed water in cubic feet, and n the number of strokes per minute. 
 
 The volume of the single-acting pump = 2*74 ( J. 
 
 The volume of the double-acting pump = 4 ( J. 
 
 The Capacity of Air-pumps for a surface-condenser is not easy to determine, 
 inasmuch as, if the condenser is tight and the stuffing-box packings of the 
 main and auxiliary engines in good working order, they have little work to 
 do beyond removing the condensed water after a vacuum has been established. 
 Quite a small pump, therefore, should suffice for the running of an engine 
 when in good working order. It is true that at starting such a small pump 
 will take longer to exhaust the condenser and form a good vacuum, but the 
 vacuum, when formed, will be maintained at less cost than with a bigger pump. 
 But in e very-day practice it would be found that with very small air-pumps 
 there are considerable, though on]y temporary, variations in the vacuum 
 due to leakages of the. glands of the L.P. members, to admission of fresh 
 make-up feed-water or to leakage from inattention to the steering engine, or 
 some other auxiliary exhausting to the condenser. If, therefore, the air- 
 pump is worked by the main engine, it must be of sufficient size to form a 
 good vacuum in a short time, and to restore it quickly if temporarily reduced. 
 It should be remembered also that the pump which is barely sufficient for 
 that purpose may be ample at the reduced speed of service, inasmuch as 
 the capacity of the pump is falling off only in direct proportion to the number 
 of revolutions, while the steam consumption is being reduced as the cube 
 of the revolutions, so that at § speed while the capacity is only 25 per cent, 
 less the steam condensed is 58 per cent. less. On the other hand, it is quite 
 likely that the pump is not so efficient at the lower speed, due to air leakages 
 in it and at the glands of the main engines. If the air-pump is driven by 
 its own independent steam engine or motor, it may automatically work 
 at the lowest speed compatable with good vacuum, and, therefore, this 
 system is the better one for all installations of large reciprocators, notwith- 
 standing that power derived from them is generated much cheaper than is 
 possible with any auxiliary. With turbines the separately driven air-pump 
 is a necessity. 
 
 The Air-pump for a Jet-Condenser could be calculated from the conditions 
 under which it was to work and on first principles. It was, however, the 
 common practice of the early marine engineers to make the vertical single- 
 acting air-pumps of paddle and geared-screw engines with a diameter half 
 that of the steam cylinder, and a stroke half that of the engine, so that the 
 rule really was — 
 
 Capacity of a single-acting air-pump = one-eighth that of the steam 
 
 cylinder. 
 
 With the advent of the horizontal engine, and its double-acting pump, 
 engine builders varied in their practice somewhat, but on the average the — 
 
 Capacity of horizontal double-acting air-pump = 0*09 the capacity of the 
 
 steam cylinder. 
 
AIR-PUMPS. 385 
 
 The diameter being about 0*3 that of the engine piston, and the stroke the 
 same as it was driven direct from it. 
 
 The Air-pump of Surface-condensers was made at first almost as large 
 as that of a jet-condenser partly for fear of leakage, but chiefly because 
 these early surface-condensers were fitted so as to be worked on the jet or 
 direct-contact principle. Confidence in the surface-condenser, begotten from 
 experience, has caused all these unnecessary appurtenances and provisions 
 to be swept away, and with them went the big air-pump, so that now, with 
 pumps operated by the main engines, their capacity when single-acting and 
 vertical is very much less than formerly. Hence 
 
 The capacity of S.A. air-pump = O04 that of the L.P. cylinder. 
 
 That is, the ratio of L.P. to pump is 25. 
 
 For express steamers making short voyages, such as crossing channels, 
 or trips along rivers and coasts, where high speed must be attained quickly 
 and maintained for short periods, it may be 21. For cargo steamers* or 
 passenger steamers making long voyages there is no need of so large pumps, 
 and the ratio in them may be as high as 30. 
 
 Air-pumps operated by Independent Means should be such that, when 
 working at full speed, they can quickly form a high vacuum, and when formed 
 may slow down to that only necessary to maintain it and carry away the 
 condensed water. The guide, to arrive at their proper capacity, as really 
 also for determining that of the engine driving them, is based on the quantity 
 of water consumed. Under good conditions a vacuum of 26 inches can be 
 maintained by a pump whose capacity is equal to 0*3 cubic foot for each 
 pound of water condensed and passing through it per stroke for general 
 working ; but to provide for casual leakages 05 to 06 cubic foot is a more 
 satisfactory allowance, while for higher vacuum, say 28 inches, 0\8 cubic 
 foot should be provided, and for higher vacua and economic engines as much 
 as 1 cubic foot will not be too much for rapid formation and steady maintenance. 
 
 Example. — An engine has a 50-inch L.P. cylinder, and when working 
 at 100 revolutions per minute develops 1,200 I.H.P., condenses 15 lbs. 
 of steam per I.H.P. per hour; the stroke of the piston is 3 feet, and that 
 of the lever-driven air-pump is 1 foot, what diameter should it be for a 
 vacuum of 28 inches ? • 
 
 Here the consumption per stroke = -^- —— = 3 lbs. 
 
 11 60 X 100 
 
 The capacity of the pump = 3 x 0'8, or 2*4 cubic feet. 
 The diameter is, therefore 
 
 y 
 
 2*4 square feet X 144 _. . , 
 
 ^ n^oE-, = 21 inches. 
 
 0- 1 8o4 
 
 If there had been three self-acting pumps independently driven at 150 
 revolutions with a stroke of 9 inches — 
 
 The capacity of each pump = -5 r-p— = 0533 cubic foot. 
 
 Area of each bucket = - _. . or 0"71 sq. foot, or 102 sq. inches. 
 
 Owo ^ ^ 
 
 The diameter of each is, therefore, 11-5 inches. 
 
 • N.E. Coast Inst. E. and S. recommend 20 when a steam ejector is fitted to the condenser 
 
 25 
 
386 MANUAL OF MARINE ENGINEERING. 
 
 The stroke of the vertical air-pump is usually from one-third to one-half 
 that of the engine, but it should be such that the velocity of bucket does 
 not exceed 300 feet per minute, and for continuous running 275 feet. 
 
 The vacuum in the condenser is liable to be spoiled by air leakage at the 
 expansion joint of the exhaust pipe. It is usual to provide for the expansion 
 of this pipe, which, in the Navy, is of brass or copper, by fitting a stuffing- 
 box, gland, etc., on the condenser or cylinder into which the pipe end is 
 free to work ; unless this box is deep, and has a good taper at the bottom, 
 and the part of the pipe fitting in the stuffing-box of brass and turned to 
 fit, it is difficult to keep it quite tight. Some engineers find it better to 
 fit a brass or copper bellows joint. In compound engines some trouble is 
 often experienced from leakage through the drain "cocks of the low-pressure 
 cylinder, and since, when the engine is working below full speed, the pressure 
 in this cylinder is less than that of the atmosphere as a rule; to prevent 
 this, pipes should be led from them to the condenser. 
 
 The Function of a Perfect Air-pump System is to effect, at the cost of a 
 minimum expenditure of energy : — 
 
 («) The highest possible degree of rarefication of the vapour space of a 
 condenser under all conditions of air leakage likely to be met with in practice. 
 
 (b) The extraction from the condenser, without thermal loss, of the water 
 resulting from the condensation of the steam. 
 
 In the Kinetic System the air and non-condensable gases are removed from 
 the condenser by the action of a steam jet followed by a special system of 
 water jets known as the " Kinetic Ejector" (v. fig. 127). 
 
 Jots of water moving at a high velocity through suitably shaped orifices 
 have been used for many years for the purpose of producing a partial vacuum 
 for various purposes, including the rarefication of the vapour spaces of 
 condensers. 
 
 A water jet has of itself, however, a relatively low capacity for assimilating 
 and withdrawing air ; but the air-withdrawing capacity of any water jet 
 device is greatly increased if the air to be extracted is previously mixed 
 with steam. 
 
 In the case of the kinetic plant the steam is introduced into the air suction 
 pipe through* a high-velocity nozzle, thus entraining the air and intimately 
 mixing with it. This steam is condensed on. the secondary sprays of the 
 kinetic ejectors, and the resulting liquid carrying with it all occluded air 
 and gases is ejected to the atmosphere by the main water jet. 
 
 The steam jet is fed with live steam in the case of a condenser fitted 
 with electrically-driven pumps, or with exhaust steam at a pressure of about 
 20 lbs. above the atmosphere, if this be available. The quantity of steam 
 required varies according to the amount of air leakage and the size of the 
 engines, but from h to £ of 1 per cent, of the steam to be condensed may 
 be taken as an average figure for installations of considerable dimensions. 
 
 The water for the kinetic jet is that of condensation which has already 
 been removed from the condenser, and the whole of the latent heat contained 
 in the steam used in the steam nozzle is absorbed by this water, which is 
 subsequently discharged to the feed tank at a correspondingly higher tem- 
 perature than when it left the condenser. 
 
 The water of condensation is withdrawn from the condenser and discharged 
 
AIR-PUMPS. 
 
 3S7 
 
 against the pressure of the atmosphere by the action of two pumps of the 
 centrifugal type, known as the " head " pump and " pressure " pump respec- 
 tively. The " head " pump works under the condenser pressure both on 
 the suction and on the delivery sides, and is designed so that it will pass the 
 required quantity of water with an extremely low head on the suction side 
 of the pump. The water is discharged from this pump into a stand pipe or 
 receiver, which provides a natural head of water on the inlet side of the 
 " pressure " pump, by which the water is finally ejected against atmospheric 
 pressure. This arrangement makes it possible to place the pumps only a 
 few inches below the level of the condenser bottom, and to maintain a per- 
 fectly regular discharge at all loads, the amount corresponding to the quantity 
 of steam condensed. 
 
 Ordinary centrifugal pumps have been successfully used for the with- 
 drawal of the water of condensation from condensers both of the surface 
 and jet types, but for satisfactory operation these pumps must invariably 
 
 9. Pressure pump discharge to tank. 
 
 10. Non-return valve. 
 
 11. Feed-water delivery pipe. 
 
 Fig. 127.— Condenser with Kinetic Air-pumps 
 
 1. Air suction orifice on condenser. 
 
 2. Exhaust steam jet. 
 
 3. Air-pipe to kinetic ejector. 
 
 4. Kinetic ejector. 
 
 5. Suction pipe to kinetic pump. 
 
 6. Kinetic pump to discharge pipe. 
 
 7. Condensed water pipe to head pump. 
 
 8. Stand pipe between head and pressure 
 
 pumps. 
 
 12. 
 13. 
 14. 
 15. 
 16. 
 
 Float-controlled feed delivery valve. 
 
 Kinetic tank. 
 
 Pressure equalising pipe. 
 
 Exhaust steam to jot. 
 
 Surplus exhaust steam 
 
 be placed at a considerable distance below the level of the condenser, a 
 condition which is difficult, sometimes impossible, of attainment on ship- 
 board. 
 
 For marine use single centrifugal pumps are apt to work very irregularly, 
 due to changes in the relative levels of the condenser and pump consequent 
 on the motion of the vessel. 
 
388 
 
 MANUAL OF MARINE ENGINEERING. 
 
 It is claimed for the kinetic plant that the vertical stand pipe between 
 the "head " and "pressure" pump maintains a practically constant head 
 on the latter irrespective of the motion of the vessel, and that the energy 
 lost from the system in a normally-designed installation does not exceed 
 0-0003 (or three ten-thousandths) of the energy developed by the total steam 
 which is condensed. This being so, the thermal efficiency of the kinetic 
 system of air and water-extracting pumps is excellent. 
 
 It will be noted that the whole of the energy in the steam jet, also that 
 required to drive the air and water-extraction pumps, is returned, to the 
 system, except that required for the extraction of the water, and that for 
 the compression of the extracted air against atmospheric pressure. 
 
 With these exceptions, which have been proved by experiment to be 
 practically negligible, the whole of the energy expended reappears in the 
 form of heat in the kinetic tank, whence it is returned to the boilers, subject, 
 of course, to such small losses as arise through radiation from exposed surfaces. 
 
 By referring to fig. 127, it will be seen that rarefication of the condenser 
 is effected by steam jet (2) followed by the action of kinetic ejector (4) supplied 
 with water by the kinetic pump through pipe (5), and discharging into the 
 kinetic tank (13). 
 
 Fig 127a. 
 
 Water of condensation flows through pipe (7) into the head pump, whence 
 through stand pipe (8) to the pressure pump, and thence through pipe (9) 
 and non-return valve (10) into the kinetic tank. The excess of water in the 
 tank corresponding to the feed-water is delivered by the kinetic pump through 
 pipe (11) and float controlled valve (12) into the feed tank, which may be 
 placed overhead. 
 
 Several Types of Rotary Air-pumps have been introduced to steam users 
 with varying success from time to time, the majority of such pumps making 
 use of the action of centrifugal force on water particles moving in small 
 channels in the rotating part of the pump, or projected at a high velocity 
 in thin films across the air suction chamber, the air beino; removed by becommg 
 
PUMP RODS. 389 
 
 imprisoned between these moving particles or films. The disadvantages of 
 many examples of this type of pump are said to be that : — 
 
 The tendency to rapid clogging and choking of the small passages with dirt or oil, 
 necessitating the frequent dismantling of the pumps for cleaning. 
 
 Extreme sensitiveness of the pumps to sudden changes of pressure in the condenser. 
 
 Great sensitiveness to changes of the water temperature. 
 
 Inability to deal satisfactorily with any serious increase of air leakage above the 
 normal figure. 
 
 Low thermal efficiency, all frictional expenditure of energy in the moving parts 
 being generally lost to the system. 
 
 Parsons' Vacuum Augmentor was the means devised by Sir C. Parsons 
 to obtain the high vacua so conducive to the high efficiency of the turbine, 
 and so necessary when a L.P. turbine is added to a reciprocating engine to 
 increase the economy of power generation. 
 
 The vacuum for brake purposes on a railway train is produced by thef 
 same means as here provided for getting rid of the gaseous products of a 
 condenser ; and Mr. Rodger thirty years ago introduced the " vacuum 
 blower " for the purpose of forming a vacuum in the condenser of a com- 
 pound engine, to ensure the easy handling of that engine. The adaptation of 
 these means for exhausting a condenser was made by Parsons with success, 
 and he is able to obtain quickly, and maintain steadily, quite a high vacuum, 
 and higher than possible for the same air-pump to attain without it. 
 
 By reference to fig. 127a it will be seen that the air-pump is placed well 
 below the condenser, so that there is a column of water between them, which 
 acts as an air seal. The air ejector A draws from the condenser bottom 
 free from condensed water, and discharges into a small auxiliary condenser 
 C, whose surface is only about 5 per cent, that of the main condenser. In 
 this condenser the gaseous products are cooled and freed from water vapour 
 which is thrown down and drains away to the air-pump, to which also the 
 cooled gases are projected. The difference in level of the pump bottom and 
 condenser bottom, 3 to 4 feet, causes a difference of 1| to 2 inches of vacuum. 
 The ejector being always in operation, the condenser is kept constantly free 
 from gaseous matter, which interfere so seriously with the efficiency of the 
 cooling surface when present with water vapour. By such means, not only 
 is the vacuum maintained high, but the cooling water required is the minimum 
 quantity when the efficiency of the surface is a maximum. 
 
 Pump Rods. — The vertical single-acting pump in oscillating paddle-wheel 
 engines is generally worked by a connecting-rod from a crank, or by means 
 of a large eccentric, whose rod passes through a trunk cast with the bucket, 
 and connected to a socket fitted to the bucket, and secured by a brass cap- 
 nut underneath, or by means of a lever, as in fig. 20. 
 
 Area of section of the rod = 0-01 x area of pump bucket. 
 
 Or in case of a round rod 
 
 The diameter of rod = 0*1 X diameter of pump. 
 
 When two bolts are fitted to connect the brasses, etc.. at the end, 
 
 The diameter of each bolt in the body = *056 x diameter of pump. 
 
 When the air-pump is of the single-acting type in screw engines, it is 
 generally worked by means of a rod, either of one of the zinc bronzes 
 
390 MANUAL OF MARINE ENGINEERING. 
 
 rolled or iron cased with brass.* having a tapered end fitting into the piston 
 or bucket, and secured by a nut on the top side, or tapered the other way 
 and secured by a cap-nut of brass (fig. 116) on the bottom side. 
 
 The diameter of the rod = *15 X diameter of pump. 
 
 It is no unfrequent thing, though, in vertical engines, to find the pump 
 worked by a connecting-rod and trunk as in paddle-wheel engines, the motion 
 being derived from levers worked from the piston-rod crosshead in the same 
 way as for pumps with fixed rods. 
 
 The arrangement of the pumps is various as regards detail, but for vertical 
 screw engines is generally alike in principle, and may be divided into three 
 systems — (1) Air-pump and circulating pump, both single-acting, and each 
 worked separately by levers, etc., from a piston-rod crosshead ; or air-pump 
 and circulating pump side by side (the latter being either single or double- 
 acting), and worked by levers from one piston-rod crosshead ; or a single- 
 acting plunger circulating pump inverted over the air-pump, and having a 
 common rod, crosshead, levers, etc. (2) Air-pump and circulating pump 
 worked each from a piston direct, or from the same piston-rod crosshead. 
 (3) Both pumps worked by eccentrics on the crank-shaft, or by a crank-pin 
 on the forward end of the crank-shaft. (4) The air-pump worked in one of the 
 ways above described, and the circulating pump a centrifugal driven by an 
 independent engine. (5) The air-pump detached and worked by an inde- 
 pendent cylinder or engine, and the circulating pump centrifugal and worked 
 by another independent engine. (6) The # air-pump and circulating pump 
 combined and worked by the same independent engine, the latter being 
 centrifugal direct and the former by gearing. 
 
 The first and fourth systems are the ones most generally adopted, the 
 sixth is specially suitable with turbines. 
 
 When the two pumps are worked independently of each other, each from 
 a crosshead by levers, the strain from the pumps is distributed, and their 
 weight tends to balance the pistons ; but per contra the expense of working 
 parts is increased, and the number of parts requiring attention and liable 
 to accident is doubled ; more space is required, and, as the low-pressure 
 piston requires the greater counterbalance, it is an advantage to have both 
 pumps worked from it. 
 
 If the two pumps are side by side, their rods are connected to a common 
 crosshead, etc., thus leaving the space on the forward or after side of them 
 free for the condenser, and admitting of more elasticity in its design ; and, 
 as is generally the case, the pumps are worked from the after cylinder, the 
 condenser is easy of access, and there is plenty of space for the tubes to be 
 drawn in a fore and aft direction. 
 
 The circulating pump inverted over the air-pump is a very convenient 
 arrangement for small engines ; but as the strains from the circulating pump 
 have to be taken by the stand pipes from the pumps and the condenser top, 
 it is found to give trouble in large engines ; also, when the ship is light, and the 
 valves out of order, it fails to pump well ; but a large amount of space is saved 
 by this plan, and it was at one time very frequently found in small ships. 
 
 The piston-buckets are made of bronze (in one thickness generally) 
 stiffened by from 4 to 8 ribs ; the circumference is either recessed down, 
 * Non-corrodible steel can now be obtained quite suitable for pump rods. 
 
PUMP BUCKETS. 391 
 
 so as to admit of gasket being coiled around it, or else it is formed like a 
 stuffing-box, and fitted with a gland or junk-ring, secured by studs and nuts, 
 which pass through lugs cast with it, so as to jam the packing tight after 
 it has been placed in. 
 
 Air-pump3 are sometimes fitted with spring rings and junk-rings, similar 
 to those of a steam-piston, but made of gun- metal instead of cast iron ; this, 
 however, has been but rarely done, and then only in very large pumps, as in 
 ordinary cases hemp packing serves the purpose very well, and is much 
 cheaper. Vertical pumps will work sufficiently well without packing ; but, 
 when required, bronze Ramsbottom rings answer the purpose quite well. 
 
 The diameter of the horizontal air-pump rod = -15 X diameter of the 
 pump -f J inch. 
 
 The rods working in this pump are of hard rolled bronze or Muntz metal; 
 the Admiralty now prefer the hard rolled bronze, and have, therefore, struck 
 the latter out of their specifications as being too soft for continued wear. 
 When the rods are of large size, they are sometimes made of wrought iron, 
 cased with gun-metal, but unless the rods are large and long, it does not pay 
 to case them, except for the sake of the harder surface (v. footnote, p. 390). 
 
 The rod fitting into the steam-piston is, of course, of wrought steel, and. 
 is connected to the brass rod by a box-coupling and cotters, the socket being 
 formed with the steel rod, in order that the piston-bucket may be drawn 
 out from behind for examination without removing the rod. 
 
 The pump-rod is usually fitted into the bucket in a similar manner to that 
 of a piston-rod, the taper at the end being about 1 in 12. 
 
 Pump Buckets. — The following Tules give the dimensions of an ordinary 
 pump bucket, of which fig. 128 is an example : — 
 
 x =0-3 x VD-\- 0-15 inch. 
 
 D is the diameter of the pump in inches. 
 
 The thickness of the disc when solid = l'O X r. 
 
 ,, ,, ,, perforated = 1*7 X x. 
 
 ,, „ flanges at edge = 1*1 X x. 
 
 „ „ metal around rod end = 1'5 X i. 
 
 „ „ ,, the rim = 1*0 X x. 
 
 ,. „ packing - H X i 
 
 „ ribs = 0-8 X x. 
 
 The breadth of the packing = 4-0 X x. 
 
 „ depth of bucket at the middle = 6-0 X x. 
 
 ,, number of ribs, one for each 4 inches of diameter. 
 
 The bronze liner is usually from h inch to f inch thick, or its thickness 
 = 1-1 x — 0-2 inch. 
 
 This liner is let into the cast-iron bottom, and made to fit tight at both 
 ends, facings on both liner and bottom being formed for that purpose ; these 
 facings are made with a very slight taper, so as to allow of a little draw on 
 fitting in ; the liner is then secured by brass screws passing through and 
 through. The pump-barrels of vertical pumps are usually of bronze, f to 
 1 inch thick, secured to the foundation or pump bottom by flanges cast with 
 them ; the head valve-boxes, etc., are also borne on them, and secured by 
 flanges in a similar manner (v. fig. 122). 
 
392 MANUAL OF MARINE ENGINEERING. 
 
 Pump Valves. — The arrangement of the valves of an air-pump is of the 
 highest importance, for, as has been stated, the efficiency of the pump to a 
 very large degree depends on it ; especially is this the case with modern 
 compound engines, where but a small amount of water is pumped at each 
 stroke. A triple-compound engine indicating 2,000 horse-power uses about 
 8 cubic feet of water per minute, so that for a single-acting pump, with 
 100 revolutions of the engine, the amount per stroke is only 138 cubic inches ; 
 in practice, however, the pump fails frequently to take this amount, and 
 so, until it accumulates behind the foot valves, no water enters the pump. 
 Aga'n, a double-acting pump is seldom so efficient at the front end as it is 
 at the back, and, consequently, until there is a glut of water, the front ceases 
 to pump. From causes of this kind trouble was often experienced with the 
 double-acting pump of surface-condensers, and a remedy that has answered 
 very well indeed in obviating this is to connect the space between the valves 
 to the hot-well by a pipe about - 15 the diameter of the pump, having a suction 
 or non-return valve fitted to it, so that the pump is to some extent continu- 
 ously charged with water. 
 
 There are certain conditions that should be most carefully observed in 
 arranging the valves, and they are — 
 
 (1) The foot valve should be in such a position that the water readily 
 drains by gravity from the condenser into the pump. 
 
 (2) The valves should be designed so as to open with the least possible 
 pressure under them, and still readily close on the return stroke of the pump. 
 
 (3) The flow from the condenser to the hot-well should be as direct as 
 possible — that is, there should be as little change as possible in the direction 
 — and no obstacles, so that the velocity generated at leaving the condenser 
 should not be checked until the water is in the hot-well. 
 
 (4) The space between the head and foot valves of a double-acting and 
 the head and bucket valves of a vertical single-acting pump should be a 
 minimum, and all recesses, etc., where air can lodge or eddies can form, 
 should be avoided without fail. 
 
 To drain the condenser efficiently, the foot valve should be below its 
 lowest level, and, when possible, the channel from it to the pump inclined so 
 that the water naturally flows in that direction ; all stiffening ribs, etc., 
 should be placed on the outside, so as to avoid gutters in which the water 
 can lie. 
 
 The foot valves can only open by the pressure under them being greater 
 than that on them, and when, as in the case of surface-condensers, the head 
 of water is not sufficient to do so, it depends entirely on the pump forming a 
 better vacuum between the foot and head valves than there is in the con- 
 denser. Now, if there be a good vacuum in the condenser, it requires the 
 full efficiency of the pump to produce a better one between the foot and head 
 valves. At most there can be but little difference between the pressure in 
 the condenser and that in the pump-chamber, and if the valves have much 
 resistance in themselves they will cease to act. The full force of this is 
 better appreciated when it is remembered that the pressure in the pump- 
 chamber must vary from slightly above atmospheric pressure, when the head 
 valves are opened, to below that in the condenser in order to open the foot 
 valves. For this reason, if the space between the head and foot valves is 
 
PUMP VALVES. 393 
 
 more than one-fifteenth the capacity of the pump, it is impossible to obtain 
 a good vacuum in the condenser. This clearance space is virtually reduced 
 by filling it with water ; without such virtual reduction very few horizontal 
 air-pumps would be capable of maintaining over 20 to 22 inches of vacuum 
 in the condenser. 
 
 The makers of the horizontal air-pump usually arrange the foot valves 
 so that they are either on a vertical plane seating or on one inclined so as 
 to allow of the valves opening easily. A brass box, having valves on five 
 of its sides, and bolted to the condenser by the sixth, so as to cover the outlet 
 orifice from the condenser to the pump, was sometimes used. In this way 
 the clearance space is partly filled, and five valves can work in the space 
 usually oc cupied by one ; three of the valves are in vertical planes and open 
 horizontally, the two others are in horizontal planes, one opening upward 
 and one downward. 
 
 The head valves should also be carefully arranged, for although not so 
 difficult to place as are the foot valves, a little want of forethought may 
 cause a badly working pump. They should always be so arranged that water 
 lodges on them, so that in case they leak only -water, and not air, escapes 
 back into the pump ; the valve seats and parts adjacent should be free from 
 pockets underneath in which air can collect, and for this purpose they should 
 be in the highest part of the pump-chamber, and if any part is partitioned 
 off by a stiffening web or fillet, a communication should be made from its 
 top to the underside of the head valves. 
 
 Since mineral oil has been used as the lubricant for cylinders, ordinary 
 india-rubber ceased to be used for air-pump valves, inasmuch as that oil is 
 a solvent of the rubber. Many attempts have been made to manufacture a 
 rubber which will withstand the oil, but none of them have been perfectly 
 successful, and only a few have succeeded in preventing the rapid action 
 noticeable on nearly pure rubber. Many engines are now kept going with 
 no supply of oil internally beyond what enters with the piston-rods. 
 
 Vulcanite Valves. — India-rubber can, by a certain process, be converted 
 into a hard black substance resembling ebony, and hence sometimes called 
 ebonite ; this substance is light and very strong, and quite impervious to oil, 
 and has been used with great success for air-pump valves. The valves 
 made of vulcanite (fig. 128a) are generally circular and flat, strung on a 
 stud through the middle, on which is a flat brass guard of the same diameter 
 as the valve, and against which the valve lifts bodily. The wear of these 
 valves is very little indeed. Valves of fibre are also fitted in this way. 
 
 Metallic Valves. — In the old days the air-pump valves were usually of 
 cast brass, and of large size, the foot and delivery valves being hinged or 
 " flap " valves, and the bucket valve annular. The flap valve is still employed 
 in cargo ships where the foot valve is not accessible or easily got at, and 
 answers the purpose very well indeed ; it should be of ample size, and have 
 very little lift. 
 
 Coe & Kinghorn's Patent (fig. 129) consists of tongues made of very thin 
 rolled sheet phosphor bronze. These tongues or flaps cover a grating in the 
 same way as india-rubber, and are fitted with curved guards to admit of a 
 gradual bend. These valves work very well, but great care is necessary in 
 making and setting the guard, so that when the valve is open there is no 
 change of flexure or angle on which the flap can work and gradually break. 
 
394 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Thompson's Patent (fig. 130). — A large number of small saucer-shaped 
 valves, of phosphor or manganese bronze, cast very thin indeed, are some- 
 times fitted, and, being very strong and comparatively light, do the work 
 very well. Single discs of thin sheet-brass or phosphor bronze have also 
 
 Fig 12S. 
 
 Fig. 128a. 
 
 Fig. 130 —Annular Valve. 
 
 Fig. 13-: 
 
 Fig.t. 128-132. — Air-pump Buckets, Valves, etc. 
 
PUMP VALVES. 395 
 
 been used for this purpose with success. Thin discs of brass are often used ; 
 they are inexpensive and work very well. Discs of fibre also answer very well. 
 
 The valve shown in fig. 131 is simply a thin, metallic disc without stiffening 
 of any kind. When very little lift is given to this valve it works very well, 
 and rarely gets out of shape sufficiently to be seriously inefficient. It is 
 sometimes fortified by a second and a third thin disc, each one smaller than 
 the one on which it is superposed, and having holes through it on a circle 
 of less diameter than those in the disc below it. In this way the lower disc 
 permits water to flow around its edge and through the holes in it, and when 
 closed the holes in it are covered by the next disc above it. 
 
 Beldam's Patent Valve (fig. 132). — In this case the valves are made of 
 very thin sheet metal and stiffened by a series of circular corrugations. This 
 valve has answered the purpose very well, for, while mantaining an even 
 flat surface to cover the sea, it is elastic and can give under an unfair pressure 
 and return to its normal form without permanent distortion. 
 
 Area through Valve Seats. — The area of opening past the valves depends 
 on the size and velocity of the pump, and should not be less than will admit 
 the full quantity of water for jet-condensation at a velocity not exceeding 
 400 feet per minute. In actual practice the area was generally in excess of 
 this. If the foot valves are large they will be sluggish in action ; if they 
 are small the velocity of water, etc., will be sufficiently high to raise the 
 valves and keep them open by the energy of the particles striking them ; 
 this argument especially applies to the pumps of surface condensers. It 
 may be noted that the vertical pumps of the old jet-condensing engines 
 were frequently without foot valves. 
 
 Suction Pipes from the Condenser to the Air-pump of marine engines have 
 in the past been made often unnecessarily large. Experiments have shown 
 that very, little reduction in vacuum is made when they have been throttled 
 down to quite small areas of cross-section, in one case to one-tenth of the 
 original without any reduction. If D is the diameter of the air-pump in 
 inches, and S the stroke in feet, and R the number of revolutions — that is. 
 the number of lifts per minute — then 
 
 Diameter of suction pipe in inches = - \/S X R. 
 
 Diameter of delivery pipe in inches = — J$ x R. 
 
 If there were no fear of a glut of water coming over at any time from the 
 air-pump its delivery pipe might be much smaller, but there is always that 
 danger from the tubes splitting, corroding, or drawing out of place, or in 
 a seaway with a list there might easily be such a glut from many condensers. 
 The mall suction pipe is, therefore, a means of safety at all times. 
 
 D 2 x Z . 
 The area through the foot valves when fitted = „ in square inches. 
 
 D 2 x Z 
 
 Area through the bucket valves „ — -p,->; - » 
 
 Area through the head valves 
 
 D 2 xZ 
 
 "1.050 " 
 
 Z is the speed of pump in feet per minute = 2S X R. 
 
396 MANUAL OF MARINE ENGINEERING. 
 
 An air-pipe should always be fitted to the hot-well, as high up as possible, 
 and its diameter should be £ that given above. 
 
 The bucket of a single-acting pump must be sufficiently large to admit 
 of a valve area through it not less than given above. To obtain this, they 
 are better designed with a short stroke, although the perimeter and the 
 source of leakage is thereby greater. It is usual now and certainly advisable 
 to deliver the water from the air-pump into a tank about 12 times the capacity 
 of the pump ; the delivery pipe is in diameter (>7 that given above. The 
 air-pipe of the tank should be about half the area in section. The tank 
 should have an overflow pipe and a glass gauge. The suction for the feed- 
 pumps should be near the bottom and fitted with a strainer. 
 
 Circulating Pumps. — Two kinds of pumps are employed to circulate the 
 cooling water in the condenser of the mercantile marine — the ordinary single- 
 or double-acting reciprocating pump, and the rotary pump. The single- 
 acting pump is usually fitted to small engines, and the double-acting to 
 larger ones. The latter is preferable to the former, but more expensive. 
 They are generally worked by the main engines, but sometimes by inde- 
 pendent ones. 
 
 Single-acting Pump. — This is used in vertical engines, and is similar tc 
 the single-acting air-pump already described. It works very well, and 
 when provided with an efficient air-vessel, the flow is fairly steady. In very 
 small engines a plunger-pump is used, and formerly inverted plunger-pumps 
 were fitted to engines of considerable power. A good pet-valve, which will 
 admit air to the pump, but not allow the water to pass out, should be fitted ; 
 and likewise a pass-valve, which opens a communication between the delivery 
 and suction, is a requisite ; the former prevents noise by providing an air 
 cushion for the water, and the latter checks the supply without straining 
 the pump. 
 
 Double-acting Pumps. — These give a steadier How of water, and cause 
 less shock and strain on all the working parts, pipes, etc.. than do the single- 
 acting pumps; but even these should be fitted with pet-valves and pass- 
 cocks or valves. 
 
 Size of Circulating Pump. — The capacity of this pump depends on the 
 quantity of cooling water and the number of strokes per minute. 
 
 Let Q be the quantity of cooling water in cubic feet, and n the number 
 of strokes per minute, and S the length of stroke in feet. 
 
 Capacity of circulating pump = — cubic feet. 
 
 Diameter „ „ = 13-55 x / — finches. 
 
 Example. — To find the diameter of a double-acting circulating pump of 
 an engine condensing 2 cubic feet of water per minute, and requiring 40 times 
 the amount of cooling water ; the stroke of the pump is 18 inches, and the 
 number of revolutions 120 per minute. 
 
 Here Q = 40 x 2 or 80 cubic feet ; n = 120 X 2 or 240. 
 
 80 
 
 -rr- — = 6-4 inches. 
 
 240 x l'o 
 
CIRCULATING PUMP VALVES. 
 
 397 
 
 The size of the circulating pump is to a large extent dependent on the 
 same conditions that determine the size of the air-pump, and may, therefore, 
 bear a constant relation to the size of the air-pump ; and since the size of 
 the air-pump is often determined by the size of the cylinders, that of the 
 circulating pump may be found in a similar manner. When the circulating 
 pump is single-acting, the capacity should be 0'038 of that of the low-pressure 
 cylinder, and when the circulating pump is double -acting, O021. 
 
 TABLE XLV. 
 
 Description of Pump. 
 
 Description of Engine. 
 
 Ratio. 
 
 Single-acting, 
 
 Expansive 1£ to 2 times, . 
 
 13 to 16 
 
 »» >> * 
 
 „ 3 to 5 ,, . . 
 
 20 to 25 
 
 t» .. 
 
 Compound, triple, and quadruple, . 
 
 25 to 30 
 
 Double- ,, 
 
 Expansive 14 to 2 times, - 
 
 25 to 30 
 
 i> i> ■ 
 
 ,, o tO , , - - - • 
 
 36 to 46 
 
 »j ?> * 
 
 Compound, triple, and quadruple, - 
 
 46 to 56 
 
 For high vacuum in the tropics the sizes should be 50 per cent, larger. 
 
 When the air-pump is double-acting, the capacity of the double-acting 
 circulating pump should be 052 of that of the air-pump, the double-acting 
 circulating pump being more efficient than the double-acting air-pump. 
 
 Table xlv. gives the ratio of capacity of cylinder or cylinders to that of 
 the circulating pump. 
 
 Circulating Pump-Rods are made of the same materials, and in the same 
 way as for the air-pump, and when possible the rods of both pumps are 
 made identically alike, so that one spare rod serves for both. 
 
 Diameter of circulating pump-rod = 022 X diameter of pump. 
 
 When the pump is double-acting, or of comparatively long stroke, 
 = 022 X diameter of pump -f- \ inch. 
 
 Circulating Pump Bucket. — When single-acting it is similar to that of 
 the air-pump, and when double-acting it is simply a piston of brass. 
 
 Although it is usual to pack these pumps with hemp gasket or bronze 
 rings, there is no necessity for this, since the water flows freely into the 
 pump by gravity, and the pump moves too quickly to allow of much leakage 
 past the piston. Many engineers now dispense with packing, and simply 
 make the piston a fairly good fit in the barrel : while others turn either a 
 spiral groove, or a series of parallel grooves, on the edge of the piston, which 
 has the effect of keeping the surface well lubricated and preventing leakage 
 when working. The friction of the unpacked pump is considerably less than 
 that of a packed one, and in fast-running engines this is no slight considera- 
 tion. It may be taken now as certain that it is better not to pack the circu- 
 lating pump of a fast-working screw engine, and packing is of doubtful 
 advantage when the engine is slow-working, provided the pump is below 
 the water-line. 
 
 Circulating Pump Valves. — These are almost always of the best india- 
 rubber, and of the quality known as " floating," from the fact of the specific 
 gravity of rubber, with only such slight admixture of foreign matter as to 
 render it usable, being less than that of water. 
 
398 MANUAL OF MARINE ENGINEERING. 
 
 Valve Area. — The clear area through the valve seats and past the valves 
 should be such that the mean velocity of flow does not exceed 450 feet per 
 minute. It is not always easy to obtain so large an area, but when the 
 velocity of flow is high the valves wear out more quickly, and the resistance 
 of the pump is considerable, so that every effort should be made in this 
 direction. 
 
 The pipes should be of such a size that the mean velocity of flow* ihrough 
 them does not exceed 500 feet per minute when comparatively small, and 
 long for their size ; when large, and having fairly easy leads, the allowance 
 may be as much as 600 feet. The suction pipe of a double-acting pump should 
 be of the same size as the delivery pipe, but it may be considerably smaller 
 when the pump is single-acting and below the water-level. 
 
 Diameter of Suction Pipes for circulating pumps may be obtained as 
 follows, when the cooling water per I.H.P. is not more than 8 lbs. per 
 minute, which is sufficient for 26 inches in the tropics and 28 inches in 
 temperate zones, and the velocity of flow as above : — 
 
 Diameter in inches = »/ * * + \ inch. 
 
 This may be used for centrifugal pumps, as also for good double-acting 
 pumps. 
 
 Let A be the area of the pump, D its diameter in inches, S the mean 
 speed of movement in feet per minute ; then 
 
 Area past the valves is not less than — r --^— • 
 r 4o0 
 
 Diameter of pipes = ^ J&. 
 
 c 
 
 . F = 22. 
 
 . F = 23. 
 
 . F = 24. 
 
 . F = 25. 
 
 . F = 26. 
 
 . F = 22. 
 
 . F - 27. 
 
 . F = 24. 
 
 Both kinds of pumps should have air-vessels, and the single-acting pump 
 should have one twice the capacity of the pump when possible, and never less 
 than one-and-a-half times its capacity. When the water is pumped through 
 the tubes, the doors of the condenser may be made with pockets, which serve 
 as air-vessels in forming a cushion. Air-pipes should be fitted to the highest 
 points of the waterways, when the water is pumped into the condenser, to 
 allow the air to escape, so that it may run full, and always allow water to 
 be in contact with the tubes. 
 
 Rotary Pumps. — One or two forms of rotary pump have been tried for the 
 purpose of circulating the cooling water, but only the centrifugal pump 
 achieved perfect success. 
 
 • The velocity of flow should not exceed 350 -^diameter of pipe. 
 
 Suction pipe 
 
 of small double-acting pump, 
 
 
 Delivery 
 
 >' 
 
 SI !» 
 
 J> 
 
 
 Suction 
 
 )> 
 
 large 
 
 5? 
 
 
 Delivery 
 
 »> 
 
 ?? ?!» 
 
 JJ 
 
 
 Suction 
 
 ii 
 
 small single-acting 
 
 »• 
 
 
 Delivery 
 
 «> 
 
 •5 ») 
 
 »J 
 
 • 
 
 Suction 
 
 »J 
 
 large 
 
 »» 
 
 • 
 
 Delivery 
 
 J> 
 
 ;> ;» 
 
 »> 
 
 • 
 
CENTRIFUGAL PUMP. 
 
 399 
 
 The advantages of the rotary pump are — 
 
 (1) There are no valves, etc., to interfere with the How of water, or to 
 get out of order. 
 
 (2) Being easily worked by an independent engine of small size, or a 
 motor, it is usually so provided, and can be then started before the main 
 engines are moved, and so keep the condenser cool during the process of 
 l> warming through " the cylinders, etc., and also while standing. 
 
 (3) Having this independent engine, the supply of cooling water is varied 
 to suit the varying circumstances, and the power required to work the pump 
 varies then with the quantity of water. 
 
 (4) The efficiency for low lifts is greater than that of a reciprocating 
 pump. 
 
 (5) The supply of water is continuous, and enters the condenser without 
 shock, thereby putting no stress on the castings, pipes, etc., beyond that 
 due to the " head." 
 
 Fig. 133.— Wheel of a Centrifugal Pump. 
 
 (6) It is easily placed in the engine-room, and the absence of a recipro- 
 cating pump worked by the engine, especially in a horizontal engine, admits 
 of a better design and arrangement of condenser and air-pump. 
 
 On the other hand, the centrifugal pump is somewhat more expensive 
 than an ordinary pump, and requires some attention when at work ; the 
 latter objection is, however, now considerably lessened, although only by 
 increasing the former. 
 
 Centrifugal Pump. — This essentially consists of a wheel (fig. 133) having 
 thin vanes as arms, which act on the water so as to give it a circular motion 
 in a cylindrical case enclosing the wheel ; this case is provided with an 
 enlarged chamber around it, into which the water from the wheel is whirled, 
 and from which it escapes through a branch tangential to it (v. fig. 134). 
 
 The principle on which this pump works is, that a particle moving in a 
 circular path is under the action of two forces, one tangential, and the other 
 
400 
 
 MAXL'AL OF MARINE ENGINEERING. 
 
 normal to its path, or at right angles to the tangential one ; when this latter 
 force ceases to act. the particle moves away in a path tangential to the circle 
 at the point where the retaining force ceased to act. In the centrifugal 
 pump, the particles of water flowing into the centre of the pump are gradu- 
 ally put into motion and whirled round until, after a spiral course, they 
 
 arrive in the outer channel, and there are retained in a circular course till 
 they reach the outlet, where the retaining or normal force ceases, and they 
 fly away tangentially through the outlet. The vanes of the wheel are some- 
 times enclosed between two discs of thin metal, to which they are attached. 
 or they fit closely to the sides of the case in which they work. 
 
CENTRIFUGAL PUMP. 401 
 
 Experience has shown that the best iorm of wheel is one whose vanes 
 are sickle-shaped, as shown in figs. 133 and 134, and there is less resistance 
 if the vanes fit the pump-case instead of having the enclosing discs, inasmuch 
 as these latter have considerable friction on their surfaces, especially on that 
 next the pump-case. 
 
 The outer passage, or whirl chamber, is usually formed like a snail, so 
 that its sectional area gradually increases from nearly nothing to the full 
 area of the discharge pipe. 
 
 The inlet pipe leads from the outer rim to the centre of the pump, having 
 generally a passage on either side, so that the water is delivered on both 
 sides of the wheel. For convenience these pumps are sometimes designed 
 now with the inlet at one side only. 
 
 The diameter of the inlet and outlet pipes of a centrifugal pump for 
 circulating purposes should be such that the velocity of flow does not exceed 
 350 ^diameter in inches, feet per minute. Hence, if W is the quantity of 
 water in gallons per minute, 
 
 Diameter of pipes in inches = A / ~j-= or * / — ; — 
 
 11 V 12 ^D V 0-0343 velocity 
 
 For triple and quadruple engines the following rule gives ample size : — 
 
 Diameter of pipe in inches = */ ' ' • 
 
 C = 16 for all ships liable to be employed in the tropics, and requiring 
 
 high vacuum, 28 inches. 
 C = 25 for ships employed only in temperate zones. 
 
 The diameter of the fan-wheel or impeller is usually from 24; to 3 times 
 that of the pipes, and always is such that the velocity at its periphery at full 
 speed is in excess of 400 feet per minute, and generally from 500 to 800 feet. 
 
 The blades are curved as shown, that the water on entry may be gradu- 
 ally set into circular motion, and caused to flow into the whirl chamber with 
 as little force as possible, and to leave the impeller also gently. 
 
 The size of the cylinder for driving the fan can be calculated from the usual 
 
 conditions ; in practice its diameter is generally about 2*3 vdiameter of pipes, 
 and its stroke 0*25 the diameter of the impeller.* When the velocity of 
 flow is high, as it often is nowadays for the purpose of keeping high vacuum 
 in summer time, the cylinder must be larger even to the extent of 20 per 
 cent, in diameter. 
 
 The impeller should be always of bronze, and made as thin as possible in 
 the blades ; the spindle should be of strong bronze or incorrodible steel. The 
 Admiralty require the spindle to be cast solid with the impeller. No doubt 
 there is considerable danger of it working loose if not well fitted and secured 
 to the spindle. This, however, can be done with care when a forged or other 
 strong form of spindle can be used, whereas by the Admiralty method only 
 a casting, with possible spongy places, is allowed. 
 
 As these pumps are sometimes required to pump out the bilges and 
 ballast tanks, it is usual to fit some means of exhausting the wheel-case of 
 
 * Large pumps should have compound engines with a low-pressure cylinder of the 
 size thus calculated ; the high-pressure cylinder should be half the diameter when the 
 boiler pressure is 120 lbs. and upwards. 
 
 26 
 
402 MANUAL OF MARINE ENGINEERING. 
 
 air, so as to cause the water to flow up into it, as the fan has no sucking 
 power, and will not draw until fully charged with water. For this purpose 
 some engineers fit a small rotary exhauster, worked by means of a belt from 
 the spindle, while others fit a small steam-ejector on the same principle as 
 Rodgers' vacuum blower. When no such special means is provided, the 
 pump may be charged from the sea, and the sea inlet left slightly open until 
 the pump begins to draw from the bilge or tanks. In the Navy the circu- 
 lating pumps are arranged to draw from the bilges ; for this purpose throttle- 
 valves are fitted to the suctions, so that the sea supply may be instantly 
 shut off and the bilge turned on, this method having been found the most 
 successful for this purpose. 
 
 Centrifugal Circulating Pumps of R.M.S. " Mauritania " are each 32 inches 
 in diameter of delivery pipes ; there are a pair of such pumps to each set, and 
 a set in each engine-room capable of delivering 45,000 gallons per minute, 
 or 450.000 lbs. The total steam consumption at full power is about 1,000,000 
 per hour, or 16,667 per minute in that ship — that is. 8,334 lbs. is condensed 
 in each condenser — there is, therefore, pump power tor 54 times the quantity 
 of water condensed. In modern naval ships the engine circulating pumps 
 are capable to supplying 45 to 60 times the weight of steam produced fo: 
 full power, cargo steamers 40 to 50. 
 
 Feed-pumps. — The duly of the feed-pump is to supply the boiler with 
 sufficient water to meet its requirements. It is obtained from the hot-well, 
 so that when there is a surface condenser it is practically pure water, and the 
 amount, under ordinary circumstances, is the same as that evaporated in 
 the boiler ; but owing to leakage and waste by blowing the steam whistle, 
 or using an auxiliary engine which exhausts into the air, the quantity of 
 water condensed is not alwa3's sufficient to make up for that evaporated. 
 It is found necessary also occasionally to blow some of the water out of the 
 boiler, to get rid of scum floating on the surface of the water, and this waste 
 must be made good by a supply from the sea or elsewhere. To prevent 
 a deposit of salt, etc., in the boiler, it was customary to blow out some of 
 the very dense water from the boiler at fixed intervals, and as the blow-ofl 
 cock was situated near the bottom of the boiler a considerable amount oi 
 solid matter was thus got rid of ; the quantity of water blown out was made 
 up by an extra supply from the sea. Now, since the surface condenser 
 might leak, the pumps of engines having surface condensers were formerly 
 made sufficiently large when worked by the main engines to do duty under 
 such circumstances. 
 
 Sea-water contains about ^ of its weight of solid matter* in solution; 
 its density is measured by a hydrometer, which in the Navy is marked with 
 degrees, so that when floating in pure water the zero point is at the surface, 
 and when in clean sea-water it marks 10°. In the merchant service, engineers 
 are accustomed to speak of the density by the number of ounces of solid 
 matter to the gallon, sea-water containing 5 ounces to the gallon. 
 
 Net Feed-water. — The net feed-water is that quantity required to make 
 up for what has been used as steam in the engine and its auxiliaries ; let 
 this be denoted by Q ; let n be the number of times the saltness of the water 
 
 * On the British coasts the sea-water contains only about one fortieth of its weight 
 of solid matter in solution. 
 
CENTRIFUGAL CIRCULATING PUMP?. 
 
 403 
 
 (a) 1 Set for a Battleship 27.000 S.H.P. (W. H. Alien). 
 
 (b) 1 Set for R.M.S. " Mauritania," 65,000 S.H.P 
 Fig. 134a.— Centrifugal Circulating Pumps (Allen). 
 
404 
 
 MANUAL OF MARINE ENGINEERING. 
 
 in the boiler is to that of sea-water, then 
 
 The gross feed-water = 
 
 1 
 
 x Q. 
 
 This is the amount of feed-water which must be pumped into the boiler when 
 salt water only is used, to maintain the saltness of n times that of the sea. 
 
 If the pumps were only of sufficient size to pump this amount of water, 
 a considerable time would elapse before the boiler would be filled to the 
 working level after " blowing off " ; to meet this objection it was usual to 
 make each feed-pump capable of pumping twice this quantity, therefore 
 
 2 n 
 Quantity of water for each pump to supply = t X Q. 
 
 Since a surface condenser supplies pure water for feeding the boiler, and 
 the waste is now made up with fresh water, either carried in the tanks or 
 distilled from sea-water, there is not the same need for blowing off the boilers, 
 the feed-pumps may be, therefore, very much smaller than when jet con- 
 densation is practised, and are generally of such a size when worked by the 
 main engine that each is capable of delivering three times the net feed-water, 
 assuming them to have an efficiency of TO; the pumps, when both are 
 working, can deliver six times the net feed, which is sufficient to satisfy 
 any extraordinary demands. 
 
 If Q be the quantity of net feed-water per minute in cubic feet, / the 
 length of stroke of feed-pump in feet, and n the number of strokes per 
 
 minute. 
 
 / 550 X Q 
 " V n x 
 
 Diameter of each feed-pump plunger in inches 
 If W be the net feed-water in pounds — 
 
 Diameter of each feed-pump plunger in inches = */ 
 
 X I 
 
 8-9 x W 
 n X I 
 
 The following empirical formula will give such sizes as will be found in 
 practice, and which will closely approximate to those given by the above 
 rule : — 
 
 Capacity of each feed-pump = 
 
 capacity of L.P. cylinder 
 
 Ge.nebal Description. 
 
 
 Steam Consumption. 
 
 Feed. 
 
 Value 
 of C. 
 
 
 
 
 
 
 
 
 Ship. 
 
 Service. 
 
 Engines. 
 
 Main. 
 
 Aux., &c. 
 
 Total. 
 
 Allowance 
 
 
 
 
 
 Lbs. 
 
 Lbs. 
 
 Lbs. 
 
 Lbs. 
 
 
 Ordinary cargo, . 
 
 General. 
 
 Quadruple. 
 
 140 
 
 0-70 
 
 14-70 
 
 17-5 
 
 650 
 
 ?» • 
 
 »? 
 
 Triples. 
 
 14-7 
 
 0-74 
 
 15-44 
 
 18-5 
 
 600 
 
 Passenger & cargo, 
 
 Long voyage. 
 
 Quadruple. 
 
 140 
 
 1-25 
 
 15-25 
 
 1S-5 
 
 600 
 
 Express, . 
 
 Short „ 
 
 Triple. 
 
 L5-5 
 
 1-50 
 
 17-00 
 
 20 
 
 550 
 
 „ Cold zone, 
 
 )> ?> 
 
 Turbines. 
 
 130 
 
 1-65 
 
 14-65 
 
 17-5 
 
 # , 
 
 »> • • 
 
 Atlantic. 
 
 Turbines. 
 
 12-5 
 
 2-50 
 
 1500 
 
 18-5 
 
 
 Net feed-water in pounds per stroke approximately = area of piston in 
 inches X stroke in feet X absolute pressure at release ■*■ 53,600. 
 
 Example. — To find the net feed-water in pounds for an engine whose 
 
VALVES AND VALVE-BOXES. 405 
 
 cylinder is 50 inches diameter, and length of stroke 3 feet, the pressure at 
 release being 12 lbs., and the revolutions 100 per minute. 
 
 Net feed-water ver minute = 200 X 1,96a * !L X U = 263 lbs. 
 
 53,600 
 
 All engines over 80 N.H.P. should have two feed-pumps, each capable of 
 supplying the boilers when the engines are at full speed ; and each^pump 
 should be so arranged that it may be worked quite independently of the 
 other, and easily put out of action when not required. This latter condition 
 is not always complied with in practice ; but if it were so there would be 
 then practically a spare pump, and only one in danger of derangement when 
 the engine is at work. 
 
 The feed-pumps of vertical engines are usually moved by the same parts 
 as operate the air and circulating pumps. They are sometimes fixed to the 
 crosshead of these latter pumps, and sometimes driven from studs in the 
 sides of the rocking levers. When there is only one set of rocking levers 
 this latter plan should be adopted ; for by placing one pump on each side 
 of the lever centre they may both be used together and deliver alternately, 
 and give a steady flow of water. In the Navy the use of feed-pumps worked 
 by the main engines is entirely discontinued, and in the mercantile marine 
 it is fast becoming the practice, especially in large ships, to feed the boilers 
 by an independent pump or pumps, which are generally self-regulating. 
 In fact, with all fast-running engines the feed-pumps are better detached 
 from them and worked with absolutely independent automatically-con- 
 trolled engines, so that the supply of water is steady and not intermittent, 
 and exactly in proportion to the demand. The usual plans for treating 
 the feed-water are the same in principle, though differing in detail. The 
 water is delivered from the hot-well to a receiving tank by the air-pump in 
 the mercantile marine, by a special pump in the Navy. The feed-pumps, 
 two in number and in duplicate, draw from this tank at a uniform rate under 
 normal circumstances, but if the supply in the tank fails the pumps are 
 slowed down and stopped by a float arrangement connected to their stop 
 valves. If the feed is wholly or partially shut off in the boiler-room the 
 extra load that then comes on the pumps slows them down, and eventually 
 stops them if " feed " is wholly shut off. As a rule, only one of these pumps 
 works at a time, and is capable of supplying two to three times the net feed 
 when working full speed (v. fig. 186). 
 
 Relief Valves should be fitted to each pump when worked by the main 
 engine, and they are so arranged that each may work separately ; when 
 they both deliver into a common pipe, one relief valve is sufficient ; these 
 valves should be loaded to 1§ times the boiler pressure. 
 
 Valves and Valve-boxes should always be of bronze ; and since the seats 
 as well as the valves wear out rapidly, they should be made separately from 
 the box casting of harder bronze. The valves should have a seating area 
 equal at least to 20 per cent, of the valve, and it is better to make them 
 flat rather than conical. Some engineers use bronze balls fitting into conical 
 seats ; since they are constantly changing their position on the seats these 
 balls wear very slightly and keep very tight. When the boiler pressure 
 did not exceed 30 lbs., the valves of the feed-pumps were usually made of 
 india-rubber ; and some American engineers employed this material 
 
406 MANUAL OF MARINE ENGINEERING. 
 
 notwithstanding the increased pressure. If the rubber is thick and capable 
 of withstanding the action of mineral oil. no doubt it will work well ; but 
 so much reliance cannot be placed on it as on the metal, especially at high 
 pressures. 
 
 Bronze valves (fig. 119) make a noise when working, owing to the quick 
 return to their seats on the pump ceasing to deliver : attempts have been 
 made 4o reduce the noise and wear in various ways. The noise of the valves 
 may be stopped by loading them with light springs made of steel nickel- 
 plated, or of hard brass. Several small valves with little lift is found to be 
 more satisfactory than one large one ; for example, to obtain sufficient area 
 through the seats of six valves to take the place of one 4 inches in diameter, 
 each would be 1| inches, giving the sum of the perimeters 30'6 inches against 
 the 12'56 inches of the 4-inch valve. It follows, then, that if the single 
 valve lifted half an inch, the small one need only lift one-fifth of an inch. 
 Such small valves must be forged of hard, tough bronze, carefully formed 
 so that there is no groove or nick to form a starting point for a crack, as the 
 tails quickly break off by the " jar " on seats and stops. 
 
 Air-vessels. — Each pump should be furnished with an air-vessel which 
 may serve the double purpose of providing a cushion, and intercepting and 
 collecting the free air from the new feed-water, which, when fresh or sea- 
 water is used, is the active agent in producing corrosion when admitted 
 to the boiler. To increase its usefulness in the latter capacity, a fine grating 
 should be fitted to the inlet orifice, which will " spray " the water, and so 
 separate the air. and a relief valve should be fitted to the top of the vessel, 
 loaded to a pressure slightly below that obtaining when the pump is delivering 
 at its maximum rate, but arranged to close when the vessel is nearly full of 
 water. Air-vessels, however, are practically useless with the high pressures 
 of steam now employed, except as a means of getting rid of gases and 
 air. 
 
 Feed-pipes. — The pipes leading to and from the feed-pumps should be 
 such that the average working velocity of flow does not exceed 500 feet per 
 minute, and small pumps should have larger pipes in proportion, so that 
 the flow through them does not exceed 350 ^diameter in inches- 
 
 Since the amount ot water actually flowing through the pipes is gener- 
 ally very considerably less than the pumps are capable of discharging, the 
 velocity is seldom more than half the above allowances ; but as the pumps 
 do occasionally deliver their full amount, the pipes must be large enough 
 for that purpose. 
 
 If d is the diameter of the feed-pump plunger, and *' its mean velocity in 
 feet per minute, then 
 
 Diameter of feed-pipe = ^ Js -f- 0"25 inch. 
 
 or, taking I.H.P. as the basis, 
 
 28 
 
 v/l.H.P. 
 
 Diameter of feed-pipe = .'■ ' •■ ' 4- 0"25 inch. 
 
 lo 
 
 Example. — To find the diameter of the feed-pipes for a pump whose 
 diameter is 6 inches, and the length of stroke 2 feet, worked from the levers 
 
FEED-PUMP ROD. 407 
 
 of an engine, making 60 revolutions per minute. Here s = 2 x 60 x 2 
 or 240 feet. 
 
 c 
 
 Diameter of pipe = — V240. or 4 inches. 
 
 Example. — To find the diameter of feed-pipe for an engine of 3,600 I.H.P. 
 
 e ■ ^3^600 . , 
 
 Diameter of pipe = — — 1- 0"25 inch = 3 -58 inches. 
 
 If there are two pumps which deliver alternatively, the pipes will be the 
 same size throughout ; but if the two pumps may deliver at the same time, 
 the pipe beyond the junction of the two from the pumps must be nearly 
 double the sectional area of one. As the resistance of pipes is due greatly 
 to friction at the surface, and will consequently vary as the diameter, while 
 the area of section varies as the square of the diameter, the resistance in the 
 single pipe will be considerably less than the combined resistance in the 
 two, and for this reason the sectional area may be less. In practice this area 
 may be 0"8 of the combined area of the two. Hence, when there are two 
 pumps delivering together — 
 
 Diameter of main pipe = 1 "265 x diameter of each of the branches. 
 
 If there were two pumps, as in the first example, delivering together, 
 the diameter of the main pipe would be 4 X 1"265, or 5*06 inches. 
 
 " Pet Valves." — Although it is prejudicial to admit air to the feed-water, 
 it is sometimes necessary for the good working of all pumps to allow a little 
 air to enter between the valves to form a cushion : they are. however, of no 
 use to a modern feed-pump. 
 
 Feed Tank. — To avoid any waste of water through the overflow or air 
 pipe of the hot-well when the feed-pumps are temporarily stopped, it is well 
 to provide a tank, into which the water is discharged from the hot-well, and 
 from which the feed-pumps draw. Such an arrangement is very beneficial 
 in all engines, and especially in those having small hot -wells. 
 
 Feed-pump Rod. — This is of iron, and, in the case of hollow plunger 
 pumps worked from a pin having a circular motion, it is jointed within the 
 pump. As the plungers of vertical pumps are seldom without water in 
 them, and difficult to empty, the joint should be such as will work with 
 water as a lubricant. This is accomplished by bushing the joint with lignum 
 vitse, and casing the pins with brass ; or white-metal (Fenton's) bushes, 
 with steel pins, will do. The rod-end within the pump should be galvanised, 
 or otherwise protected from being corroded. 
 
 When the rod is long and of iron, p being the boiler pressure. 
 
 tv 1 1 j j diameter of plunger *- 
 Diameter of feed-pump rod — j^ s — X Vp + 0*b. 
 
 If of brass, divide by 35 instead of 40. 
 
 When the rods are short, the diameter may be 0*7 of that given by the 
 above rules. 
 
 The plungers, valve-boxes, and valves are always of best bronze, and the 
 pump-barrel or case may be of bronze, but in the merchant service this is 
 generally of cast iron. The same remark applies to the air-vessels, escape- 
 
408 MANUAL OF MARINE ENGINEERING. 
 
 valves, etc. The capacity of the air-vessel should be from 15 to 2 X the 
 capacity of the pump. 
 
 Bilge Pumps. — These pumps, which are for the purpose of freeing the 
 bilge of water, are somewhat similar in construction and method of working 
 to the feed-pumps. Although the Board of Trade and Lloyds require that 
 all steamships shall have two bilge pumps worked by the main engines, 
 except in the case of engines under 70 N.H.P., they do not specify the 
 capacity of these pumps. There is no definite basis of calculation for the 
 size of these pumps, and it is generally at the caprice of individual engineers, 
 many of whom still adhere to the old practice of making them of the same 
 size as the feed-pump. They should not be, however, made less than given 
 by the following rule : — 
 
 _ . .... capacity of L.P. cylinder 
 Capacity of bilge pump = — — ^kK • 
 
 The Capacity of the Bilge Pumps should really bear some reasonable 
 relation to the displacement of the ship rather than to the size of the engines. 
 The following rule, therefore, should be followed with this object in view : — 
 
 Total bilge pump capacity in cubic inches = 3-5 X D 5 > 
 
 D being the displacement in tons. 
 
 Example. — A ship of 5,000 tons' displacement should have bilge pumps 
 whose capacity = 5,000* X 3 5, or 1,022 cubic inches. Such a ship, if a 
 cargo steamer, might have a low-pressure cylinder 60 inches diameter x 42 
 inches stroke. Each pump would be by rule 340, and the two 680 ; on the 
 other hand, the low-pressure of a faster ship might be 72 inches, and the two 
 be then 1,000 cubic inches. 
 
 The plungers are usually of bronze, but in the merchant service are often 
 of cast iron, and as this is harder, especially after the hard skin is formed by 
 rubbing, they w 7 ear longer when of this material. With cast iron the neck 
 and gland bushes should be of Fenton's metal to prevent corrosions setting 
 it fast when not at work. 
 
 In the Navy the bilge pumps are now like the feed-pumps, quite detached 
 from the main engines ; in the mercantile marine, when the engines are run 
 at a high number of revolutions, the same practice is followed ; but when 
 the engines are not quick running, as in cargo steamers, the bilge pumps 
 are, as formerly, worked from the main engines. The Admiralty have at 
 least one pump set apart for the bilges in each engine-room, and this is gene- 
 rally done in passenger steamers. There is also in each engine-room another 
 pump which may be used for pumping out the bilges or for discharging water 
 on deck ; or one to do this service and the other to pump from the sea on 
 deck or from the sea, and tanks to the boilers. 
 
 In the mercantile marine, the valve-boxes are of cast iron, and the valves 
 often hinged "clacks," which are easy removed for cleaning. The covers 
 of these boxes should be so made that they may be easily and quickly removed 
 and replaced, and for that purpose are sometimes hinged, but better still, 
 held down by two hinged bolts fitting into recesses in the cover. 
 
 The Board of Trade require that one engine bilge pump shall be arranged 
 to draw water from the sea, and pump it on deck in case of fire. When this 
 is done, the suction pipes should be fitted with a three-way cock, whose 
 
POWER TO DRIVE CIRCULATING PUMPS. 409 
 
 plug has only one port, so that it cannot be open to the sea and bilge at the 
 same time, and so flood the ship. When the pipes are very large, it is not 
 always convenient to fit such a cock, then a double valve-box with self- 
 acting non-return valves is substituted ; when possible, the cock should be 
 Btted as the better and safer plan. 
 
 Directing Boxes. — It is required that the bilge pumps shall draw from 
 each compartment of the ship. For this purpose, the suction from the 
 pump is connected to the top of a box containing a series of valves by opening 
 any one of which a communication is made to a separate compartment ; 
 the cover of each valve should have on it a label signifying to which com- 
 partment the valve opens a communication. These directing boxes should 
 be placed in such a position that they are easily got at, and above the flooi 
 plating when possible, that the labels may be seen and the valves operated. 
 
 Mud Boxes. — Between the directing box and the pump should be fitted 
 a box with a strainer, which shall intercept such solid matter as would derange 
 the pump valves, if allowed to enter among them. They should have covers 
 similar to those of the pump valve-boxes, and placed in such a position as 
 to be easily got at for cleaning or examining. 
 
 Sanitary Pump. — In passenger steamers, it was usual to have a pump 
 of about a half or one-third the capacity of the bilge pump, which could be 
 put into gear, and worked by the engines, to discharge water on deck for 
 sanitary purposes ; this duty is now generally performed by a separate 
 donkey pump. 
 
 Ejectors. — In the Navy, as in the merchant service, steam ejectors are 
 often used for clearing the engine and boiler-room bilges. They are in 
 principle very like the well-known feed injectors, without valves or other 
 internal fittings liable to derangement from ashes, dirt, etc. They are 
 started and stopped by merely opening and shutting a cock or valve, and 
 require no attention when working, and are, consequently, most suitable 
 for such a service. 
 
 The Power required to drive the Circulating Pumps may be estimated by 
 
 taking the friction of the tubes and inlet and discharge pipes, allow for the 
 
 loss of head due to entering and leaving the tubes, and assume the efficiency 
 
 of the pump to be 0*6, and the engine driving it 0*65, so that the combined 
 
 efficiency is 0'39. The " head " to overcome the friction in the tubes may 
 
 L x v 2 L x v 2 
 
 be taken as = — , 0'60 = ,„„ 7 feet, where L is the aggregate length of 
 
 2 g X d 107 d o& & o 
 
 tubes in feet, d the diameter in inches, and g is 32 at sea level ; v the velocity 
 
 in feet per second. 
 
 For Example. — What head of water is necessary to overcome the frictional 
 
 resistance of 30 feet of f-inch tubes, the velocity of flow being 8 feet per 
 
 second. The internal diameter = - 64 inch ? 
 
 „ 30 X 64 ' , 
 
 H = ior^64 = 28 feet - 
 
 The sum of the other resistances may be taken as a quarter of tne above, 
 so that the total " head " required to force water through a condenser with 
 f-inch tubes so long would be 35 feet, or equivalent to a pressure of 15*2 lbs. 
 per square inch. 
 
410 
 
 MANUAL OF MARINE ENGINEERING. 
 
 
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POWER TO DRIVE CIRCULATING PUMPS. 
 
 411 
 
 The work done will be weight of water x 35 foot-lbs. 
 Ike net horse-power = ta <w ~- 
 
 rri T XT T» W X 35 W 
 
 The gross or I.H.P. = Q . 39 x 33^ = ^ 
 
 Example. — If the above condenser requires 12.000 lbs. of cooling water 
 
 per minute. 
 
 12 000 
 Then I.H.P. = -=£P, or 326. 
 
 000 
 
 TABLE XLVIa. — Flow through Pipes not exceeding 240 V Diameter 
 Long under Ordinary Working Conditions per Minute. 
 
 
 Watc: 
 
 • 60° F. 
 
 Live Steam. 
 
 Exhaust St 
 
 earn. 
 
 Air Atmosphr. 
 
 Press., 
 
 
 200 lbs. Press 
 
 , Abs. 
 
 3 lbs. Press. 
 
 Abs. 
 
 60° F. 
 
 
 Diameter 
 
 of 
 
 Pipe. 
 
 
 
 
 
 
 
 
 
 Ve- 
 locity. 
 
 Quan- 
 tity. 
 
 Ve- 
 locity. 
 
 Quantity. 
 
 Ve- 
 locity. 
 
 Quantity. 
 
 lodty. *>"*«* 
 
 Ins. 
 
 Ft. 
 
 C Ft. 
 
 Ft. 
 
 C. Ft. 
 
 Lbs. 
 
 Ft. 
 
 C. Ft. 
 
 Lbs. 
 
 Ft. C. Ft, 
 
 Lbs. 
 
 1-0 
 
 350 
 
 1-909 
 
 5.000 
 
 27-19 
 
 11-71 
 
 3,300 
 
 17-95 
 
 0-152 
 
 1.000 
 
 5-438 
 
 0-418 
 
 1-5 
 
 401 
 
 4-917 
 
 5 535 
 
 69-51 
 
 30-00 
 
 3,653 
 
 45-94 
 
 0-389 
 
 1,107 
 
 13-90 
 
 1-070 
 
 2-0 
 
 441 
 
 9-600 
 
 5,945 
 
 129-7 
 
 54-12 
 
 3,923 
 
 75-60 
 
 0-646 
 
 1.189 25-94 
 
 1-995 
 
 2-5 
 
 475 
 50* 
 
 16-19 
 
 6.285 
 
 213-9 
 
 92-12 
 
 4,148 
 
 141-2 
 
 1-196 
 
 1,257 42-78 
 
 3-290 
 
 3-0 
 
 24-80 
 
 6,575 
 
 322-6 
 
 138-9 
 
 4,340 
 
 212-9 
 
 1-804 
 
 1,315 
 
 64-52 
 
 4-963 
 
 3-5 
 
 531 
 
 35-45 
 
 6,835 
 
 456-3 
 
 196-6 
 
 4,511 
 
 301-2 
 
 2-552 
 
 1,367 
 
 91-26 
 
 7-020 
 
 4-0 
 
 556 
 
 47-93 
 
 7,070 
 
 616-8 
 
 265-6 
 
 4.666 
 
 407-7 
 
 3-456 
 
 1,414 
 
 121-3 
 
 9-331 
 
 5-0 
 
 598 
 
 99-00 
 
 7.475 
 
 1.019 
 
 438-9 
 
 4,933 
 
 672-5 
 
 5-700 
 
 1,495 
 
 203-8 
 
 15-68 
 
 6-0 
 
 636 
 
 126-7 
 
 7,825 
 
 1,536 
 
 661-5 
 
 5,165 
 
 1,014 
 
 8-590 
 
 1,565 
 
 307-2 
 
 23-63 
 
 7-0 
 
 670 
 
 179-2 
 
 8,130 
 
 2,180 
 
 939-0 
 
 5,362 
 
 1,419 
 
 12-02 
 
 1,626 
 
 436-0 
 
 33-54 
 
 8-0 
 
 700 
 
 244-3 
 
 8,410 
 
 2,938 
 
 1,263 
 
 5,550 
 
 1,939 
 
 16-43 
 
 1,682 
 
 587-6 
 
 45-20 
 
 9-0 
 
 728 
 
 321-7 
 
 8,660 
 
 3,826 
 
 1,649 
 
 5,715 
 
 2,525 
 
 21-40 
 
 1,732 
 
 765-3 
 
 58-87 
 
 10-0 
 
 754 
 
 431-6 
 
 8,890 
 
 4,852 
 
 2,100 
 
 5,867 
 
 3,202 
 
 27-14 
 
 1,778 
 
 970-4 
 
 76-65 
 
 12-0 
 
 800 
 
 628-0 
 
 9,305 
 
 7,300 
 
 3,144 
 
 6.116 
 
 4,818 
 
 40-83 
 
 1,861 
 
 1,460 
 
 112-3 
 
 14-0 
 
 844 
 
 903-6 
 
 9 670 
 
 10,357 
 
 4,012 
 
 6,382 
 
 6,836 
 
 57-91 
 
 1,934 
 
 2,071 
 
 159-3 
 
 16-0 
 
 882 
 
 1,229 
 
 10,000 
 
 13,950 
 
 6,000 
 
 6,600 
 
 9,207 
 
 78-00 
 
 2,000 
 
 2.790 
 
 214-6 
 
 18-0 
 
 917 
 
 1,550 
 
 10,295 
 
 18,018 
 
 7,760 
 
 6,795 
 
 12,000 
 
 101-7 
 
 2,059 
 
 3,646 
 
 280-5 
 
 20-0 
 
 900 
 
 2,096 
 
 •10,570 
 
 23,050 
 
 9,930 
 
 6,976 
 
 15,220 
 
 129-0 
 
 2,114 
 
 4,611 
 
 354-7 
 
 25-0 
 
 1,023 
 
 3,485 
 
 
 . , 
 
 
 7,378 
 
 25,210 
 
 213-7 
 
 2,236 
 
 7,626 
 
 586-6 
 
 30-0 
 
 1,087 
 
 5,345 
 
 
 . . 
 
 
 7,725 
 
 38,000 
 
 321-8 
 
 2,341 
 
 11,033 
 
 848-7 
 
 35-0 
 
 1,145 
 
 7.620 
 
 . . 
 
 
 . , 
 
 8.030 
 
 53,600 
 
 453-6 
 
 2,433 
 
 16,480 
 
 1,268 
 
 40-0 
 
 1,197 
 
 10.470 
 
 
 •• 
 
 •• 
 
 8.290 
 
 71,800 
 
 608-5 
 
 2,514 22,500 
 
 1,731 
 
412 MANUAL OF MARINE ENGINEERING. 
 
 CHAPTER XVI. 
 
 VALVES AND VALVE-GEAR. 
 
 Steam was admitted to and released from the cylinders of the early land 
 engines by means of conical valves, operated by tappet gear in such a way 
 that the steam valve was suddenly opened and as suddenly closed at the 
 proper times, and the valve which allowed the steam to escape to the con- 
 densers and closed before steam was admitted worked with the same precision 
 and by tie same methods. Such an arrangement permits of a high state of 
 efficiency for the steam, but is open to the objection that motion so sudden 
 is liable to cause shock with much wear and tear of the working parts. In 
 those early days the pressure of steam was only a little above that of the 
 atmosphere, and the number of strokes per minute comparatively few, so that 
 leakage past the valves was important, and the wear of the tappet gear 
 not so great as might be expected ; on the whole, the early engineers had 
 every reason to be satisfied with their valve-gear. 
 
 That such gear can be made to work well and give general satisfaction 
 is evident from the fact that it is still employed in large pumping engines, 
 which work at 60 lbs. pressure and move at much higher speeds than 
 formerly obtained, and in modern times in the marine type large electric 
 generating engines at central stations, as well as in all oil engines. 
 
 Modifications of this form of valve and valve-gearing have been adopted 
 for marine purposes, but not with that degree of success, at least in this 
 country, as to commend themselves to engineers for extended adoption. In 
 America, however, they are still largely employed for marine engines, 
 and most successfully so in the various engines of river and lake steamers 
 moving at comparatively low rates of revolution. ' 
 
 Such gearing was impossible in the locomotive engine on many grounds, 
 so that an entirely different valve was adopted to suit it. This valve 
 differed from previous valves in many ways, but chiefly in respect to its 
 motion, which was a sliding one. From it this form of valve is called a 
 slide-velve, and frequently, when in its simple form, the locomotive slide-valve, 
 to distinguish it from the long and short D-valves and piston-valves so 
 generally used in the engines on land and sea in the early part of the nine- 
 teenth century, which also had a sliding motion. 
 
 The locomotive slide-valve (fig. 135), in its simple and extended forms, is 
 the one most generally used in the marine engine. It consists essentially 
 of a rectangular" block having a central cavity, the flat bars between the 
 outer edge at each end and the central cavity being sufficiently broad to 
 cover the cylinder ports, and when in its mean position both ports are 
 covered. The amount by which the outer edges overlap the ports when in 
 the mean position is called the lap, and the amount by which the valve is 
 
LOCOMOTIVE SLIDE-VALVE. 
 
 413 
 
 open when the piston is at the commencement of its stroke is called the 
 lead ; the space through which the valve is moved during half a revolution 
 of the engine is called the travel. The amount by which the inside edges of 
 the valve overlap the port is called the inside lap, and when,- as often happens, 
 instead of overlapping inside, the port is slightly open to the cavity of the 
 valve, the valve is said to have negative inside lap. 
 
 The early locomotives made by Stephenson had, like modern steering 
 engines, little or no lap, and little or no lead. Timothy Hackworth, by 
 giving the valves of his engines lead and lap, effected an earlier cut-off, and 
 consequently obtained some expansion of the steam, thereby saving the 
 locomotive from threatened failure, and making it a commercial success. 
 The effects of lead and lap will, however, be shown later on. 
 
 Seawards Valves. — To get a more perfect arrangement of cut-off and 
 release than is. possible with one valve, Seaward fitted a separate valve and 
 ports for steam and exhaust, and to avoid the large amount of clearance 
 which such an arrangement would entail, he used four valves, a steam and 
 an exhaust-valve for each end of the cylinder, and by placing the ports close 
 to the ends of the cylinders, reduced the passages to a minimum. Each 
 valve was simply a flat plate, and worked by cams on the shaft, and how- 
 
 Fig. 135. — Common Locomotive Slide-valve. 
 
 ever early the steam valves cut off, the exhaust- valves always opened just 
 at the end of the stroke. 
 
 Common or Locomotive Slide-valve. — Fig. 135 shows the modern form of 
 the valve, and it is so well known as to require no further description. So 
 long as the cut-off is later than half the stroke of the piston, this valve 
 answers very well for engines of moderate size ; but when an earlier cut-off 
 is required, sufficient opening to steam can only be obtained by excessive 
 travel, early leads, or very broad ports, neither of which is desirable when 
 it can be avoided. Large travel of valve means large power to drive it ; 
 broad ports produce the same result by increasing the area, and consequently 
 the load on the valve ; and early leads are apt to produce severe shocks on 
 the rods, framing, etc., and to unduly check the piston. 
 
 Trick Valve. — Fig. 136 shows an ingenious plan for obtaining a double 
 opening to steam by means of a passage around the valve, the entrance to 
 which is at the end remote from that at which steam usually enters, and 
 whose exit is through the lap or cover of the valve. It will be seen that 
 the effective surface of the valve exposed to steam pressure is considerably 
 reduced below that of the common slide-valve. 
 
 Double-ported Valve.— This valve (fig. 137) has a system of ports and 
 passages which is addecl to the locomotive slide-valve, to allow of admission 
 
414 
 
 MANUAL OF MARINE ENGINEERING. 
 
 and emission of steam through a second port in the cylinder face, so that 
 with the same travel as the common valve, there is double the area of opening 
 for steam, and double the area for exhaust. This form of valve is very 
 generally adopted for both the medium-pressure and low-pressure cylinders 
 of compound engines of large size, and for the low-pressure cylinder of even 
 quite small engines. Care must be taken in designing such a valve that 
 
 m 
 
 
 Fig. 130.— The Trick Valve. 
 
 there is the requisite area for steam to enter into the cavity leading to the 
 steam port, and also that there is ample room at its back for the exhaust 
 steam to pass from the outer port of the cylinder. 
 
 Treble-ported Valves (v. fig. 75). — To obtain still larger opening for steam, 
 the lap of the double-ported valve is extended so as to cover a third port 
 at each end ; a portway is made through the cover lap, which admits steam 
 
 j — ~-~~i n 
 
 Fig. 137. — Common Double- ported Valve. 
 
 to the second port of the cylinder, while steam enters the third port of the 
 cylinder through the opening beyond the valve in the usual way. But 
 since the length of face and valve required for this form of valve is very 
 nearly as much as if it had a port and passage to admit of exhaust, it has 
 been discarded in favour of a treble-ported valve, similar in all respects to 
 the double-ported. Some very large engines have four-ported valves, similar 
 
PISTON VALVES. 
 
 415 
 
 in design to the double-ported. Such valves are very large and heavy, and 
 require a large amount of power to move them ; but they work very satis- 
 factorily, and so far have not any successful rivals for steam of low-pressure. 
 When used for the medium-pressure cylinder of compound engines, both 
 treble- and double-ported valves require relief plates or frames to diminish 
 the large area they expose to steam pressure. 
 
 The Travel of Flat Valves should be such that their mean speed is not 
 more than 200 feet per minute for the H.P. cylinder, nor 250 for L.P. cylinders. 
 
 Piston Valves. — No system of relief frames or plates has yet been tried 
 which has given universal and entire satisfaction ; some indeed produce 
 more resistance than they have been designed to reduce, and the best cannot 
 be depended on for any very long period when exposed to the temperature 
 of high-pressure steam. The area of opening of port is restricted when 
 only a common locomotive slide-valve is used, and its extensions magnify 
 the evil which relief frames are supposed to cure. It has been stated that 
 circular valves of the mushroom type do not work well in fast-running engines, 
 although they give a good opening to steam. To combine the advantages 
 of the two systems the piston valve is designed (fig. 138). The port area 
 is nearly three times that of a flat valve of the same dimension transversely, 
 
 Fig. 138.— The Piston Slide-valve. 
 
 and the pressure on the sides due to the steam is nil. Essentially, the piston 
 valve consists of two pistons, the face of each being equal in length to the bars 
 of a locomotive slide-valve, and connected by a rod. These pistons are 
 fitted into a cylindrical chamber having ports corresponding to those in 
 the cylinder face ; the faces of the piston cover these ports, and have the 
 same amount of lap, etc., as a common valve. Steam is sometimes admitted 
 outside the pistons, and it exhausts from the cylinder into the space between 
 them, and from there into the exhaust passage in the usual way. But the 
 piston valve permits of the steam being between the pistons and the exhaust 
 steam passing their outside. In this case the valves are said to be with 
 "inside cut-off." Piston valves were used very early in the nineteenth 
 century, especially in marine engines, and found to be superior to the 
 : ' Long D " and " Short D " valves ; they were always of the " inside cut- 
 off " type, and then, as now, had the advantage of confining the steam of 
 high pressure to the circular casing surrounding the middle of the valve, 
 so that no joint or stuffing-box was exposed to its action. This quality is 
 to-day, with steam of really high pressure and high temperature, of great 
 value, especially when the steam is rendered very dry by superheating, etc. 
 
416 MANUAL OF MARINE ENGINEERING. 
 
 When the pistons are sufficiently large they are connected by a pipe or 
 hollow casting (as shown in fig. 138), through which steam can pass from one 
 end to the other : if this cannot be accomplished, the two ends of the valve- 
 case are connected by a pipe cast with or connected to it. 
 
 Small engines, when fitted with such valves, have them in their simple 
 form, the pistons being plain bronze or cast-iron discs of the required thick- 
 ness, generally cast in one piece. Such a form would suit all sizes of engines, 
 if always working at high speed ; but when standing or running slow, the 
 leakage past the valve, especially when it was worn, would soon be so con- 
 siderable as to cause serious loss, besides making the engine very unhandy. 
 To avoid this it is usual to pack the pistons much in the way that ordinary 
 pistons are packed, except that the junk-rings and flanges are chamfered 
 away, and the packing rings are made to project from them so as to allow 
 free passage to the steam. The packing rings are also made sometimes with 
 an extension so as to entirely mask the junk-ring and flange, so that the 
 acting face is at least as long as the body of the piston portion of the valve. 
 The spring-rings are made of strong cast iron, and since, owing to the very 
 slight velocity at which the valve moves, the wear is small, the rings should 
 have very little net or spring. When steam of high pressure is in contact 
 with the valves of large size the rings should be restrained from pressing 
 on the walls, as in the case of pistons (fig. 88). The liners in the valve- 
 box are usually made of cast iron, fitted tightly in and secured by flanges. 
 With superheated steam, and also to prevent rusting fast when not at work 
 the packing rings may be of a very hard bronze like bell-metal. 
 
 There are diagonal bars across the ports to act as retaining guides to the 
 packing rings ; these bars are usually from f to 1J inches broad, and take 
 away about a third of the .gross portway. The passage way around the liner 
 must be so designed as to allow due area of section for the passage of steam ; 
 and to economise space, and reduce the clearance volume, they are eccentric 
 to the liner and valve. To avoid the chief defect in these valves — viz.. large 
 clearance space — the valve should be long, so that its ports are nearlv in 
 line with those at the cylinder bore. 
 
 Piston valves are now very general, and prolonged experience of them 
 has given the necessary confidence, even for their more extended use. For 
 steam of a pressure over 100 lbs. they are a necessity. Some manufacturers 
 use them for the low-pressure cylinder, but few engineers will care to incur 
 the expense of them for a purpose where they are generally quite unnecessarv. 
 They are, however, necessary for the fast-running engines, such as fitted 
 in torpedo boats, destroyers, and very high-revolution express steamers, 
 etc., where safe continuous running is of more importance than extreme 
 economy. They also have some advantages for very large engines, when 
 the port area is so great that two valves are required to avoid cutting away 
 and disconnecting the cylinder ends so much from the body. 
 
 Relief Frames. — The resistance of a flat slide-valve is almost wholly due 
 to the friction on its face, and the greater the pressure on the valve the 
 greater will be the friction. This is due to the fact that the pressure on 
 the back of the valve is only very partially balanced by that on the front. 
 The common locomotive slide-valve may, during certain portions of its 
 stroke, have no pressure at all tending to press it oft" its face. Just at cut-off 
 it is relieved, however, to the extent of the area of one port. > but this relief 
 
DOUBLE VALVES. 
 
 417 
 
 decreases as soon as the steam in the cylinder expands : and if it exhausts 
 before opening at the outer end for steam, its whole area is exposed again 
 to the full amount pressing it to its face. It will be seen from this that the 
 resistance is ever varying, and this condition is true also of double- and 
 treble-ported valves, so that if a definite and fixed area of the back of the 
 valve is so covered as not to be exposed to steam pressure, it does not quite 
 meet the case ; this, however, is the usual way in which the valve is relieved, 
 the area being such that by the balance of pressure the valve is always pressed 
 to its face. An exception to the rule occurs in the relief arrangement of 
 the valves designed by Messrs. J. I. Thornycroft & Co. These valves have 
 a face on their back similar to their front, with recesses corresponding to the 
 ports ; on the back of the valve there is a piate, which has ports like those 
 
 Fig. 139.— Self-balanced Valve. 
 
 on the cylinder face, bearing gently, in one respect, therefore, it is as evenly 
 balanced as a piston valve. 
 
 Double Valves. — Fig. 139 shows an ingenious arrangement by which one 
 valve is made to balance another by attaching one to the other, and as the 
 two valves operate for the same purpose, each may be half the size of an 
 ordinary valve. 
 
 The two valves are faced up, and so set that they are a slack fit between 
 the two faces when cold. The port-ways are, as shown on the diagram, 
 bolted to the main casting on each side, and from their form the pressure 
 of the steam in the valve-box causes them to spring and so approach one 
 another, making the combined valve tight on each face. At the same time, 
 the pressure is not sufficient to cause very great resistance to the motion 
 
418 
 
 MANUAL OF MARINE ENGINEERING. 
 
 of the valve. This form of valve has been found in practice to work very 
 well indeed, and many months of successful use have produced little or no 
 wear on the valve faces. 
 
 Fig. 140.— Thorn's Patent Piston Valve. 
 
 Fig. 140 shows a piston valve adapted by Mr. Thorn with the same 
 principle as the Trick valve. 
 
 This arrangement rather complicates the piston valve and renders it 
 far less simple than in the ordinary form, but there are no practical 
 
 n^Cover Face 
 O O O O 
 
 Fig. 141. — Dawe & Holt's Patent Relief Frame. 
 
 difficulties in the way of accomplishing what Mr. Thorn aims at — viz., to 
 admit some steam for the purpose of cushioning at one end of the cylinder 
 from the exhaust just commencing at the other end. This, when quiet 
 
RELIEF VALVES. 
 
 419 
 
 fin 
 
 Fig. 142. — Common Relief Valve. 
 
 Back or VaZve- 
 
 Fig. 143. — Spiral-spring Packing to Relief Frame 
 
420 
 
 MANUAL OF MARINE ENGINEERING. 
 
 running is necessary, it is hardly needful to say, is of considerable advantage 
 with the low-pressure cylinder,' but it is not so necessary to the medium-, 
 and not at all necessary to the high-pressure cylinder of a triple-expansion 
 
 engine. 
 
 Dawe & Holt's Patent.— This consists of a rectangular cast-iron frame, 
 fitting steam-tight to the face on the back of the valve, and riveted to 
 a steel or bronze sheet, which is itself secured steam-tight to the valve 
 (fig. 141). The diaphragm is of steel hard copper or phosphor bronze, and only- 
 
 Fig. 144. — Martin and Andrews Valve Relief. 
 
 Fig. 144a. — Transverse Section of Martin and Andrews Valve. 
 
 18 B.W.Gr. thick, and so allows the frame to move slightly to suit the 
 valve. This is so made that the frame is pressed back about ^ inch when 
 the valve is in place, and the frame itself, being exposed to steam pressure, 
 is always pressed against the valve, the area enclosed by the frame beiug 
 the amount of relief given to the valve. 
 
 Common Relief Frame. — The ordinary method of relieving the back of 
 the slide-valve from steam pressure is by means of a frame of circular or 
 rectangular form, fitting steam-tight, but freely, into a recess in the valve- 
 
RELIEF FRAME. 
 
 421 
 
 Side Elevation. 
 
 Section through Middle of 
 Valve. 
 
 Fig. 145.— Church's Patent Relief Frame. 
 A. Plan of Relief Ring. | B. Plan with Ring Removed. 
 
422 MAX GAL OF MARINE ENGINEERING. 
 
 box cover, and pressed against the smooth face at the back of the valve by 
 set screws, an elastic material, or spiral metallic springs, as shown in fig. 143, 
 being interposed between the screws and the frame, in order to give greater 
 freedom to it. 
 
 Fig. 142 is a modification of the common plan, and is especially applicable 
 to valves of nearly square form. In this case the relief frame is circular, 
 fitted into a recess in the back of the valve, and pressed out by a star-shaped 
 steel spring. 
 
 Martin and Andrews Valve. — Thornycroft's frame consisted of a casting 
 with recesses and faces corresponding to the ports and bars of the cylinder 
 face ; the valve was with flats and recesses on its back corresponding to its 
 face. The frame rested on legs touching the cylinder face, and just touched 
 the valve back so as to keep the steam off the surfaces of contact. Figs. 144 
 and 144a represent a modification of this plan so improved that what appears 
 as an ordinary single-ported loco valve really acts as a double-ported one 
 by allowing steam to enter at both front and back edge, and to pass to exhaust 
 at two places in the same fashion — viz., by way of the little port crossing 
 from front to back of the valve at each end. This invention is patented 
 by Messrs. Martin and Andrews, and much used by the former engineer 
 for the cylinders of the large and successful quadruple engines turned out 
 by him from the Eoyal Schelde Dockyard at Flushing. 
 
 There are many other plans for preventing slide-valves from pressing 
 unduly on the cylinder face, some by means of springs, and some by pistons 
 acting perpendicularly to the direction of motion of the valves, but none of 
 them are free from objection, nor have any given greater satisfaction than 
 Church's (fig. 145) very ingenious arrangement. 
 
 Through Exhaust Valve. — Fig. 146 shows a low-pressure double-ported 
 valve arranged to exhaust through its back direct ; provided the packing- 
 ring is well designed, well made, and kept in good order, this is a safe and 
 useful one. as it permits of the face being much closer to the cylinder than 
 ordinarily, and a perfectly unobstructed lead to the condenser. 
 
 Back Guides and Springs. — To prevent slide-valves from being forced 
 from the cylinder face, it is necessary to have guides behind the valves, with 
 suitable rubbing surface on the valve itself ; but as it is not always easy to 
 provide such guides, and since, when the valve and face have worn some- 
 what, they fail to keep the va^ve to its face, it is generally preferable to fit 
 a pair of flat bar springs, whose backs press on rubbing strips on the valve, 
 and the ends on a fixed part, one end of each being secured, and the other, 
 free to slide on the flattening of the arc. 
 
 Balance Pistons. — Since the weight of the valves, their rods, and gearing, 
 in the case of vertical engines, is taken by the eccentric straps, and if it were 
 unbalanced the top half of the straps would be subject to more wear than the 
 lower, it is advisable to provide some means of avoiding this. The readiest 
 is by fitting to the top end of the valve-spindle a piston (figs. 78 and 138), 
 which works in a cylinder provided in the valve-box cover or top ; this 
 piston is of sufficient area that the steam-pressure on it affords the balance 
 required, and it also acts the part of a guide for the spindle. When the 
 eccentric-rods are very long, it is better to give an excess of area to the balance 
 piston, so that the rods may be always in tension, instead of alternately in 
 tension and compression. Since the pressure in the valve-box of the low- 
 
BALANCE PISTONS. 
 
 423 
 
 and medium-pressure cylinder of a compound engine is constantly varying 
 it is better, especially when the above object is aimed at, to place the balancing 
 cylinder outside the box, and supply steam from the boiler to the under side 
 of the piston, so that the balancing force may be constant ; but, on the 
 
 a 
 o 
 
 TO 
 
 1-*' 
 
 O as 
 
 ■a « 
 
 O » 
 
 a -a 
 
 43 
 
 em 
 
 s 
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 •s 
 
 -u 
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 9 
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 4= 
 M 
 
 w 
 
 ► 
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 • 60 
 S ;•" 
 
 .2 ^ 
 
 u 
 
 a, 
 CO 
 
 '+3 
 
 other hand, since the resistance of the valve varies with the pressure in the 
 valve-box, there is not the same necessity for so much balancing force as 
 would be the case were it otherwise ; so that, on the whole, the piston exposed 
 to the pressure of the valve-box is sufficient for the needs of most engines. 
 The top of the balance cylinder of the high- and medium-pressure valves 
 
424 
 
 MANUAL OF MARINE ENGINEERING. 
 
 should be connected to the next receiver by a small pipe, and that of the 
 low-pressure valve to the condenser. 
 
 Joy's Assistant Cylinder. — In the hands of the late David Joy the balance 
 piston with its cylinder has been transformed by a very ingenious and simple 
 device, from merely passively supporting the weight of the valve and its 
 gear, to actively aid the link motion in moving the valve up and down. 
 Fig. 147 shows the method by which this was accomplished at the outset ; 
 here the steam is admitted to the top and bottom of the piston through 
 
 Pass Port 
 
 Steam 
 
 Fig. 147a. — Joy's Improved Assistant Cylinder. 
 
 To P Bottom, 
 
 Section ThroughA B. 
 
 Fig. 147.— Toy's Assistant Cylinder. Fig. 1476. — Indicator Diagram from Joy's 
 
 Assistant Cylinder (fig. 147a). 
 
 passages in the piston itself which open to steam ports in the sides of the 
 cylinder at or near the end of the stroke. The exhaust is also through ports 
 in the side of the cylinder which are opened and closed by the piston itself 
 towards the end of the stroke. Fig. 147a shows another method, which 
 Mr. Joy called the compound system, inasmuch as the steam contained in 
 the body of the piston, after partially expanding on the upstroke, is passed 
 by means of a recess near the top of the cylinder into the space above the 
 piston so as to expand there and so impel the piston downward. The steam 
 
VALVE-RODS OR SPINDLES. 425 
 
 obtained access from the underside to the interior of the piston by means 
 of another recess near the bottom. The valve spindle rod is reduced in 
 diameter for some little distance below the piston, and, passing through a 
 sleeve with port holes in it, does duty as a valve for the admission and cut- 
 off of steam. Just before cut-off the piston passes the lower recess so as 
 to allow its interior to fill. The piston face in way of the exhaust ports is, 
 of course, plain, so as to completely cover them and prevent the steam inside 
 the piston from passing to exhaust. 
 
 Fig. 1476 shows a set of indicator diagrams taken from an assistant 
 cylinder working under these conditions, and they certainly are remarkable, 
 considering the simple method by which the steam is distributed and released. 
 These assistant cylinders have been fitted by several firms to fast-running 
 engines which have of necessity heavy slide-valves, especially those of the 
 low-pressure cylinder. The results have been very satisfactory, and the 
 headgoing, eccentrics, etc., relieved to a very considerable extent from the 
 heavy work entailed in operating such valves at a high speed. 
 
 * Valve-rods or Spindles. — Since it is possible for a slide-valve to be exposed 
 to the pressure of steam on its whole area, without any relief due to the 
 ports, etc., and this may occur even when the valve is fitted with relief frames, 
 it is better to assume this in making all calculations for determining the 
 sizes of the parts to move it. 
 
 If L be the length of a valve, and B the breadth in inches, p the maximum 
 absolute pressure to which it is exposed in pounds per square inch, then, 
 
 Maximum pressure on the valve = L x B x p lbs. 
 
 The coefficient of friction should be taken at 0*2, or that of metallic 
 surfaces rubbing together dry, as this is the worst condition likely to occur, 
 then, 
 
 Load on valve-rod = 0*2 (L X B X p) lbs. 
 
 Since the spindle has to take stresses similar to those on a piston-rod, it 
 may be dealt with in the same way. 
 
 A stress of 3,000 lbs. per square inch should be allowed in the case of 
 comparatively long rods, and 3,600 when very short and well-guided. Hence 
 
 Diameter of slide-valve rod 
 
 When the rod is long 
 ,, ,, short 
 
 In the case of the valve of the low-pressure cylinder, the pressure should 
 be taken at 25 lbs. absolute, as the pressure in the receiver is seldom as high 
 as 15 lbs. above that of the atmosphere. For convenience, the rods for 
 both high-, medium-, and low-pressure cylinders should be alike, and con- 
 sequently the size should be the largest given by calculation from the above 
 rules. 
 
 When the guides of the valve-rod are above the rod-end {v. fig. 148), 
 so that in reversing the link a severe bending strain would come on the rod, 
 it should be somewhat larger in diameter in the guide. Tn cases of this kind, F 
 should be taken at 20 per cent, less than given by the above rules for that part. 
 
 *»r^^. ..i..^. jo , c . •.!• diameter piston-rod ... . 
 * X.L. Coast Inst. E. and S. rule for valve spindles is , 1- J inch. 
 
426 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Valve-rod Bolts. — When the joint at the valve-rod end is made capable 
 of adjustment by means of two bolts (fig. 152) ; then, 
 
 /T v T^ v ti 
 _ TL -f 0-25 inch. 
 
 When the bolts are of mild steel, H = 48,000. This same rule applies 
 
 ■ Y//M/////////A 
 
 Fig. 148.— Valve-rod Guide. 
 
 to the belts of eccentric-rod ends, and to the bolts at the butt-end of 
 eccentric-rods. 
 
 For the bolts of eccentric straps, H = 40,000 for mild steel. 
 
 Valve-rod Guides. — The valve-rod end should always have a guide, so 
 that the side strain due to the obliquity of the eccentric-rods, or to the action 
 
VALVE-ROD GUIDES. 
 
 427 
 
 of the link motion, shall not only not bend the rod, but shall not come on 
 the gland and so cause leakage, etc., through the stuffing-box. The kind 
 of guide depends very much on the link-motion used, as what suits one form 
 is not applicable when another kind is used. The simplest form consists 
 of a cast-iron bracket with an eye bushed with bronze, through which the 
 rod passes ; but since the rod soon wears away the bush, if subject to much 
 side pressure, this is not a satisfactory plan for large engines, although a 
 
 Fig. 149. — Proportions of a Common Valve. 
 
 Fig. 149a. — Proportions of a Trick Valve. 
 
 convenient and fairly trustworthy one for small ones. If the rod is made 
 of larger diameter in wake of the guide, and planed flat on the sides which 
 are exposed to pressure, and the bracket fitted with brass liners, shaped 
 like a common key, so as to be secured by screws through the gib ends, a 
 simple and efficient means is provided for taking all side-pressure, but the 
 drip from the gland is apt to wash out the oil. A better but more costly 
 
428 MANUAL OF MARINE ENGINEERING. 
 
 one is to fit a bracket to the valve-rod above the link, having a shoe like 
 that of a piston-rod, and running on a guide like that of a piston-rod end. 
 In large engines it is fitted with a slipper, etc., for adjustment (fig. 148). 
 
 Slot links require a different kind of guide when they work on over- 
 hung pins, which was often the case, of necessity, in horizontal engines, to 
 obtain eccentric-rods of sufficient length. Then the valve-rod is fitted into 
 the socket of a guide-piece, shaped somewhat like a bayonet, having a shank 
 of square section and working in a guide bracket fitted to the framing. 
 
 Proportions of Slide-valves are found as follows : — 
 
 Let x be the outside lap of the valve at the front end. and y that at the 
 back end. Then, 
 
 c a 
 
 H = T + x ; and K = -^ + y. 
 
 Let z be the inside lap at the front, and iv that at the back. Then 
 
 B = - - — z ; and C = « — w. 
 
 Ji 1 
 
 To keep the valve as small as- possible, E need be only a little wider than 
 the steam ports, provided that the half travel does not materially exceed 
 the width of bar. 
 
 The Double-ported Valve dimensions follow similar rules to the above. 
 
 Let x, y, z, and w be the laps as before. Then 
 
 Also, 
 
 H 
 
 G _L 
 
 
 and K = - + y. 
 
 B 
 
 F 
 
 = 2 - * ; 
 
 
 F 
 
 and C = — w. 
 
 a 
 
 A 
 
 G . 
 
 inch ; 
 
 and D = % f 4 
 
 2 8 
 
 The openings through the valve laps or covers must be as large as possible, 
 but need not exceed the ordinary opening of the valve to steam at the outer 
 edge ; then 
 
 G+ P = K+ N; 
 
 and G + Q = H + M. 
 
 Valve Gear : Link Motion. — The common form of motion employed to work 
 the valves of a screw engine consists essentially of two eccentrics keyed on 
 the crank-shaft in such a position relatively to the crank that when one is 
 operating on the valve, the engine will propel the ship ahead, and is said to 
 be in head-gear, and when the other, the engine will propel the ship stern 
 first, and is said to be in stem-gear, their rods being connected by a bar or 
 bars on which is a sliding block to which the valve spindle is attached. This 
 bar connection is called the link, and is of such a form that by sliding it 
 through the block, the head or stem eccentric may at pleasure be brought 
 to operate on the valve. 
 
 In designing a link gear, the most important objects are to give the valve 
 such motions as shall cause it to open to steam at or slightly before the piston 
 
LINK MOTION. 429 
 
 is at the end of its stroke, the amount by which it i* open at the end of the. 
 stroke, or commencement of the next stroke, being called the lead of the 
 valve ; then, to open fully, and close quickly at the required period of the 
 stroke of the piston called cut-off ; to confine the steam during he remaining 
 portion of the stroke, that it may expand in the cylinder, and at or just before 
 the end of the stroke to allow the steam to escape freely from the cylinder, 
 called exhaust ; to close the port again some time before the end of the stroke, 
 so that the piston comjyresses the steam remaining in the cylinder and port, 
 to form an elastic cushion for the piston, etc. These operations should 
 be effected with the expenditure of as little power as possible, and with 
 this end in view the motion of the link should be, as far as can be, limited 
 to moving the valve only ; consequently the link itself should have, wher: 
 in full gear, no sliding motion longitudinally, called slotting motion, in the 
 block, but only the angular motion due to the two eccentrics. A perfect 
 valve motion is such that the valve opens to steam wide soon after the crank 
 has passed the dead centre, and remains wide open during the admission 
 of steam, so that there is no wire-drawing ; the valve should close suddenly, 
 and remain closed during expansion ; at the end of the stroke it should open 
 wide to exhaust, and remain in that state during the greater part of the 
 period of exhaust, when it should close suddenly, and remain closed till 
 opening again to steam. This is not obtainable with the ordinary link motion, 
 nor to its full extent with any motion when one valve only is employed foi 
 both ends of the cylinder, because the period of cut-off at one end does not, 
 as a rule, correspond to the period of compression at the. other end ; but 
 there are valve gears which have two periods of very quick motion, and two 
 of very slow in each cycle which very closely fulfils the above conditions, 
 and which will be noted later on. 
 
 Slot Link. — This, which is one of the oldest forms of link, is still retained 
 by many engineers, and is well adapted to the circumstances of several forms 
 of engine, such as the oscillating paddle-wheel engine, and all engines in 
 which there is not direct connection between the eccentric-rods and the 
 valve-rod, as in some of the horizontal engines when it is either impossible 
 or inconvenient to have direct connection. It is also much liked by those 
 responsible for quick-revolution engines, as there are no nuts to slack back 
 or bolts to break. 
 
 Fig. 150 is an illustration of the ordinary slot link, having means of adjust 
 tnent for the sliding block and eccentric-pin brasses. Locomotive engineers 
 prefer, as a rule, to have the pin-holes fitted with hard bushes rather than 
 adjustable brasses ; but this opinion is not shared by marine engineers, 
 and chiefly on the ground that in a foreign port it is seldom possible and 
 never convenient to engage the services of workmen and tools to renew 
 these bushes when so badly worn as to require adjustment and therefore 
 renewal. 
 
 This kind of link is generally suspended from the end next the head- 
 going eccentric-rod, at a point in line with the arc through the block-pin ; 
 and if the pin in the lever, which operates on the link to reverse it, is placed 
 in the proper position, there is very little slotting motion indeed when working 
 in head-gear. The same remark applies to the position of the pin when in 
 stern-gear, except that the amount of slotting motion is somewhat greater 
 of necessity ; but since a marine engine, as a rule, works very little in stern- 
 
430 
 
 MANUAL OF MARINE ENGINEERING 
 
 gear, and its efficiency there is of small consideration comparatively, this 
 defect is of little moment. Locomotive engineers still keep to the slot link, 
 as it has no small bolts and nuts to rattle loose and give trouble, as might 
 be the case with other gears ; but in the locomotive the eccentric-rod ends 
 are often connected to the ends of link at the place where the drag-rod is 
 in fig. 150. 
 
 Position of Suspension Pin. — To obtain the best position of reversing 
 ievei pin, it is necessary to delineate the path of the centre of pin in the 
 link end through one revolution of the engine when the link has no slotting 
 motion. The path so found is like an attenuated figure 8 in head-gear, and 
 somewhat more pronounced in stern-gear. The arc of a circle of radius 
 equal to the length of the suspension or bridle-rods, is then drawn through 
 each of these figures in such, a way that there is the least possible deviation 
 
 Fig. 150.— The Slot Link. 
 
 from the figure on either side — that is, the arc is the centre line of the figure, 
 if the . deviation on one side equals that on the other. The centres of the 
 circles to which these arcs belong should be the centres of suspension of the 
 bridle-rods, or position of pin in rever ing lever end. By drawing arcs of 
 circles of radius equal to the length of the reversing lever from these two 
 centres, the points of intersection are the two possible positions for the centre 
 of weigh-shaft. 
 
 This same method of construction is suitable to all kinds of links, and 
 for all positions of the point of suspension of the link. 
 
 When it is necessary that the valve motion shall be as satisfactory in 
 head-gear as in stem-gear, the link should be suspended from a point in the 
 line dividing it symmetrically, and by preference at the intersection of this 
 line with the arc through the centre of block-pin, so that the centre of 
 
DOUBLE-BAR LINKS. 431 
 
 suspension is in line with the centre of block-pin when in mid-gear. When 
 this is so, the pins for suspending the link are on side plates bolted to the 
 sides of the link. 
 
 The distance from centre to centre of eccentric-rod pins should not be 
 less than two and a half times the throw of the eccentrics, and is usually, 
 when space permits, two and three-quarters to three times. The throw of 
 the eccentrics in this case is, of course, equal to the travel of the valve when 
 in full gear. 
 
 Size of Slot Link. — Let D be the diameter of the valve spindle, R being 
 the revolutions per minute, and F = (13,500— 10 R), from the following 
 calculation 
 
 then 
 
 Diameter of block-pin when overhung . . 
 
 ,, ,, ,, secured at both ends 
 
 „ eccentric-rod pins .... 
 
 ,, suspension-rod pins .... 
 
 ,, suspension-rod pin when overhung . 
 
 Breadth of link ....... 
 
 Length of block ....... 
 
 Thickness of bars of link at middle .... 
 
 If a single suspension-rod of round section, its diameter 
 If two suspension-rods of round section, their diameter 
 
 The objections to the slot link are, that it is an expensive one to make, and 
 that, owing to the eccentric-pins and the block-pins being out of line, there 
 is always an uneasy motion about the block-pin, and more slotting motion of 
 the block. The former objection is valid, especially when the link is made 
 of wrought material, but when the link is a steel or bronze casting it is not so 
 expensive as some other forms. The uneasy motion is often due to bad 
 design, for when well designed and carefully hung, it will work quite satis- 
 factorily. 
 
 Single-bar Link. — This kind of link consists of a single solid bar, ol 
 rectangular section generally, and having the eccentric-rods connected to 
 each end, and a sliding block between, to which the valve spindle is connected. 
 The form of link, although tried by more than one eminent firm of engineers, 
 was gradually dropped, except by the firm that introduced it. 
 
 Double-bar Links. — There are two kinds of double-bar links ; one (fig. 151) 
 having the eccentric-rod ends, as well as the valve-spindle end between the 
 bars, so that the travel of the valve is less than the throw of the eccentrics ; 
 the other (fig. 152) has the eccentric-rods formed with fork ends, so as to 
 connect to studs on the outside of the bars, and thus admits of the block 
 sliding to the end of the link, so that the centres of the eccentric-rod ends 
 and the block-pin are in line when in full gear. 
 
 The former plan is cheaper to make, is simpler in construction, and has 
 fewer parts to get out of order and adjust ; and when adjustment is required, 
 it is easier to make, and there is less chance of its being done improperly. 
 When in head-ge&x, part of the work of moving the valve is done by the 
 stern-going eccentric, so that the wear is not limited to the one eccentric 
 
 
 D. 
 
 
 
 =1 
 
 0-75 
 
 X D. 
 
 
 = 
 
 0-7 
 
 X D. 
 
 
 — 
 
 0-55 
 
 x D. 
 
 
 =.: 
 
 0-75 
 
 X D. 
 
 
 
 0-8 to 0-9 > 
 
 D 
 
 i= 
 
 1-8 to 1-6 > 
 
 D 
 
 
 0-7 
 
 X D. 
 
 
 
 0-7 
 
 X D. 
 
 
 Z.I 
 
 0-55 
 
 X D. 
 
 
432 
 
 MANUAL OF MARINE ENGINEERING. 
 
 strap. When properly hung, the slotting motion is exceedingly small, and 
 the valve motion is as perfect as with the other form. The objection to it 
 is that the eccentrics are larger in diameter than those with the other links, 
 and the links themselves are longer, and more space is required for the 
 eccentric-rods to move in. 
 
 Size of Bar Links. — Let D be the diameter of valve spindle, found as before. 
 
 Fig. 151. — Double-bar Link with Rods inside. 
 
 Fig. 151, distance between centres of eccentric-pins, 3 to 4 times thro* 
 of eccentrics. 
 
 Depth of bars, . 
 Thickness of bars 
 Length of sliding block 
 Diameter of eccentric-rod pins 
 
 centre of sliding block 
 
 = 1-25 xD + | inch. 
 = 0-65 X D + £ inch. 
 = 2-5 to 3 x D. 
 = 0-8 x D -f- £ inch. 
 = 13 x D. 
 
 Fig. 152. — Double-bar Link with Rods outside. 
 
 Fig. 152, distance between eccentric-rod pins 2J to 2| times throw of 
 eccentrics. 
 
 Depth of bars . . . 
 
 Thickness of bars 
 Length of sliding block 
 Diameter of eccentric-rod pins 
 Length 
 
 = 1-25 x D + A inch. 
 = 0-65 X D + | inch. 
 = 2-5 to 3 x D. 
 = 0-75 x D. 
 = 0-80 xD + J inch. 
 
 Diameter of eccentric bolts (top end) at bottom of thread — 04 X D when 
 of mild steel. 
 
VALVE GEARS. 433 
 
 These bars should be of a very good description of iron and case-hardened, 
 or of hard steel : the latter is, of course, free from seaminess and stronger. 
 The eccentric-rod pins of the kind of link (fig. 152) are usually forged solid 
 with the bars, but there is no absolute need of this, and it adds very much 
 to the cost, both of manufacture and renewal when worn. Since the wear 
 on these pins is limited to a very small portion of their circumference, it is 
 not unusual to file away the parts which are not subject to wear, so as to 
 admit of the brasses being closed when worn. When loose pins are fitted, 
 they should be steel, and hardened, so that all wear may come on the brasses 
 which are capable of adjustment. 
 
 In another arrangement of bar-link motion, the sliding block is divided, 
 and on the outside, while the eccentric-rod ends are between the bars. This, 
 while having some slight advantages, is on the whole very clumsy, and the 
 block-pins wear badly ; besides which, the link can be only suspended from 
 the extreme end. Mr. Martin, of Flushing, adopts an ingenious combina- 
 tion of these bar links for the large engines he turns out. He fits three 
 eccentrics to each valve ; the rods of the two outer ones are connected each 
 to a pin on the outside of the link, as in fig. 152, and act in " ahead " gear. 
 The rod of the centre eccentric is connected to a pin between the bars, as in 
 fig. 151, and acts in " astern " gear. 
 
 Single Eccentric Gear. — The valves of slow-working engines can be worked 
 by a single eccentric, which is free to move round on the shaft from the 
 position for head-geav to that for stem-gear ; it is driven by a key or stop fixed 
 to the shaft, pressing against a shoulder on the side of the eccentric. The 
 eccentric sheave is balanced so that it will keep in any position, and only 
 move when driven by the stop on the shaft. The eccentric-rod is so fitted 
 that it may be disconnected from the valve-rod or its gear. This is generally 
 effected by providing a gab or gap in the eccentric-rod end instead of a pin- 
 hole, which allows the eccentric-rod to be lifted from its pin by suitable 
 gearing. When the eccentric-rod is disconnected from the valve gear, the 
 valve ceases to move, and the engine comes to rest. To restart the engine 
 the valve must be worked by hand until the gab can be brought in line with 
 the pin : if the motion of the engine is reversed the shaft will move around, 
 the eccentric remaining motionless until the stop on the shaft comes in con- 
 tact with the other side of the projection on the sheave ; it is then in the 
 right position for driving the valves, and the eccentric-rod may now be 
 dropped into gear. This method admitted of the paddle engine being 
 handled very dexterously when the valves were not so large and the steam 
 pressure so high as to prevent their being moved by hand, and was the one 
 generally adopted in engines of moderate power long after link motion was 
 invented and in general use on locomotives. 
 
 There have been many other methods of moving a single eccentric from 
 ahead to astern position; some of which by sliding wedges were at one 
 time received favourably, but are now almost forgotten. 
 
 Another favourite method is that shown in fig. 69, where by means 
 of a spiral slot in the eccentric sleeve and a straight one in the shaft a sliding 
 bolt in the latter is made to twist the eccentric round. 
 
 Hackworth's Dynamic Valve Gear. — The locus of a point on a rod. one 
 end of which moves in a circle, and the other on a straight line passing thiough 
 the centre of that circle, is an ellipse whose major axis coincides with the 
 
 28 
 
434 MANUAL OF MARINE ENGINEERING. 
 
 straight line. If, however, the end of the rod slides on a line inclined to 
 this centre line, the major axis of the ellipse will be inclined. 
 
 Hackworth's valve motion worked on this principle, for this gear had 
 a single eccentric keyed to the shaft exactly opposite to the crank ; the eccentric- 
 rod had its end attached to a block, which slid on a guide-bar inclined to 
 the line through the centre of shaft ; and a rod from about the middle of the 
 eccentric-rod was connected to the valve-rod, which works in a line at right 
 angles to the line through the eccentric-rod when in mean position. If the 
 guide-bar on which the eccentric-rod block slides be moved so as to reverse 
 its inclination, the inclination of the axis of the ellipse was reversed, and 
 the motion of the engine reversed. Here the motion of the valve is due 
 to the eccentric, and the lead to the inclination of the slide-bar. 
 
 The great advantage of this gear, beyond the saving of an eccentric, is 
 the better motion imparted to the valve, inasmuch as there are two quick 
 and two slow motions in a revolution ; the quick ones occurring at opening 
 and cut-off, and the slow ones during entry and at exhaust previous to opening. 
 A large variation in the amount of cut-off is possible with this arrangement, 
 without wire-drawing from small opening and slow closing of the port, as is 
 the case with the common link motion. The chief objections urged against 
 this gear are the excessive friction, and consequently wear on the sliding 
 blocks, and the liability of so many pins to derangement. The first of these 
 is the most valid, and it has been overcome by fitting rollers instead of sliding 
 blocks. In both ways, however, this gear has worked fairly well ; and for 
 engines of small power it is a very convenient arrangement, especially when 
 much variation in cut-ofi is required. 
 
 Since the valve-rods with this gear do not come in line with the piston- 
 rod.^, or immediately over the line of shafting, the valves and their chests 
 are removed from the usual positions, so as to admit of the cylinders being 
 closer together ; the engine is consequently shorter, cheaper, and occupies 
 less space. This suited the expansive screw engine and paddle engines of 
 the diagonal type admirably ; for the two valves of an expansive engine 
 being in one common chest, could be examined through a door in front, of 
 ample dimensions to admit a man ; for the compound engine the valve- 
 boxes are side by side with short-pipe connections (vide fig. 21) ; but generally 
 it allows of little space for the receiver, and the exhaust from the low-pressure 
 cylinder of screw engines must pass through a belt within the receiver, 01 
 else this belt has to traverse more than half the circumference of the 
 cylinder. 
 
 Marshall's Valve Getfi is a modification of Hackworth's, and differs from 
 it in the method of getting the oblique motion of the rod end. Fig. 153 
 shows the plan adopted by the late F. C. Marshall, and also illustrates generally 
 what has been said of Hackworth's. Here the eccentric-rod is hung, by 
 means of a rod from the end of a lever on a reversing shaft, in such a way 
 that it moves on the arc of a circle inclined to the centre line. The motion 
 is not quite so perfect as with the inclined sliding bar, and necessitates double 
 ports to the bottom end of the slide-valve, in order to get as much opening 
 to steam there as at the top end ; but there is less friction, and on the whole, 
 it works most satisfactorily. The pins require to be of good size, and they 
 should all have adjustable brasses to provide for the large amount of wear 
 which of necessity comes on them. 
 
VALVE GEARS. 
 
 435 
 
 Joy's Valve Gear.— Fig. 154 shows the ingenious arrangement whereby 
 the late David Joy has avoided altogether the use of eccentrics. He obtained 
 
 Fig. 153. — Marshall's Patent Valve Gear. 
 
 the motion from the connecting-rod and qualified it by a sliding block 
 like Hackworth, or a suspension-rod like Marshall. The motion thus 
 
436 
 
 MANUAL OF MARINE ENGINEERING. 
 
 obtained is a very perfect one for a slide-valve, as the two quick and two 
 slow periods are just when required, the amount of opening is equal at 
 
 both ends of the valve, and early 
 i > i r- — >_ cut-offs can be effected without ex- 
 
 cessive leads and compressions or 
 premature exhaustings. This gear 
 has been applied with great success 
 to locomotives, where the saving of 
 space for the eccentrics admits of 
 longer crank-pins and journals ; it 
 was also taken up by a few marine 
 engineers. The chief objections to 
 this gear are that it comes in the 
 way of the principal working parts, 
 and makes them a little difficult 
 to get at when working, and also 
 that a small amount of wear on the 
 joints of the gear would cause a 
 defect on the valve motion, and pro- 
 duce a serious amount of rattle of 
 the gear. These, however, can 
 be got over by making the pins 
 substantial, and all the joints 
 adjustable. 
 
 Sell's Valve Gear. — Here the 
 valves are worked by an inde- 
 pendent shaft, which derives its 
 motion from the main shaft by 
 means of wheel-gearing ; between 
 these two shafts are two inter- 
 mediate wheels gearing into one 
 another — one of them gears into the wheel on the main shaft, and the other 
 into that on the valve shaft. These two intermediate wheels are carried on 
 a frame which can be lifted up and down, and is guided in such a way as to 
 keep the wheels always in proper gear with their corresponding wheels. If 
 the gearing is set so that the valves are moved by their crank-shaft, and 
 drive the engine ahead when the frame is at the bottom of its traverse, the 
 moving of the frame to the top of its traverse so alters the relative angular 
 position of the shafts, that the valves will be set to drive the engine astern. 
 This is easily understood by supposing the engine at rest, and the frame 
 raised ; the wheel geared into the one on the crank-shaft, in moving round 
 the shaft also turns on its own axis, and in so doing turns its companion- 
 wheel, which turns the wheel on the valve shaft, and consequently the valve 
 shaft itself. 
 
 This gear was fitted to many engines in the British and foreign navies, 
 and was especially suitable for the three-cylinder horizontal engines for 
 which it was originally designed. It had, however, the objectionable feature 
 of wheel-gearing, which, when new, made a great noise, and was liable to 
 accident from racinq of the engines, but when worn, these are magnified 
 very considerably by the hack lash which results. Had the wheels been 
 
 Fig. 154. — Joy's Valve Gear. 
 
• EXPANSION VALVES. 437 
 
 machine-cut accurately to shape, with helical teeth of sufficient size, this 
 gear might have been in use to-day, as it was a very convenient one for hori- 
 zontal engines, as also it is for large vertical ones. A similar arrangement 
 has been used by certain engineers for vertical triple engines, but the auxiliary 
 or valve shaft was in this case driven by an auxiliary engine, which had a 
 sensitive-ball governor to regulate it. In rough weather the wheels connecting 
 the two shafts were thrown out of gear, and the main engine was said to be 
 satisfactorily governed by the small one controlled by its own ball governor. 
 
 Expansion Valves. — When ordinary link motion is employed, and a much 
 earlier cut-off than half-stroke is required as a normal condition, it is neces- 
 sary to effect it by means of an independent valve, usually called the expan- 
 tion valve. These valves are generally common plates, sliding either on a 
 face with ports on the side of -the valve-box, or on a face on the back of the 
 ordinary slide-valve. In the latter case there are two methods, commonly 
 known as the inside cut-off and the outside cut-off. 
 
 Gridiron Expansion Valves. — When the valve worked on an independent 
 face on the side of the valve-box, it consisted of a plain plate with steam 
 ports in it, corresponding to the steam ports on the face ; there was sufficient 
 lap. and the gear was so set that it remained closed after it had cut-off, 
 until after the main valve was closed to the cylinder, when it might open 
 again so as to fill the valve-box, and be ready to supply and cut off steam 
 from the other end of the cylinder. The variation in cut-off is effected by 
 varying the travel, and the equalisation of cut-off at each end is effected by 
 varying the proportion of lap on each side of the ports. The variation in 
 travel is obtained by means of a link on which the block to which the ex- 
 pansion eccentric-rod is attached slides, so as to vary the length of lever ; 
 the larger the travel of the expansion valve, the quicker is the steam cut 
 off, and the smaller the travel, the later is the cut-off, until, finally, the 
 travel is so small that the expansion valve does not close at all. The varia- 
 tion in the lap of the expansion valve is effected by means of the curvature 
 of the link. 
 
 Some engineers preferred the mean position of the expansion valve to be 
 with the ports closed ; in this case the quickest cut-off is effected with the 
 least travel, and consequently with least opening ; but since the piston 
 speed is comparatively slow when receiving steam with the early cut-off, a 
 reduction of opening will not be felt ; whereas the slow cutting-off when 
 carrying steam to mid-stroke, effected by the valve, when the late cut-off is 
 with least travel, produces considerable wire-drawing. 
 
 If the engine is generally to work with a cut-off at from ^ stroke to | 
 stroke, then the former valve arrangement is the better ; if the cut-off is to 
 be, as a rule, from f to § the stroke, then the latter is better, except that 
 it is somewhat inconvenient to have the ports closed when the valve has 
 least travel, and also that when the expansion valve is not required it must 
 be moving at its highest speed. 
 
 The chief objection to this kind of expansion valve is the large amount 
 of clearance space between it and the piston, for the steam expands in the 
 valve-box, as well as in the cylinder, until the main valve closes. It was, 
 however, very generally used in the Navy for the cut-off for cruising speeds, 
 and had the advantage of not interfering with or in any way affecting the 
 main valves in case of anything happening to it. 
 
438 MANUAL OF MARINE ENGINEERING. 
 
 Outside Cut-off Valves. — When the expansion valve works on a face on 
 the back of the main valve, the latter is provided with steamways through 
 it from the ports on its back to the ports on its face. If the main valve 
 is single-ported, it is like a common locomotive slide-valve with enclosing 
 ends, so that steam passes to the cylinder only through the passages thus 
 formed. There are then two ports on the back of the valve, one of which 
 is covered at cut-off by the expansion valve, and remains covered until the 
 main valve cuts off. The expansion valve may consist of a single flat plate, 
 which, when in mean position with respect to the main valve, is between the 
 j)orts on the back of the main valve. If any variation of the cut-off is required 
 in this case, it is effected by varying the travel of the expansion valve with 
 respect to the main valve, or by altering the lead of the expansion eccentric. 
 The latter plan would generally be very inconvenient, and, consequently, 
 when such a valve is fitted, the variation is effected by varying the travel. 
 
 On account of the cut-off being effected by the outside edges of the 
 expansion valve and the ports in the main valve, these are called outside 
 cut-off valves. Since, however, the motion of the valve will be slow when 
 cutting off with a small relative travel, it is unusual to fit a single plate, 
 but to divide it and arrange the gear so that the two plates may be moved 
 apart from one another so as to virtually increase their lap, and cause them 
 to cut off earlier with the same travel. The result of this arrangement is 
 that the expansion valve moves always at the same velocity and through 
 the same space ; the cut-off occurs when moving at or near the maximum 
 velocity, whatever be the period of cut-off, and there is consequently little 
 or no wirc-draiving at the ports. The plates should be so designed that 
 when cutting off at the earliest required period, they do not overrun the 
 port so as to pass steam through at the inner edges, and that they may 
 approach close enough together as to allow of a late cut-off if required. 
 
 The usual method of altering the position of the plates is by securing 
 them by nuts to a common spindie, on which is cut a right-handed thread 
 for the one plate, and a left-handed thread for the other, so that when the 
 spindle is revolved the plates move in opposite directions. The nuts should, 
 of course, be made of bronze, and provided with collars or other suitable 
 device, so that when the spindle is turned round they do not revolve with 
 it, but only move the plates. The valve spindle usually passes through 
 both ends of the valve box, and is connected to the eccentric-rod by a 
 swivel-joint, which permits it to be turned round ; the other end has a 
 featberway cut in it, and passes through a socket, which is held in place 
 by a suitable bracket, provided with a fixed feather which fits into the 
 groove in the spindle, and a wheel by which it can be turned ; by these 
 means the spindle can be turned while the engine is at work. 
 
 This plan suffices for small engines, but the power required to move the 
 spindle rapidly increases with the size of the engine, and this not so much 
 owing to the friction of the valves or external fittings, as to that of the 
 screwed parts of the spindle in the nuts on the back of the valve from want 
 of lubrication. This difficulty is so great that in engines of 50 N.H.P., 
 only after a few days of work without altering the position of expansion 
 valve, it has been found impossible to turn the spindle round with the means 
 at command. To avoid this difficulty the gear for varying the cut-off should 
 be outside the valve-box, and under the inspection of the engineer. A 
 
EXPANSION VALVFS. 439 
 
 good plan for carrying out this is to provide each valve plate with a separate 
 rod ; each rod passes through a stuffing-box, and is connected to a cross- 
 head by a nut in it, which is free to turn round so as to adjust the position 
 of the spindle. These nuts may have wheels keyed to them gearing into 
 one another, or they may have worm-wheels operated on by a worm between 
 them ; in either case the nuts will turn round in opposite directions, so 
 that if the threads on both the spindles are right-handed, the turning of the 
 nuts will cause the valves to move in opposite directions. The spindles 
 are secured to the valve plates, so that they cannot turn round when the 
 nuts are turned ; they are also of necessity out of centre with the valves, 
 which must consequently be guided sideways on the main valve. The only 
 objectionable feature in this arrangement is, that the spindles are not secured 
 to the middle of the valves ; but practice has shown that this is no detri- 
 ment, and the whole system works exceedingly well. 
 
 If the engine to which this kind of expansion valve is fitted, is required 
 to work with the same efficiency in stern-gear as in head-gear, the eccentric 
 should be in line with and on the side of the shaft opposite the crank, and 
 its throw should not be less than that of the main valve eccentrics. If, 
 however, the engine will seldom, and for only short periods, work in stern- 
 gear, the eccentric should be on the same side as the crank, but nearer the 
 stem-going eccentric than the head-going one ; then the relative travel of 
 the valves is greater in head- than stern-gear, and the cut-off is also earlier 
 in head-gear, so that if a sudden order were given to astern, the engine is 
 better prepared for carrying out that order. 
 
 The equalisation of cut-off at top and bottom of cylinder, for all positions 
 of cut-off, is effected by causing one plate to move more per revolution of 
 the spindle than the other ; and in the case of the double spindle arrange- 
 ment, the wheels on the nuts are so proportioned as to obtain the same result. 
 
 Inside Cut-off Valves. — So called, because steam is cut off at the inside 
 edges of the expansion valve. The eccentric is in this case, also, nearly in 
 line and on the same side as the crank, so that the relative travel with the 
 same throw of the eccentrics is greater than in the case when the expansion 
 eccentric is opposite the crank, and consequently on the same side as the 
 eccentrics of the main valve. 
 
 The variation in cut-off can, in this case, be effected in a manner similar 
 to the last ; but since the relative travel is so much greater, and since it 
 would be inconvenient to spread out the valve so much, it is usually effected 
 by varying the travel by means of a link, as already described, or by a link 
 having one end connected to an eccentric and the other to a rod jointed to 
 a fixed point on the foundation, or more usually on the shaft, and called the 
 dumb rod, because of its similarity in position to the eccentric-rod, without 
 the longitudinal motion of the latter. 
 
 Piston Expansion Valves. — When the main valve is a piston valve, the 
 expansion valve may be also a piston valve working within the main valve, 
 and on the same principles as described for slide valves. This plan has been 
 put into practice by several engineers, and although somewhat complex in 
 structure, does, no doubt, work satisfactorily, and with little friction. The 
 spindles are sometimes out of centre ; but not always, as the expansion 
 valve-rod passes through the top of the valve-box, while the main valve-rod 
 is through the bottom as usual. 
 
440 MANUAL OF MARINE ENGIHEERINS. 
 
 Expansion Valves for Compound Engines. — It has been stated that to 
 effect a cut-off earlier than half-stroke efficiently, a separate cut-off valve is 
 necessary, but this is only strictly true when the full speed of the engine is 
 maintained with such a cut-off. If full speed is attained at a cut-off some- 
 what later than half-stroke, a reduced speed can be obtained with efficiency 
 with an earlier cut-off by simply " linking up ; " for the port area being 
 suited to the higher speed, the reduction of opening on " linking up " is 
 not materially felt at the reduced speed, the consequent compression also 
 adds to the economic working of the engine at high grades of expansion, and 
 if the valves have been so set that at full speed the work is evenly divided 
 between the cylinders, so it will be, with but slight variation, on " notching 
 up " the links of all the cylinders (v. fig. 155). Expansion valves are needless 
 additions to the modern marine engine, and should be avoided, as all source? 
 of possible breakdown or derangement ought. These valves have never beer 
 fitted to triple-expansion engines. 
 
 Eccentrics. — The sheaves, or, as they are sometimes called, pulleys, are 
 made usually of cast iron, and, when possible, each is in one piece, bored out 
 so as to fit the shaft on to which it is keyed. It is esseutial that it shall fit 
 the shaft quite tightly, or otherwise it. will soon become loose from the con- 
 tinual sudden and intermittent application of the load. When the couplings 
 or flanges on the shaft do not admit of the sheaves being fitted in this way, 
 it is usual to divide them on a line through the centre of the shaft at right 
 angles to the line passing through the centres of shaft and eccentric. The 
 two unequal parts are securely bolted together, and keyed to the shaft on 
 the line through the centres, so that the whole of the strain comes on the 
 larger and stronger half, the lesser half acting only the part of a clamp to 
 hold it to the shaft. When it is desirable to keep the diameter of the sheaves 
 as small as possible, this clamp piece should be of steel. Great care should 
 be exercised in fitting the two parts of the sheave together, and also in 
 " bedding " them on the shaft. Some engineers make all eccentric sheaves 
 in parts, owing to the difficulty of fitting them tightly to the shaft when 
 in one casting. When the division is made through the centres there should 
 be a connecting bolt as close to the shaft as possible, as well as one at the 
 extremity, and also there should be a key to each half of the sheave and 
 the shaft. When the eccentrics are very large, and have to drive heavy 
 valves, it is also advisable to make the small part, when divided, as above 
 described, of wrought or cast steel. 
 
 The diameter of an eccentric sheave = 1*2 (throw of eccentric -j- diameter 
 of shaft). 
 
 Breadth of the sheave at the shaft . . = 1*25 X D -f- 0"65 inch. 
 
 strap . 
 
 = 1-2 
 
 X D + 0-63 
 
 Thickness of metal around the shaft . 
 
 = 0-7 
 
 X D -j- 0-5 
 
 ., ,, at circumference . 
 
 = 0-6 
 
 X D + 0-4 
 
 Breadth of key ..... 
 
 = 0-7 
 
 xD-f0'5 
 
 Thickness of key ..... 
 
 = 0-25 
 
 XD + O'5 
 
 Diameter of bolts connecting parts of strap 
 
 = 0-6. 
 
 X D + 01 
 
 The headgoing sheave should be 33 p 
 
 er cent. 
 
 broader. 
 
 D is found as before, and is = */ 
 
 Bx p 
 F 
 
 (v. p. 425). 
 
S.S. "Kovno"' Cylinders, 25 and 50 x 45. 
 
 441 
 
 . 
 
 Steam. 
 
 Vacuum. 
 
 Revs. 
 
 Mean Pressure. 
 
 Indicated Horse Power. 
 
 H.P. 
 
 L.P. 
 
 H.P. 
 
 L.P. 
 
 Total. 
 
 1st Grade, 
 2nd Grade, 
 3rd Grade, 
 
 87 
 90 
 88 
 
 251 
 26| 
 27 
 
 73 
 
 64 
 39 
 
 59-92 
 
 35 8 
 180 
 
 11-62 
 
 9 07 
 
 4-57 
 
 430 
 256 
 
 78-3 
 
 379 
 259 
 79 5 
 
 809 
 515 
 1578 
 
 85 
 60 
 70 
 60 
 SO 
 40 
 30 
 20 
 10 
 
 Bottom 
 
 Atmospheric Line 
 
 Bottom 
 
 Fig. 155. — Effect of Notching up Links Simultaneously 
 
442 MANUAL OF MARINE ENGINEERING. 
 
 Eccentric Straps are made generally in the form of a hoop, with lugs 
 to connect the halves together, and with a base for the rod, if it is not formed 
 with the strap. When the face of the sheave is formed with small flanges 
 on each side, the section of the strap is rectangular ; but the more common 
 practice is to make the strap with small flanges, so as to overlap the face 
 of the sheave, and bear on parts on each side. The advantage of the latter 
 plan is that the sheave is not necessarily broader than the strap, and the 
 oil does not run off so readily when the engine is at work. 
 
 Some engineers used to turn the face of the sheave with a V-shaped 
 groove, and form the strap to fit into this ; but it is not a good plan, owing 
 to a small amount of wear permitting of very considerable side play (this 
 again, though generally a fault, becomes a virtue with badly-fitted link- 
 motions). 
 
 Eccentric sheaves have, been made with a narrow projecting collar in 
 the middle of the face, the straps having a groove turned to suit it. By 
 this plan the straps are prevented from having side play, but the oil is not 
 retained, as is the case with all the other methods. 
 
 The straps are usually made of bronze, or of wrought steel lined with 
 bronze or white metal ; but cast-iron straps give very great satisfaction, 
 especially when of large size ; cast-iron and steel castings lined with white 
 metal have been adopted by some engineers with success ; the latter, while 
 having a strength beyond that of the cast iron, do not cost nearly so much 
 as bronze or wrought steel, and are now being largely employed in the mer- 
 cantile marine. 
 
 When of bronze or best cast iron : — 
 
 The thickness of eccentric strap at the middle = 04 X D + 0*6 inch. 
 „ „ ., sides = 03 x D -f- 05 inch. 
 
 When of wrought or cast steel : — 
 
 Thickness of eccentric strap at the middle = 0'4 xD-f OS inch. 
 „ „ „ sides = 0-27 X D + 04 inch. 
 
 Eccentric-rods. — Unless these are very short indeed, it is better to make 
 them separate from the straps, as then the breakage of either does not con- 
 demn both. They are made of wrought iron or steel. 
 
 The diameter of the rod in the body and lower end may be calculated 
 in the same way as that of a connecting-rod, the length being taken from 
 centre of strap to centre of pin. 
 
 The diameter of eccentric-rod at the link end = 08 D -f- 0*2 inch. 
 
 Eccentric-rods are, however, often made of rectangular section, which 
 is the correct form for the stresses it has to withstand, but it is not so economic 
 to manufacture. 
 
 Reversing Gear should be so designed as to have more than sufficient 
 strength to withstand the resistance both of the valves and their gear at the 
 same time under the most unfavourable circumstances ; it will then have 
 the stiffness requisite for good working. 
 
 Assuming the work done in reversing the link-motion, W, to be only 
 that due to overcoming the friction of the valves themselves through their 
 
DIAMETER OF WEIGH-SHAFT. 443 
 
 whole travel, then if T be the travel of valves in inches ; for a compound 
 engine 
 
 and for an expansive engine with two cylinders 
 
 w _ T/LxBxA T /T _ 
 
 To provide for the friction of link-motion, eccentrics, and other gear, 
 and for abnormal conditions of the same, take the work at one and a-half 
 times the above amount. 
 
 To find the stress at any part of the gear having motion when reversing, 
 divide the work so found by the space moved through by that part in feet, 
 the quotient is the stress in pounds ; and the size may be found from the 
 ordinary rules of construction for any of the parts of the gear. 
 
 In modern high-revolution engines the weigh-shafts, levers, and all other 
 parts of the reversing gear must be much more substantial than formerly 
 to withstand the effect of inertia forces and to avoid excessive " give " and 
 vibration when the engine is at full speed. 
 
 The Diameter of Weigh-shaft of compound engines may be fixed by 
 the following rule : — 
 
 _. /N.H.P. X revolutions 
 
 Diameter 
 
 =v^ 
 
 F 
 
 For two-crank engines F = 1,000, three-crank 900, four 800; very large 
 ones may be hollow tubes. 
 
 5 E. Coast Tost. E. and S. rule is diameter weigh shaft not Iese thaD 
 (diameter ot piston-rod — 1 inch). 
 
444 
 
 MANUAL UF MARINE ENGINEERING. 
 
 CHAPTER XYIL 
 
 VALVE DIAGRAMS. 
 
 Motion of the Piston.— Fig. 156 illustrates the ingenious method introduced 
 by Professor Zeuner for finding the position of the piston at any position of 
 the crank, and should be used always when constructing a valve diagram, in 
 order to find the points of cut-off, release, compression, &c. 
 
 Fig. 156.— Diagram of the Piston Path. 
 
 The following is the construction of such a diagram : — Draw a line T F, 
 and on it take a middle point ; cut off a part C D, equal to the radius of 
 the crank (to any convenient scale), and a part D T equal to the length of the 
 connecting-rod. 
 
 With D as centre, and D T as radius, draw a circle cutting the initial line 
 at E. With C as centre, and T as radius, draw another circle ; and with 
 C as centre, and C E as radius, draw a third circle. To find the position of 
 
DIAGRAM FOE COMMON SLIDE-VALVES. 
 
 445 
 
 the piston, with respect to its extreme position, for a position of crank C R 
 when it has moved through an angle DCE from the " dead " point C D ; 
 produce C R to cut the circles at B'PT'; then the piston has moved through 
 a space T P in turning the crank through the angle DCR, and it is distant 
 from the other end of its stroke by a space P B'. The correctness of this 
 construction is easily seen by supposing an engine whose piston is con- 
 nected directly to the crank-pin (such as the trunk engine) to be turned 
 around about its shaft, while the crank remains stationary in the posi- 
 tion D. The circle TT'F represents the path of the back end of the 
 cylinder, and BB'E that of the front end, since the cylinder is turned about 
 the point C ; the path of the piston will be on the circle T P E, since it is 
 held at the same distance from the crank-pin D. 
 
 Diagram for the Common Slide-valves. — (1) Oiven the travel of the valve, 
 amount of the lead or opening of valve at commencement of stroke, and the 
 point of cut-off, to find the lap of the valve and the position of the eccentric, 
 <fcc. These are the conditions generally predetermined in every-day practice. 
 
 front. B\ 
 
 Fig. 157. — Zeuners Diagram for the Common Valve Motion. 
 
 Fig. 157. — Draw a straight line, whose length A B is equal to the travel 
 of the valve, and on it as diameter draw a circle, the centre of which is at 
 C. Draw a line CE, so that ACE is the angle through which the crank has 
 to move to arrive at the position of cut-off (this is to be obtained by drawing 
 the piston diagram outside the circle A E B). With A as centre, and a 
 radius A F equal to the lead, draw part of a circle, and through E draw a 
 line touching the circle and cutting the original circle at K. 
 
446 MANUAL OF MARINE ENGINEERING. 
 
 Through C draw a line C D perpendicular to E K, and cutting it at L. 
 This line C D bisects the angle KCE. 
 
 C L is the amount of lap required, and B C D is the angle between the 
 crank and eccentric to obtain the cut-off required ; and since D is the 
 half-travel of the valve, L D is the maximum amount of opening of the port. 
 
 To extend the usefulness of the diagram, on CD as diameter, draw a 
 circle, and with C as centre, and L as distance, draw the arc of a circle 
 GLM, which is called the lap circle. The part G H intercepted by these 
 circles, is equal to A F, and represents the lead or opening when the crank 
 is in position CA; XY likewise represents the opening of the port, when 
 the crank is in position Y. 
 
 To determine the operation of the valve at the other end, it is necessary 
 to produce D C to D' and on C D' as diameter, draw a circle. Let H'G' be 
 the "lead" at the other end of the valve; with C as centre, and CG' as 
 radius, draw another lap circle, cutting the circle D' at M'. Through 
 C M' draw a line C E'. Then E' is the position of crank at cut-off, it 
 having moved through the angle BOE' and C G' is the amount of lap 
 required. 
 
 (2) If, however, the travel, lap, and lead are given to find the position of 
 eccentric and cut-off, the following is the construction : — 
 
 Draw on A B as diameter, the circle as before. Cut off C G equal to the 
 lap, and G H equal to the lead. At H, erect a perpendicular to A B, cutting 
 the circle at D. Join C D, and on it as diameter, draw a circle ; with as 
 centre, and C G as radius, draw the lap circle, cutting the eircle C D at M. 
 Draw a line through C M, cutting circle A D B at E. 
 
 Then C E is the position of cut-off, and B D is the angle between the 
 crank and eccentric. 
 
 (3) Given the travel, lap, and position of eccentric, to find the cut-off, 
 lead, <fec. Let B C D be the angle between the crank and the eccentric. 
 Draw the travel circle as before, and on CD as diameter draw the circle 
 cutting A B at H. Draw the lap circle GLM, then C M E is the position 
 of cut-off, and G H is the lead. 
 
 The position of the crank when the valve commences to open is C K, or 
 at the angle A K, with its initial position. 
 
 If the valve has no inside lap — that is, the ports are both just closed to 
 exhaust when the valve is in mid position — the position of release is at C R, 
 a line at right angles to CD; and the position at which compression takes 
 place is C B,', also at right angles to C D. 
 
 If, however, the valve has inside lap amounting to, say, C G', release will 
 not take place till the crank is at C K'. 
 
 If, on the other hand, the valve is cut away on the inside so as to have 
 negative lap to the amount of C M, then release will take place at C E, and 
 compression at C K. 
 
 If the inside lap is less than these amounts, the position of release and 
 compression can be found by drawing an inside lap circle with a radius equal 
 to the lap, and through the points where it cuts the circle on C D and C D', 
 the lines drawn through C and those points of intersection will give the 
 positions of crank. 
 
 Travel of Valve. — It is usual to fix the travel of the valve before deter- 
 mining any further particulars, as so many things depending on this have 
 often to be considered before there is leisure to finally decide the lead, 
 iap, «tc. 
 
c< 
 
 LEAD." 447 
 
 It will be seen, on reference to the diagram, that the opening to steam 
 will vary with the travel, and if the area of opening is fixed from certain 
 considerations before-mentioned, it is an easy matter to calculate how much 
 the valve must open to give that area. Now the opening, together with the 
 lap, is equal to half the travel, and with the same -positions of " lead " and 
 cut-off, the opening will be a constant ratio of the travel. This ratio can be 
 determined by drawing a preliminary diagram with an assumed travel of 
 valve. Let R be this ratio, and Q the amount of opening desired ; then 
 
 Travel of valve = Q -5- R. 
 
 Since it is usual to design the ports of a steam cylinder so that the flow of 
 steam when exhausting may not exceed a certain velocity, it is evident that 
 the port should open fully for that purpose ; hence the travel of the valve, 
 when there is no inside lap, should not be less than twice the length of the 
 port, and is generally about 2| to 3 times the length of the ports. 
 
 The less travel the valve makes, the less work is absorbed in moving it, 
 as the work is very nearly proportional to the travel. Double- and treble- 
 ported valves are resorted to with the object of reducing the travel, as by 
 them double and treble openings are obtained, and the travel may, therefore, 
 be one-half and one-third that of the common valve with single ports. 
 
 "Lead." — The amount of lead given to the valve is generally decided 
 arbitrarily and according to judgment or prejudice. It will be seen by 
 referring to fig. 157 that with the same " lap " an earlier cut-off is obtained 
 by moving the eccentric farther from the crank, but at the same time increasing 
 the " lead ; " if the earlier cut-off is to be obtained without altering the 
 " lead," the lap must be increased and the eccentric moved ; but if the lap 
 be increased, the opening is decreased, and the resistance at the port thereby 
 increased. If, therefore, an early cut-off is required without the aid of an 
 expansion valve, the " lead " must be great or the travel great, to get suffi- 
 cient opening to steam, or the lead and travel must be both larger than 
 would be the case with a later cut-off. 
 
 The considerations which should operate in deciding the amount of lead 
 are the speed of the engine and the inertia of moving parts compared with 
 the piston area. The momentum of the piston and rods should, if possible, 
 be absorbed in compressing the steam remaining in the cylinder when the 
 valve closes to exhaust ; and were this always the case, there would be no 
 need for fresh steam to enter the cylinder until the piston was at the end of 
 its stroke ; but this, under ordinary circumstances, seldom occurs, and if no 
 fresh steam were admitted, " to form a cushion " for the piston, there would 
 come a considerable jar on the bearings, etc., at the end of every stroke, 
 owing to the strain increasing the displacement of the shaft in its bearings, 
 and the sudden application of the load when the valve opens, causing its 
 replacement with considerable force. This action is also observable when 
 there is considerable " lead " without adequate " compression ; " and although 
 the "hammering" is usually attributed to excessive "lead," it is really 
 due to want of adequate compression, for on " notching-up " the link, it 
 generally ceases, notwithstanding that the lead is then thereby very con- 
 siderably increased. For really quiet and sweet running there should be 
 plenty of compression and no lead — even negative lead is of advantage at 
 high rate of revolution, provided the valve eventually opens wide enough. 
 
448 MANUAL OP MARINE ENGINEERING. 
 
 Long-stroke engines may have considerably more lead than those of the 
 same cylinder capacity with shorter stroke, as the weight of moving parts 
 is not less and sometimes more, the piston velocity is more (with the same 
 number of revolutions), while the piston area is less. Fast-running engines 
 without considerable compression may also have more " lead " than slow- 
 running ones, as the element of time bears on the operation of the steam. 
 Again, if owing to the valve motion the valve opens slowly, the period of 
 admission may be earlier than should be the case with a quick opening one. 
 as the. wire- drawing at the commencement of admission will produce a similar 
 effect, to compression. 
 
 The amount of lead at each end will vary for two reasons ; first, the 
 cut-off at each end being effected by the same eccentric, any variation in 
 position of cut-off must be obtained by varying the lap, and any variation 
 in the lap will cause an inverse variation in the lead ; secondly, the lead 
 should be less at that end of the valve remote from the gear, as the adjust- 
 ment after wear of the latter tends to increase the " lead " at that end. If 
 the cut-off is equal at both ends the laps will vary, that at the end of a direct- 
 acting engine remote from the shaft being more than at the other. With 
 the ports as generally found in every-day practice, and with such travels 
 as are practicable, the cut-off at the end near the shaft should be earlier 
 than at the end remote from it, as the larger opening got by the decreased 
 lap causes a fuller indicator-diagram at that end. For these reasons many 
 engineers arrange so that the lead at the back or top end is only half that 
 at the front or bottom end. 
 
 Inside Lap.— The effect of positive lap on the exhaust side of the valve 
 is to retard the period of release, and to increase the compression by accel- 
 erating the closing to exhaust. The effect of negative lap is, of course, the 
 reverse of this, and is also to give a full opening to exhaust sooner than 
 would be the case with positive lap. A large amount of lap, either negative 
 or positive, is required on the inside of a valve to materially change the 
 periods of release and closing to exhaust, because just then the valve is 
 moving at its quickest speed ; but a small amount of negative lap makes 
 a considerable difference to the amount of opening to exhaust at the end 
 of the stroke, when the steam should exhaust wholly from the cylinder. 
 There is no advantage in retaining steam in the cylinder to the very end of 
 the stroke, as the effort of the piston on the crank is then ineffective to produce 
 good results, and any forward pressure only serves to increase the difficulty 
 of bringing the piston gently to rest by means of cushioning, etc., and this 
 is especially true of quick-moving engines, whose steam ports are compara- 
 tively small. 
 
 If the valve has negative lap on the exhaust side, it is manifest that at a 
 certain period there is a momentary communication between both ends ol 
 the cylinder ; when the negative lap is considerable, and the ports small, 
 the effect of such a communication is to fill the exhausted cylinder on ont 
 side of the piston with the steam released from the other side, just before the 
 valve closes to exhaust, and is seen very distinctly, on the indicator-diagrams, 
 in the form of a sudden rise of pressure at the commencement of the com- 
 pression. In cases of this kind, ample compensation is made for the retarding 
 of the compression by negative lap in the increase of back pressure at the 
 commencement of compression. 
 
EFFECT OF "NOTCHING UP. 449 
 
 The high- and medium-pressure valves of compound screw engines should 
 always have a considerable amount of negative inside lap, to allow of a 
 continuous and easy flow of steam during exhaust. The amount of negative 
 lap on the inside should be always less than the outside lap of the valve, 
 otherwise there will not be enough bar to cover the port, so that steam 
 can pass from the valve-box to the exhaust just as the valve is opening 
 and shutting ; for a quick-running compound engine there should be such 
 iap that exhaust commences from the high- and medium-pressure cylindei 
 at 0*9 of the stroke; and with a slow-working paddle engine not later than 
 95 the stroke. 
 
 Effect of " Notching Up." — When an engine is fitted with link-motion 
 or other means of reversing it without throwing " out of gear," so that 
 the mechanism which would produce stern motion, can be made to effect 
 that producing ahead motion, a certain variation in the cut-off, release, 
 &c, can be effected by what is called " notching up." The origin of this 
 term is clearly traceable to the locomotive, whose reversing lever is held 
 in place by a sliding-bolt fitting into a notch in a quadrant provided for the 
 purpose. 
 
 When the link is notched up so that the block works the valve from a 
 point between its two extreme positions on the link, the motion is one due 
 to the combined effort of both eccentrics, and in order to examine clearly 
 the operation of the valve under these circumstances, it is necessary to find 
 the position and throw of the equivalent eccentric. To determine this 
 exactly is somewhat difficult, but a very close approximation can be made 
 by the method suggested by the late J. Macfarlane Gray, as follows : — 
 
 Suppose the link (fig. 159) to be notched up to a point M ; or, in other 
 words, the link moved so that the link block is distant MT from the 
 point at which the eccentric-rod is attached to the link. 
 
 Draw the valve diagram (fig. 158) due to the position and throw of the 
 eccentrics of fig. 159; and through DD' draw the arc of a circle with a 
 radius found as follows : — 
 
 Radius = length of eccentric-rod from centre to centre x half the distance 
 between the centres of the two eccentric sheaves -i- the distance between the 
 centres of eccentric-rod pins on the link. 
 
 DT x DD' 
 
 2"TN 
 
 Referred to fig. 159, Radius = 
 
 On this arc, take a point Z, so that ^-^-, = ^-vf • 
 r D D T N 
 
 Join C Z, and on it as diameter draw a circle, cutting the lap circle at 
 points L and E ; through L and E draw lines which represent the 
 position of opening and closing respectively at one end of the valve when 
 the link is notched up to the point M. 
 
 It will be seen that the effect of notching up is the same as would be 
 produced by an eccentric whose position on the shaft is at the angle B C Z 
 with the crank, and its eccentricity C Z; that the opening to lead is earlier, 
 and the magnitude of the lead considerably increased ; that the reduced 
 travel gives a smaller opening to steam, and that the cut-off is earlier; 
 likewise, by examining further, it can be seen that release takes place sooner 
 and compression commences earlier than when in full gear. 
 
 When the eccentric-rods are arranged as shown in fig. 159, they are said 
 to be open, or it is an open rod link-motion ; if, on the other hand, D N and 
 
 29 
 
450 
 
 MANUAL OF MARINE ENGINEERING. 
 
 D' T be joined, it is said to be a crossed rod link-motion. When the link 
 motion has crossed rods, the arc D Z D' should be convex to the point 0, 
 that is, the centre of curvature is on the side A. 
 
 When the rods are crossed, an earlier cut-off is effected by notching up 
 without materially altering the lead, and, in most cases, with a very early 
 cut-off there is only a very small lead, smaller even than when in full gear. 
 But the travel is reduced very rapidly by notching up, giving a correspond- 
 ing reduction of opening to both steam and exhaust ; so that, altogether, 
 crossed rods are not so convenient as open ones for working expansively 
 by notching up the link, and are very seldom so fitted. 
 
 Fig. 15S.— Diagram showing the Effect of " Notching up." 
 
 Fig. 159.— Link Motion "Notched up." 
 
 The valve diagrams, when the link-motion is such as shown in fig. 159, 
 are constructed on the same principle as for notching up, and the position 
 of the eccentrics and their throw are determined by producing the arc D D' 
 (tig. 158), and taking a point Z' beyond D, so that 
 
 Z'D TM 
 
 DD' + 2Z'D" TN* 
 Join C Z'. 
 
EXPANSION VALVE. 
 
 451 
 
 Then C Z* is the eccentricity of the sheaves, and BCZ' is the angle 
 between them and the crank. 
 
 Expansion Valve on an Independent Face, Central Position Ports Closed. — 
 The valve in this case is working under precisely the same conditions as a 
 common slide-valve, and the same construction of diagram as fig. 157 holds 
 good. 
 
 Expansion Valve on an Independent Face, Central Position Ports Open. 
 — Let C O' (fig. 160) be the position at opening of valve when set at earliest 
 cut-off, which should be somewhat before that of the main valve, and C E 
 the position at the earliest cut-off which is required. Bisect the angle O'C E 
 by the line C T, and make C T equal to half the maximum travel of the 
 valve. 
 
 »»»- 
 
 Fig. 1 60. — Diagram for an Independent Expansion Valve. 
 
 On C T as diameter, draw a circle cutting E at H. 
 
 Produce C T to T', making CT' = CT. 
 
 On C T' as diameter, draw a circle cutting O at G. 
 
 Then C H is the distance of the cutting-off edge of the valve from the 
 edge of the steam port ; that is, C H = b h ; and the lap of valve b d = C H 
 — length of port. 
 
 The width of the bar a b must not be less than C T - b d, and should 
 be = (C T - b d) + \ inch. 
 
 For the other end of the valve, a similar construction will give the 
 required results. 
 
1 > MANUAL OF MARINE ENGINEERING. 
 
 Let E' be the position of cut-off at the other end of the cylinder, 
 cutting the circle T' at H'. 
 
 With as centre, and H' radius, draw a circle cutting the circle C T" 
 at G ; through C G draw a line, then C G is the position at which the valve- 
 opens again. 
 
 Hence ag = CH'; and the lap a c = C H' - length of port. 
 
 Again, the width of bar a b must not be less also than C T — ac : and 
 should be = (0 T - a c) + { inch. 
 
 The bar cd = ab + lap b d + lap a c. 
 
 To find the reduction in travel for a later cut-off at position F cutting 
 the port circle at K. Draw K R perpendicular to F cutting C T at R. 
 On C R as diameter, draw a circle cutting the lap circle at L. C R is the 
 half travel of valve. L is now the position at which the valve opens 
 again, which should not be before the main valve is closed to steam. 
 
 In arranging an eccentric and gear for this valve, care must be taken in 
 setting it that the expansion valve will open at the earliest cut-off before 
 the main valve opens, and that at the latest cut-off it does not open again 
 till the main valve is closed. When the expansion valve spindle is in the 
 same plane with the piston-rod, and the gear is direct from the eccentric to 
 the valve-rod, the eccentric should be set at the angle B T with the crank 
 on the same side as the head-going eccentric. 
 
 Expansion Valve Working on the Back of the Main Valve, Cutting off at 
 the inside edge, and Variation in Cut-off made by Varying the Travel. — Let 
 A (fig. 161) be the dead point of the crank at the back of the stroke, and 
 C E the position of earliest cut-off. B C D is the angle between the crank 
 and head-going eccentric, and C D its eccentricity. Out off C H equal to 
 the lap of the expansion valve, or distance of the inner edge of expansion 
 valve from inner edge of steam port in the back of the main valve, when 
 the expansion valve is in its mean position with respect to the main valve. 
 Draw the lap circle H G as before. From H draw a line H O perpendicular 
 to C E with D as centre and the half travel (absolute) of the expansion 
 valve as radius, draw the arc of a circle cutting H O at O. Join D O and 
 0; on O as diameter, draw a circle, which will pass through H, since 
 C H is a right angle. Complete the parallelogram D by drawing D T 
 parallel to O C, and C T parallel to D O. Then C O is the half-travel of the 
 expansion valve relative to the main valve ; B C T is the angle between the 
 crank and the expansion eccentric, or T C D is the angle between the ex- 
 pansion eccentric and the head-going eccentric. 
 
 By producing O C to O', and on C O' drawing a circle, &c, the cut-off, 
 lap, &c, may be investigated for the other end of the valve. 
 
 To find the effect of shortening the absolute half-travel to C R. Since 
 the position of the eccentric remains unchanged, join R D, and through 0' 
 draw N parallel to R D, and cutting D in N. On C N as diameter 
 draw a circle cutting the lap circle at L and K. 
 
 Through L and C K draw lines which are respectively the positions 
 of opening and cut-off of the expansion valve for the back end, and C N is 
 the relative half-travel. 
 
 It will be seen that when the travel of the expansion valve is nothing, 
 that the relative half-travel is C D, and that if the lap is the same as tliat 
 of the main valve, the cut-off and lead will be the same also. For this 
 reason it is customary when designing a valve of this kind, which may have 
 to go out of gear when at full speed, or for some particular purpose, to> 
 
EXPANSION VALVE. 
 
 453 
 
 make the lap the same as that of the main valve, and for practical reasons 
 to make the travel the same also. 
 
 To find the effect of the expansion valve when in stern-gear, join T to 
 D' of the stern-going eccentric, and complete the parallelogram OD' as before. 
 
 It will be seen also, that with this kind of expansion valve, the opening 
 to steam is never less at the expansion valve ports than at the cylinder ports. 
 
 Fig. ltjl. — Diagram for an Expansion Vah'e. 
 
 Expansion Valve Working on Back of Main Valve, and Cutting off at 
 Outside Edge — Let A (fig. 162) be the dead point of the crank, as before, 
 at back of stroke, and C E the position of earliest cut-off required, BCDis 
 the angle between the crank and the head-going eccentric, and C D its 
 eccentricity. From C E cut off a part, C H, equal to the opening to steam 
 of the main valvje at that end, or such as would give sufficient opening by 
 the rules as laid down already. From H draw a line, H O, perpendicular 
 to E. With D as centre, and D O a radius equal to the eccentricity of 
 the expansion eccentric, draw the arc of a circle cutting H O at O. Join 
 C 0, and on it as diameter draw a circle, which will pass through H, since 
 C H O is a right angle. 
 
 Complete the parallelogram CD by drawing DT parallel to 0, and 
 C T parallel to O L>. 
 
 Then C is the half-travel of the expansion valve with respect to the 
 main valve ; B T is the angle between the crank and the expansion 
 
454 
 
 MANUAL OF MARINE ENGINEERING. 
 
 eccentric, and TCD the angle between the expansion eccentric and the 
 head-going one of the main valve. C H is the distance of the edge of the 
 expansion valve (fig. 163) from the outer or cutting-off edge of the port 
 on the back of the main valve, when the expansion valve is in its mean 
 position with respect to the main valve. 
 
 If the expansion eccentric is set exactly opposite to the crank, so that the 
 point T is on the line A B, then the cut-off is the same in stern-going as in 
 head going gear. If, however, it is set as shown in fig. 162 so as to be nearer 
 the stern-going eccentric, the cut-off is later in stern-going than in head-going 
 gear, and the relative travel of the expansion valve is greater. 
 
 Fig. 163. — Expansion Valve. Fig. 162. — Diagram for an Expansion Valve. 
 
 If it be required to cut-off at a later period, as at C K, the expansion 
 valve must be so moved that when in mean position with respect to the 
 main valve, the cutting-off edges are apart by a distance equal to CK; that 
 is, the valve is moved towards the middle by a distance M N or C K - C H. 
 
 The laps, cut-off, &c, at the other end of the valve may be investigated 
 in a similar way by producing O C to O', making CO' equal to CO, and on 
 it as diameter drawing a circle cutting the required position of earliest cut- 
 off E' at H'. 
 
EXPANSION VALVE. 
 
 455 
 
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456 MANUAL OF MARINE ENGINEERING. 
 
 C H' is then the distance apart of the cutting-off edges of the valves at 
 the other end. Also, if a later cut-off C K' is required when K is the cut- 
 off at the other end, the valve must be moved towards the middle by a space 
 equal toOK' - H'. 
 
 If C K be the position of latest cut-off, care must be taken that the 
 arrangement is so designed that the re-opening of the port at C L is after 
 the main valve has cut off to steam. 
 
 In order to obtain as large a relative travel as possible to the expansion 
 valve, it is customary to make the angle BCD about 100*, so that the main 
 valve has of necessity very little lap and small lead, and consequently there 
 is a late release and a very small amount of compression. 
 
 In speaking of C D as the eccentricity, and BCD the angle of the head- 
 going eccentric, it must be understood that an eccentric acting directly and 
 in line with the main valve spindle is meant. If the valve gearing is such 
 that the eccentric is otherwise, then C D is the eccentricity and B C D is the 
 angle of the equivalent eccentric. 
 
 This remark applies likewise to the case of the expansion valve (fig. 163) 
 cutting-off at the inside edge. 
 
 Construction of Valve Diagrams. — In making the diagram for an engine, 
 the piston diagram (fig. 156) should be drawn first to such a scale as to be 
 well clear of the valve diagrams. The main-valve diagram should then be 
 drawn full size in black ink, and the expansion- valve diagram on it in red, 
 or other distinctive colour. The effect of each valve can then be traced both 
 separately and conjointly. 
 
 Effect on Indicator Diagrams of Crank Sequence. — Fig. 164 shows the 
 effect of high-pressure crank leading the M.P. crank, and also the effect of 
 the low-pressure crank leading the mean-pressure. It will be seen that in 
 the former case the loads on the pistons are greater at the beginning of the 
 stroke than is the case when the low-pressure leads. This is a good thing, 
 especially in fast-running engines where the inertia of the moving parts is 
 considerable ; the ratio of maximum to mean load is greater in this case, 
 and the variation in temperature is greater than when the low-pressure 
 leads. 
 
PROPELLERS. 457 
 
 CHAPTER XV HI. 
 
 PROPELLERS. 
 
 The Fundamental Principle on which all Marine Propellers act is that the 
 reaction from a stream of water projected from a floating body is equal to 
 its action in magnitude but opposite to it in direction : and that reacting 
 force called thrust applied to the body tends to proauce motion of it in that 
 ■opposite direction. 
 
 The stream of water may be produced by any instrument either within 
 the ship, as by some form ol pump, or without as by a paddle-wheel or screw. 
 The object of the engineer is, therefore, to obtain in any case as large a thrust 
 ■as possible with the least amount of water disturbance, and the efficiency of 
 the means he employs to propel a ship is measured by such a comparison, 
 whereas the maker of pumps aims at disturbing as much water as possible 
 with the least amount of thrust, inasmuch as to produce the thrust he has 
 to expend energy just as the marine engineer has in disturbing the water ; 
 the end and the means being transposed, the aim is therefore different, 
 •and in considering propeller problems it is desirable to free the mind from 
 too much analogy to pumps otherwise grave misconceptions may arise which 
 will lead to fallacious conclusions. It may also be impressed on the mind, 
 while dealing with the question of efficiency of propeller, that mere thrust 
 in the gross is not the true measure of the success of a propeller, for a screw 
 may, and often does, create a serious amount of augmented resistance, so that 
 its actual or net propelling effect is less than another screw which does not 
 produce nearly so much gross thrust. In former times, when the tendency 
 was to supply every steamship with a propeller as large as her draft of water 
 ■and construction permitted, and to make the surface and breadth of blade 
 proportional to the diameter, it was no uncommon thing to find the propulsive 
 •efficiency very poor notwithstanding exceedingly high thrusts. The pheno- 
 menon of negative slip generally showed itself as an accompaniment to such 
 screws ; and although the so-called negative slip is not absolutely incom- 
 patible with good results, satisfactory ones seldom do accompany it. 
 
 If a steamship is at rest with respect to the water in which it floats, and 
 is to be set in motion by the stream of water projected from it, the inertia 
 must be overcome, and until motion actually ensues the velocity of the 
 stream is the same with respect to the still water as that with which it flows. 
 Then the energy is for a time used in accelerating the motion of the ship, 
 ■and with the increase in its motion there is a corresponding decrease in the 
 flow relative to surrounding water. Moreover, as the ship increases in speed 
 her resistance increases as the square of the speed, and following that law 
 the increments lessen gradually until the maximum speed is reached, when 
 
458 MANUAL OF MARINE ENGINEERING. 
 
 the whole of the effective propelling effort is absorbed in overcoming the 
 resistance — the net thrust is then equal to the gross resistance. 
 
 If V is the velocity of the issuing stream from the ship, and v is the velocity 
 of the ship itself, both in feet per second, the velocity of the stream with respect 
 to still water is (V — v). This is called the slip, and to distinguish it from 
 other features of marine propulsion, it is named real slip. It is generally 
 expressed as a percentage of the velocity of flow, hence — 
 
 Real slip per cent. = I — - — J 
 
 100. 
 
 The propelling effect of a stream of water depends on its mass and velocity ; 
 the mass depends on its quantity, and that quantity on the sectional area 
 of the stream and its velocity. If, therefore, M is the mass of water projected 
 with a section of A square feet at a velocity V from a ship moving at a velocity 
 v, and taking the weight of a cubic foot of sea water at 64 lbs., and the effect 
 of gravity at sea level as 32 — 
 
 The momentum of the stream = M (V — v). 
 
 M = W -*- g ; W = A x V x 64. 
 
 64 x A x V 
 Then, momentum of stream = ^ (V — v) = 2 A x V (V — v). 
 
 The thrust or propelling force due to this stream is, therefore, 
 
 2A x V (V - v). 
 
 All attempts at what may be called propulsion by internal propellers 
 have proved failures in practice, however efficient the instruments them- 
 selves may have been. The losses from friction in the pipes, orifices, etc., 
 the great inconvenience experienced by the presence of pipes, channels, 
 etc., within the ship, in addition to the incumberance of the propeller itself, 
 form an impassable bar to the use of such apparatus in any but small ships 
 of quite low speeds, whatever the efficiency as propellers may be. 
 
 The Stream Water in Every-day Practice proceeds now from some form 
 of propeller without the ship, and it is fed by a flow to it from the surrounding 
 water mostly acting under the influence of gravity. With a screw deeply 
 immersed, and well away from the body of the ship, the supply of water 
 is ample, uniform, and free from air bubbles ; it is, however, not free alto- 
 gether from air, inasmuch as there is always a considerable volume of it 
 held by the water in solution, otherwise fishes could not remain alive in it ; 
 if the propeller, from any cause, demands such a supply as to produce 
 decrease in pressure at the back of the blades to be considerable, some of this 
 air is liberated from solution, and appears as bubbles mechanically mixed 
 with the water, and following the stream to the front or active face of the 
 screw, it thereby materially reduces the thrust. This phenomenon is 
 known as cavitation, from the fact that the action of propellers moving at 
 high velocities with insufficient blade area is to make cavities in the stream 
 flowing through them in which the freed air escapes and remains enveloped 
 with water as bubbles. These must not be confounded with the air bubbles 
 formed by the suck of the screw from above, producing air eddies which act 
 as air " spouts," nor with those formed by the blade tips breaking the surface 
 
PROPELLERS. 459 
 
 when there is not sufficient immersion. In both these cases the air mixes 
 at once with the feed stream, producing disastrous results with the thrust, 
 so that it may even at times disappear altogether for a moment. 
 
 When any Submerged Screw is working it is drawing to itself what was 
 previously still water, and discharging in rear of it a mass of water which 
 has to come to rest again sooner or later by the resistance of the surrounding 
 water. It is certain now that the flow from it is a column nearly cylindrical, 
 and that in the dispersion it retains the same form, although of a somewhat 
 larger diameter than that of the screw. There is no circumferential disper- 
 sion, and so no need, in fact, for any means of retention, such as so many 
 well-meaning inventors have patented " to save the loss due to centrifugal 
 force action." The otherwise obnoxious water bubbles tell too truly the 
 course of flow through and from the screw for any mistake to be made on that 
 score. Prof. Flamm and Sir Charles Parsons have shown by exhaustive 
 experiments and photography exactly what takes place, and having seen 
 the performance of their model screws, it is easy to confirm their conclusions 
 by making careful observations on real working screws in smooth, clear water. 
 In addition to the " head " of water itself, there is the further equivalent 
 " head " due to atmospheric pressure (equal to 33-7 feet of water) imposed 
 on the screw race to keep up a supply to the propeller to permit it to work 
 without cavitation ensuing ; but this latter pressure is also the cause of such 
 a ready and plentiful supply of air to it when an eddy forms. It also conduces 
 to the formation of the cavity or depression of the water surface just over 
 the screw, into which the water in advance is always flowing under the 
 influence of gravity as the ship proceeds. Abaft the screw there is always 
 the tendency to pile the water up as the easiest way to disperse the columu, 
 the resistance being less near the surface than lower down ; hence the wave 
 or lump of water so often seen in the wake of a screw. 
 
 The Effects of Wake Currents on a Screw, particularly on the central one, 
 is considerable, especially in ships of high speed with rather full " buttock " 
 lines. In a sailing ship — that is, one without a screw — the flow of water into 
 the void made by the ship is not limited to the streams following its lines or 
 form ; and when the speed is increased the tendency for such streams from 
 either side to meet abaft the rudder is great ; the space left by the ship 
 must be filled, therefore, either from below or by water flowing in from behind 
 under the action of gravity. The latter is the easier method, and the one 
 usually observed. If a screw of considerable diameter be set to work at the 
 end of the ship, it requires to be fed with abundance of water, and so it causes 
 a most distant suck on either side of the stern in front of the screw ; gravity 
 again forces the flow along the skin to be accelerated, and induces a further 
 supply to come from abreast. The shape of the ship at the stern causes 
 the stream to follow a certain course, after which it and the screw's action 
 divert it outwards again, so that the axial flow through the screw is less than 
 that of the undisturbed water of the sailing ship — in fact, with respect to still 
 water, the feed has forward motion with the ship. Inasmuch, then, as every 
 screw is working in more or less disturbed water, so that no one can possibly 
 say for certain what is the real value of V, the real slip is a matter of conjecture, 
 but a fairly close approximation can now be made to it, thanks to the investi« 
 gations and experiments of Dr. R. E. Froude, Prof. Taylor, and other patient 
 and pertinacious men of genius and opportunity. 
 
4 60 MANUAL OF MARINE ENGINEERING. 
 
 With the Paddle-wheel the Problem of Propulsion is simpler, inasmuch 
 as the side wheel is working in water comparatively undisturbed, it produces 
 but little augmented resistance, and the acting surfaces are generally normal 
 to their action. On the other hand, in its course it breaks some of the water 
 into foam, so that the surface stream has not nearly the density of still water, 
 and a considerable portion is at times dispersed into the air. When, how- 
 ever, the ship is fairly underway the paddle-wheel is much more efficient 
 than it appears, as the amount of foam and broken water then is more apparent 
 than real. . There is, however, the same tendency to formation of hollow 
 due to sluggishness of feed, and there is the piling up of water abaft, which 
 is inevitable, as there is only a small amount of side dispersal — in fact, the 
 stream is rectangular from the wheel to its race end more or less, as that of 
 the screw is cylindrical — and it is quite as difficult to estimate the real slip 
 of a paddle-wheel as it is of a screw, and so the practical engineer has to be 
 content to take the nominal slip. 
 
 The Nominal or Apparent Slip of a Propeller is what the slip should be 
 if there were no interference, and the propeller worked as by theory it should 
 do. The velocity of the stream is assumed to be that calculated by multi- 
 plying the lineal effect of the propeller by the number of revolutions. In the 
 case of a paddle-wheel, the circumference at the centre of effort of the floats 
 is taken, and with a screw the pitch or axial distance apart of two threads 
 of one convolution. 
 
 If D is the diameter of the circle through the float effective centres, which 
 is generally now taken as the centres on which the floats oscillate in a feather- 
 ing wheel ; P is the pitch of a screw, and R the revolutions per second of 
 either ; then in case of 
 
 (1) The paddle-wheel, V=iDx Rfeet. 
 
 Apparent slip per cent. = ( — rr — I 100 = 100 — ^ =r. 
 
 V V ./ D x R 
 
 (2) The screw, V = P x R feet. 
 
 V — v\ 100 v 
 
 Apparent slip per cent. = ( — ^— - J 100 = 100 
 
 PxR 
 
 The Actual or Real Slip must be in every ship in excess of the nominal 
 or apparent. In some ships the excess is great, not only because the wake 
 currents are great, but because the stream of water projected is not really 
 a solid cylindrical column. It is, of course, highly desirable that the stream 
 shall flow at an uniform rate throughout its transverse section — that is, 
 there shall be no difference in the rate of flow from the tip than from that 
 at the middle or root of the blade — there should be no envelopes or cylinders 
 of water passing through or sliding over one another. In practice such a 
 consummation is scarcely possible, hence the real slip probably varies very 
 considerably from root to tip, and there will be always a core following the 
 boss or hub with no imparted velocity whatever. 
 
 The Spiral Path of the Blade Tips, as of every other portion of the blade, 
 does no doubt by skin friction cause to be induced a more or less spiral flow 
 to the wake ; and probably as the feed approaches the screw it begins to 
 assume a somewhat spiral path. Hence the column flowing from the screw 
 is more like a common wire rope than a solid liar ; so that if there are differ 
 
PADDLE-WHEELS. 461 
 
 ences of flow and slip there will be layers or cylinders revolving as well as- 
 sliding in one another. The screw that avoids this action is a desideratum. 
 
 The Selection of Propeller Dimensions and Form is. from these 
 disturbing causes, rather a matter of experiment and determination by 
 empiric formula derived from successful practice, than direct from mere 
 considerations of first principles and theory, involving more or less com- 
 plicated mathematical calculations, as well as some necessary assumptions- 
 Dr. R. E. Froude and Professor Taylor have, however, done so much to clear 
 the problems of complications, and by exhaustive experiments produced 
 data of inestimable benefit, that a good propeller can be designed which 
 shall suit the conditions under which it has to work, and permit of at least 
 good efficiency of its motor ; further, by their help, the best propeller 
 neglecting such limitations can be indicated. But, after all, it is only by 
 actual trial on the real ship at sea under average working conditions of weather 
 that one can say for certain that the screw is quite satisfactory. It is, of 
 course, the prime duty of the designer to provide a screw that will work 
 well and with the highest efficiency on service and under service conditions,. 
 rather than one which on trial in smooth water will produce the highest 
 speed for a limited period under mere trial conditions, in spite of the ever- 
 present temptation to this latter course. It is not an uncommon thing 
 nowadays to hear that turbine-driven steamers on service have not fulfilled 
 the expectations of their promoters, or done so well as on the trial trips. 
 Such things have happened frequently in the past, when the trial trip has 
 been the crucial test of the ship's performance in the minds of those who 
 determined the proportions of the screws. Even turbine steamers may suffer 
 no actual loss on service if the screw is of somewhat coarser pitch than requisite 
 for the speed of the turbine itself to produce maximum joint efficiency. 
 
 The Common Paddle-wheel with Fixed Floats is now seldom seen*; it 
 was the oldest and simplest form of marine propeller, and consisted of a 
 framework of wrought-iron bars, braced together and stayed to prevent 
 racking, attached to a cast-iron hub or boss keyed (though originally staked) 
 to the shaft end. The boards or floats were secured to the radial arms just 
 inside the outer rim by means of bolts having a hook end, which fitted on 
 the arm. No doubt it was a development of the " under shot " mill wheel 
 of the old millwrights. It was very simple in construction, strong and easily 
 repaired if damaged, and replaced if worn out. For river craft carrying only 
 passengers and parcels with little variation of draft of water, the radial 
 wheel was fairly efficient ; owing to the comparative ease with which the floats 
 could be shifted or " reefed," it had advantages on a sea-going craft making 
 fairly long voyages, for, in spite of a somewhat reduced efficiency under the 
 circumstance's as the "dip" of the float was reduced thereby, the engine- 
 was permitted to make the full number of revolutions for maximum power. 
 These wheels to-day are used only on small, light draft, river craft in tropical 
 regions, where a feathering wheel would be liable to derangement, and with 
 no means for effecting repairs to it. 
 
 The Feathering Float Paddle-wheel, introduced by Elijah Galloway into 
 general use in 1830, has continued to the present time as the accepted and 
 best form of paddle propeller. The general idea of such a contrivance is 
 
 * It ia, however, still in general use on American rivers, and preferred by U.S. 
 engineers as permitting of a timber construction to ? l?rge extent, and consequently 
 easy to repair en route. 
 
462 
 
 MANUAL OF MARINE ENGINEERING. 
 
 to cause the floats to assume a position whereby they shall enter the water 
 without shock and undue disturbance ; to act in the most effective way 
 during the passage through the wheel race ; and finally to emerge without 
 disturbance of water with as little resistance to the wheel torque as possible. 
 
 ■S 
 
 o 
 
 i 
 
 60 
 
 a 
 '3 
 
 a 
 
 I 
 
 t 
 
 .5° 
 
 This is carried out in a feathering wheel, by giving the floats when passing 
 through and leaving the water the same angular inclination to the vertical 
 that would obtain with a fixed float on a radial wheel of much larger diameter. 
 What that equivalent diameter should be is a matter of geometrical construc- 
 tion based on the facts of each case as to speed of ship, fraction of slip, and 
 
PADDLS-WHBiiJLS. 
 
 463 
 
 qeo/j 
 
 6U/M0//0J 
 
 is 
 
 s> 
 
 =3 
 
 is 
 
 60 
 
 s 
 
 u 
 
 o 
 
 
 in 
 e 
 
 ■S 
 S 
 
 9 
 
 •w 
 <B 
 O 
 
 ' S 
 
 -a 
 
 a 
 
 A 
 
 a 
 
 o 
 o 
 
 
 bo 
 
464 MANUAL OF MARINE ENGINEERING. 
 
 actual diameter of wheel. To ascertain the proper position of a float at 
 entry and the equivalent diameter of the wheel, take a point P (fig. 165) 
 at the edge of a float just touching the water level. Draw line P A as repre- 
 senting the water level, and cut a part P A to represent the speed of the 
 ship on a convenient scale ; let P B be the tangent to the circle passing 
 through P, and make P B as the speed of the wheel on the same scale. Com- 
 plete the parallelogram A P B, whose diagonal P R represents the direction 
 and rate of velocity of the point P at that particular moment of its course, 
 and if the float face lies on the line made by continuing P E to A, it will 
 start to enter the water edgeways with least resistance, and Aj is the centre 
 of the equivalent radial wheel. If the wheel is to work with considerable 
 slip, the speed, as measured by P B, will be large compared with P A, this 
 will mean an increase in value to A R, and a reduction in the angle made 
 by A x R with the vertical and the raising of A,. In other words, with greater 
 slip the equivalent wheel must be greater. Fig. 166 shows the locus of the 
 floats and float centres of a feathering wheel. 
 
 The Means provided for constraining the Floats are quite simple, and consist 
 essentially of a fixed gudgeon, or its equivalent, eccentric to the wheel axis, 
 its centre being in front of the shaft centre. On it revolves a ring, which 
 receives its motion by an arm fixed to it and jointed to one of the float levers, 
 so that it turns with the wheel ; connected to it is a rod from each float lever,. 
 so that as the wheel turns round the eccentricity of the gudgeon causes the 
 lever to vibrate through the angle necessary for efficiency. 
 
 The position of the gudgeon may be ascertained by placing a float with 
 its lever in the proper position for immersion and emersion ; with the lever 
 pin centres as centre, and the radius of the pitch circle as radius, two inter- 
 secting circles can be described ; then the point of upper intersection is the 
 position of the centre of the feathering gudgeon. Sometimes, however, 
 from practical considerations, the levers are at a less angle than 90° with the 
 float face (v. fig. 167) ; then the length ot the describing radii is less than 
 that of the distance of float centres from wheel centre. Formerly with the 
 slower-running engines an equivalent diameter twice that of the actual was 
 common practice and sufficient. Nowadays, with the smaller wheels and 
 the higher percentage of slip to give greater revolution rate, the equivalent 
 diameter is much greater, and may be even three times, as shown in fig. 167. 
 
 The Diameter of a Feathering Paddle-wheel depends on the speed of the 
 ship, the revolution of the engines, and the amount of slip. It is, of course, 
 evident that if, for any reason, the sectional area of the projected stream 
 of water from a propeller must be restricted or reduced there must be an 
 increase in the rate of flow, so that the effect may be the same. It has been 
 already shown that when (V — v) is the slip — 
 
 Thrust or momentum of water = 2 A X V (V — v) pounds. 
 
 The increase in (V — v) can be made only by an increase in V, ina much as 
 the speed of the ship v is fixed 
 
 This means that the float velocity must increase as the reduction of area 
 takes place, but A will vary inversely as V 2 . With the modern paddle 
 steamer of high speed it is, of course, desirable and even absolutely necessary 
 to have the floats, and consequently the wheel, as small as possible ; on the 
 other hand, in order that the machinery may be of minimum weight and 
 
465 
 
 Fig. 167.— Feathering Paddle-wheel. 
 
 30 
 
466 MANUAL OF MARINE ENGINEERING. 
 
 occupy the least space, the rate of revolution should be as high as practicable. 
 A high rate of slip is, therefore, altogether an advantage. In practice, 
 therefore, it is quite usual to find the rate 20 per cent., and sometimes even 
 as high as 25. For tug boats, however, 15 per cent, is general practice, as 
 the float is not designed for the speed of the ship, but the power of the engine, 
 although even in such ships the engines must not be so muzzled as to be 
 unable to develop good power. 
 
 If D be the effective diameter of a paddle-wheel, which may be taken 
 at the float centres of a feathering wheel, and at the centres of floats of a radial 
 one. 
 
 A is the area of one float. 
 
 V ,, the velocity of float centres in feet per second = 2 nY) X ^k- 
 
 E ,, the revolutions of the wheel per minute. 
 S ,, the speed of ship in knots per hour. 
 
 v ,, the velocity of ship in feet per second = -^- — '-oft — 1*689 S. 
 
 . V — v 
 
 f „ the fraction the slip is of V — that is, — == — 
 
 e ,, the efficiency of the engines. 
 
 E „ the efficiency combined of engines and wheels. 
 
 T R ,, the resistance of the ship in pounds. 
 
 Then, Power transmitted to wheels = I.H.P. X e. 
 
 Power delivered by the wheels = I.H.P. X E. 
 
 The stream of water projected by each wheel, A X V X 64 lbs. 
 
 The mass of water, — = 2A X V. 
 
 o'2i 
 
 The acceleration given to the water is (V — v). 
 
 Their pressure on the float = 2 A X V (V — v). 
 
 (a) The net work done (thrust X speed of ship in feet per minute) = 
 
 2A X V (V - v) X 60t; = 120 A x V (V - v) v. 
 
 .But the work done is also measured by T K X v, and by I.H.P. x E. 
 
 That is, ^ ^ = I.H.P. X E, or the tow rope horse-power (P R H.P.). 
 oo, 000 m Y p 
 
 The speed of the wheel, V = — — — ; it is also equal the speed of the 
 
 ship plus the slip — that is, v + (V — v), from which may be deduced the 
 
 rule that — - 
 
 V *» 
 
 Diameter of wheel in feet = 19 x _• 
 
 R 
 
 The Area of the Paddle Float can be calculated from first principles if 
 it is true that the transverse section of the wheel race or stream of water 
 projected from it is equal to the area of a float, and that the observed or 
 apparent slip is the real slip. Assuming they are so, then — 
 
PADDLE-WHEELS. 467 
 
 Quantity of water projected per second = — — - — _ ~ cubic feet 
 
 60 
 
 The acceleration may be taken as fX =- f X tcDR . 
 
 „, { , AxjDxE 64 Ax^DxR 
 Then mass of water = _ X ^= _ 
 
 ,_, , AxiDxE v /X)[DxR A/ , ^ _._ 
 Thrust = 3Q X 6Q = T ^ ) (-D X E) 2 . 
 
 A _ thrust X 1,800 _ thrust 182 
 
 ° r ' ■ /(«D X R) 2 (D X R) 2 X 7~' 
 
 This thrust is that due to the effort of one wheel ; so that if the thrust is 
 calculated from the ship resistance, and there are two wheels, half the value 
 so found must be inserted in the above equation. 
 
 Applying this theoretical formula to establish a rule to agree with the 
 best practice of the day, the following holds good : — 
 
 Area of one feathering float = -j — '—' x ( ^ — ^ 1 . 
 
 If E is 0'6, the value of C for a pair wheel ship in which I.H.P. is the 
 gross power developed by the engine, 83 - 2. 
 
 If there is a single wheel at the stern, C = 109. 
 
 As some paddle-wheel steamers have an efficiency value as high as 0*66, 
 the value of C is in that case 85 -6 ; this, however, is not common, while, on 
 the other hand, there are those whose efficiency is below - 6. 
 
 Example (i.). — What should be the area of float of a side- wheel steamer 
 having wheels 22 feet diameter revolving at 35 per minute, and a gross 
 I.H.P. 4,150, the slip percentage being 15 and speed 20 knots ? 
 
 Here area of float= ^^^, -X (^x^s)^ 41 ' 5 S 1 Uare feet - 
 
 Example (ii.). — What float area should a stern wheeler have with a wheel 
 10 feet diameter running at 50 revolutions with a speed of 12 knots, the slip 
 20 per cent., and I.H.P. 400 ? 
 
 400 / 109 \ 3 
 
 Area float = Q . 2 _ (02)2 * ( jo x 50 j = 25 ' 8 s q ua re feet. 
 
 The Number of Floats on a Feathering-wheel should be such that when 
 revolving at slow speeds there is not too much thud owing to irregular angular 
 advance — that is to say, the floats must be near enough together to maintain 
 something like an uniform area of float surface acting on the water. 
 
 D + 2 
 Rule in Practice. — Number of floats = — ^ — • 
 
 Radial wheels usually had a float for each foot of diameter, which means 
 they were at the tips 3"14 feet apart. The tendency of modern practice 
 
468 
 
 MANUAL OF MARINE ENGINEERING. 
 
PADDLE-WHEELS. 469 
 
 with the higher rate of revolution is to have as few floats as consistent with 
 good working. 
 
 The Proportions of Paddle Floats are quite arbitrary within reasonable 
 bounds. In practice the ratio of length to breadth on a radial fixed float 
 is 4 to 5, while for a feathering-wheel it is 2"6 to 3 — that is, A = r B X B, 
 or r B 2 , where B is the breadth in feet and r is the ratio ; hence 
 
 /A 
 
 Breadth of float = */ — . 
 
 Since length is r B, then 
 
 /A 
 Length of float = r*/ — • 
 
 For example, take the float at 41*5 square feet, and r as 3 — 
 
 Then, Breadth of float = yj^~ = 372 feet. 
 
 Length = 3 X 372 = 11-16 feet. 
 
 If the steamer has to navigate narrow passages, as dock entrances, etc., 
 the ratio will not exceed 2*6, in that case — 
 
 Breadth of float =\/ ~k^ = 4*0 feet. 
 Length „ =4x2-6 = 10-4 feet. 
 
 The Thickness of the Elm Floats is usually about T \r the breadth, but 
 steel floats, as shown in fig. 168, made of plate stiffened by slightly curving 
 it. The edge resistance of steel plates is, of course, less than that of elm 
 boards, but, if damaged by contact with wreckage, they are not so easy to 
 remove in a seaway, although when removed they can be generally restored 
 to proper shape and used again. The steel plates are also cheaper, as there 
 is considerable labour invo'ved in preparing, shaping, and bolting together 
 the elm when of large size. 
 
 Paddle-Wheels are usually made now without outer rims, as shown in 
 fig. 168 ; they are thereby less liable to damage, and the limit of extreme 
 diameter is less ; moreover, the friction in the water is less, although to a 
 large extent any resistance to the passage of the wheel should go towards 
 thrust. On the other hand, the floats have not the same protection as they 
 enjoy with the outer rim, and the arms beyond the inner rim being canti- 
 levers, must be much stouter than when the arm has the double support 
 of the two rims. The structure of wheel is now wholly of steel, as being 
 cheaper and, on the whole, stronger than wrought iron ; and the hub or boss, 
 which was formerly of cast iron hooped with wrought iron, is now often a 
 steel casting. 
 
 The Feathering Centre and Gear is, for convenience, often secured to the 
 sponson beam, as shown on fig. 168. but, as this beam is liable to be sprung 
 considerably when coming alongside piers, especially in rough weather, it 
 is safer and snugger to fit an eccentric on the outer bearing, as in fig. 169. 
 This arrangement possesses another advantage that, being accessible and 
 somewhat shielded from water wash, it can be attended to and treated with 
 
470 
 
 MANUAL OF MARINE ENGINEERING. 
 
 O 
 
 SP 
 
 •c 
 
 a> 
 
 ■** 
 
 eg 
 S3 
 
 1 
 •S 
 
 .3 
 
 t- 
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 41 
 
 OS 
 
 
PADDLE-WHEELS; 471 
 
 oil as a lubricant. The eccentric is rigid in its connection, and has no spring 
 in itself, and liable to no damage as the outer gear is. 
 
 The Immersion of the Paddle Float must be greater in sea-going ships 
 than in those limited to smooth water. In a general way ships which are 
 liable to ply among waves should have in smooth water an immersion of 
 the inner edge of float when at lowest position equal to one-half the breadth, 
 while if for service in comparatively smooth water an immersion of a quarter 
 is sufficient. For light-draft steamers as little as one -eighth suffices, but it 
 must always be borne in mind that the greater the immersion the larger is 
 the section of wheel race ; moreover, as the immersion is increased by the 
 additions to the deadweight carried, some regard must be given to the cir- 
 cumstances of the ship's lading. 
 
 Paddle-wheel Bosses are made of the form shown in figs. 167, 168, and 169, 
 and formerly of the toughest cast iron ; they were, when of large size, further 
 strengthened against splitting by shrinking on wrought-iron hoops, as shown 
 in fig. 168, both at the centred body and at the edges of the flanges. To-day 
 such castings can be obtained in steel which require no such safeguards, 
 and are much lighter as well as stronger. The earliest practice was to stake 
 the boss on the square section end of the paddle-shaft ; that, however, is a 
 bygone practice, which, while rough and ready, had the merit of safety. 
 Since then the boss has been bored out to fit the shaft, and keyed to it with 
 two keys 90° apart. It was, however, most difficult to withdraw bosses 
 fitted in this way ; therefore, the better plan is to fit them with a taper, 
 as is done with screw propeller bosses. 
 
 Taking the diameter of the shaft, which is sufficient for the torque of 
 the engines, as a criterion of size from which to calculate the wheel scantlings, 
 or take the diameter of the inner journal of the actual shaft calculated by the 
 following, viz. : — 
 
 yT XJ p 
 ' ' ' x F. 
 
 For a side-wheel steamer with a single cylinder, . . . F = 80. 
 
 For side wheels with two-cylinders, cranks at 90°, . . F = 58. 
 
 „ „ ,, ,, and an intermediate shaft, F = 50. 
 
 „ „ „ 2 solid cranks, 90°, . . F = 55. 
 
 „ three- „ 3 cranks, 120°, . . F = 50. 
 
 Thickness of cast-iron boss round shaft = 0*28 X d. 
 
 „ steel „ „ = 0-20 X d. 
 
 Diameter of outer flanges, = 4'0 X d. 
 
 Thickness ,, cast iron = 0-15 X d. 
 
 ,, ,, cast steel = Oil X d. 
 
 Breadth of two hard steel keys each =0*18 X d -f- \ inch. 
 
 Thickness „ „ „ = 0-09 X d + \ inch. 
 
 All shafts over 8 inches diameter should have two keys. 
 
 Paddle Arms of small radial wheels with two rims are generally made of 
 flat rolled bars. In large feathering wheels they are always forgings and 
 now generally of steel. 
 
 Let n be the number of floats on a wheel, with a pair of arms to each 
 float, then if the framework is braced and tied together as a compact structure, 
 
472 MANUAL OF MARINE ENGINEERING. 
 
 and thereby distribute the load, and b is the breadth and t the thickness 
 of each ami at the boss — 
 
 The resistance to bending of each arm is =/{t X b 2 ). 
 The total resistance of the structure = 2nJ(t X b 2 ). 
 
 The shaft journal has to work against this resistance, which may be 
 
 designated as P. 
 
 m , P X diameter of wheel . , , ., , . 
 
 Then = is the torsion on the shaft, and equal to the 
 
 L 
 
 bending moment on the wheel structure. 
 
 i n d z 
 
 Now, the tore | ue = -^- X/i- 
 
 D -f/ s 
 Hence P X ■= = — X / 2 = 2w/ a (i X 2> 2 ). 
 
 If/j be taken as practically equal to 08/ for ordinary mild steel, then 
 
 txb 2 = -0982 d? x 0-8 = 0-07854 -. 
 
 In practice the ratio of b to t is 5 at the boss and 3*5 with same thickness 
 
 at the rim ; when there is only an inner ring the ratio of b to t just outside 
 
 the ring is from 6 to 7. 
 
 b z 
 Taking the ratio at boss as 5, then t X b 2 = -=-• 
 
 Then % = 0-07854 - ; and b = /J'---. 
 o n V n 
 
 Then breadth of arm = 3 - • 
 
 \'n 
 
 For example, if n is 8, then b = 0-365(7. 
 
 For tug boats and ships making sea passages in winter time, where the 
 wheels are liable to heavy shocks from the sea, the structure should be a 
 somewhat heavier one ; moreover, in ordinary work the wheel is liable to 
 rough usage, and the load is not evenly distributed. To provide for such 
 contingencies, the divisor should be (n — 2) rather than n. 
 
 The Rims of Paddle-wheels are of rolled bar iron or mild steel, and when 
 there are two rims, as in fig. 167 — 
 
 Breadth of outer ring = 0*4 X d. 
 ,, inner ,, = 4 X d. 
 Thickness of outer ,, = 0-08 X d. 
 ,, inner ,, = - 10 X d. 
 
 When there is only an inner ring the section should be 0'5d X 0-14(7 with 
 bolts 0*15(2 in diameter, while the bolts with two rings may be smaller, or 
 0*12(2. The diagonal tie-rods are (0*18(2 +i inch) in diameter, while the 
 cross bars are (0*18(2+ | inch), and made hollow. 
 
 It should be noted that these rules are laid down for guidance, for in actual 
 practice the nearest standard section should be selected for use, as also the 
 bolts used should be of store dimensions. 
 
 Of the Feathering Gear, the Moat gudgeons should be in diameter equal 
 to (0*1 the breadth of Hoat -f- » inch), and the length of bush 1*40 the diameter. 
 
THE SCREW PROPELLER. 473 
 
 The radius-rod pins should be in diameter = 0*50 that of the gudgeons, 
 and fitted as shown in fig. 169. The pins, gudgeons, etc., are usually cased 
 in hard bronze, and the holes in which they fit bushed with lignum vitce, 
 inasmuch as the only lubricant they get is water. It is, however, not 
 uncommon now to use hard steel or case-hardened pins, etc., and employ 
 white metal bushes, which wears well in sandy water. 
 
 The Position of the Gudgeons on the Floats is not always the middle, but 
 nearer to the centre of pressure, as shown in fig. 167. About two-fifths the 
 breadth of the float from the outer edge is a good position for the turning 
 axis, as then the twisting load on entry is not so great, and during the progress 
 through the water there is a better balance than if hung centrally. 
 
 The Outer Bearing of the Paddle-shaft has to take the weight of the shaft 
 and wheel and the horizontal thrust of the propeller ; it ought, therefore, 
 to be designed with the centre line diagonal instead of vertical, as it usually 
 is for convenience of getting the shaft into place. The resultant pressure 
 on the bearing is diagonal in direction and of considerable magnitude, so that 
 the -bearing must be large and strong. It is usually two diameters of the 
 shaft- journal in length, and lined with white metal. It is largely treated 
 with water, so that it cannot get hot, and supplied with a central chamber 
 filled with tallow, which keeps out the water and lubricates the bearing if 
 it gets warm. It is desirable that it shall be also lubricated with oil, 
 as is any other bearing, if good results are to be expected. The cap of the 
 bearing is generally a plain iron casting like that on a tunnel bearing in a 
 screw ship. It takes no load, and only prevents the shaft from lifting or 
 jumping in the bearing, as it might do in a seaway. 
 
 The Screw Propeller is a more important instrument of propulsion, as 
 it is in universal use for every kind of ship and on every service, having 
 ousted the paddle in places where at one time it seemed impossible to be 
 employed with advantage. The inventions of Sir John Thornycroft and 
 Sir A. Yarrow, whereby screws can be employed in the shallowest of 
 water — so shallow, indeed, that when at rest the water level is very near 
 the screw axis — have revolutionised the practice of marine engineers. More- 
 over, the covering in of the propeller races in the ways adopted by these 
 gentlemen has immensely improved the efficiency of all screw ships where 
 the screw was quite but only just immersed. 
 
 For Purposes of Calculation it is sufficient to assume that the stream 
 of water from a screw is a hollow column of the same external diameter as 
 the screw, and the internal is as the diameter of the boss. In the mercantile 
 marine the boss is so small, compared with the diameter of the screw, especi- 
 ally with solid ones, that no great discrepancy arises if the column is assumed 
 to be solid. In the naval service, and with all ships having large power and 
 screws of small diameter, this course cannot be adopted, for the boss disc 
 is from 5 to 12 per cent, of the screw disc. 
 
 The Surface itself is now generally a portion of a true helix, although there 
 are still not wanting those who hold that there is some virtue in a screw of 
 variable pitch, and while certain incline to a variation from the boss to the 
 tip, others prefer the increase to be from the leading edge to the following, 
 so that the water at contact with the blade may have only a very slight 
 displacement, but a gradual increase as it passes along the face with the 
 final acceleration just as it leaves at the following edge. If such propellers 
 
474 
 
 MANUAL OF MARINE ENGINEERING. 
 
 continue to be made, the effective pitch, or that taken for calculation, must 
 not be a mean, but the final pitch. 
 
 The Original Screw of F. P. Smith {v. fig. 170), as adopted in the Navy 
 and mercantile marine, was a true screw of two convolutions cut off by parallel 
 planes one-third to one-eighth of the pitch apart ; in fact, it was a two-bladed 
 screw with very wide tips, but quite a moderate amount of acting surface, 
 so that in H.M.S. " Rattler " it was very efficient, much more so than most 
 of the screws made for many years afterwards and called " common screws," 
 but of very large diameter, and their length only one-eighth the pitch. 
 
 Theoretically, the Two-bladed Screw is the most efficient, and in practice 
 in smooth water on a ship with fine lines it is likewise so ; the vibration 
 set up by it, however, militates so much against its use that to-day it is only 
 found in quite small craft, and in those ships that still have to make use of 
 sail power as well as steam. 
 
 The Developed Surface of a screw is that actually acting on the water 
 
 Fig. 170. — Smith's Screw Propeller as first fitted in the Navy. 
 
 and capable of being measured as a part of the blade. It is, therefore, some- 
 times called the acting surface. 
 
 The Projected Surface is that of the blade as projected on a plane at right 
 angles to the axis, or the view of the blade from abaft as drawn on paper. 
 This is sometimes spoken of as the active surface, and taken as the criterion of 
 the screw's ability to produce thrust, being the transverse component of 
 each portion of the real surface. 
 
 The Diameter of the Screw is that of the circle moved through by the 
 extreme tip of a blade. 
 
 The behaviour of a screw and the conditions prevailing in its neighbour- 
 hood when at work have been already dealt with (Chap, ii.), it will, therefore, 
 be assumed here that the screw is working in practically undisturbed water, 
 which before entering the screw race is moving with respect to the screw at 
 the same speed as the ship, but in the opposite direction. This is really the 
 
INDICATED THRUST. 475 
 
 condition under which model screws generally have been tried in the past. 
 To-day, with the modern tank and its elaborate apparatus, the model screw 
 is tried behind the model ship, which is presumed to have similar wake currents 
 to that of the real ship. 
 
 The Thrust of a Screw is the resultant of all the pressures on it acting in a 
 direction parallel to the axis applied axially through the shafting to the thrust 
 block. 
 
 Indicated Thrust is an approximation to the actual thrust arrived at by 
 means of a formula proposed by the late Dr. W. Froude, and very useful 
 for guidance in dealing with propeller problems when rightly understood 
 and properly applied. It can be used with advantage in treating compara- 
 tively the results of trials, but for absolute results it cannot be relied on. 
 Froude's rule is as follows : — 
 
 t a- 4. i 4.x. j I.H.P. x 33,000 
 Indicated thrust in pounds = ^ ~ , 
 
 where V is the velocity of the stream in feet per second — that is 
 = pitch X revolutions. 
 
 With the 80 per cent, efficiency of the engines in Dr. Froude's day the 
 results obtained by this formula were more nearly true than now with better 
 engines. 
 
 Real thrust X 60 v = I.H.P. X E X 33,000. 
 
 v. 
 If ^ = E, then Froude's formula holds good — that is, when the slip per 
 
 cent, is the same as the per cent, of loss of engines and screw it is true. 
 
 The efficiency of the screw may be examined on the system proposed 
 by him — viz., by constructing curves of indicated thrust on the same principle 
 as that described in Chap, ii., for curves of I.H.P. 
 
 In this case, however, the ordinates represent the thrust as calculated 
 from the I.H.P., or from the pressure on the pistons, and this for convenience 
 is generally expressed in tons. 
 
 It is assumed that the pressure on the pistons multiplied by twice their 
 stroke is equal to the thrust multiplied by the pitch, and if there were no 
 loss by friction of machinery, etc., by the principle of work this would be true. 
 
 Let p be the mean effective referred pressure on the low-pressure pistons 
 in pounds per square inch, n their number, A their area in square inches, 
 L the length of stroke in feet, and P the pitch of the screw in feet ; then, 
 
 Thrust xP = pxAxwx2L, 
 
 or Thrust = 5 
 
 If both numerator and denominator of the fraction be multiplied by R, 
 the number of revolutions per minute, 
 
 m , pXAxwX2LxR. 
 
 Thrust = pint • 
 
 but P xAxnx2LxR 
 
 33,000 
 
 Therefore Thrust = - — J , p ' • 
 
 r X K 
 
476 MANUAL OF MARINE ENGINEERING. 
 
 This is called the indicated thrust, and it was by constructing a curve of 
 indicated thrust that the inefficiency of the original screws of H.M.S. ' ; Iris" 
 was discovered. 
 
 Dr. Froude* also explained the method by which he estimated the 
 " initial friction," or " the equivalent of friction of the engines due to the 
 working load." He said — " When decomposed into its constituent parts 
 indicated thrust is resolved into several elements, which must be enumerated 
 and kept in view. These elements are — 1, the useful thrust, or ship's true 
 resistance ; 2, the augment of resistance, which is due to the diminution 
 which the action of the propeller creates in the pressure of the water against 
 the stern end of the ship ; 3, the equivalent of the friction of the screw blades 
 in their edge way motion through the water ; 4, the equivalent of the friction 
 due to the dead weight of the working parts, piston packings, and the like, 
 which constitute the initial or slow-speed friction of the engine ; 5, the 
 equivalent of friction of the engines due to the working load ; 6, the equi- 
 valent of air-pump and feed-pump duty. 
 
 " It is probable that 2, 3, and 4 of the above list are all very nearly pro- 
 portional to the useful thrust ; 6 is probably nearly proportional to the square 
 of the number of revolutions, and thus, at least at the lower speeds, approxi- 
 mately to the useful thrust : 5 probably remains constant at all speeds, and 
 for convenience it may be regarded as constant, though perhaps in strict 
 truth it should be termed ' initial friction.' If, then, we could separate the 
 quasi-constant friction from the indicated thrust throughout, the remainder 
 would be approximately proportional to the ship's true resistance. 
 
 " Now, on drawing a curve (of indicated thrust) ... it becomes at 
 ■once manifest in every case, that at its low-speed end the curve refuses to 
 descend to the thrust zero, but tends towards a point representing a con- 
 siderable amount of thrust, and it is impossible to doubt that this apparent 
 thrust at the zero of speed, when there can be no real thrust, is the equiva- 
 lent of what I have termed initial friction; so that if we could determine 
 correctly the point at which the curve, if prolonged to the speed zero, would 
 intersect the axis Y (fig. 171), and if we were to draw a line through the 
 intersection parallel to the base, the height which would be thus cut off from 
 the thrust ordinates would represent the deduction to be made from them in 
 respect of constant or initial friction, and the remainders of the ordinates 
 between this new base and the curve would be approximately proportional to 
 the ship's true resistance." 
 
 Dr. Froude then explains his method, which is substantially as follows : — 
 
 Let OB, C, D represent the three progressive speeds, observed in 
 the usual way, and BE, C F, D G the indicated thrust, calculated from the 
 data observed at those speeds ; through the points G, F, E draw the curve of 
 indicated thrust. Let A represent a low speed, which should not exceed 
 5 knots, and A H the corresponding indicated thrust ; continue the curve to 
 H. and at H draw a tangent to the curve cutting Y. Take a point M 
 
 between A, so that ^-^ = yrr- Draw M K parallel to A H, and cutting 
 
 ■the tangent line at K. Through K draw a line K T parallel to D, cutting 
 
 * Transactions of the Institution of Naval Architects, vol. xvii., 1876 
 
RANKINE S RULE FOR THRUST. 
 
 477 
 
 Y at T. Then T is the part cut off Y by the curve, and represents the 
 initial friction. 
 
 Dr. Froude deduced, from careful investigation, that only 37 to 40 per 
 cent, of the whole power delivered is usefully employed, and that " the 
 constant friction is equivalent to from one-eighth to one-sixth of the gross 
 load on the engine when working at its maximum speed and power." 
 
 Rankine's Rule for Thrust is summarised as follows : — 
 
 Professor Rankine gives (Rules and Tables, p. 275) this same «ule in 
 another and more convenient form for practical use, viz. : — 
 
 Rule V. — To calculate the thrust of a propelling instrument (jet, paddle, 
 or screw) in pounds ; multiply together the transverse sectional area, in square 
 feet, of the stream driven astern by the propeller ; the speed of the stream relatively 
 to the ship in knots ; the real slip, or part of that speed which is impressed on 
 
 Speed in Knots 
 Fig. 171. — Curve of Indicated Thrust. 
 
 that stream by the propeller, also in knots ; and the constant 5-66 for sea-water, 
 or 55 for fresh water. 
 
 That is, if S is the speed of the screw in knots, s the speed of the ship 
 in knots, A the area of the stream in square feet (of sea-water), 
 
 Thrust in pounds = A X S (S - s) X 5-66, or D 2 X S (S - s) X 4'45. 
 
 The effective slip or the actual mean velocity imparted to the water 
 cannot be found accurately any more than can the mass of water set in 
 motion. The effective thrust, however, may be found approximately, thus — 
 
 Rule.— Effective thrust = (D X S) 2 -=- F. 
 
 D the diameter in feet, S the speed of screw in feet per second, and F varies 
 from. 3-3 with small quick propellers to 5 with large slow ones. Ordinary 
 merchant steamers, 5 to 45 ; fast-running merchant and naval ships. 40 ; 
 very high speed, as with turbines, 33. 
 
473 
 
 MANUAL OF MARINE ENGINEERING. 
 
 
 
 TABLE 
 
 XLVIL- 
 
 -Particulars of Scr; 
 
 ew Propellers. 
 
 
 
 
 
 
 Solid Cast Iron. 
 
 
 
 
 -! 
 
 Ship. 
 
 
 
 Blades. 
 
 Length 
 of 
 
 Diaru. 
 of 
 
 
 Mat 
 
 srlaL 
 
 
 B = SSngl« 
 
 Screw. 
 
 Diam. 
 
 Pitch. 
 
 
 ja 
 
 
 ■_■ a> 
 
 I.H.P. 
 
 
 
 Weight. 
 
 1 = Twin 
 Screw. 
 
 
 
 No. 
 
 a ca 
 
 ** - 
 
 pa 
 
 j4 
 
 
 
 H 
 Ins. 
 
 CS 
 
 jo 
 Sq. ft. 
 
 Boss. 
 
 Shaft. 
 
 it 
 
 Bom. 
 
 Blades. 
 
 
 
 Ft. Ins. 
 
 Ft. Ins. 
 
 
 Ft. Ins. 
 
 Ft. Ins. 
 
 Ins. 
 
 
 
 
 Cnts. 
 
 Dn., S., 
 
 19 6 
 
 26 6 
 
 4 
 
 3 104 
 
 10 
 
 112 
 
 3 9 
 
 171 
 
 64-7 
 
 • • * 
 
 ... 
 
 250-5 
 
 Nno., S., 
 
 18 3 
 
 26 
 
 4 
 
 3 9 
 
 9 
 
 96 
 
 3 
 
 13 
 
 36-1 
 
 ••a 
 
 ... 
 
 202-5 
 
 Sir., S, 
 
 16 8 
 
 23 6 
 
 4 
 
 4 3 
 
 84 
 
 92 
 
 3 6 
 
 13| 
 
 367 
 
 •■• 
 
 ... 
 
 189-0 
 
 Ddo., S., 
 
 17 6 
 
 17 6 
 
 4 
 
 3 6 
 
 74 
 
 90 
 
 2 9 
 
 134 
 
 342 
 
 •M 
 
 •«. 
 
 124-7 
 
 Ard., S., 
 
 16 
 
 20 
 
 4 
 
 3 9* 
 
 74 
 
 85 
 
 2 9 
 
 13| 
 
 33-0 
 
 • •• 
 
 •M 
 
 121-0 
 
 Arl., S., 
 
 10 
 
 18 
 
 4 
 
 2 9* 
 
 n 
 
 74 
 
 2 9 
 
 121 
 
 218 
 
 • •• 
 
 .•• 
 
 110-75 
 
 Eld., S., 
 
 13 6 
 
 21 
 
 4 
 
 3 3| 
 
 n 
 
 68 
 
 2 9 
 
 114 
 
 28-8 
 
 BM 
 
 »«« 
 
 98 90 
 
 Hyd.,S., 
 
 12 9 
 
 16 
 
 4 
 
 3 34 
 
 H 
 
 58 
 
 2 5 
 
 11 
 
 14-8 
 
 • •• 
 
 ... 
 
 75-65 
 
 Esw., S., 
 
 13 
 
 16 4 
 
 4 
 
 2 54 
 
 5| 
 
 48 
 
 2 3 
 
 9| 
 
 10-3 
 
 • • • 
 
 
 6140 
 
 Alt., S., 
 
 10 3 
 
 17 3 
 
 4 
 
 2 11^ 
 
 5 
 
 43 
 
 2 1 
 
 H 
 
 8-7 
 
 • * • 
 
 ... 
 
 44-10 
 
 Bin., T., 
 
 12 
 
 16 6 
 
 3 
 
 3 3 
 
 8 
 
 40-5 
 
 2 6 
 
 HI 
 
 20-0 
 
 • •* 
 
 • *• 
 
 75-50 
 
 Qnw.,S., 
 
 9 9 
 
 13 6 
 
 4 
 
 2 4 
 
 4 
 
 33-5 
 
 1 94 
 
 6| 
 
 6-10 
 
 ... 
 
 ... 
 
 29 60 
 
 Hwh.,T., 
 
 9 
 
 14 
 
 3 
 
 2 10| 
 
 y 4 
 
 26 
 
 1 10 
 
 n 
 
 5 69 
 
 ... 
 
 
 27 50 
 
 Erp., S., 
 
 9 
 
 13 
 
 3 
 
 91 
 
 44 
 
 2 10 
 
 1 6 
 
 6£ 
 
 4-00 
 
 
 ... 
 
 20 30 
 
 Wtm.,S., 
 
 7 8 
 
 9 
 
 3 
 
 2 14 
 
 34 
 
 175 
 
 1 3 
 
 54 
 
 230 
 
 ... 
 
 • •• 
 
 12-80 
 
 Skb., S., 
 
 6 3 
 
 7 
 
 3 
 
 1 9| 
 
 H 
 
 125 
 
 1 14 
 
 4| 
 
 115 
 
 ... 
 
 ... 
 
 8-50 
 
 Fbc, S., 
 
 4 6 
 
 5 
 
 3 
 
 1 lj 
 
 24 
 
 5-5 
 
 9 
 
 31 
 
 42 
 
 • *« 
 
 • •• 
 
 2-90 
 
 
 
 
 Looi 
 
 m Blades Boi 
 
 .TED O 
 
 N. 
 
 
 
 
 
 Ft. Ins. 
 
 Ft. Ins. 
 
 
 Ft, Ins. 
 
 Ins. 
 
 Sq. ft. 
 
 Ft. Ins. 
 
 Ins. 
 
 
 
 
 
 Cpt., S., 
 
 20 3 
 
 25 
 
 4 
 
 4 5J 
 
 84 
 
 120 
 
 3 9 
 
 181 
 
 70-0 
 
 C. steel 
 
 M. bronze 
 
 311-5 
 
 Mtl., S., 
 
 19 
 
 22 
 
 4 
 
 3 84 
 
 6 
 
 96-5 
 
 3 6 
 
 154 
 
 40 
 
 C. steel 
 
 Steel 
 
 196-7 
 
 Str., S., 
 
 19 
 
 23 
 
 4 
 
 4 44 
 
 Si 
 
 960 
 
 3 74 
 
 154 
 
 41-2 
 
 C. iron 
 
 C. iron 
 
 28S-0 
 
 Czg., S., 
 
 17 
 
 21 6 
 
 4 
 
 4 2 
 
 H 
 
 !)0-0 
 
 3 6 
 
 15 
 
 41 
 
 C. iron 
 
 C. iron 
 
 223-9 
 
 Ldn., T., 
 
 17 6 
 
 19 9 
 
 4 
 
 4 
 
 7£ 
 
 86 -0 
 
 4 
 
 174 
 
 72 
 
 M. bronze 
 
 M. bronze 
 
 261 -0 
 
 Std.. T, 
 
 16 
 
 25 
 
 3 
 
 3 If 
 
 6i 
 
 770 
 
 3 9 
 
 18 
 
 75 
 
 D. metal 
 
 D. metal 
 
 216-5 
 
 Stm., S., 
 
 17 
 
 21 
 
 4 
 
 3 5 3 
 
 71 
 
 72 
 
 2 8 
 
 134 
 
 301 
 
 C. iron 
 
 C. iron 
 
 183 7 
 
 Rmo..S., 
 
 15 
 
 23 
 
 4 
 
 3 8£ 
 
 6| 
 
 63-6 
 
 2 9 
 
 134 
 
 25-0 
 
 C. iron 
 
 C. iron 
 
 170-3 
 
 End.,T., 
 
 16 6 
 
 22 
 
 3 
 
 3 10 
 
 7i 
 
 60 
 
 3 44 
 
 m 
 
 54 5 
 
 A. bronze 
 
 A. bronze 
 
 166-6 
 
 Kvn., S., 
 
 14 
 
 17 
 
 4 
 
 2 9 
 
 6 
 
 55-0 
 
 2 1 
 
 93 
 
 11 2 
 
 C. iron 
 
 C. iron 
 
 99 5 
 
 Apl., T., 
 
 13 
 
 17 
 
 3 
 
 3 34 
 
 61 
 
 40-0 
 
 3 
 
 144 
 
 32 9 
 
 A. bronze 
 
 A. bronze 
 
 107 7 
 
 Clm., T., 
 
 12 
 
 16 
 
 3 
 
 3 2£ 
 
 5 
 
 39 
 
 2 6 
 
 hi 
 
 201 
 
 C. steel 
 
 S. bronze 
 
 71 80 
 
 Cbg.. T., 
 
 11 6 
 
 18 6 
 
 3 
 
 3 1J 
 
 *1 
 
 34 
 
 2 2 
 
 101 
 
 133 
 
 C. iron 
 
 S. bronze 
 
 62 -85 
 
 Prl., V.. 
 
 12 6 
 
 14 
 
 3 
 
 2 9 
 
 6 
 
 33 
 
 2 6 
 
 12* 
 
 22-8 
 
 A. bronze 
 
 A. bronze 
 
 76 5 
 
 Pis., T., 
 
 10 G 
 
 12 
 
 3 
 
 2 5£ 
 
 4£ 
 
 24-0 
 
 2 6 
 
 12g 
 
 16 3 
 
 A. bronze 
 
 S. bronze 
 
 64-7 
 
 Blc, T., 
 
 9 3 
 
 10 
 
 3 
 
 2 5^ 
 
 3| 
 
 213 
 
 2 14 
 
 9 
 
 7-5 
 
 A. bronze 
 
 A. bronze 
 
 36-0 
 
 Sim., T., 
 
 6 7 
 
 8 9 
 
 3 
 
 1 10| 
 
 2 
 
 12-7 
 
 1 Of 
 
 64 
 
 5 
 
 S. bronze 
 
 S. bronze 
 
 7-45 
 
 The Pressure per Square Inch on a Propeller Blade at any position and 
 at any rate of revolution may be found by referring to the diagram on p. 479 
 (fig. 172). If, for example, it is required to know the pressure at a radius 
 of 6 feet when the revolutions are 80, all that is necessary is to make the 
 intersection of the vertical line through 6 with the curve for 80 revolutions, 
 
SPEED A,ND PRESSURE DIAGRAM. 
 
 479 
 
 REVOLUTIONARY SPEED IN FEET per SECOND 1 
 
 S^s^s u u * * 01 o> © 01 
 O01 o eg o 01 o 
 
 
 5) 
 
 o 
 
 PRESSURE in LBS per SQ-INCH. 
 
 and the horizontal line through it to the right-hand side shows it to be a 
 little over 17 lbs. 
 
 The Diameter of Screw best suited to a Ship is a matter for much 
 
480 MANUAL OF MARINE ENGINEERING. 
 
 consideration, and to arrive at it all the circumstances of the particular case 
 must be taken into account. The three leading features of a screw — viz., 
 diameter, pitch, and surface of blades — are interdependent, one on each of the 
 other two, or two together on one. The shape of the ship and her draught 
 of water are governing factors ; and, last but not least, there must be some 
 relation between the power developed per revolution by the engine or motor 
 and the area of disc of the propeller, as also the area of blades themselves. 
 There are some other considerations extraneous to ship and engine, such as 
 the fear of cavitation with a coarse pitch and insufficient blade surface ; the 
 tendency to swamp the screw with air if the blade tips are too near the 
 surface ; and the risk of damage if the blades of multiple- screw ships project 
 so far out sideways that the quarters of the ship cannot protect them from 
 contact with dock and quay walls. 
 
 In a general way it may be taken as an axiom that the screw of highest 
 efficiency is the one which with the smallest diameter is sufficient for the 
 power transmitted to it. The objection to large diameter screws for high 
 efficiency of propulsion as against high thrust is well set forth by Dr. W. 
 Froude, who says — * 
 
 " Take the case of a screw 20 feet diameter, making 80 revolutions per 
 minute ; the tips of the blades are travelling at a speed of about 50 knots ; 
 now, the resistance of a surface so short in the line of motion as a screw 
 blade, even ivhen its surface is quite smooth, is as much as l£ lbs. per foot 
 at 10 knots, and is nearly as the square of the speed, and as each square foot 
 of blade area involves 2 square feet of skin, the resistance of each is over 
 60 lbs. ; thus, making some allowance for thickness and bluntness, there 
 is involved in driving it at 50 knots at least 10 I.H.P., and collectively the 
 outmost foot of four such blades, each 3 feet wide, would absorb fully 120 
 l.H.P. in surface friction ; and though the parts nearer to the root move 
 with proportionally less speed, and therefore with less resistance, yet, on the 
 other hand, screw blades are generally rough from the sand, and have pro- 
 bably a still higher coefficient of frictional resistance. " 
 
 If a screw is of sufficient diameter, and has ample blade area, any addition 
 beyond the tip will act then quite as a brake and seriously reduce its 
 efficiency; and that such screws produce augmented resistance is also a 
 common experience. 
 
 The loss by surface friction of blades in these days of high rates 
 of revolution is more pronounced than ever. But perhaps the most 
 attractive reason for adopting the small screw is that greater immersion of 
 upper-blade tip is possible ; all other tilings being equal, the small screw 
 thoroughly well immersed must be a more efficient one. in a sea-going ship 
 especially, than a larger one with small immersion. 
 
 That the Limit to Screw Diameter is the draught of water of the ship 
 at the stern no longer holds good as it did when draft of water was the 
 governing factor. As already shown, a screw in a shallow-draft ship may 
 be largely in excess of it if means are taken to keep the water massed up 
 by the screw when working from the pressure of the atmosphere and from 
 dispersal. Professor Flamm has shown how, by means of a stream of water 
 at high velocity, directed diagonally along a surface, the relief of air pressure 
 is greater than the normal effect of the impinging stream on that surface. 
 
LIMIT OF SCREW DIAMETER. 481 
 
 It is also a matter of common knowledge that, if the stern lines of a ship 
 are full, so that the now of water into the screw race is restricted, a small 
 diameter screw is a very inefficient one to produce thrust, so the ordinary- 
 cargo steamer must have one of large diameter compared with the engine 
 size to work effectually. On the other hand, a twin-screw ship having its 
 propellers in the open, and largely free from the influence of ordinary wake 
 currents, may have screws of smaller diameter in proportion to the engine 
 power. 
 
 It is evident then the diameter must be such as to suit both the ship 
 and the engine power, and that it cannot be arrived at satisfactorily by 
 considering only the needs of the one, as was the case when the ratio of area 
 of screw disc to that of immersed mid-section was the criterion. The following 
 rules may be used to give an appropriate diameter of screw under the varying 
 circumstances experienced in ship design :— 
 
 D is the diameter of the L.P. cylinder of the engine whose piston stroke 
 
 is S, both in feet. 
 
 ^ . n . . „ ■ , t ,1 i • displacement X 35 
 
 P c is the prismatic coefficient of the ship = -, -r— =-j -. — 
 
 r r length X mid-section 
 
 Z a multiplier = (2'4 + P c ) for twin screws, 
 
 and (2-8 + P„) for single screws. 
 
 R the revolutions per minute of the screw. 
 
 Rule I. — Diameter of screw in feet = Z X n/l- X t 5 . 
 
 _ -r, 3 /l.H.P. 
 
 Rule II.— ., „ = x X P c ^/ R ♦ . 
 
 For single screws, . •. . a? = 7*25. Ocean-going express, 7*61. 
 
 ,, twin screws, . . . x — 6-55. „ ,, 6-88. 
 
 ,, quadruple screws, . . x = 6-25. „ „ 6-51. 
 
 * ,, turbine-driven centre screw, x = 6-55. „ ,, 6-88. 
 
 wing ,, x = 5-75. „ „ 6-04. 
 
 N.B. — In no case, however, must P c have a less value than 0-55. 
 
 Example (i.). — What should be the diameter of the screw of a cargo boat 
 having engines with cylinders 25, 38, and 64 inches diameter and 45 inches 
 length of stroke, the prismatic coefficient of hull being 085 ? 
 
 /64 45 
 12 X 12 ~ 16 ' 3 feet 
 
 Example (ii.). — An express steamer for cross-channel service is intended 
 to be twin screw and to have engines of 6,000 I.H.P. aggregate. The rate 
 of revolution is 180, and the prismatic coefficient of hull is - 6. 
 
 3 / 5 000 
 
 Diameter of screw (Rule II.) = 6-55 x0^^= 10-02 feet. 
 
 Example (iii.).— An ocean-going steamer driven by turbines is to have 
 three screws, the total S.H.P. is 24,000 at 250 revolutions, the hull prismatic 
 
 * With geared turbines and electrically driven screws the ordinary rules hold good 
 
482 MANUAL OF MARINE ENGINEERING. 
 
 coefficient is 065, what diameter should be the screws, the power being 
 evenly divided ? 
 
 yo 000 
 -V^TT = 14*17 feet. 
 250 
 
 wing „ „ = 6-04 X 0-65^^ =12-44 feet. 
 
 The Diameter of a Screw Propeller may be determined, however, by 
 means of a formula based on quite correct principles of theory, but with a 
 factor or constant multiplier modified by applying the rule in its purely 
 theoretical form to the actual successful practice of to-day with all kinds 
 of ships. 
 
 (a) The thrust of a screw in pounds = 2 A X V (V — v). 
 
 (b) The work done per mimite = 2Ax V (V — v) 60 v foot-lbs. 
 
 „, , 2 A x V (V - v) 60v AxV(V-o)t) 
 The horse-power = ^-1 = m 
 
 If, as before, E is the efficiency of screw and engine — 
 
 (C) LHR " 275E 
 
 Assuming A = -j— , or 07854 D 2 ; 
 
 and (V — v) -j- V as the fraction of V for the slip, say/, so that — 
 
 ^-=^=J; or(V-t>)=/V, 
 and v = V - / V, or V (1 -f). 
 
 Then, substituting these values in equation (c) — 
 
 I.H.P. = 0-7854 P2>< ^ VXV (1 ~f) 
 
 D 2 X V» (1 -/)/ 
 350 E 
 
 But V - P X B; then I.H.P. = P* * ff * W~/»). 
 
 ooO hi * 
 
 In actual practice there are disturbing causes which prevent the factor 
 from being so low as 350 ; for example, with very large bosses the column 
 of water is a hollow one, and the equivalent diameter is then less than D. 
 Moreover, the apparent slip / is less than the real or actual slip — that is, 
 less than the true measure of the acceleration, for there are wake and other 
 currents ; and consequently to know how large a real screw should be for 
 the needs in practice another factor is necessary ; hence 
 
 The Rule for the Diameter of Screw for good work and high efficiency 
 can be found by making the formula as follows : — 
 
 TV * * / I.H.P. / C V> /I.H.P. / C \ 3 
 
 D 1 ameterofscrew=^_ T] -^ X (p^^j,or^^-^x( F ^). 
 
 x = 0-2 P c +y, or an approximation to the real slip. 
 
PITCH OF THE SCREW. 483 
 
 For single-screw ships and the centre screw of triple screws, x = 0-18 P c -f /, 
 where/ is the apparent slip fraction. Then the value of C is 450. 
 
 Example (i.).— What should be the diameter of the screw of a tramp 
 steamer whose prismatic coefficient is 0-85, the I.H.P. 1,000, the speed is 
 10 knots, and the apparent slip to be 10 per cent. ? 
 
 Here x = 0-3 x 0-85 + 0-10 = 0-355. 
 
 10 X 6,080 
 ^ X K ~ 60 x 0-9 ~ 1,12b ' 
 
 D- / 1,000 J450\» 
 
 ~V 0-253 -0-063 Vl,126/ ~ * 
 
 Example (ii.). — Determine the diameter of the twin screw of a ship whose 
 P c is 0-6, the I.H.P. 6,000, and the speed 20 knots, the apparent slip to be 
 15 per cent. - 
 
 Here P x R = 2°_^080 + . 85 = %fj ^ 
 
 and x = 02 X 06 -f 0-15, or 027. 
 
 Diameter of screws = J^^Z x (_^) 8 = ifrl feet. 
 
 Example (iii.). — What should be the diameter of the screws of the triple- 
 screw turbine-driven scout, whose P c = 0-55 ; her speed is to be 27 knots 
 with 15,000 H.P., and the apparent slip 25 per cent. ? 
 
 In this case x = 0-18 X 0-55 -f- - 25, or 0-34 for the centre screw, 
 
 and x = 0-20 x 0*55 -+- 0-25, or 0-36 for each wing screw. 
 
 ^ -, 27x6,080 9ato . 
 PXR= 60 x0-75 ' Or3 ' 648feet - 
 
 - I 5^00 / 450 \ 3 a C1 
 
 Diameter of centre screw = ^ ^ _ ^ X f^^) = 651. 
 
 * • / °T000 / 450 \ 3 c . , 
 
 Diameter of wing screw = */ „ „„ „ , » X I o~cTq ) = "'* ieet - 
 
 Example (iv.). — A four-screw ship of 60,000 H.P. has a P c of 0-7; her 
 speed is 25 knots, and the slip per cent. 16, of what diameter are the screws ? 
 
 Here x = 0-2 X 0-7 + 0-16, or 0-30. 
 
 p v -r „ 25 X 6, 080 __ 
 
 P x R - Wxttl ~ 3 ' 017 * 
 
 ^. t I 15,000 / 450 V -.koi 
 
 Diameter of screw = w _ X ( jjQjy ) = I 5 " 8 feet « 
 
 If, however, the slip is to be only 12-5 per cent. — 
 
 x = 0-14 + 0-125 = 0-265. 
 
 Then the diameter is 17 feet. 
 
 The Pitch of the Screw depends on the speed of the ship and amount of 
 acceleration required to be given to the stream of water to produce the 
 
484 MANUAL OF MARINE ENGINEERING. 
 
 required thrust, or perhaps it should be said to produce the maximum amount 
 of thrust possible with the shaft horse-power transmitted to it. That is to 
 say, the pitch of screw is arrived at by dividing the sum of the speed of the 
 ship and the " slip " in feet per minute by the revolutions. 
 
 The amount of slip required will depend on the size of the column of water 
 projected, as in the case of the paddle-wheel (p. 464) — that is, as the disc 
 area is decreased so that the column is diminished in volume the acceleration 
 given it to produce the same thrust as before must be increased. Taking 
 the same notation as before for speed and velocity, and A as the area of 
 screw disc, its diameter D and its pitch P — 
 
 A = ?-=- = 07854 D 2 . 
 4 
 
 Then momentum or thrust = 2 A X V (V — v) lbs. 
 
 = 1-57 D 2 x V (V - v) lbs. 
 Taking (V — v) as/V (a fraction of V). Then 
 
 Thrust = 157 D 2 x/V 2 lbs. 
 As for given speed and size of ship, etc., thrust may be assumed constant, 
 then 
 
 /D 2 . V 2 = constant. 
 
 That is, D X V vary inversely as J/. 
 
 Thus, if the fraction of slip is increased, the value of (D . V) 2 decreases, 
 and as V = K X P, any variation in velocity V may be made by a variation 
 in pitch or revolution, or both. 
 
 . Pitch of Screw. — Lei S be the speed of the ship in knots, R the number 
 of revolutions per minute, and s the slip in knots ; then 
 
 (S+s)x 6,080 
 
 Pitch = 
 
 60 x R 
 
 If the slip is expressed as so much per cent, of the speed of screw, which 
 is the common way, and is x per cent. Then 
 
 S = speed of screw ( 1 — ^kf] )' 
 
 or Speed of screw =S-f ( 1- r^ 1 =Sx 10fi _ — 
 
 ^ . S X 100 x 6,080 S 10,133 
 Then Pitch = nnTt „ nn r = ™ X 
 
 60R(100-z) R 100- a; 
 
 Example. — To find the pitch of the screw for a ship whose speed is 15 knots, 
 the slip is 10 per cent., and the number of revolutions 60 per minute. 
 
 ^ , 15 10,133 __ ■ 
 
 Pitch = 6 q-X 10Q _ 1Q > or 28-14 feet. 
 
 The apparent slip of a well-proportioned screw on an ordinary cargo ship 
 of fairly good form should be about 8 per cent, when at sea full speed, and 
 10 per cent, when on trial full speed. High-speed ships have higher rates — 
 viz., from 12A to 16 — with reciprocators and 15 to 25 with turbines. The 
 
ACTING SURFACE OP SCREW BLADES. 485 
 
 large propellers of bluff cargo boats already alluded to, however, seldom 
 exceed 5 per cent., and to them the remark does not apply. An excessive 
 amount of slip does not of necessity imply waste of power, as it may arise 
 irom smallness of diameter ; and also when a screw of larger diameter gives 
 a better speed than that obtained by the original small one, it is often due 
 to the increased efficiency of machinery at a reduced number of revolutions. 
 
 An abnormally large amount of slip may, however, be due to want of area 
 of screw blade, and when this is the case the curve of indicated thrust exposes 
 it, as it will then exhibit want of proper augmentation of thrust at the higher- 
 speeds of revolution. 
 
 The pitch of a screw should never be less than the diameter if it can be 
 avoided, as in practice it seldom happens that such screws produce satis- 
 factory results, although, since the turbine has come into use, designers have 
 been compelled to supply screws whose pitch ratio is less than 1*0 to suit 
 the revolutions necessary to high efficiency of the motor. It really means 
 that the angle of blade at tip should not be less than that whose tangent is 
 0-32, or 18°. 
 
 Pitch Ratio is the ratio of the pitch of a screw to its diameter and an 
 expression in common use and convenient in algebraic expressions. It is 
 now sometimes as low as 075, but generally is over 1*0, and best results are 
 obtained with values of 1*2 to 1'6. Screws are seldom found now with such 
 high pitch ratios as once were common when high revolution was obtained 
 by gearing to the slow-running engine. Such high-pitch ratios were also 
 necessary and quite appropriate to vessels using large sail power with steam. 
 It was no easy matter with such ships to keep the engines from racing badly 
 — that is, from too rapid revolution during a squall. 
 
 Surface Ratio is also an equally popular and useful expression for the 
 ratio of the actual blade area to that of a disc of equal diameter. It is quite 
 as important that the surface of the blades shall be sufficient for the power 
 transmitted or the thrust to be produced as that the diameter and pitch are 
 right. It is claimed that by taking the total thrust and dividing it by the 
 area of the blades as projected on a plane transverse to the axis the quotient 
 should not exceed a certain figure depending on the immersion, otherwise 
 cavitation is liable to ensue. The figure in question is said to be from 14 lbs. 
 per square inch with shallow fast ships, such as scouts and destroyers, and 
 16 to 20 in the better immersed screws of the merchant ship. It is obvious 
 that, with the small diameter screws of turbine steamers the surface ratio 
 must be very much larger than that with similar ships of equal power with 
 screws of larger diameter, although the. actual surface of the latter may be 
 the greater. Surface ratio, therefore, cannot be taken as any guide, except 
 as a comparison with similarly driven ships. 
 
 The Acting Surface of the Blades of a Screw can be best determined by 
 referring to successful practice as the guide, and experiments with models 
 in tanks as a light. Prof. Taylor, U.S.A., as well as Dr. R. E. Froude, has 
 done yeoman service for engineers in making such elaborate, careful, and 
 exhaustive experiments, and publishing them, together with useful rules, 
 etc., for the service of designers, etc.* 
 
 Screws of the same diameter may vary in form of blade in amount of 
 
 * Which may be studied with advantage in Prof. Biles' book on the Design and 
 Construction of Ships, vol. ii., C. Griffin & Co 
 
486 MANUAL OF MARINE ENGINEERING. 
 
 acting surface and in pitch ; the first may be eliminated, inasmuch as modern 
 screws do not differ much in that respect, there remains their surface ratio 
 and pitch ratio as variants when making comparisons. With variations 
 in pitch, however, there must be differences of slip ratio, involving questions 
 of revolution. To examine the performance of a screw, there must be only 
 one variable at a time, and the test or criterion of performance is the ratio 
 of thrust to the torque producing it — that is, 
 
 B the efficiency = thrust X travel of ship -5- torque X 2 x x revolutions. 
 If, therefore, one screw may be examined by running it at different revolu- 
 tions, a base line of revolutions may be taken, and with ordinates of the 
 thrust and torque a pair of curves may be drawn, and with the various values 
 of E a curve of efficiency. Also, a curve of slip may be constructed ; or 
 with a varying slip ratio as a base line curves of power efficiency can be 
 constructed, and the results of the trials critically examined. The same 
 screw may be tested with pitch ratio as a variant, and the power constant 
 or the thrust constant and curves drawn with pitch ratios as the base or 
 abscissae. Another screw with a different surface ratio may be put through 
 a similar set of curves constructed. Any one of these latter may be compared 
 with the corresponding one of the former by means of their curves of effi- 
 ciency. In this way it is possible to select the screw which shall best suit 
 the conditions of the motor driving it by superimposing on the curves of 
 efficiency that of the motor at various rates of revolution, and combining 
 them so as to produce a new curve of general or combined efficiency. 
 
 Prof. Taylor made a long series of experiments with screws 16 inches 
 diameter, the pitch ratio varying from 04 to 1*5 in six steps, and the surface 
 ratio from (H207 to 05635 by five steps. He used as abscissae the slip per 
 cent., and constructed curves of efficiency and power for each of the pitch 
 ratios of a particular screw. He also experimented with screws having 
 two, three, and four blades, so that the surface ratio was varied by variation 
 in number of blades, as well as by the breadth of them. It will be under- 
 stood, then, that the number of such experiments was very great, especially 
 as many of them had to be repeated ; the figures involved enormous ; the 
 investigations of the results and their general application to practical engin- 
 eering is a matter of special study, which may be made of them in Prof. 
 Taylor's own book, or in that of Prof. Biles', which contains the chief part 
 of them. 
 
 Prof. Taylor's investigations, however, were not limited to pitch ratios, 
 slip ratios, and surface ratios ; he has tested screws of different transverse 
 sections, and shown that many of the old ideas respecting such were erroneous. 
 
 The Acting Surface of a Screw Propeller may be calculated by taking the 
 power developed per revolution, and a factor which is governed by the fine- 
 ness of the ship's water lines and the number of blades ; it is also dependent 
 on the position in the ship whether central or on the sides, as with twin screws. 
 
 yT XT p 
 ' — * . 
 
 Here K = P c X M, where 
 
 For four-bladed single screws, M = 20 ; for twin screws, 15*0. 
 ,, three-bladed „ M = 19 ; „ 143. 
 
 „ two-bladed „ M = 175 ; „ 13-1. 
 
ACTING SURFACE AREA OF A SCREW. 487 
 
 • 
 
 The Thrust of a Screw may be deduced with a fair approximation to 
 accuracy by the following formula, where A g is the acting surface ; P r the 
 pitch ratio ; D the diameter in feet ; V the velocity of screw in feet per 
 
 P x R 
 
 second, or — ^ — ; G a coefficient depending on the disposition of the 
 
 surface of the blades, and which varies in value from - 42 of the mercantile 
 ship to 0*5 of the broad-tipped small diameter screws with high-power 
 turbine steamers. 
 
 -. 4 . , DXn/A,xV 2 n 
 
 Thrust in pounds = ~ X G. 
 
 "t 
 
 For ordinary leaf -shaped blades, . • • G = 0*4:0. 
 
 ,, circular blades, . . . • . G = - 42. 
 
 „ broad-tipped, as with turbines, . . G = - 45 to 050. 
 
 ,, shaped as generally found in merchant ship, G = 0-42. 
 
 The Area of Acting Surface of a Screw may also be calculated by means 
 of this formula for thrust. 
 
 From the above the following is deduced : — 
 
 Area of acting surface = ( t) v V 2 v C ) ' 
 
 Taking the slip as a fraction of Y, so that it is/ X V, then — 
 
 Speed of ship v is Y — /V, or V (1 — /). 
 
 The efficiency of engine and propeller is, as before, E ; then — 
 
 mi ,. I.H.P.X 33,000 x E I.H.P.x550xE 
 Thrust in pounds = ^ = Y (1 — ) / * 
 
 Substituting this value of f— 
 
 . . . /I.H.P. X (550 X E) x P,\ 2 
 Area ot acting surface =< f ) v v3 m _ f\ r< f * 
 
 For the ordinary merchant cargo steamer, . 550 E = 330. 
 „ express and naval reciprocators, . 550 E = 360. 
 ,, turbine direct-driven ship, . . 550 E = 380. 
 
 Example (i.). — What should be the surface of each screw of a twin-screw 
 reciprocator ship of 6,000 I.H.P. ; R X P = 2,400 ; P r = 1'4 ; and slip 
 20 per cent. ; the diameter is 10 feet ? 
 
 360 X 3,000 X 1-4 . . 
 
 Area = 10 X 403 (1 _ . 2) Q. 45 = 42 ' 8 S * Uare feet - 
 
 Example (ii.). — \Yhat surface should the screw of a tramp steamer have 
 which has engines indicating 1,200 H.P. at 75 revolutions ; it has to have 
 four blades, and the prismatic coefficient is - 85 ? 
 
 yj 200 
 -^-— - = 68 square feet. 
 < o 
 
 Example (iii.). — A twin-screw steamer, whose prismatic coefficient is - 6, 
 
488 MANUAL OF MARINE ENGINEERING. 
 
 indicates 6,000 horse-power at 130 revolutions. The screws are to be three- 
 bladed. and for smooth-water service. 
 
 y3 000 
 T30~ = 41 2 ' 
 
 Example (iv.). — A turbine-driven steamer crossing the Atlantic develops 
 24,000 S.H.P. at 250 revolutions ; her prismatic coefficient is 0*66, and she 
 has three screws dividing the power evenly. 
 
 Here surface of blades = 0-66 X 14-3 x \/ ^q~ = 53 " 8 square feet. 
 
 Thickness Of Blade. — The strength of a blade at the root, or part near 
 the boss, will vary nearly directly as the breadth and as the square of the 
 thickness, and for practical purposes may be assumed to do so. The strain 
 on the propeller will vary as the cube of the diameter of the tunnel shafts, 
 and very nearly as the cube of the diameter of the propeller shaft. In 
 similar engines/ the twisting power of the engines has been shown to vary 
 as I.H.P. -*- revolutions. 
 
 Let d be the diameter of the screw shaft close to the propeller, b the 
 breadth of blade in inches at a distance of 1| X d from the centre, and n the 
 number of blades : R. the number of revolutions per minute. Then 
 
 Thickness of blade (at 1J X d from centre) = */ , X K, 
 
 / I.H.P. X / . 
 
 also „ „ „ •• . - V R x n x i x * > 
 
 f may be taken as 100 -for compound engines of the ordinary two-cylindei 
 type ; 90 for three-cylinder compound, four-crank triple, and quadruple 
 engines ; 85 for three-crank triple-compound engines, and 75 for turbines. 
 The value of K is 4 for cast iron, 2 for ordinary gun-metal, 1*6 for cast steel, 
 and 1 - 5 for forged steel and bronzes of superior make. For ships engaged in 
 the North Atlantic or other seas where rough weather largely prevails, these 
 factors may with advantage be increased by 20 to 25 per cent. This is 
 especially necessary for cargo vessels, which may often have to run " light." 
 
 The thickness of metal at the tip should be - 2 of that at the root. 
 
 Propellers made of steel and bronze are very superior in strength to those 
 of cast iron, even if made in accordance with these rules ; for if made of the 
 same strength only, the blades would not be stiff enough, and would vibrate 
 very considerably, especially when racing. 
 
 Fig. 173 is a typical design of solid cast-iron propeller. It is also typical 
 of the form of the blade of a loose-bladed propeller, and the proportions are 
 such as would be found generally with those having three blades. The 
 Admiralty, as well as many private engineers, prefer to have the blades 
 rounded at the top as shown, but blades in the past, as well as many of them 
 in the present, are shaped as shown in dotted lines at M. 
 
 The surface of propellers designed approximately to fig. 173 may be 
 found by the following rule : — 
 
 Aggregate surface = D x B X n ■*- F. 
 
 D being the diameter in feet, B the maximum breadth in feet, n the 
 number of blades, and F for rounded tips, 3, and for square ended, 285. 
 
489 
 
 Yi,r 173.— Solid Cast-iron Screw Propeller. 
 
490 MANUAL OF MARINE ENGINEERING. 
 
 The blade sections are obtained by taking a point E, so that 
 
 cl being the diameter of the screw shaft at the inner end or that of the tunnel 
 shafting, n the number of blades, /; their maximum breadth in inches. The 
 value of K 1 is, for cast iron, 6*3 ; for ordinary gun-metal or bronze, 32 ; 
 for cast steel, 2*5 ; and for bronzes of superior strength, 235. 
 
 Taking T as the thickness at|-th of the diameter, or l'od from the centre, 
 the thickness at f ths should be 065 T. at fths = 0-4 T, and at f-ths = 0-23 T. 
 
 Propeller Boss. — When for a loose bladed propeller, this is usually spherical 
 in general form, with flats or recesses for the blades. The diameter of the 
 sphere is from |- the diameter of the screw for small propellers, to i the diameter 
 for large ones, and may be taken at 
 
 = 0*9 \/diameter screw. 
 
 The length of the boss depends on the size of the base of the blades, and 
 is generally about - 85 the diameter of sphere for a two-bladed screw, and 
 0*75 when four-bladed. 
 
 The bosses of some very large screws have been made oval in fore and 
 aft section, and the base of the blades made oval to fit them. This form 
 is not so good for many reasons as the spherical, and was adopted more on 
 account of the boss being a forging than from any other cause. 
 
 In H.M. Navy, as well as in the navies of most countries, bronze of some 
 kind is the material employed for both boss and blades. The Admiralty used 
 to specify that it must be composed of 87*65 best selected copper, 8*32 tin, and 
 4*03 best Silesian spelter.* This mixture, when carefully made, should have 
 an ultimate tensile strength of 16 tons per square inch, and be very tou<rh. 
 Unless very carefully melted, however, test pieces from large castings seldom 
 exceed 14 tons. 
 
 The Admiralty have also used phosphor-bronze, Stone's and Parson's 
 manganese-bronze propellers, which are more suitable to the service under 
 present conditions. Bronze is a necessity with some naval ships, on account 
 of the copper on the bottoms and the brass elsewhere in the submerged parts 
 affecting cast iron or steel, but generally it is employed because its strength 
 is superior to that of cast iron, and steel is viewed with some considerable 
 amount of distrust, which has not been lessened by the action of some of 
 the larger steamship companies, who have now generally adopted bronzes in 
 lieu of steel. The zinc bronzes do not affect the hulls as tin bronzes did. 
 
 In the mercantile marine, the boss is generally of cast iron, of as strong 
 a mixture as can be made, or of cast steel. In large steamers of full power, 
 and all ships which make long voyages to distant parts, the bosses should 
 be of cast steel, whether the blades be of iron, steel, or bronze. 
 
 The studs for securing the blades to the boss are of bronze for bronze 
 blades, and sometimes for iron and steel blades, but for the latter best steel 
 is preferable, the nuts being of brass with closed ends to protect the studs. 
 
 The bosses of solid cast propellers are oval in fore and aft section as a 
 rule, and are from £ to £ the diameter of the screw in diameter, and about 
 2h to 2 £ times the diameter of the shaft in length. 
 
 * Of late years the Admiralty have decreased the amount of zinc (v. Chap. xxx.). 
 
SCREW PROPELLER BLADES. 
 
 491 
 
 Screw Propeller Blades. — The screws of cargo ships of the mercantile 
 marine are generally of cast iron, and np to moderate sizes are almost invari- 
 ably cast solid, although generally there is a decided tendency to adopt those 
 with loose blades. A solid screw is rather more efficient than one whose 
 blades are bolted on, and is only about half the cost of the latter. The cause 
 of inefficiency arises from the resistance of the projecting nuts, and often this 
 is added to by the clumsy flanges of the blades, which project beyond the 
 boss itself. A well-designed and carefully made screw should have the base 
 of the blade conforming to the general outline of the boss, and the nuts or 
 bolt heads recessed into the blade base, and covered in with a metal case or 
 
 174. — Screw Propeller Blades and Boss Bronze. 
 
 cement flush with the surface (fig. 174). The screws in H.M. Navy are 
 generally made in this way, and are then quite as efficient as a solid screw. 
 
 If one of the blades of a solid screw is broken off, the whole screw is 
 practically ruined,* and the expense of a new one thereby entailed ; also, 
 what is of more serious consequence, the ship must go into a dry dock, or on 
 a slip or " hard " to have it removed, which operation is long and tedious, as 
 the boss must be forced off, the tunnel-shaft taken down, and the screw-shaft 
 withdrawn before the new one can be fitted. On the other hand, if the 
 screw has its blades bolted on, and one is damaged, it can be removed by a 
 * A clever founder can burn-on new ends to blades both with iron and bronze. 
 
4.92 MANUAL OF MARINE ENGINEERING. 
 
 diver, and a new one fitted in its place at a very small expense, and in a very 
 short time ; and even when an experienced diver cannot be obtained, the 
 ship can be tipped by discharging cargo from aft, etc., or shifting it forward. 
 
 The forms of propeller blades are so numerous and varied as to be beyond 
 description here. The well-known blade (fig. 180) introduced by Mr. 
 Griffiths, and now bearing his name, was the one most generally adopted 
 by engineers, and generally gave satisfaction. Only a few carry out the 
 plan he usually so strongly recommended, of bending the blade slightly 
 forward, perhaps principally because it then came too near to the stern-post. 
 Some have tried blades bent in the reverse way, and satisfied themselves 
 that improved performance is obtained. The data published by inventors of 
 screws are generally very misleading, as the failures are never mentioned, 
 and seldom is it observable that an old screw has been replaced by a patent 
 one of the same diameter, pitch, and area. There is little doubt that the 
 better results with the " improved screw " are due rather to better propor- 
 tions than to the particular shape of blade. The number of patents relating 
 solely to the form of blade is endless, and some special ones, introducing 
 additions " to prevent loss from the centrifugal action," reappear periodically 
 as novelties. This latter is a great bugbear with many engineers ; for after 
 all, the loss from this cause is non-existent, while it always happens that 
 means adopted to prevent it are themselves causes of a considerable loss 
 from frictional resistance. 
 
 As a rule, the greatest breadth of blade should be beyond one-third of 
 the diameter of the disc from the centre, and should be approximately as 
 given by the following rule : — 
 
 Maximum breadth of blade in inches = K A / - t '"* -— . 
 
 evolutions 
 
 = K 3 / I-H. 
 
 V revolu 
 
 For a four-bladed screw, K = 14 ; for three-bladed, K = 17 ; and for 
 two-bladed, K = 22. For modern high-speed screws with turbines K is 
 higher. 
 
 The breadth of blade at the tip should be from £ to § the maximum. 
 
 Propellers are sometimes made with the blades of finer pitch near the boss 
 than at the tip, partly that the angle shall not be so coarse that this part of 
 the blade only churns the water, and partly that the hold on the boss due to 
 the increased breadth of the blade may be greater. A decrease of pitch of 
 10 per cent, has given good results, and when the propeller is of small dia- 
 meter, with a very coarse pitch, as much as 15 per cent, decrease may be 
 adopted ; as a rule, however, a true helix gives best propelling results, especi- 
 ally if the tail part of the sections is rounded off ship shape. 
 
 The Number of Blades is arbitrary ; the designer may have two, three, 
 four, and even six, for experiments both with models and real screws have 
 shown that the efficiency is .not seriously effected by the number. It is, 
 however, certain that in smooth water, and neglecting vibration as a factor, 
 the two- or even the one-bladed screw can show the highest efficiency, and 
 that three is better for twin-screw ships of all kinds ; but for rough water 
 with moderate immersion four blades are found to be generally most effective 
 in large single- and twin-screw ships. The efficiency of propellers with a 
 larger number of blades is not so good as with four, while the weight and 
 cost is greater. In practice the four-bladed screw is universally employed 
 
SCREW PROPELLER BLADES. 
 
 493 
 
 in the deep-water mercantile marine, and three-bladed ones in the Navy 
 and for smooth-water express steamers. 
 
 It remains for future experimental trials with turbine-driven ships to 
 decide whether the necessary surface is better distributed over five or even 
 six moderately wide blades than over three or four with abnormally wide 
 tips. The experiments made by the Admiralty with six-bladed screws in 
 
 Fig. 175. — Modern Fast-running Bronze Screw for Cruiser. 
 
 HtM.S. " Shannon " showed a decrease in efficiency of 5-6 per cent, with the 
 six as against lour blades ; but in that case the surface was proportional to 
 the number of blades, and, therefore, excessive with*the six. A six-bladed 
 screw in H.M.S. " Emerald," having a surface only 4 per cent, greater than 
 that of the four-bladed of the same diameter and pitch, gave an indicated 
 thrust of only G per cent, less than did the four-bladed. But the speed 
 
494 
 
 MANUAL OF MARINE ENGINEERING. 
 
SHAPE OF PROPELLER BLADES. 
 
 495 
 
 coefficient with the six-bladed screw was 175*1 against 171*7 with the four- 
 bladed, and the slip per cent. 11*26 against 12*28 ; the speed, however, was 
 12*003 with the four-bladed screw as against 11*726 with the six, and on 
 this the case appears to have been decided in favour of the former. 
 
 The Shape of Propeller Blades approximate more or less to that introduced 
 by Mr. Griffiths, and used in the Navy from 1860, as shown by the dotted 
 line in fig. 173. To-day naval screws, as well as a large number of those 
 
 Fig. 177. — Hirsch's Screw. 
 
 in the mercantile marine, are shaped as shown in the full lines of that figure, 
 but they generally are made with loose or fitted blades, as shown in fig. 174. 
 With the increase in rate of revolution in the full-powered high-speed twin- 
 screw express steamer and cruiser, the surface ratio has been largely increased, 
 so that the blades have assumed almost a circular form, as shown in fig. 175. 
 While the demand for surface for the very small diameter screws of the 
 turbine-driven ship has caused a further increase in the breadth of blade, 
 
496 MANUAL OF MARINE ENGINEERING. 
 
 so as to approximate to an ellipse with the minor axis radial, the surface 
 ratio in their case is very high, even as much as - 56 to O60. In the ordinary 
 cargo boat with its four-bladed screw of large diameter the surface ratio is 
 only about 0*37, and the blades generally retain the squareness of tip as 
 formerly obtained, inasmuch as the outer part of the blade is the most effec- 
 tive, especially when running in ballast trim. Fig. 176 is a good example of 
 such a screw with the blades inclined so as to be as far removed from the body 
 of ship as possible, and thereby work in better water. 
 
 Fig. 178.— One of the new Propellers of R.M.S. " Mauretania," 68,000 S.H.P. 
 
 (Of "Parsons" Special Turbadinm Alloy.) 
 
 There have been from time to time new forms of screw which have estab- 
 lished a reputation for a while, but not finally survived. One of the most 
 noteworthy of these, Ilirsch's Patent 60° Screw, as shown in fig. 177, was 
 a favourite one, both on the Continent and in Great Britain, and deservedly 
 so, inasmuch as it was quite successful in competition with even good screws, 
 
DIAMETER OF BLADE BOLTS. 497 
 
 and when replacing inferior ones marked improvement in propelling effi- 
 ciency was always obtained ; but it was not superior to the best designs 
 with the ordinary leaf-shaped blade, as in fig. 174, of true pitch. As a stern- 
 going propeller, Hirsch's screw was very successful, and at the time it was 
 tried vibration with it was less than with the screws then obtaining. 
 
 The Section of Screw Blades is usually an approximation to a segment of 
 a circle, as shown in fig. 175, but it is the fashion with some engineers, who 
 claim for their screws a higher efficiency thereby, to make them of a trans- 
 verse section, shown in fig. 176a, with either the hump-back or the rounded 
 one. * In both cases, what may be called the " forebody " is not so fine as the 
 after or following portion. Experiments by Prof. Taylor have proved what 
 had been for some time a surmise — viz., that the blade with a ship-shape 
 cross-section moves with the least resistance and highest efficiency. When 
 the stereotyped segmental form near the boss is departed from, and the 
 sections there become ship-shaped, there is always an improvement, especially 
 at high speed of revolution. The mere rounding of the tail or following 
 portion of the blade sections seems to influence the efficiency more than a 
 similar modification with the leading half does. The effect of such changes, 
 however, is virtually to decrease the pitch, especially near the root. 
 
 The Studs or Bolts of a Screw Blade should be of the strongest material, 
 especially in these days of high rate of revolution, when the centrifugal 
 forces which they have to resist are very high. Under ordinary working 
 conditions the bolts on the side of the acting face are in tension, as they 
 tend to keep the blade from tipping under the thrust on the surface ; those 
 on the back exert no action on them, unless the blades are simply attached 
 to the boss without being recessed, when they are subject to sheer from the 
 same pressure. But all the bolts help to hold the blade against centrifugal 
 
 force, which is expressed by — X — , where W is the weight in pounds of 
 
 the blade, v the velocity in feet per second of the centre of gravity whose 
 distance is r feet from the axis, and g, gravity = 32. 
 
 Wir 
 That is. Tension on the bolts due to C. Force = xz- • 
 
 oJi r 
 
 If R be the revolutions per minute, and pitch is P, then 
 
 v = 5 VP 2 + (2 * r)\ 
 
 The Diameter of the Blade Bolts, when made of high tensile bronze 
 or steel, whose number is n, can be found from the following : — 
 
 ~. , , , -i d X Z 
 Diameter of bolts or studs = , 
 
 H 
 
 where d is the diameter of the solid shaft of sufficient size for the torque, 
 
 and Z is a factor which, for a three-bladed screw, is 1*6, and for a four-bladed 
 
 1-3. The sizes obtained in this way are generally sufficient, but for high 
 
 rate of revolution should be checked by estimating the tension on them 
 
 due to centrifugal force, and allowing a stress not exceeding 4 tons per square 
 
 inch of section at bottom of thread. 
 
 32 
 
498 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Material for Screw Blades. — Cast iron is, of course, the cheapest material 
 for this purpose, and also possesses another advantage not so much appreci- 
 ated until experience teaches — viz., that when struck violently against an 
 obstacle like a jetty or wreckage, it breaks clean off, without, as a rule, 
 damaging the shaft ; on the other hand, steel or bronze blades, which are 
 strong enough and tough enough to resist fracture, are sometimes bent so 
 as to prevent the central screw from turning, and may cause the shaft to be 
 bent or even broken (v. fig. 179). Cast iron when used should be of the very 
 best description, twice cast and cooled slowly ; hematite, and even steel, 
 is often added to strengthen the mixture ; the former is now generally used 
 by moulders for this purpose, and has a tensile strength of 15 tons. 
 
 Fig. 179. — Damaged Bronze Screws. 
 
 Steel blades could be bought at about two and a-half times the cost of 
 cast-iron ones ; they are very much stronger, even when the section is con- 
 siderably reduced, and consequently were more reliable, especially for engines 
 of very large power, which must work at high speeds in rough weather. The 
 efficiency of the propeller would be somewhat increased in consequence of 
 the reduction in thickness, but steel blades were seldom true to pitch ; hence 
 when the best possible results are necessary blades of the strong bronzes 
 should be fitted with their surfaces ground smooth and made true to pitch. 
 
 The chief objection to steel, especially to cast steel, is its liability to rapid 
 corrosion on the backs of the blades. Cast iron corrodes into pits, but steel 
 goes much more rapidly, and in some instances the tips are " honeycombed " 
 
BEVIS' PATENT FEATHERING SCREW. 499 
 
 in a few weeks. Corrosion, however, is not unknown with some of the bronzes, 
 especially with screws of high revolution ; lurbadium, however, and some 
 other alloys are free from this defect as well as from erosion (v. Chap. xxx.). 
 
 Paint provides little or no protection, and even nickel plating, which has 
 been resorted to on steel, has not proved very efficacious. Muntz metal 
 sheathing has been tried, but with not very satisfactory results. 
 
 Phosphor-bronze, Parsons' special bronze, Bull's metal, Stone's bronze, 
 Delta metal, etc., have taken the place of steel in most quarters, notwith- 
 standing that the cost is about twelve times that of cast iron. Blades made 
 of these materials can be cast very thin — thinner, in most cases, than if made 
 of steel — because there is no loss of strength by corrosion ; in fact, the very 
 high speed of certain ships is attributed largely to the high efficiency of the 
 propeller from the thinness and truth of the blades. Fig. 178 is a notable 
 example of this in a huge solid screw. 
 
 Bronze propellers are objected to on the ground that injury is often done 
 to the iron work near them by galvanic action ; but the Admiralty does not 
 hesitate to continue the practice of fitting them, nor do the large steamship 
 companies. Corrosion does sometimes take place in such a way as to cause 
 the source of the evil to be attributed to bronze screws ; but the fitting of 
 some slabs of zinc in way of the propeller prevents it, and removes all fear of 
 ill consequences. The zinc bronzes are less liable to cause corrosion. 
 
 Weight of Screw Propellers may be calculated with a close approximation 
 to truth by the following rule : — 
 
 Surface x thickness 
 
 Weight in cwt. = 
 
 Ji. 
 
 The surface is taken in square feet, the thickness in inches at the root of 
 the blade. 
 
 K is for solid cast-iron screws, 4*5 
 ,, ,, bronze ,, 38 
 
 ,, built cast-iron ,, 3'0 
 „ „ cast-steel ,, 2*8 
 
 „ ,, bronze ,, 2 - 5 
 
 Feathering Screws. — Yachts and ships which are required to sail as well 
 as steam cannot well do the former when the screw is stopped, unless some 
 means be adopted of feathering the blades, so that they are nearly in a fore 
 and aft plane, or else by withdrawing the propeller altogether from the water. 
 The late Bennett Woodcroft patented, in 1844, a plan for feathering the 
 blades, which in a modified form was fitted by Messrs. Maudslay, Sons & 
 Field to several ships. The blades, of which there are two, have shanks 
 fitting into the boss, to which short levers are secured inside the boss ; these 
 levers are connected by links to a sliding collar outside the boss, which is 
 carried round with the shaft, but is capable of being moved " fore and aft" 
 on it by means of a pair of bell-crank levers actuated by a screw from on 
 deck. When it was desired to sail, the blades were moved round into the 
 fore and aft position by sliding the collar from the boss. 
 
 Bevis' Patent Feathering Screw. — Many patents were taken out for 
 methods of effecting a similar movement of the blades, without the objec- 
 tionable feature of external bell cranks, and to the late Mr. Bevis is due the 
 
51)0 
 
 MANUAL OF MARINE ENGINEERING. 
 
 perfecting of this idea. In 1858 Gregory and Craymer patented a feathering 
 screw of which " the screw propeller shaft is made hollow, and a second shaft 
 goes through it, carrying worm threads, which act on the propeller blades, so 
 as to feather them to the angle desirable." In 1866 H. B. Young patented a 
 
 TZT 
 
 somewhat similar idea as to the hollow shaft, but says " levers may be attached 
 to the shanks." Fig. 180 shows the Bevis plan, which needs no description, 
 and answers its purpose admirably. 
 
 The feathering screw can be converted into a reversing one by moving 
 the blades through a greater angle, and, although their efficiency in that 
 
LIFTING SCREWS. 501 
 
 ■state will be very low, it will be good enough for casual astern going with a 
 non-reversible oil engine of moderate power. It would be quite a satisfactory 
 way of reversing small craft, and it will avoid the use of wheel gearing. * 
 
 Lifting Screws. — The screws of warships were formerly nearly always 
 made and fitted, so that when desired they could be raised to the level of the 
 deck for examination and repair when necessary, and to prevent obstruction 
 when sailing. This plan is a very costly one, and not so efficient as the 
 feathering blade, inasmuch as the ship steers badly owing to the gap in the 
 •deadwood, but it admits of examination and repair, which is of the utmost 
 importance in a warship. A hole, however, is also necessary through the 
 stern to admit of the screw coming on deck, which very much weakens what 
 is already a somewhat weak part of the hull. 
 
 The lifting screw has a short piece of shaft cast with the boss or fitted to 
 it in the usual way ; the forward end of this shaft is provided with a driving 
 piece, which is so formed as to fit into a slot across the cheese coupling keyed 
 -to the outer end of the stern shaft. The propeller is carried in a frame, called 
 the banjo frame, which is arranged to slide up and down in grooved metal 
 guides secured to the stern and rudder posts, and supported, when in working 
 position, by two strong brackets or chairs also secured to these posts, and 
 held down by two strong wooden sampson posts, whose upper ends are fitted 
 with jam screws to the metal guides ; such an arrangement is made wholly of 
 gun-metaL , _ . 
 
 Number Of Screws on a ship depends on various circumstances. Multi- 
 plication arose first with ships of light draught and good speed, inasmuch 
 .as before Yarrow and Thornycroft's discoveries it was the rule to have the 
 screw immersed when at rest. Twin screws were found to have the advantage 
 of safety and power to manoeuvre. A screw at each end was the invention of 
 James Howden for tugs, to provide for one being immersed when pitching, as 
 such short boats do when towing. Three screws were used in foreign naval 
 ships to permit of more elastic subdivision of power than obtains with 
 twins. Three screivs also were found an advantage with turbine direct-driven 
 ships of large power, and with greater power still four scenes were necessary 
 for the draught of water possible, and in naval ships for the subdivision of 
 power desirable when cruising and manoeuvring. With the reciprocators 
 and turbine combination three screws obtain. 
 
 * This method of reversing is coming into use with ships driven by oil engines of 
 moderate power. The screw blades are, of course, formed so as to be fairly efficient when 
 in "astern gear." 
 
5G2 MANUAL OF MARINE ENGINEERING. 
 
 CHAPTER XIX. 
 
 SEA-COCKS AND VALVES. 
 
 The importance of having efficient connections to the skin of the ship for 
 inlet and outlet purpose has long been recognised, since neglect in the design 
 or working of these fittings has caused damage, and sometimes the loss of a 
 ship. Next to absolute safety, simplicity should be most aimed at in all 
 parts of the machinery of a ship ; but in no place, perhaps, is there greater 
 need of careful consideration than in the arranging and designing of the sea 
 and bilge fittings, so that neither a careless nor an ignorant engineer can 
 endanger the safety of the ship ; it must be "fool proof." 
 
 Lloyd's Committee have issued special rules on this sub] ct, and the 
 Board of Trade are no less vigilant in endeavouring to eliminate every 
 source of danger, as indeed are all the other Register Societies. 
 
 It is needless here to give any illustration of how easily a ship may be 
 flooded by an ill-considered arrangement of sea-cocks in ignorant hands, nor 
 to give examples of ships that have foundered, for every engineer has heard 
 of them. But even good arrangements may cause mischief, if, from want of 
 simplicity, they are ill understood. 
 
 In the first place, the cocks and valves themselves should be of simple 
 construction, and carefully marked, so as to be easily seen whether they are 
 open or shut. The parts exposed should be either strongly made or carefully 
 guarded, so that they are not liable to injury or derangement ; and, when 
 possible, so designed that they can be packed and examined without having 
 to dock the ship. It is also necessary that the very large openings should 
 have a second valve, as a safeguard in case of accident to the first one. 
 
 Sea-cocks and valves should be so placed that they are accessible at all 
 times ; and, as far as possible, within sight from the working platforms. 
 
 Kingston Valve. — For all large inlets the Kingston valve is preferable, as 
 it acts as a non-return valve in case of the spindle breaking, and can then 
 always be worked by simply forcing it outwards, either with the spindle 
 itself, or by a rod substituted for it. 
 
 The Kingston valve is usually made in the form of a frustrum of a cone, 
 with a taper of about 8 in 12; the length of the seat is generally about 
 
 1 inch a =-= . The object of this form of seat is to allow the valve to 
 
 close itself tight in place, in case of the spindle breaking, by the pressure of 
 the water. The Admiralty require that these valves shall be made in one 
 with their spindles, of best gun-metal, and the spindle tested by a load of 
 half a ton for each square inch of section of valve. For example, a valve 
 4 inches diameter, having an area of 12-5G square inches, should have a 
 spindle sufficiently large to be tested to 6-|- tons. 
 
 Hence, if the proof stress of good gun-metal be taken at 6 tons (which is 
 quite high enough) per square inch, then, 
 
KINGSTON VALVES. 
 
 503 
 
 Diameter of Kingston valve spindle at bottom 1 
 of thread, . . I 
 
 diameter of valve 
 3^6 
 
 so that a 4-inch valve should have a spindle 1-15 inch diameter at its smallest 
 section. 
 
 But the Admiralty do not require any valve to be tested above 12 tons ; 
 so that for mere test purposes, no spindle need have more than 2 square 
 inches in section, or be more than If inch diameter at its smallest part; but 
 since for very large valves a spindle of this size would not be stiff enough, 
 the following rule for all valves above 5| inches diameter holds good : — 
 
 diameter in inches — 5^ inches. 
 T6 
 
 Diameter of spindle = If inch + 
 
 Kingston valve-boxes and 
 tubes are generally made of 
 gun-metal ; but this is not a 
 necessity, except where hot 
 water or steam is blown 
 through them, when cast iron 
 would be dangerous. If the 
 body is of cast iron, of course 
 the valve, seat should be of 
 brass, and the working parts 
 bushed with brass. 
 
 Fig. 181 shows a good 
 arrangement of Kingston 
 valve for large sizes. It has 
 a liftiug nut secured in a 
 bridge, as well as a handle on 
 the spindle end; the former 
 is used to start the valve or 
 jam it in its seat, and the latter 
 merely to open or shut it ; the 
 lock nut with handle is to 
 secure it in any required posi- 
 tion. In this figure the head 
 is shown screwed on to the 
 tube, and is such as was neces- 
 sary in wooden or composite 
 ships ; but when fitted to the 
 skin of an iron ship, a flange 
 is formed at the bottom of 
 the conical part, as shown by 
 dotted lines and in Fig. 182 
 (A, B, C, D). A much simpler 
 and less expensive plan is to 
 form the valve like an ordinary 
 stop-valve with four wings, and 
 the spindle inverted — that is, 
 on the same side as the wings ; 
 the bottom part of the box is then only slightly conical, anil much shorter 
 —in fact, only sufficiently long to allow of the valve lifting a distance equal 
 
 Fig. 181.— Kingston Valve (in a Wooden Ship). 
 
504 
 
 MANUAL OF MARINE ENGINEERING. 
 
 to one quarter of its diameter, and in this position leaving space between 
 it and the grating for the free flow of water. 
 
 All inlet valves should be fitted with a brass grating, whose meshes 
 should not exceed half an inch in breadth, the total area through then: 
 being at least 20 per cent, larger than the net area of the tube. In the 
 Navy, Kingston valves are fitted to all inlets and blow-off pipes, and are 
 always supplemented with a cock or valve attached to them. 
 
 
 it! 
 
 Tvbcofh 
 
 Fb/yed luff-pieeer-\ 
 riveted to tube*) 
 
 Grating\~\ 
 
 Outer 
 'tto/a 
 
 Zowerpart 
 of Kingston 
 
 1). 
 Fig. 182.— Details of Inlet Valves (in Steel Sh'ips). 
 
 Valves are fast taking the place of cocks for all general purposes, so thafe 
 whenever convenient a valve may be fitted in lieu of a cock to even the 
 smallest Kingston, but to large Kingstons a cock could not be fitted, and 
 so it is usual to find an ordinary sluice valve. The spindles of these 
 supplementary valves should always be within easy reach of the platforms, 
 and in case of the very large ones, when possible, they should be carried so 
 high as to admit of the valve being worked when the engine-room is flooded. 
 
 In the merchant service, whether valves or cocks are fitted to the skin 
 of the ship direct, the same precautions should be taken as enumerated 
 above, and extra care taken to protect them f»om injury, as they are very 
 
DISCHARGE VALVES. 505 
 
 liable to damage from the flooring plates washing about when a consider- 
 able quantity of water is in the bilge and the ship rolling heavily. 
 
 For the larger inlets of a merchant ship an ordinary stop -valve, opening 
 inwards, is usually fitted, the box being of cast iron. A good plan, so as to 
 (avoid a number of openings in the ship's skin, is to fit a single box with one 
 I stop- valve in it, and that by preference an inverted one, and to this box 
 fix the various cocks or valves necessary for the different requirements. 
 For ships destined for a cold climate, a small cock can be fitted near the 
 bottom of the box to admit steam to thaw any ice that may be blocking 
 the orifice, it is also useful to blow away seaweed, fish, etc., which may be 
 choking the meshes, and may be fitted with advantage to every ship. 
 
 Discharge Valves should be fitted to all outlets through the ship's side, 
 and they consist, as a rule, of simple non-return valves having spindles 
 passing through the covers, so that the valves may be lifted or pressed down 
 as required. Here again a number of holes through the skin of the ship 
 may be avoided by connecting the smaller valves to the box of a large one 
 above the valve itself. In many cases this method is a very good one, 
 having advantages beyond that already named — viz., that the pipes can be 
 shorter, and the valves standing clear of the frames of the ship are easy of 
 access. 
 
 The Board of Trade* require that all inlet and discharge valves shall 
 be connected directly to the skin of the ship, and have no pipe or joint inter- 
 vening (this rule, however, does not apply to the case of the smaller valves 
 when attached to the box of a large one), and also that all discharge valves 
 shall be placed above the load water-line. 
 
 The Admiralty, on the other hand, for obvious reasons, require all the 
 discharge orifices to be below the water-line ; and while in the merchant 
 service the discharge valve-boxes are nearly always of cast iron, in the Navy 
 they are invariably of bronze. 
 
 It is not unusual to fit the smaller discharge valves as simple non-return 
 valves having no external spindle whatever ; but this has the disadvantage 
 that one is never certain if the valve is shut, and, if not shut, without the 
 spindle it is beyond control. 
 
 It is sometimes found convenient to fit, for large discharges, a straigld- 
 ■ through valve in lieu of the ordinary mushroom-valve. This straight- 
 through valve is an ordinary flap valve resting on a vertical or nearly vertical 
 seat, and lifted by means of a horizontal spindle, which passes through a 
 stuffing in the side of the valve-box, having inside a slotted lever connected 
 by a pin to a pair of lugs on the back of the valve, and, outside, a hand lever 
 for controlling the valve. 
 
 * The Board of Trade regulations referring to sea connections are as follows : — 
 
 137. All inlets or outlets in the bottom of side of a vessel near to, at, or below the 
 deep-load water line other than the outlets of water-closets, soil, scupper, lavatory, and 
 urinal pipes, should have cocks or valves fitted between the pipes and the ship's side or 
 bottom ; such cocks or valves should be attached to the skin of the ship, and be so 
 arranged that they can be easily and expeditiously opened or closed at any time ; and 
 the cocks, valves, and the whole length of the pipes should be accessible at all times. 
 
 Cocks or valves standing exceptional distances from the ship's plating, that is where 
 the necks are longer than is necessary for making the joint, should not be passed without 
 the sanction of the Board of Trade, and one condition of their being passed is that they 
 should be made of gun-metal and well bracketed. 
 
506 MANUAL OF MARINE ENGINEERING. 
 
 Bilge Valves. — As has been already said, too much care cannot be devoted 
 to avoid all risk of flooding the ship by carelessness. An easy method of 
 overcoming the difficulty in the case of pipes not greater than 4 inches dia- 
 meter, is by means of the switch or three-way cock — that is, a cock having 
 two side branches and a hollow plug with one orifice in the side and the 
 other in the bottom, so that all water must pass by way of the bottom and 
 one side only at one time. Suppose that an auxiliary engine is required to 
 draw water from the bilge as well as from the sea, a three-way cock should 
 be fitted so that its bottom is connected to the pump suction, one side branch 
 connected to the sea inlet, and the other side branch to the bilge piping ; 
 it will then be easily seen that the two branches cannot be connected by 
 the plug, and consequently, however careless the engineer may be, water 
 cannot pass from the sea to the bilges. When, owing to the largeness of 
 the piping, a cock cannot be fitted, self-closing non-return valves should be 
 connected to the tail pipes leading to the bilge. 
 
 Communication Boxes. — All the bilge suctions, and all ballast-tank suctions, 
 should be led to a communication box placed in some convenient position in 
 the engine-room, above the level of the flooring, so that the requisite changes 
 of service of pumping may be easily and quickly effected. 
 
 The communication box (called sometimes a " directing " box) consists 
 of a rectangular chest having as many valves in it as there are pumping 
 stations or parts from which water is to be drawn ; under each valve is a 
 separate tail, to which the suction pipes are attached ; the valves themselves 
 are of the ordinary mushroom type, non-return, and arranged so that each 
 may be shut close by screwing down a spindle on it. The box cover is fitted 
 by hinge bolts, so as to be easily and quickly opened for examination, and 
 the joint made again. 
 
 To the upper part of the box the suction pipes of the pumps are fitted, 
 so that the auxiliary pumping engines and engine bilge-pumps may draw 
 from the same set of pipes. Both bilge-pumps should not be connected 
 direct to one box, as it not unfrequently happens that it is required to pump 
 water separately from two compartments at the same time, or that one of 
 the bilge-pumps may be required when no sanitary pump is fitted to delivei 
 on deck. The lid of the communication box should have inscribed legibly 
 on it the lead of each valve. 
 
 Bilge Suction Piping is sometimes of cast iron and sometimes of lead, but 
 the tails leading directly into the bilge should be of iron ; cast iron is now 
 often used instead of lead as less liable to damage, but it cannot be so easily 
 cut in case of necessity. The Admiralty require all tail pipes to be of 
 galvanised wrought iron ; and all copper piping laid in the bilges to be 
 covered with waterproof canvas varnished over, and otherwise insulated to 
 avoid deleterious effects on the ironwork from galvanic action. All suction 
 pipes to the bilge should fit into rose-boxes of iron galvanised, and having 
 an aggregate area through the holes at least twice that of the pipe. The 
 cover of the box should be hinged, or so fitted that it can be opened and 
 shut again very quickly. The boxes should also be so placed as to be easy 
 of access. Bilge pump suction pipes should also be fitted with mud traps, 
 consisting simply of a cast-iron box with a strainer and lid, in which the 
 slight cessation of flow allows the deposit of the heavy dirt carried on so far 
 with the water. All copper or brass-work subject to the action of bilge- 
 
EXPANSION JOINTS. 507 
 
 water, should have the connections made with bronze bolts and nuts. Iron 
 quickly corrodes unless well protected ; and Muntz metal has been found 
 to decay in a very peculiar way. Naval brass, however, which is Muntz 
 metal improved with a little tin, is suitable for this purpose. Every care 
 should be taken to protect the metallic surfaces from the corrosive action 
 of bilge-water ; for this purpose a good coat of Portland cement wash answers 
 even better than paint or varnish. 
 
 When it is required to pass pipes, especially copper ones, through a water- 
 tight bulkhead, a good plan is to fit a short length of cast brass, having a 
 flange at each end to connect to the pipes on either side, and a collar in the 
 middle, about 4 inches larger in diameter than the flanges, to secure it to 
 the bulkhead. 
 
 If pipes are likely to expand considerably, or to be in such a position 
 that the working of the ship in a sea-way would affect them, it is better to 
 pass them through stuffing-boxes in the bulkhead. Very small pipes can 
 be connected by union nuts to the bulkhead joint above described, thereby 
 decreasing the size of the hole to be cut and the collar for closing it. 
 
 The Board of Trade require that one of the bilge-pumps shall be so fitted 
 that it can pump water on deck in case of fire. 
 
 The engine-room pumps should have the necessary piping, valves, etc., 
 that water may be drawn from each hold or compartment separately, from 
 each side of the engine and boiler rooms, and from any other part of the ship 
 liable to leakage or lodgment of water. 
 
 In addition to the above-mentioned means of freeing the ship from water, 
 it is usual to have a bilge suction, so that the circulating pump may be utilised. 
 The chief danger to be apprehended when the circulating pump draws from 
 the bilge is, that in cases of flooding of the boiler compartment, the small 
 coal is washed out of the bunkers, and soon chokes the condenser tubes as 
 well as the pumps themselves, unless the rose boxes have sufficiently small 
 holes to prevent it : and when the rose-box holes are so small as to prevent 
 small coal passing through them, they soon choke, and render the pumps 
 useless for the time. With a centrifugal circulating pump there is no danger 
 of the pump choking or getting out of order, and by having a direct discharge 
 overboard, the condenser cannot get choked. 
 
 Water Service. — An arrangement of water service is provided in every ship, 
 so that water can be applied to all bearings and slides, and also that a hose 
 may be used in case of fire or any other emergency. This generally consists 
 of a main pipe from a sea-cock leading into a convenient position, and having 
 branches with a stand-pipe and brackets at each main bearing, crank-pin, 
 and group of eccentrics. The water for the tunnel-bearings and thrust- 
 bearing is usually taken from the inside of the stern-tube, thereby serving 
 the additional purpose of circulation in the tube. 
 
 Expansion Joints. — Great care should be taken that all pipes conveying 
 hot water or steam, and thereby liable to great change of temperature, 
 should have provision made for the consequent expansion. Expansion 
 joints of all kinds are, however, objectionable on one or two grounds, and 
 therefore should be avoided unless absolutely necessary. If small pipes 
 have bends between the rigid connections, the expansion is allowed for by 
 the variation in their curvature. If the rigid connections are far away, care 
 should be taken to prevent the pipe from assuming full curvature to the 
 
508 MANUAL OF MARINE ENGINEERING. 
 
 circle, and thereby tear away from the flanges ; but in all large pipes, and 
 wherever the bends are but very slight, expansion joints should be fitted, 
 for very serious accidents have occurred in consequence of their absence. 
 When the length of piping to undergo expansion is not great, the ordinary 
 bellows joint is preferable, as being free from any liability to leak ; but when 
 the expansion is very great an ordinary stuffing-box and gland, or Fawcett 
 joint, must be resorted to.* This latter should be avoided for exhaust pipes to 
 a condenser, as they are very liable in this case to leakage which is not easy 
 to detect. A thick india-rubber jointing ring will generally serve this purpose 
 with eduction pipes, which in the case of large ones may be 2 inches thick. 
 
 Again, in ships of light construction, which undergo a considerable amount 
 of racking in a sea-way, no length of pipes should be without some provision 
 for expansion and contraction, and all extra rigidity should be avoided, 
 that they meet the changes in a ship. Great care should be taken to provide 
 means for resisting the thrust of pipes which are subject to internal pressure 
 or heat, otherwise the joints are severely strained, and the connection of 
 pipe to flange and of valve-box to boiler in danger of destruction. This is 
 the more necessaiy when bends are expected to do duty for expansion- joints, 
 etc. Loose pipes should be always carefully " anchored." 
 
 Safety Collars. — It has sometimes happened that pipes have blown out of 
 a Fawcett joint, owing to there being a bend near it on which the steam 
 pressure acted ; in all cases of this kind a collar should be brazed to the 
 pipe, not less than 18 inches from the bend, through which bolts pass con- 
 necting it to the flange of the stuffing-box on the other pipe. 
 
 Flanges. — Few things look worse in an engine-room than heavy broad 
 flanges to the copper pipes, especially when they are quite unnecessary. 
 The breadth of the flange need not exceed three times the diameter of the 
 bolts through it, and its thickness should be proportional to the pitch of the 
 bolts. The pitch of the bolts should be from four to six times their diameter, 
 depending on the steam pressure and thickness of flange. 
 
 Table xlviii. has been drawn up as generally indicating the best thickness 
 of pipes for various pressures and purposes. 
 
 In the merchant service the main steam pipes, when of copper, are usually 
 from No. 1 to 4 B.W.G. thick, depending on their diameter and the pressure 
 to which they are exposed. 
 
 Both in the Navy and often in the mercantile marine all steam pipes 
 over 2 inches diameter are of steel ; in the former it is solid cold-drawn, 
 in the latter hot-drawn, or welded with riveted cover-straps for large sizes. 
 
 Board of Trade Rules for Pipes. 
 
 j.17. The working pressure of well-made copper pipes when the joints 
 are brazed is found by the following formula : — 
 
 6,000 X (T - &) . . 
 
 -^ S2.C = working pressure ; 
 
 T = thickness in inches ; 
 
 D = inside diameter in inches. 
 
 • These arc not always satisfactory, for if the packing is tight enough to prevent leakage of high- 
 pressure steam it grips the pipe and prevents sliding. It is found now that a spigot end carefully fitted 
 into a smooth bored socket and made so thin that pressure tightens it is much more satisfactory. (Vtde 
 fig. 182a, p. 510.) 
 
BOARD OF TRADE RULES FOR PTPES. 
 
 5C9 
 
 When the pipes are solid drawn and not over l"0 inches diameter, substitute 
 in the foregoing formula -g\ for ^. 
 
 118. The internal pressure on steel pipes made of good material, which 
 are lap- welded and are a sound job, may be dete mined by the following 
 formula, provided that the thickness is not less than J-inch : — 
 
 T = thickness in inches ; 
 
 D = diameter inside in inches ; 
 
 6,000 X T 
 
 — = working pressure. 
 
 TABLE XLVIIL— Thickness of Copper Pipes (L.S.G.). 
 
 & 
 
 O ! 
 
 a - 
 « .3 
 
 22 
 
 21 
 
 20 
 
 19 
 
 18 
 
 17 
 
 16 
 
 15 
 
 14 
 
 13 
 
 12 
 
 11 
 
 10 
 
 9 
 
 8 
 
 7 
 
 6 
 
 5 
 
 4 
 
 3 
 
 2 
 
 1 
 
 7/0 
 
 6/0 
 
 5/0 
 
 4/0 
 
 3/0 
 
 2/0 
 
 1 
 
 2 
 
 3 
 
 5 
 
 6 
 
 S 
 
 10 
 
 13 
 
 Steam Pipes. 
 
 Boiler Pressures in Lbs. 
 
 180 
 
 155 
 
 125 
 
 85 
 
 
 
 
 
 
 • • • 
 
 1 
 
 ... 
 
 ... 
 
 ... 
 
 1 
 2 
 
 • • * 
 
 
 4/0 
 
 2 
 
 
 6/6 
 
 3/0 
 
 3 
 
 7/0 
 
 5/0 
 
 2/0 
 
 3 
 
 6/0 
 
 4/0 
 
 
 
 4 
 
 5/0 
 
 3/0 
 
 
 
 i 
 
 4/0 
 
 2/0 
 
 1 
 
 5 
 
 3/0 
 
 
 
 2 
 
 5 
 
 2/0 
 
 1 
 
 3 
 
 6 
 
 
 
 1 
 
 3 
 
 6 
 
 1 
 
 2 
 
 4 
 
 7 
 
 2 
 
 3 
 
 5 
 
 8 
 
 3 
 
 4 
 
 6 
 
 8 
 
 4 
 
 5 
 
 7 
 
 9 
 
 6 
 
 6 
 
 8 
 
 9 
 
 7 
 
 7 
 
 9 
 
 10 
 
 9 
 
 9 
 
 10 
 
 11 
 
 11 
 
 11 
 
 11 
 
 12 
 
 13 
 
 13 
 
 13 
 
 13 
 
 50 
 
 7 
 
 7 
 
 7 
 
 8 
 
 8 
 
 8 
 
 8 
 
 8 
 
 9 
 
 9 
 
 9 
 
 9 
 
 10 
 
 10 
 
 10 
 
 11 
 
 11 
 
 11 
 
 12 
 
 12 
 
 12 
 
 13 
 
 3 
 X 
 
 2 A 
 
 ■< 
 
 11 
 11 
 11 
 11 
 11 
 11 
 
 12 
 12 
 12 
 12 
 12 
 32 
 12 
 13 
 
 a 
 £ 
 S 
 
 <o 
 an 
 v 
 
 13 
 13 
 14 
 14 
 14 
 14 
 14 
 15 
 15 
 15 
 15 
 15 
 15 
 15 
 15 
 15 
 15 
 15 
 
 a 
 
 3 
 
 4 
 
 4 
 5 
 5 
 5 
 6 
 6 
 7 
 7 
 7 
 8 
 8 
 9 
 9 
 9 
 
 10 
 10 
 11 
 11 
 12 
 12 
 13 
 
 a 
 .2 
 
 BflJ 
 
 a 
 
 9 
 
 10 
 10 
 11 
 11 
 12 
 13 
 
 
 
 ■o S 
 
 
 
 c o 
 
 
 o 
 
 a! 2 
 
 
 Ri 
 
 - '-< 
 
 
 (k 
 
 .25 
 
 
 O w 
 
 43 _ 
 
 
 
 5 E 
 
 
 e" 
 
 - — 
 
 
 
 
 sa.S 
 
 rf -j- 
 
 
 Q 
 
 s 
 
 35 
 
 3 
 
 
 33 
 
 3 
 
 
 32 
 
 4 
 
 
 29 
 
 4 
 
 
 28 
 
 5 
 
 
 25 
 
 5 
 
 
 24 
 
 6 
 
 
 21 
 
 6 
 
 
 20 
 
 7 
 
 
 17 
 
 7 
 
 
 16 
 
 8 
 
 
 15 
 
 8 
 
 
 14 
 
 9 
 
 
 13 
 
 9 
 
 
 12 
 
 10 
 
 
 11 
 
 10 
 
 
 10 
 
 11 
 
 
 9 
 
 11 
 
 
 8 
 
 12 
 
 
 7 
 
 12 
 
 
 6 
 
 13 
 
 
 5 
 
 13 
 
 Diameter of pipe in inches, 
 
 Thickness, L.S.G., 
 
 Blow-off and Scum Pipes. 
 
 10 
 
 li 
 
 10 
 
 H 
 
 10 
 
 If 
 
 2 
 9 
 
 21 
 
 8 
 
 24 
 
 Feed discharge pipes to be as steam pipes for 30 per cent, higher pressure ; but in 
 no case to be taken lower than 125 lbs. 
 
 Receiver pipes to be as steam pipes for half the test pressure of the cylinder to which 
 they lead steam ; but in no case to be taken lower than 50 lbs. 
 
 The above gauges refer to straight pipes only ; bends to be suitably strengthened. 
 
510 
 
 MANUAL OF MARINE ENGINEERING. 
 
 119. Feed pipes should be made sufficient for a pressure 20 per cent, in 
 excess of the boiler pressure. 
 
 The Bureau Veritas Rules for steam pipes are — 
 
 D x P 
 
 < 1 ) If solid drawn copper, t = ^ - -f- 0"05 inch. 
 
 D x P 
 
 . (2) Steel pipes = ; but no pipe to be less than /„ -inch thick. 
 
 Lloyds Register Rule for wrought-iron and steel pipes is — 
 
 w i- ptf- 0-125) 
 Working pressure = C ! =r -• 
 
 t is the thickness, and D the internal diameter in inches. 
 C = 9,000 for lap-welded and 12,000 for seamless pipes. 
 
 They are to be tested to three times the working pressure. 
 
 The Admiralty specifications allow of steel pipes being uged very freely 
 for steam pipes, being solid drawn up to 15 inches, and the larger ones solid 
 drawn or welded and fitted with a riveted cover strap on the weld outside. 
 The flanges are generally riveted on to the large pipes. 
 
 In the merchant service cast iron was sometimes used for large pipes 
 conveying water and exhaust steam, and occasionally for the main steam 
 and feed pipes, and where cost and stiffness are of more consideration than 
 weight, they may be used with advantage, for although there seems an 
 element of danger in them, experience proved them to be reliable and safe 
 with even comparatively high pressures of steam (100 lbs.). With the higher 
 steam pressures, however, this material has been dropped for steam and feed 
 pipes, and copper, solid drawn steel, or welded wrought iron substituted. The 
 flanges are sometimes riveted and sometimes brazed or screwed. Some pipes 
 are made with flanges in one piece by an ingenious system of press. 
 
 The joints of all pipes should be in such positions as to be easy of access. 
 
 
 ^^w 
 
 JS^L 
 
 te 
 
 ZS22 
 
 Fig. 182a. — Steam and Feed Pipe Expansion Joint. 
 
AUXILIARY MACHINERY. 511 
 
 CHAPTER XX. 
 
 AUXILIARY MACHINERY. 
 
 In the Early Days of Marine Engineering there was practically no auxiliary 
 
 machinery whatever, and until quite modern times it has been restricted 
 to such moderate limitations as to require little or no comment, and the 
 demand from such as there was for steam to work it so small as to need no 
 addition to the boiler capacity. Thirty-five years ago the merchant ship 
 had a donkey pump which served almost every purpose, and it was only 
 in large ships that a second general purposes pump was found, besides the 
 ballast pump in ships having water ballast. On deck there were the steam 
 winches, a steam steering gear, and generally a steam windlass in all but 
 quite small ships. In passenger ships there was an electric light engine and 
 sometimes ventilating fans. To-day the auxiliary machinery is both numerous 
 and powerful, so that the demand for steam to supply them all requires con- 
 siderable additions to the boiler installations, and the consumption of fuel in 
 such ships is an imperfect indication of the economic quality of the main 
 engines. 
 
 Auxiliary Machinery on Shipboard to-day requires a knowledge to deal 
 with it and a capacity to take care of it quite special, and its supervision is 
 as important as, and rather more exacting than, that of the main engines. 
 The marine engineer in charge of it must have a technical knowledge suffi- 
 ciently good and general to enable him to overhaul, adjust, or repair all such 
 mechanisms and fittings, but it is not necessary that he should be able to 
 design and construct them, that is the metier of the specialists engaged 
 in their manufacture. 
 
 There are Three Distinct Classes of Auxiliary Machinery: — 
 
 (1) Those things which are absolutely necessary to and are practically 
 part of the main engines, such as — 
 
 The centrifugal circulating pump and its engine. 
 
 The special air-pumps and engines or motors when separate from 
 
 the main engine. 
 The feed-pumps and auxiliary or supplementary feed-pumps. 
 The steam reversing and stop valve gears. 
 
 (2) Those machines which are incidental to the machinery whereby the 
 working is rendered more efficient and economical, such as — 
 
 The steam-turning gear, and any winches or hoists in the engine 
 
 room. 
 Ventilating fans and forced-draught fans. 
 Steam ash hoists. 
 
 Electric light and power engines and dynamos. 
 Steam gear for operating fire-bars, etc. 
 
512 MANUAL OF MARINE ENGINEERING. 
 
 (3) Tho6e engines and machines which are incidental to and requisite 
 for the ship, such as — 
 
 The steam steering-gear. 
 
 Steam winches. 
 
 Steam or hydraulic cranes. 
 
 Steam windlass and capstans. 
 
 Baggage and boat hoists. 
 
 Refrigerating machinery and plant. 
 
 Hydraulic engine and pumps for hydraulic cranes and hoists. 
 
 Bilge pumps and ballast pumps. 
 
 Ventilating fans. 
 
 Sanitary pumps. 
 
 (a) On Warships there are in addition to most of the preceding — 
 Air compressing machines for torpedoes, etc. 
 Ammunition hoists. Coaling machinery. 
 
 Gun training machinery. Boat hoisting winches. 
 
 (6) Where Oil Fuel is used there must be, in addition — 
 Pumps for dealing with the storing of the fuel. 
 Pumps for supplying it to the boilers. 
 
 In addition to all these things there are other fittings which may come 
 under the category of auxiliary machinery, as — 
 Feed filters. 
 
 Feed heaters. Lubricating oil pumps. 
 
 Feed distillers and fresh drinking water distillers. 
 Auxiliary and winch condensers. Auxiliary boilers. 
 
 The consumption of steam by the auxiliaries in a warship when under 
 working conditions is enormous, for in a ship such as the second-class 
 cruiser " Diana," with ordinary cylindrical boilers, in 1899, it was at 
 the rate of 4,544 lbs. per hour when cruising at slow speed, while at high 
 speed (i full power) it was as much as 16,384 lbs. per hour, or 12*5 per cent, 
 of all the steam generated, and at full power it was 10 per cent. The waste 
 in this ship necessitated the distillation of 1,792 lbs. of water per hour to make 
 it up. For the ship's use for domestic purposes 560 lbs. per hour were required, 
 so that the distillation of 2,352 lbs. of water per hour has to go on regularly 
 to keep such a ship going when working at high speed. 
 
 The consumption of steam for auxiliary machinery on a modern express 
 turbine-driven steamer is enormous, and when on service with a full comple- 
 ment of passengers the amount required for heating and distilling fresh 
 water for washing and domestic purposes is very large. Mr. Bell has shown 
 that on the " Lusitania " during her trials, and, therefore, practically without 
 passengers, when developing her full sea power of 65,000 S.H.P., there were 
 pumped to the boilers no less than 998,000 lbs. of water per hour, and he 
 found as follows : — 
 
 The main turbine engines used 851,500 lbs. of steam, or 13'] lbs. per S.H.P. hour. 
 The auxiliary machinery, 114,000 „ — 175 „ 
 
 For evaporating in distilleries, 32,500 ,, — 0-50 „ 
 
 Total consumption, 998,000 „ - 15-35 
 
AUXILIARY MACHINERY. 
 
 513 
 
 It will be seen from this that of the steam generated by the boilers, the 
 auxiliaries used no less than 11-42 per cent., and the main engines propelling 
 the ship 85-32 per cent. The following Table xlix. shows the consumption of 
 steam fuel, etc., on the "Lusitania" on service conditions at different speeds, 
 and is interesting as illustrating one of the economic arguments in favour of 
 high speed for the Atlantic service. Mr. Morison reckons that the demands 
 of the auxiliaries in the engine department (including steering gear) of an 
 ordinary merchant ship amount to lh per cent, of the total steam production, 
 and, therefore, demanding more attention than has been hitherto bestowed 
 on them in that class of ship. 
 
 TABLE XLIX. — Steam and Fuel Consumptions of R.M.S. "Lusitania" 
 
 RUNNING ON SEA-SERVICE CONDITIONS AT VARIOUS SPEEDS. 
 
 Speed of ship in knots per hour, 
 
 15-77 
 
 18-00 
 
 21-00 
 
 23-00 
 
 25-40 
 
 Shaft horse-power, .... 
 
 13,400 
 
 20,500 
 
 33,000 
 
 48,000 
 
 68,850 
 
 Consumption of steam per hour, total, 
 
 
 
 
 
 
 lbs. 
 
 284,500 
 
 253,000 
 
 493,300 
 
 668,300 
 
 879,500 
 
 ,, „ „ auxiliaries, lbs. 
 
 71,000 
 
 76,400 
 
 85,700 
 
 96,700 
 
 116,500 
 
 Consumption of auxiliaries per cent, of 
 
 
 
 
 
 
 total, ...... 
 
 25-0 
 
 21-6 
 
 17-4 
 
 14-51 
 
 13-2 
 
 Steam consumption per S.H.P. hour, 
 
 
 
 
 
 
 turbines, .... lbs. 
 
 21-23 
 
 17-24 
 
 14-91 
 
 13-92 
 
 12-77 
 
 Steam consumption per S.H.P. hour, 
 
 
 
 
 
 
 auxiliaries, .... lbs. 
 
 5-30 
 
 3-72 
 
 2-60 
 
 201 
 
 1-69 
 
 Steam consumption per S.H.P. hour, 
 
 
 
 
 
 
 total, ..... lbs. 
 
 26-53 
 
 20-96 
 
 17-51 
 
 15-93 
 
 14-46 
 
 Coal consumed per S.H.P. per hour, 
 
 
 
 
 
 
 total, ..... lbs. 
 
 2-52 
 
 201 
 
 1-68 
 
 1-56 
 
 1-43 
 
 Total consumption of fuel on voyage, 
 
 
 
 
 
 
 3,100 miles, -. . . tons 
 
 2,980 
 
 3,190 
 
 3,670 
 
 4,520 
 
 5,390 
 
 It will be seen by the above that, taking auxiliaries into account, as well 
 as the increase in consumption of the main engines at slow speeds, a horse- 
 power at 25 - 4 knots costs only 57 per cent, of that at 15*77 knots and 71 per 
 cent, of that at 18 knots. If domestic demands be taken account of the 
 difference is still greater, so thao the S.H.P. then is considerably less than 
 57 per cent, in cost. 
 
 In the ordinary passenger steamer the demand for fresh water for engine 
 requirements and domestic purposes is always great, and in ships trading in 
 cold climates steam heating is resorted to, which is another drag on the boilers. 
 
 It will be understood, then, that in the modern passenger steamer and 
 warship the boilers must be considerably in excess of the requirements of 
 the main engines, even to the extent of 10 to 15 per cent., and in small 
 express ships probably greater. 
 
 The Exhaust Steam from all auxiliaries should be retained in the system, 
 and be delivered to a feed heater or else to a special condenser, where it is 
 converted into feed-water for the boilers. When the exhaust is delivered 
 to the main condenser, it is better that it shall not be done direct, so that 
 all auxiliaries may then be worked with a back pressure above instead of 
 
 33 
 
514 MANUAL OP MARINE ENGINEERING. 
 
 below that of the atmosphere, for in the latter case the vacuum may be 
 seriously affected by the air leaks from winches and such like things as are 
 not immediately under th^ engineering staff's notice, and liable to be in 
 want of adjustment of glands ; it is better, therefore, that these work with a 
 back pressure above atmospheric pressure, and their exhaust enter the L.P. 
 cylinder valve box of the main engine, or the L.P. turbine if so fitted. 
 
 Mr. D. Morison has shown that in an ordinary cargo steamer conveying 
 passengers, and having engines of 5,000 I.H.P., whereas the main engines 
 alone consumed 70,000 lbs. of steam per hour, the auxiliaries consumed 
 8.650 lbs., or 12*36 per cent, of the main engine demand, and 11 '0 per cent. 
 of the total supply to machinery, and was distributed among them roughly 
 as follows : — 
 
 Consumed by the evaporators to make up waste, 
 ,, steam-steering engine, . 
 
 ,, feed-pumps, 
 
 ,, centrifugal circulating pumps, 
 
 „ fan engines on boilers, etc., 
 
 ,, electric light engine, 
 
 2,800 lbs. 
 1,000 „ 
 
 950 „ 
 1.800 „ 
 1.300 „ 
 
 800 „ 
 
 In an ordinary cargo steamer of 1,500 I.H.P. and a consumption in the 
 main engines of 21,000 lbs., the auxiliaries require 1,660 lbs. per hour, or 
 7*9 per cent, of that of the main engines, and 7 32 per cent, of the total 
 generated. 
 
 Under these circumstances it is quite requisite that means be adopted 
 whereby the demand for these necessary parasites be most carefully con- 
 sidered and reduced to as small an amount as possible ; but what is perhaps 
 the better policy, since the supply must be always large, is to make the best 
 use of the heat rejected by these engines. Mr. Morison points out that 
 by exhausting to the condenser 80 per cent, of the available heat is lost, 
 and even if the exhaust steam from them is passed through the L.P. cylinder 
 of the main engines the loss is almost as great, whereas if it is used to heat 
 the feed-water there is no practical loss, and even with a very high vacuum 
 in the main condenser, and the consequent low temperature of the hot well, 
 the feed-water, as delivered to the boilers, will be quite hot. 
 
 It is estimated that from such sources in the 5,000 I.H.P. ship as much as 
 llf millions of thermal units are available for the purpose, and that thereby 
 the water as coming from the condenser having a vacuum of 28*25 inches, 
 with a temperature of 87° F., only will be raised to 211° F. after mixing 
 and before delivery to the boiler. Further, that in the cargo steamer of 
 1,500 I.H.P. the available heat is nearly 2| millions B.T.U., and the feed- 
 water will be raised from 87° to 175° F. by mixing before delivery to the 
 boilers. But seeing that the ordinary feed-pumps on shipboard can draw 
 water only when the temperature is below 170°, special means must, of 
 course, be provided for dealing with water so much warmer. This, however, 
 is quite a simple matter, and inexpensive considering the great* gain to be 
 made thereby. 
 
 Some of the Auxiliaries work constantly at a fairly uniform rate, and 
 when the main engines are working at full speed these auxiliaries also proceed 
 at nearly full power ; consequently they may have a compound arrangement 
 of cylinders or other means of using steam at a fairly high rate of expansion. 
 
ELECTRIC LIGHT ENGINES. 515 
 
 and provision for overload when occasion demands more than their normal 
 full output permits of their cylinders being comparatively small. The cir- 
 culating pump, for example, might be always driven by such an engine ; 
 likewise the air-pump, feed-pump, etc., draught and ventilating fans, and 
 such like machines. 
 
 Other Auxiliaries work intermittently and at full speed only for short 
 periods ; they may require then to give a big output compared with that 
 under normal conditions. The steam-steering gear, for example, may stand 
 for hours with hardly a movement, and during the day on the high seas in 
 fine weath.er do very little work altogether, and at no one time develop more 
 than 10 per cent, of its maximum power. For such work the small cylinder 
 with the late cut-off answers quite well. For the windlass, capstans, etc., 
 it is not worth while, for the short time they work, to have any complica- 
 tions for the sake of economy. The winches and cranes are abominably 
 extravagant with steam, and seeing the number of hours they work in many 
 cargo steamers, it is worth while studying means for a reduction. Attempts 
 have been made from time to time by supplying them with compound cylinders; 
 but the wear and tear on them is so great, and in the past winches are not 
 heeded very much so long as they will somehow do the work required by the 
 stevedore, that small encouragement has been given to these elementary 
 efforts. Still, it is quite worth the while of the shipowner to have a much 
 better sort of steam or electric winch, to work with less than half the steam 
 and half the wear and tear ; but such an instrument cannot be turned out 
 at the exceedingly low price of the present winches. 
 
 The Electric Light Engines also are important enough to have bestowed 
 on them more consideration than has obtained often in the past. They, too, 
 have intermittent work, but they are running every day that the ship is 
 on service, and during the early hours of the night the demand on them 
 in passenger ships is very heavy. There is every reason for their having 
 compound cylinders, that a fair economic rate of expansion may be obtained 
 from steam of the pressure of the boilers ; in fact, in large ships some few 
 of these engines might be triple compound with advantage, or else what 
 is much the same thing, two-stage compound with a L.P. turbine connected 
 to the exhaust. In most ships electric engines must be run all day for one 
 purpose or another, inasmuch as so far the storage batteries are not a success ; 
 they are heavy, costly, liable to get out of order, and the acid fumes from 
 them a nuisance. It is also imperative that in passenger ships there must 
 be provision for such a contingency as a break-down or stoppage ; nowadays 
 the enclosed forced lubricated engine (figs. 188, 189) can be relied on to 
 keep going for an indefinite period without stopping, and the wear and tear 
 on them is a negligible amount ; but the dynamo is still sensitive to water 
 and water vapour. There should be, therefore, a stand-by unit in every 
 ship,* or the full requirements divided into two or more units, rather than 
 have one engine and dynamo only. In large passenger steamers the full 
 requirements at night time necessitate the running of all but one or two units, 
 which are held in reserve ; as the lights are put out the other units are one 
 after the other stopped, until the load is easily carried by those left running. 
 In a comparatively small steamer there should be two engines and two 
 dynamos, each pair separate and capable of generating 60 to 70 per cent, 
 of the maximum demand, so that the two are run together only for a short 
 
 * By preference an oil engine placed well above water to be useful in case the boiler-rooms are 
 flooded. 
 
516 MANUAL OF MARINE ENGINEERING. 
 
 time, and in ca3e of a break-down with one the other can practically keep 
 the ship well lighted. For a larger ship there should be three units, each 
 capable of doing 55 to 60 per cent, of the maximum, so that only two are 
 required at one time, the third being quite in reserve, and one only during 
 the clay, the early and late night. The largest should have four units, each 
 capable of generating 40 per cent., and only three running at full demand. 
 Some very large and important ships have a still greater number of units, 
 and keep two in reserve. It may seem extravagant to have such a large 
 subdivision, but the thermal efficiency of these small engines at full load is 
 much higher than that of larger ones at loads less than the maximum. The 
 mechanical losses also are less in the smaller engines at the same revolutions. 
 In fact, it has to be borne in mind that electric light engines run at constant 
 rate of revolution whatever the load may be, and so instead of an engine 
 whose maximum power is 400 at 450 revolutions per minute doing, say, 
 100 B.H.P., there is one of only 200 H.P. doing the 100 B.H.P. with less 
 expenditure of fuel and oil, and as it may be running all day, the difference 
 in expenditure is very appreciable. It would be better still to have, say, 
 three units, each of 133 H.P., to do the 100 B.H.P. output, for a margin of 
 33 per cent, would be ample, and as the fall might be at times to 50 I.H.P. r 
 of which 20 only would be B.H.P. or useful power, the cost would be very 
 great compared with that of the 133 B.H.P. engine, the B.H.P. of which for 
 50 I. H.P. would be probably not less than 39. 
 
 The current on shipboard is much lower than is now common for house 
 lighting on shore, being generally only 100 volts at the switchboard, whereas 
 in London 210 is common. 
 
 The Working of all Auxiliaries by Electric Current is, no doubt, the ideal 
 thing, for thereby much that is highly objectionable in the present ordinary 
 arrangement is avoided, and all that is desirable and worth having is accom- 
 plished. The power is then generated by a highly efficient and economic 
 engine placed under the immediate observation and control of the skilled 
 and responsible engineers on watch ; the energy is transmitted by means 
 of cables, which may be led throughout the ship in the easiest way possible, 
 and when lead-covered and otherwise protected from harm are both safe 
 and efficient transmitters. The motors are simple, light, and easily fixed 
 machines, which can be placed in any position in quite a contracted space, 
 and protected, water-tight, etc. Their efficiency is much higher than that of 
 any of the steam engines employed for similar purposes on shipboard, and 
 they require little or no attention from skilled or other attendants. Winches, 
 capstans, windlasses, steering gears, pumps, etc., can all be worked by motors, 
 and avoid the transmission of steam more or less wet to distant parts of 
 the ship through pipes which, even when well lagged, leak an appreciable 
 amount of heat, and are always a nuisance to designers, ship officers, and 
 all having to do with them. Their exhausts are a worse nuisance still, for 
 if overboard the ship's deck is often smothered in steam, if to the main 
 condenser the vacuum is often affected adversely, and the only thing i* 
 to increase the back pressure, and thereby reduce their efficiency, or conduct 
 their exhaust steam to a separate condenser with much loss of energy thereby. 
 Tbere is another objection to steam for deck auxiliaries in winter time — that 
 is, that the small amount of leakage past the stop valves permits of some 
 condensation in the deck pipes, and the water so formed may in severe cold 
 
FEED-PUMPS. 
 
 517 
 
 weather freeze and prevent the working of the gear at an inopportune 
 time. 
 
 On the other hand, an electric installation is a somewhat expensive one, 
 and it is urged by some that it is dangerous from cut-outs fusing and from 
 short-circuiting due to water. This is all quite true as a liability of the 
 system, but if proper care is taken in the selection of materials, and in getting 
 the installation carried out by experienced men, the risk of any serious 
 damage arising from it is now only small. 
 
 There is One Cause of Serious Loss of Steam which is often overlooked, 
 and that is the use of it in whistles and syrens to give notice to other steamers, 
 etc. In such a ship as the " Lusitania " the whistle consumption as a per- 
 centage of the total steam generated is, of course, very small, but considering 
 
 Fig. 183.— Double Ram Flywheel Feed-Pump, 
 
 a tramp steamer with a generating power less than 2 per cent, of the " Lusi- 
 tania " has to make the same noise, and probably oftener, during the twentv- 
 four hours, the loss to her must be considerable and serious. It would seem, 
 therefore, that compressed air either in bottles or stored in a receiver from 
 a pump driven by the main engines would answer the purpose even better 
 than wet steam, and cost much less. Compare the expenditure of air by a 
 bugle in producing a noise which can be heard as far and as clearly as many 
 ships' whistle.-- with the steam required for such of them as evidenced 
 by the cloud flowing away at each blast. It must be remembered also 
 that all day long steam is being condensed in the whistle pipe, and if the 
 resultant water is not blown out, to the damage of the funnel paint and deck 
 
518 
 
 MANUAL OF MARINE ENGINEERING. 
 
 fittings, it runs back minus its latent heat. Steam whistles, therefore,, 
 waste much heat and much water, both of which cost the shipowner money. 
 
 The Feed-Pumps, both main and auxiliary, are now in naval and express 
 steamships worked independently of the main engines, but in cargo steamers 
 the feed- and bilge-pumps are still worked f om the air-pump crosshead of 
 
 Fig. 184. — External Packed Pumps (Duplex). 
 
 reciprocating engines. There are two kinds of such pumps, the flywheel 
 type, in which the pump is driven direct from the steam cylinder piston, 
 but controlled by a flywheel, etc., as seen in fig 183, and the direct-driven 
 without flywheel, as shown in figs. 184 and 185. The flywheel pump is not 
 
FEED-PUMPS. 
 
 519 
 
 3o much fancied now as formerly ; it is, however, a good working one, and 
 always makes its full stroke; the pump rams are packed externally, and, 
 therefore, kept in a good state of tightness ; the cut-off in the steam cylinder 
 
 Fig. 185.— Section of a Weir Pump. 
 
 may be comparatively early, whereas with the others it is very late, and 
 the" clearance very great compared with the flywheel one. On the other 
 hand, such pumps as shown can be worked "dead slow" or at high speed 
 
020 
 
 MANUAL OP MARINE ENG1.MKKRING. 
 
 with equal ease, and be quite automatic and under control. For quite 
 small snips, where there is only one, or at most two, auxiliary pumps for 
 general purposes, the duplex flywheel pump (fig. 183) is a very good and 
 
 Fig. 180. — Weir's D. A. Feed Pumps, with Float Tank and Automatic Gear. 
 
 reliable one. For larger sizes, and especially where the feeding is all done 
 by independent pumps, the type invented and successfully developed by 
 Mr. James Weir (fig. 185) is in general use. Fig. 18G shows a pair of such 
 
ELECTRIC LIGHT AND POWER ENGINES. 521 
 
 pumps fitted with a " float ' chamber and gear, so that their operation is 
 under complete control. By this arrangement the pumps are slowed down 
 on the supply of water from the hot well failing and stopping altogether 
 before the supply is so low as to risk the indraught of a : r. So long as the 
 water is supplied to them from the hot well or reserve tanks they will continue 
 to deliver to the boilers, and should the water tenders in the stokeholes 
 check the supply, the pumps will slow down under the increased 
 resistance. 
 
 Fig. 184 is the vertical variety of the well-known and popular duplex 
 pump introduced by the Worthington Company many years ago, and 
 improved in this country, as shown by the arrangement whereby the plungers 
 are packed externally. Such pumps as these have the merit of cheapness, 
 and when pressures are not high the ordinary piston (pump) type work very 
 well, but with tightly packed pistons or plungers there is a tendency to fail 
 making the full stroke every time. The Weir, like the Worthington types of 
 pumps, are somewhat wasteful of steam, but have high mechanical efficiency 
 as a set off ; moreover, they require little or no oil and few repairs, so that 
 altogether the Weir pump, even when without compound cylinders, is a 
 favourite with sea-going engineers who appreciate their reliability to continue 
 doing steady duty with no attention, when the demand on their time is for 
 other and more important things. In fact, all auxiliaries of this kind should 
 be automatic and reliable, and require no watching at all, but only the casual 
 attention from the greasers on their rounds. 
 
 Eng ne-room Pumps are nowadays very numerous, as may be seen on 
 the plan of R.M.S. " Olympic." There are, besides these feed and auxiliary 
 feeds, other similar pumps to deliver water on every deck in case of fire, 
 and for daily use in washing them ; pumps to supply closet and other domestic 
 tanks, baths, boilers, etc. Pumps to keep the bilges of the engine-room 
 free from dirty water, oil, etc. ; to draw from the bilges of every hold and 
 compartment ; to draw from reserve fresh-water tanks, and from ballast 
 and trimming tanks ; and most of these many pumps must necessarily 
 be in duplicate. 
 
 Steam Reversing Gears are dealt with elsewhere, and those gears now 
 used for operating the massive stop and pass valves of a turbine installa- 
 tion are in all essentials like the Brown reversing pull-and-push engine 
 (fig. 266). 
 
 Ventilat ng and Forced-draught Fans are now in very general use, and 
 are often driven by electric motors in preference to a steam engine, as they 
 are often in out of the way, inaccessible places, and require stopping and 
 starting by unskilled attendants. (For Table, see p. 522.) 
 
 Electric Light and Power Engines are now fitted to almost all ships, 
 inasmuch as the light obtained thereby is better than that of any oil lamps, it 
 is cheaper too, and when the installation is carried out in a proper way it is 
 much safer and cleaner. In the passenger steamer and warship electric power 
 transmission, as well as lighting, is of great advantage. It is, however, highly 
 desirable that such engines as are employed for this purpose shall be fairly 
 economic in steam, but. above all things, capable of running continuously 
 without variation in speed or trouble of any kind day after day, night after 
 night, without any attention from the staff beyond a casual visit. For 
 this purpose the compound-enclosed self-lubricating engines patented and 
 
522 
 
 MANUAL OF MARINE ENGINEERING. 
 
 introduced by Bellis & Morcom have proved most satisfactory, and continue 
 in demand. Fig. 187 shows in section the two-crank compound engine as 
 first brought out by that firm, having an ingenious arrangement of central 
 valve, which serves both cylinders, and answers. admirably for engines of 
 moderate size ; their improved engine, as shown in figs. 189 and 189a, is, 
 however, a better one for larger sizes, and is generally more economic of 
 steam. Fig. 188 is a tandem compound single-crank form of engine occupying 
 little space, and used largely in destroyers and small merchant steamers. 
 It is a good steady worker, and gives much satisfaction to those having 
 charge of them. The Admiralty have tried heavy oil engines for driving 
 dynamos in port when steam is not available" ; they may be of the Diesel 
 or semi-Diesel type ; they are somewhat noisy, and unless care is taken 
 when using certain oils the exhaust smells rather badly. Turbines have also 
 been tried for electric light, genera ting in certain ships, but the high revolu- 
 tions of such as have been used have proved trying to the dynamos, and, 
 therefore, not altogether satisfactory ; moreover, the small turbine is never 
 so economic in steam as the small compound reciprocator, and it is not 
 until a power of at least 1,000 horse that the turbine equals the reciprocator, 
 and further, whereas the latter remains fairly economical when exhausting 
 to the air, the turbine does not. The space occupied by a turbine is so small 
 and the head room so much less than required for a vertical reciprocator 
 that in naval ships, as also in some express steamers, turbines are of advan- 
 tage. Fig. 190 shows the combination of turbine and dynamo specially 
 designed by Belliss & Morcom for ship lighting purposes. 
 
 TABLE L.— Particulars of Admiralty Type of Fans and 
 Engines for Forced Draught. 
 
 Diameter 
 
 of 
 Impeller. 
 
 Width 
 
 of 
 
 Impeller. 
 
 Revolu- 
 tions per 
 Minute. 
 
 Cubic Feet 
 
 of Air 
 per Minute. 
 
 Air 
 
 Pressure 
 W.G. 
 
 Effective 
 
 HP. 
 required 
 to Drive. 
 
 Size, of Engine at 
 180 libs. Pressure. 
 
 Approximate 
 
 Weight 
 
 without 
 
 Fan Case. 
 
 Diameter 
 
 of 
 Cyiinder. 
 
 Stroke. 
 
 Ft. Ins. 
 2 
 
 2 6 
 
 3 
 
 3 6 
 
 4 
 
 4 6 
 
 5 
 
 5 6 
 
 6 
 
 6 6 
 
 7 
 
 7 6 
 
 8 
 
 Inches. 
 
 H 
 
 4 
 
 4i 
 
 5 
 
 6 
 
 7 
 
 8 
 
 9 
 10 
 11 
 12 
 15 
 18 
 
 950 
 800 
 650 
 550 
 500 
 450 
 400 
 350 
 320 
 300 
 275 
 250 
 250 
 
 4,500 
 6,500 
 8,800 
 11,000 
 15,000 
 20,000 
 26,000 
 32,000 
 39,000 
 46,000 
 55,000 
 74,000 
 95,000 
 
 Inches 
 2 
 2 
 2 
 2 
 2 
 o 
 2 
 2 
 2 
 2 
 2 
 2 
 
 2 
 
 3 
 
 4 
 
 6 
 
 7 
 
 9 
 
 12 
 
 14 
 
 18 
 
 22 
 
 26 
 
 30 
 
 40 
 
 50 
 
 24 
 
 2J 
 3* 
 
 34 
 
 44 
 
 ±4 
 
 5 
 
 5 
 
 6 
 6 
 
 7 
 7 
 8 
 
 24 
 2| 
 
 3 
 3 
 3i 
 
 H 
 
 4 
 4 
 5 
 5 
 6 
 6 
 7 
 
 G'wts. 
 
 n . 
 
 34 
 
 H 
 *4 
 
 6" 
 
 6* 
 
 8 
 
 8* 
 11* 
 Hi 
 loh 
 16 
 194 
 
 Some care should be always exercised in placing these high-speed engines 
 that there can be no lodgment of water from condensation, priming, or other 
 
ELECTRIC LIGHT AND roWER ENGINES. 
 
 523 
 
 so 
 
 O 
 
 0) 
 
 s 
 
 u 
 
 o 
 
 .3 
 
 a 
 
 o 
 
 H 
 
 o 
 O 
 
 Qi 
 oo 
 O 
 "o ' 
 
 S 
 
 
 
 t> 
 
 > 
 
 u 
 
 c 
 o 
 
 o 
 u 
 
 o 
 
 • to 
 
 P2 
 
 CO 
 
 bo 
 
 cause either previous to or during the working; and equal care taken that 
 no water can enter the L.P. cylinder from the exhaust pipe. In the past it 
 has been only too often the custom to suppose that an auxiliary can be placed 
 
Fig. 1SS. — Tandem Compound High-spjeed 
 Electric Light Engine. 
 
 Fig. 1S9. — Compound Two-crank 
 Electric Light Engine. (Belliss 
 & Morcom's V Type.) 
 
Fig. 189a. — Naval Electric Generating Engine. 
 Self-lubricating Two-crank Compound Inclined Valves. (Belliss & Morcom.) 
 
 Fif. 190.— Turbine-driven Naval Electric Generating Set, 200/3UO kw. 
 
526 
 
 MANUAL OF MARINE ENGINEERING. 
 
 in any convenient corner, and while this may be true of a slow working pump, 
 it is not so of a high-speed engine. 
 
 Fig. 191. — Navy Air Compressor (Belliss). 
 
 The Hydraulic Pumps and Accumulators are fitted in those ships where 
 ■energy is transmitted by water to cranes, capstans, and winches, and generally 
 
AIR COMPRESSOR. 
 
 527 
 
 placed in or near the engine-room. In certain trades such means are much 
 better served by water than by steam, and the working is much steadier 
 and safer. In the tropics and sub-tropical regions, where there is no danger 
 of freezing, and also in temperate zones with a mixture of glycerine and 
 water, hydraulic cranes, etc., are most satisfactory; they are desirable as 
 against the noisy, wasteful winch with its heavy wear and tear and com- 
 paratively short life. 
 
 In the Navy the Air Compressor is a most important machine, inasmuch 
 as by its means the deadly torpedo is endowed with life and speed. Air is 
 drawn from the atmosphere and delivered^to the necessary tubes at a pressure 
 of 2,500 lbs. per square inch at a safe and comparatively low temperature. 
 
 A 
 
 To Automatic *S Water Connection for 
 
 Separator Column X\ 'Vernal Lubrication 
 
 Of AirCjImde, 
 y — x\ ?rjfsi„ n - ■ — 
 
 /^Stoge) 
 
 cooling coirrr 
 
 2 nd Stage 
 Cooling Co 
 
 3C? 'Stage &o 
 
 4t* 'Stage C 
 
 Circulating I 
 Pump 
 
 Fig. 191a. — Section through Pumps of a Naval High-pressure Air Compressor fot 
 
 2,500 lbs. per square inch. 
 (Belliss & Morcom.) 
 
 Fig. 191 is the compressor supplied by Belliss & Morcom for battleships 
 and big cruisers worked by a steam engine. In this case the lower portion 
 is simply an enclosed self-lubricating steam engine. Superimposed on it 
 is a compound four-stage system of compressors, whose pistons are connected 
 direct to the engine piston-rods. As the air is compressed it is cooled by 
 coils containing cold sea-water which is passed through them. This machine, 
 when running at 350 revolutions per minute, can deliver 30 cubic feet of 
 air compressed to 2,500 lbs. pressure (v. fig. 191a). By replacing the steam 
 cylinders with the compressors and connecting an electric motor to the 
 
528 
 
 MANUAL OF MARINE ENGINEERING. 
 
 crank-shaft next the flywheel an electric-driven machine is provided, and 
 in this form is sometimes preferred. 
 
 Steam Steering Gears are now fitted to quite small ships, and although 
 designed originally by the late J. Macfarlane Gray for the " Great Eastern " 
 
 Fig. 192.— Vertical Combined Steam and Hand Steering Gear. 
 (Bow, M'Lachlan <fe Co. ) 
 
 Fig. 193.— Combined Hand and Steam Steering Gear as fitted in High-speed 
 
 Cross Channel Steamers. 
 (Bow, M'Lachlan <fc Co.) 
 
STEERING GEAR. 
 
 529 
 
 and other very large vessels, which could not be controlled or properly 
 navigated by hand gear, and improved by the late A. B. Brown for all classes 
 of large ships to save the hard work of a large number of men (it would he 
 the whole watch of a modern tramp steamer). Fig. 192 is an example of the 
 kind of steering engine that has been evolved in course of time for use in the 
 ordinary merchant steamer of small and moderate size ; it will be observed 
 that the same wheel with which the steam gear is operated when under 
 steam may be employed to steer by hand in case of mishap to it by simply 
 slipping a clutch into gear with the pinion geared into the spur wheel on the 
 chain or wire rope barrel shaft. In a general way in fine weather on the 
 open sea many ship captains prefer hand steering; with such an arrange- 
 
 Fig 194. — Type of Vertical Steam Steering Gear fitted in War Vessels. 
 
 ment this can be done, and on quite short notice ckanged to steam gear on 
 an emergency. The compensating gear, whereby the valve displaced by the 
 steersman is brought back to mid or normal position is seen on the front, 
 and deriving its motion from the worm drive. Fig. 193 is a new arrange- 
 ment as adopted in the modern high-speed turbine steamer, and placed at 
 the stern in direct connection with the well-known right- and left-handed 
 screw arrangement for working by hand. The engine is controlled from 
 the bridge by means of a Brown's telemotor (by preference) or mechanical 
 gearing. Fig. 194 is the vertical type of steering engine as required by the 
 Admiraltv for cruisers, and fixed to a convenient bulkhead under water, 
 J 34 
 
530 MANUAL OP MARINE ENGINEERING. 
 
 and protected from damage by shot or shell. In all these cases the speed 
 of the engine is reduced to that of the gear by means of a worm and wheel, 
 as being the most convenient if not the most efficient method. Such worms 
 should work in an oil bath, and run on specially cut wheels with the least 
 possible friction. 
 
 Brown's Patent Steam Tiller. — The importance of fitting steamships, 
 whether of high or low speed, with absolutely reliable steering gear cannot 
 be overrated ; but in many cases commercial considerations are allowed to 
 weigh in the design of this vital adjunct to a steamer's equipment. By far 
 the greatest number of steamships have their steering engines placed near 
 the bridge, the communication being made with the quadrant aft by means 
 of chains, rods, or wire ropes, with or without spring buffers to take off the 
 shock of a heavy sea. In conjunction with this, hand gear is fitted aft, 
 having double screws with nuts and cross-head, the mode of connection 
 being by pins dropping into connecting links, or by a clutch working on the 
 rudder-head and engaging the cross-head. The trouble involved in keeping 
 these steering ropes or rods properly adjusted and the various pulleys properly 
 oiled, as well as danger arising from the ropes being carried away, has brought 
 about a change in more recent applications of steering gear. The steering 
 engine is placed aft, being coupled by right- and left-hand screws, and in 
 a variety of other ways, direct to the rudder-head ; communication from the 
 steering valve being made by a line of shafting to the bridge, thus dispensing 
 with the objectionable rope or rod communication, which is, in the first- 
 mentioned system, subject to the full rudder strains. 
 
 An ideally perfect steering gear should fulfil the following conditions : — 
 
 1. The steering engine should be attached to the rudder-head without 
 the intervention of chains or ropes. 
 
 2. It should let go the rudder when unduly strained^ and, when the 
 abnormal strain has gone, return automatically to its former position. 
 
 3. The connection from steam to hand gear, and vice versd, should be 
 affected without the use of jaw chitches or the shipping of bolts into holes 
 — which operations are difficult to effect when the ship is rolling at sea with 
 the rudder adrift. 
 
 4. The communication from the bridge to the machinery aft should be 
 of a kind which dispenses with rods, chains, and shafting, all being equally 
 troublesome to the shipbuilder to arrange, and to the officers of the ship to 
 keep in order. 
 
 With reference to condition 3, it is a common practice to fit rudder brakes 
 on ships where clutches are the means of connection ; but as simplicity 
 and fewness of parts are of first importance in steering gear, it is better that 
 such a connection between the steering engine and the rudder, or the hand 
 gear and the rudder, should be one which will act both as a clutch and a 
 brake. 
 
 To meet these conditions as far as possible, the steam tiller was designed 
 by the late Andrew B. Brown. Fig. 195 shows an elevation with hand 
 steering gear and plan. 
 
 The prominent feature of this gear, in which it differs from all others, is that advan- 
 tage is taken of as long a lever as will reach from the rudder-head to the limits of the 
 poop deck, which, in the greatest number of ships, varies from 7 to 10 feet, and in the 
 
BBOWN'S PATENT STEAM TILLER. 
 
 SECTIONAL DETAIL OF WORM RNKWHtHB 
 
 i'~S Saa/e 
 
 Fig. 195.- -Brown's Steam Steering Gear. 
 
532 MANUAL OF MARINE ENGINEER]' 
 
 largest class of vessels has reached the length of 17 feet. It will be obvious that the 
 strains at the end of such a lever will be reduced to the smallest possible amount, and 
 that the gear necessary to give the requisite power to steering the ship will be of the 
 simplest form. 
 
 This tiller is shown in fig. 195, A, keyed to the rudder-head, B, and at the other 
 end a jaw, C, is fitted with gun-metal bearings, into which a driving pinion, D, works, 
 gearing into the toothed segment, E, which is bolted securely to the deck. The steering 
 engines are carried on the tiller and move round with it, receiving and exhausting their 
 steam through a double stuffing-box arrangement, F, which also contains the reversing 
 valve, and is mounted on the axis of the rudder-head. The steam cylinders, G, are of 
 the usual well-known construction, fitted with piston valves. Motion is communicated 
 to the pinion, D, which is of cast steel, through the intervention of an expanding friction 
 clutch, H, which is lined with elm wood, and engages the worm-wheel, I. This wheel, 
 to reduce friction, is carefully machined in the teeth, and made an exact fit to the worm r 
 J, which is of Admiralty gun-metal, and works in the worm-wheel without any back- 
 lash or shake. The worm is also curved to the radius of the wheel so that three teeth 
 are engaged, thus prolonging the life of the gearing by three times that of a straight 
 worm. 
 
 Motion is given to this worm by the steam engine as shown. The clutch, H, is 
 expanded by a screw bolt and worm-wheel, K, which turns in and out of the nut, L, at 
 one end, the other abutting against a series of laminated springs, M, so that by turning 
 the worm, N, by a handle (provided for the purpose) to the right or left, the steam gear 
 is engaged or disengaged at any position the rudder may be in, and at the same time 
 it forms an efficient brake to seize hold of the rudder in a sea-way. In practice it is usual 
 to expand this friction brake or clutch sufficiently tight to put the rudder hard over 
 at full-speed trials ; but the springs in any case have not sufficient force to hold the 
 connection tight enough to cause fracture of any part ot the machinery. In the event 
 of a heavy sea striking the rudder, it immediately slips, allowing the rudder to move 
 out of position : but by that act the steam valve is opened and the engines bring the 
 rudder back to its normal place. As the steam tiller is intended to work (and in most 
 cases has been so fitted) on the open deck, without any house, the whole of the machinery 
 is placed in a water-tight casing, which forms the framework of the steering engines, 
 access to which is got by the doors, 0. The oiling of the various parts is effected 
 automatically by two valveless oil pumps, P P, driven off the valve-rods of the engine. 
 These throw the oil from a well in the bottom of the casing through the hollow piston- 
 rod into the reservoir, Q, and from there a copious supply of oil is supplied to every 
 working part, as well as the piston and valve-rods. In actual practice the oil is renewed 
 once in three months, about 2 gallons being required. 
 
 The pinion end of the tiller is carried up by gun-metal slippers and spiral springs 
 under the lugs, R R, which are capable of adjustment. As there is always a tendency 
 of the tiller or quadrant to shake or chatter when there is no strain on, the rudder being 
 fore and aft, the slippers in that position work upon inclined planes, S S, which gives 
 sufficient brake action to prevent*any vibration. At the same time, when putting the 
 helm over to such an extent that the pinion bears hard on the rack, the slippers run 
 down on to the flat part of the toothed segment, when the springs slack off, and only 
 carry up the weight of that end of the tiller. 
 
 The hand gear consists of a strong standard, T, bolted to the deck, and carrying an 
 exactly similar worm-wheel, and worm with hand-wheels and friction clutch as that 
 described in the steam gear. At the lower end of the shaft there is a similar pinion to 
 D, which engages the toothed segment, U, of cast steel, which is securely bolted to the 
 steam tiller. The operation in changing from hand to steam or steam to hand by means 
 of these clutch brakes can be, and has been performed, without any undue haste, in 
 half a minute. It may here be pointed out that the result of actual experience is that, 
 with this system of hand gear, the friction is one-third of that of the double screw system 
 with nuts and connecting-rods to a cross-head on the rudder-head. Therefore, one 
 man on the worm-wheel gear is as effective as three on the double screws. 
 
 The hand wheels, it will be observed, are set to one side of the centre line, which 
 economises space fore and aft, and brings the position of the man steering immediately 
 opposite the compass. 
 
 The steering valve is operated by the lever, V, which causes the piston valve to 
 turn on its axis inside the trunnion casing, and as the tiller moves round it carries the 
 valve face with it and so closes the port. The lever, V, is connected to the motor 
 
REFRIGERATING MACHINES. 533 
 
 cylinder of the telemotor gear, leading up to the bridge by two pipes | inch in diameter. 
 In case of accident to these pipe communications to the bridge, a steering station, W, 
 is shown aft, which can be connected to the steering valve. 
 
 It is claimed for this design of steering gear that it has the fewest number of parts 
 possible — namely, one pinion, one worm-wheel and worm — which, it can easily be seen, 
 is due to the fact that the toothed segment represents in a 10-feet tiller a steering wheel 
 20 feet in diameter, and this rack being shrouded to the points of the teeth and bolted 
 at short intervals to the steel deck, is extremely secure. 
 
 From a commercial point of view, there is a distinct saving in the adoption 
 of such a design, as no space is required for a steering engine amidships, 
 and no house is required aft — a few stanchions and rods or netting being 
 placed at the extremity of the gear for the protection of passengers. 
 
 Refrigerating Machines, whereby the air in any chamber can be cooled 
 down to any degree of temperature required "and kept steadily at that tem- 
 perature, are now in great demand for the comfort and well-being of passengers 
 on mail steamers, for the safety if not the comfort of those on warships, and 
 as a means of transporting food of all kinds from the ends of the earth to 
 those countries which are unable to produce sufficient for the wants of those 
 dwelling in them. By their means also the luxuries of life are brought cheaply 
 to the doors of even poor people at very moderate cost. The original idea 
 on which such machines were projected was to abstract the heat of air after 
 compression, so that when it was freed again and expanded it robbed its 
 surroundings of the heat necessary tor that process. If, therefore, a machine 
 compressed air to, say, 100 lbs. pressure, its latent heat became sensible and 
 great, so that on passing it into a chamber with pipes through which cold 
 water was forced it was cooled down to nearly the temperature of the water ; 
 on admitting this air to a store chamber or powder magazine on expansion 
 it would freeze the water particles in its neighbourhood, producing snow, 
 and lowering the temperature of the air in them to that desired. Or if, 
 instead of admitting the free expanding air into the chamber it is used to 
 cool down the water in a system of pipes, like that of a heating apparatus 
 fitted within the chambers, the action will not be so sudden or so spasmodic, 
 and the whole much more under control. To-day other and improved 
 methods are adopted, one of which is that of Messrs. J. & E. Hall, in which 
 carbonic acid gas is used instead of air with advantage, as it can be made 
 and bottled up at comparatively little cost. 
 
 Fig. 196 shows the machine made for use on warships in order that the 
 temperature in the magazines containing the modern explosives shall be 
 kept sufficiently low to prevent their decomposition, and the dangers arising 
 from the instability of these compositions when warm. Fig. 197 is an illus- 
 tration of the larger horizontal type used on merchant ships to chill the air 
 in the cargo holds for the safe conveyance of meat, fruit, vegetable products, 
 etc., which can only be preserved by maintaining them at a low temperature 
 without freezing their fluid portions. This particular machine has an ice- 
 making rating of 30 tons per twenty-four hours. 
 
 The machine consists of two complete units, of which the compressor 
 are made from solid blocks of high carbon forged steel, coupled to the 
 tail rods of a compound steam engine, whose crank-shaft is in two parts, 
 with a flywheel mounted on each coupling, and so capable of being dis- 
 connected for the independent running of either side when required in an 
 •emergency. The C0 2 condenser coils are of solid drawn copper contained 
 
534 
 
 MANUAL OF MARINE ENGINEERING. 
 
 in the two halves of the base casting, and are arranged for withdrawal side- 
 ways into the space which has to be provided for the gangway. By this 
 arrangement the coils can be exposed for their entire length for cleaning 
 
 Fig. 196. — Refrigerating Apparatus for Naval Ships. 
 
 Fig. 197. — Refrigerating Machine for Large Merchant Ships (J. & E. Hall). 
 
 when the large door is removed, without disturbance of the gas joints. The 
 evaporator coils, of wrought-iron hydraulic tubing, electrically welded into 
 continuous lengths, are contained in built steel casings constructed to suit 
 
AUXILIARY CONDENSER. 
 
 535 
 
 the ypace available in each particular case. The coil boxes, or headers, 
 on the condensers and evaporators are machined from solid mild steel forgings, 
 which also form the stop valve chests, and thereby keep the number of gas 
 joints at a minimum. Each half of the machine is complete with its own 
 pressure lubricator, oil separator, condenser and evaporator gauges, and 
 other accessories. 
 
 Fig 198 is a section in diagrammatic form, which shows clearly the working 
 of the Hall system. It will be seen by it, and by reference to fig. 196, that 
 the apparatus supplied to the Admiralty is very compact, and occupies 
 little space, and every part kept small enough to pass through the small 
 opening available in the lower regions of warships. It is worked by an 
 
 Condenser 
 Coil 
 
 Brine return 
 Thermometer 
 
 Patent Safety^ 
 Valve in here 
 
 Steam 
 Cylinder 
 
 Compressor 
 
 Separata 
 
 Patent hollow 
 OH Gland 
 
 Condenser Gauge 
 
 Evaporator 
 Gauge 
 
 Regulator 
 
 Pot catching 
 Gil From 
 Gland 
 
 _Lvaporator 
 Cod 
 
 Insulation 
 ound 
 Evaporator 
 
 Water Circulating 'Pump Condenser Casing 
 
 Fig. 198.— Section of Hall'a Refrigerating Apparatus. 
 
 electric motor through machine-cut wheel gearing, and can be operated 
 easily by the ordinary attendant in the magazine quarters. 
 
 Auxiliary Condenser. — In these days of high pressures of steam when 
 fresh water only must be used in steam generators, it is essential to save all 
 the steam possible, and as large quantities are used by the auxiliary machinery 
 in modern steamers, it is worth while having a special condenser for this 
 purpose. In the mercantile marine it is generally called a winch condenser, 
 as primarily it was fitted to take the exhaust steam from the winches and 
 return it to the auxiliary or donkey boiler supplying them. These con- 
 densers in the Navy are similar in design and construction to the main 
 
536 
 
 MANUAL OF MARINE ENGINEERING. 
 
 condenser, but of course much smaller ; a combined air and circulating 
 pump is attached and kept working, when in use, at a constant speed ; an 
 automatic regulating valve can be fitted in such a way that a the vacuum 
 improves steam is reduced, and when 24 inches of vacuum is obtained, 
 the engine goes " dead " slow. 
 
 Fig. 199 gives a simple arrangement for the mercantile marine which 
 
 Fig. 1 09. — Winch Condenser, complete, with Air and Circulating Pumps 
 
 may be placed in any convenient position in engine or boiler room, as it is 
 self contained. Sometimes in the mercantile marine a plain cylindrical 
 condenser is fitted so that the ballast pump supplies circulating water, and 
 oue of the auxiliary feed-pumps draws away the condensed water ; in this 
 case little or no vacuum is maintained, but no water is lost. In every-day 
 
AUXILIARY CONDENSER. 537 
 
 •work the winch and other auxiliary engine glands and drain cocks leak so 
 badly that only a very poor vacuum is possible with even a good-sized air- 
 pump. 
 
 Fig. 200 is a modern high - class surface condensing plant, such as is 
 used on shipboard when a good vacuum is required for working such things 
 as turbine generators, etc. It is also a good representation of a combined 
 set of air-pumps and circulating-pumps, such as may be employed on a 
 
 Fig. 200. — Auxiliary Surface Condensing Plant (Belliss & Morcorn). 
 
 main condenser. It consists of a pair of Edwards' air-pumps worked by 
 cranks driven by a spur-wheel having helical machine-cut teeth, ijito which 
 there is a pinion on the crank-shaft end of a tandem compound enclosed 
 forced-lubricating engine ; to the other end of the crank-shaft the spindle 
 oi a centrifugal pump is coupled. In large sizes there are sometimes three 
 pumps driven by cranks at 120° apart, but really the two pumps with cranks 
 opposite one another are sufficient 
 
838 MANUAL OF MARINE ENGINEERING. 
 
 CHAPTER XXL 
 
 BOILERS, FUEL, ETC., EVAPORATION. 
 
 The boiler proper consists essentially of two parts ; the one the fire-place in 
 which the fuel is burnt and heat generated, the other in which the heat is 
 applied to boil the water and convert it into steam. 
 
 The part where the fuehis burnt is called the furnace, and that on which 
 it is laid is called the grate. 
 
 The quantity of steam generated will depend, in the first place, on the 
 quality of the fuel, and on the quantity burned ; and in the second place on 
 the capacity of the boiler for absorbing and transmitting the heat generated. 
 The quantity of solid fuel consumed depends on the area of grate, and the 
 draught or flow of air into the furnace. 
 
 The efficiency of the furnace depends on its capability of burning with as 
 little waste as possible the whole of the fuel in it, and that without super- 
 fluity of air ; also, to produce perfect combustion there must be as little 
 waste of heat as possible in obtaining the necessary draught. The portion 
 of heat generated which is applied to the production of steam is called the 
 available heat. 
 
 Experience has proved that a grate may consume a large quantity of 
 fuel without thoroughly burning it, and that even when the fuel is thoroughly 
 burnt only a comparatively small portion of its heat may be usefully employed. 
 
 The chief loss is at the chimney, which is a rough and ready way of 
 inducing the air to flow into the furnaces with sufficient velocity to cause 
 the fuel to burn ; but it is an exceedingly wasteful one, and will, some day, 
 undoubtedly be superseded by a more scientific and economical apparatus. 
 One pound of fairly good coal can be made to evaporate 14 to 15 lbs. of 
 water ; but in the best of marine boilers only 10 to 11 lbs. of water per pound 
 of such coal are evaporated in practice with every care taken, showing that 
 the efficiency of the boiler is less than 075 from this point of view. With 
 picked Welsh coal and most careful stoking 12 lbs. can be evaporated if the 
 rate of combustion and evaporation is not high. 
 
 Efficiency of the Furnace. — Loss may take place at this part of the boiler 
 from the following causes : — 
 
 (1) Bad stoking, whereby fuel is lost by falling through the bars only 
 partly consumed, and is thrown away with the ashes. This generally takes 
 place with good fuel from its friability ; but may be due in some cases to 
 carelessness in laying the fire-bars. Unevenly disposing of fresh fuel and 
 allowing clinkers to form are also causes of inefficiency. Too much cannot 
 be said of bad stoking, as from this cause alone the best-designed and made 
 boiler may prove most inefficient. • 
 
 (2) Want of air, whereby the fuel is not wholly consumed, but part of it 
 
FUEL. 539 
 
 passes off through the funnel as carbon monoxide, part in the form of smoke, 
 part is deposited as soot in the tubes, and part is burnt in the smoke- 
 box. This is sometimes due to bad design and want of proper means of 
 regulating the supply of air ; but it is more frequently due to bad stoking 
 and carelessness on the part of the fireman in using the means provided. 
 
 (3) Excess of air, whereby some of the heat is employed in raising the 
 temperature of the superfluous air, and thereby doing no good. This is 
 generally caused by bad stoking, the bars being uncovered in parts of the 
 grate, and so admitting a too large inflow of air. 
 
 (4) Radiation through the mouth when the doors are open for firing, and 
 any radiation through openings for other purposes. This cannot be avoided, 
 but should be comparatively small. 
 
 The first three of these sources of loss cannot be said to be unavoidable, 
 as with ordinary care the loss from them should be very small compared 
 with the chief loss. 
 
 Chimney Draught. — To obtain an efficient draught in a chimney or funnel, 
 the temperature at its base should be about that of melting lead, or nearly 
 GOO . Professor Rankin e said that " the best chimney draught takes place 
 when the absolute temperature of the gas in the chimney is to that of the 
 external air as 25 to 12." 
 
 That is, if T be the temperature of the air, then 
 
 Absolute temperature at the base of funnel = f § X (461° -f T), 
 
 or temperature according to thermometer (Fahrenheit) 
 
 = || (461° + T) - 461° = 2-08 T + 500°. 
 
 Taking the air temperature at 60°, the proper heat at the base of the 
 funnel is 2'08 X 60 + 500, or 625°. Now, the heat in a furnace having a 
 natural draught is usually about 2,400°,. and if the heat at the exit from 
 the boiler to obtain the necessary draught is 600°, it follows that 20 per 
 cent, of the total heat of combustion is wasted owing to this. If, instead of 
 a chimney, a draught were produced artificially, say by means of an economic 
 engine driving a fan. a very considerable portion would be saved. 
 
 F ue i. — The solid fuels used on board ship are coal of various descrip- 
 tions and qualities, wood in those parts of the world where it is plentiful 
 and coal scarcer, and patent fuels or combinations of very small coal with 
 other substances. Cole is seldom used, but oil, comparatively cheap, and 
 in some instances most convenient, has now come into general use, especially 
 in warships. Mineral oil is capable of producing a large quantity of heat, 
 is clean and convenient for stowing, and therefore suitable for warships, 
 yachts, and passenger steamers ; arrangements are now made for the regular 
 supply of a high flash point cheap oil at foreign and home ports, so that 
 ships can depend on getting their tanks filled as surely as they can their 
 bunkers. Creosote waste, as well as petroleum refuse, can be burnt quite, 
 safely and readily in the furnaces of a boiler by means of special apparatus ; 
 the one invented by Mr. James Holden, and used by him so successfully 
 for many years in the locomotives of the G.E.E. Co., London, is shown in 
 detail in fig. 201, and fig. 202 gives a very simple form used on the Japanese 
 ships. Residue, after the light illuminating oils have been abstracted from 
 
> 
 - ~ > 
 
 o 
 
 S 
 o 
 o 
 
 s 
 
 01 
 
 e8 
 
 to 
 
 C 
 
 G 
 
 2 
 
 g 
 
 1 
 
 o 
 
 60 
 
PATENT FUELS. . 54 1 
 
 the petroleum, in South Russia called astatki, has for very many years 
 been used on the steamers of the Caspian and Volga for fuel. 
 
 Of coals there are several distinct kinds, and many more qualities. There 
 are five distinct varieties, known as — 
 
 (1) Anthracite, consisting almost entirely of free carbon, generally jet 
 black in appearance, but sometimes greyish like black lead, has a specific 
 gravity generally of about 15, but sometimes as high as T9 ; it burns without 
 emitting flame or smoke, but requires a strong draught to burn at all. It 
 is capable of evaporating (theoretically) nearly 16 times its weight of water, 
 but to obtain good results from it careful stoking is necessary, as when 
 suddenly exposed to heat it is very friable, breaks up into small pieces, and 
 falls through the bar-spaces if disturbed much, as it does not cake. The 
 fires should be worked light when using it, and the coal carefully spread. 
 The heat is very intense and local, so that furnaces intended to burn it should 
 be high in the crowns and protected at the sides by bricks, fireclay, etc., or 
 else have no air-spaces down the sides. 
 
 (2) Dry bituminous coal contains from 70 to 80 per cent, of carbon, and 
 about 15 per cent, of volatilisable matter ; its specific gravity is from 13 to 
 1-45. It burns easily and swells considerably while being converted into 
 coke. The harder kinds do not burn so readily, nor do the pieces stick 
 together so easily when burning, and are generally better adapted for marine 
 boilers. 
 
 (3) Bituminous caking coal, containing from 50 to 60 per cent, of carbon, 
 is generally of about the same specific gravity as the dry bituminous ; it 
 contains, however, as much as 30 per cent, of volatilisable matter, and con- 
 sequently develops hydrocarbon gases ; it burns with a long flame, and 
 sticks together in caking, so as to lose all trace of the original forms of the 
 pieces. It requires special means to prevent smoke. 
 
 (4) Cannel coal, or long flaming coal. — This is seldom used for steam 
 purposes, as it gives off large quantities of smoke, and is very scarce. It 
 is the best coal for the manufacture of gas. 
 
 (5) Lignite, or brown coal, is of later formation than the other coals, and 
 in some instances approaches to a peaty nature. It contains, however, 
 when good, from 56 to 76 per cent, of carbon, and has a specific gravity 
 from 1'20 to 135. It also contains large quantities of oxygen, and a small 
 quantity of hydrogen. The commoner kinds of lignite are poor, and contain 
 as little as 27 per cent, of carbon, and, therefore, are not suitable for steaming 
 purposes. 
 
 In some inland parts of the world, where coal cannot be easily obtained, 
 wood is used for fuel, and the furnace specially constructed to burn it ; it 
 contains, on the average, when dry, about 50 per cent, of carbon, 41 of oxygen, 
 and 6 of hydrogen. 
 
 Patent Fuels consist principally of coal dust, and their value depends, 
 therefore, on the quality of the coal from which they are made. To utilise 
 the small coal from the mines and yards was a somewhat difficult problem, 
 as it could not be conveniently transported, and is difficult to burn in an 
 ordinary furnace. Friable coal is also equally difficult to deal with. By 
 mixing "a small quantity of pitch or tar with it, and baking the mixture in 
 moulds, a hard brick is produced, which is easily handled and burns well. 
 
542 
 
 MANUAL OF MARINE ENGINEERING. 
 
 There are many other ways of manufacturing patent fuel from small coal, 
 but the result is the same — viz , a hard compact brick. 
 
 Oil Fuel is now largely used in warships and in many of the mercantile 
 ships in various parts of the world, where it can be obtained in quantity 
 
 <w 
 
 X 
 
OIL BURNERS FOR MARINE BOlLtRS. 543 
 
 cheaply. It is used chiefly at present for raising steam, and, therefore, its 
 composition is of small moment compared with its calorific value, but in the 
 near future the demand will be considerable for use in the oil internal-com- 
 bustion engine, when the composition will be a matter of supreme importance, 
 inasmuch as those oils containing asphaltum matter in appreciable quantity 
 cannot be used in spray form in an engine cylinder, owing to the deposition 
 of coke sand of a hard refractory nature, which will not consume nor freely 
 depart from within, consequently the pistons, cylinder walls, rods, etc., get 
 badly damaged. To use such oils in an internal-combustion engine success- 
 fully the pitchy matter must be deposited outside, and only the volatile 
 constituents permitted to enter the cylinders. 
 
 Crude Oils, pumped from the wells, after exposure, can be used in boilers 
 for raising steam, and, being cheap and easily handled on shipboard, have 
 great advantages over coal, especially where coal is imported and compara- 
 tively dear. Such oil, however, contains much volatile matter, and is, there- 
 fore, not so safe as coal, nor so safe as the residue from it after refining, which 
 is quite as good for burning, and has a high flash point. 
 
 Table lia. gives the composition and heat value of all the principal oils 
 now used, as well as that of the residues and of coal tar. 
 
 Admiralty Conditions for Oil Fuel.— The flash point is not to be below 175° F. 
 The sulphur contents are not to exceed 3 per cent. Practically free from acid, 
 which must not exceed 0-05 per cent. The water with it must not exceed 0-5 
 per cent. The viscosity such that the flow at a temperature of 32° F. by the 
 Kedwood viscometer does not exceed 2,000 seconds for 50 cubic centimetres. 
 There must be no earthy or fibrous matter with it. 
 
 The other Heavy Oils used for the Diesel and other engines are the resi- 
 dues after the light volatile and inflammable constituents, such as petrol, 
 benzine, paraffin, etc., have been extracted, and the flash point is then over 
 200° F., and often as high as 250°. The Caspian astatki has a higher flash 
 point, being generally as high as 390° F. American residues are those 
 remaining after lubricating as well as lighting oils have been distilled. 
 
 Oil Burners for Marine Boilers are designed to pulverise the oil as it flows 
 through their nozzles, and to direct the spray and the consequent flame to 
 that part of the furnace desired. There are three ways of effecting this — 
 (1) By means of a steam jet, (2) by an air stream, and (3) by purely mechanical 
 means. The steam injector is the simplest, and has been used with success 
 by Mr. Holden and others on locomotives, but on shipboard the loss of water 
 is too serious a matter, and so some marine engineers prefer the air blast, 
 which at comparatively low pressures is much more economical than the 
 steam jet ; at high pressures, however, the air system is very little, if at 
 all, superior to the mechanical methods. The mechanical methods do not 
 involve so great a steam consumption nor such expensive mechanism as 
 air compressors, nor do they require quite so much space. The oil is by 
 them forced into the burners at pressures varying from 50 to 250 lbs. per 
 square inch, and there heated by steam to temperatures varying from 150° to 
 300° F. The fires thus formed may be under natural or forced draught, 
 even to such a forcing as measured by 5 inches of water, but the best results 
 are obtained at quite low rates of combustion, when with decent oil 14 to 
 16 lbs. of water (from and at 212° F.) can be evaporated per pound of oil, 
 as against 8 to 11 lbs. when using best Welsh steam coal. 
 
£44 
 
 MVNUAL OF MAtilNE ENGINEERING. 
 
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VALUE OF A FUEL. 
 
 54, 
 
 TABLE Lla.— Liquid Fuels, Composition and Calorific 
 
 Values of. 
 
 N.B. — Actual Evaporation in Practice is about 70 per cent, when highly forced, 
 and 77 per cent, when lightly forced. 
 
 
 
 
 
 Chemical Components. 
 
 
 Evapora- 
 tion per 
 
 Description a 
 
 id Origin. 
 
 Flash 
 Point. 
 
 Specific 
 Gravity. 
 
 
 
 B.T.U. 
 per Lb. 
 of Fuel. 
 
 Lb. from 
 and at, 
 
 Mi!*. 
 
 
 
 
 
 
 
 
 Car- 
 bon. 
 
 •™ r <H Oxygen. 
 
 Sul- 
 phur. 
 
 
 B.T.U. 
 + 06C. 
 
 Heavy Petroleum 
 
 W. Virginia, 
 
 F.° 
 
 0-873 
 
 83-50 
 
 13-30 
 
 3-20 
 
 
 18,324 
 
 18-97 
 
 J* »J 
 
 Pennsylvania, 
 
 
 0-886 
 
 84-90 
 
 13-70 
 
 1-00 
 
 . . 
 
 19.210 
 
 19-88 
 
 Crude ,, 
 
 Mexico, 
 
 
 0-937 
 
 83-21 
 
 11-32 
 
 2-90 
 
 2-43 
 
 17,460 
 
 18-07 
 
 J» )t 
 
 Texas, 
 
 180° 
 
 0-924 
 
 84-60 
 
 10-90 
 
 2-87 
 
 1-63 
 
 19.060 
 
 19-73 
 
 Treated 
 
 »• • 
 
 216° 
 
 0-926 
 
 83-26 
 
 12-41 
 
 3-83 
 
 0-50 
 
 19,480 
 
 20-16 
 
 Residue „ 
 
 California, . • 
 
 278° 
 
 0-864 
 
 84-00 
 
 11-07 
 
 3-93 
 
 0-87 
 
 17,400 
 
 18-01 
 
 j) ri 
 
 . 
 
 185° 
 
 0-935 
 
 84-78 
 
 11-85 
 
 2-45 
 
 0-84 
 
 17,550 
 
 18-17 
 
 Treated 
 
 it • 
 
 311° 
 
 0-966 
 
 81-52 
 
 11-01 
 
 NO 6-9 
 
 0-55 
 
 18,667 
 
 19-32 
 
 Petroleum from 
 
 Parma, 
 
 
 0-786 
 
 84-00 
 
 13-40 
 
 1-80 
 
 
 18.218 
 
 18-96 
 
 
 Pechelbronn, 
 
 
 0-892 
 
 85-70 
 
 12-00 
 
 2-30 
 
 , . 
 
 18,036 
 
 18-67 
 
 - 
 
 E. Galicia, . 
 
 
 0-870 
 
 82-20 
 
 12-10 
 
 5-70 
 
 . . 
 
 18,153 
 
 18-79 
 
 
 W. Galicia, . 
 
 
 0-885 
 
 85-30 
 
 12-60 
 
 2-10 
 
 
 18,416 
 
 19-06 
 
 ,, (Residue) 
 
 Roumania, . 
 
 284° 
 
 0-930 
 
 
 , . 
 
 0-43 
 
 0-20 
 
 18,900 
 
 19-56 
 
 j) »> 
 
 Balakhnny, . 
 
 % t 
 
 •0-882 
 
 87-40 
 
 12-50 
 
 0-10 
 
 
 21,060 
 
 21-80 
 
 „ (Light) 
 
 Baku, 
 
 . , 
 
 0-884 
 
 86-30 
 
 13-60 
 
 0-10 
 
 
 20,628 
 
 21-35 
 
 „ (Heavy) 
 
 j) • 
 
 . # 
 
 0-938 
 
 86-60 
 
 12-30 
 
 1-10 
 
 , . 
 
 19,440 
 
 20-12 
 
 ,, (Residue) 
 
 »» • 
 
 , , 
 
 0-928 
 
 87-10 
 
 11-70 
 
 1-20 
 
 , , 
 
 19,260 
 
 19-93 
 
 »> »j 
 
 Java, 
 
 4 , 
 
 0-923 
 
 87-10 
 
 12-00 
 
 0-90 
 
 , . 
 
 19,496 
 
 20-18 
 
 Shale Oil, 
 
 France, 
 
 t t 
 
 0-911 
 
 80-30 
 
 11-50 
 
 NO 8-7 
 
 
 16,283 
 
 17-41 
 
 
 Scotland, 
 
 
 0-862 
 
 85-35 
 
 12-44 
 
 . , 
 
 0-29 
 
 18,317 
 
 18-96 
 
 Coal-tar oil (light) 
 
 , British, 
 
 120° 
 
 0-958 
 
 86-16 
 
 9-05 
 
 ., . 
 
 0-30 
 
 18,062 
 
 18-70 
 
 „ (heavy) 
 
 , British, 
 
 
 1-038 
 
 84-50 
 
 7-62 
 
 6-28 
 
 0-19 
 
 16,325 
 
 16-90 
 
 ; J t y 
 
 France, 
 
 
 1-044 
 
 82-00 
 
 7-60 
 
 10-40 
 
 
 16,050 
 
 16-61 
 
 »» *» 
 
 Germany, 
 
 208° 
 
 1-065 
 
 90-12 
 
 7-09 
 
 1-66 
 
 0-63 
 
 15,950 
 
 16-51 
 
 Gas oil (light). 
 
 British, 
 
 176° 
 
 0-850 
 
 86-98 
 
 12-15 
 
 0-16 
 
 0-71 
 
 18,000 
 
 18-63 
 
 ,, (heavy), 
 
 )» • 
 
 
 1-067 
 
 87-62 
 
 5-95 
 
 
 0-67 
 
 16,153 
 
 16-72 
 
 N.H. — For evaporation per gallon multiply by (10 x specific gravity). 
 
 The Value of a Fuel is determined by its chemical composition. All 
 fuels contain more or less carbon, and must have hydrogen and oxygen 
 in various proportions, also some small quantities of nitrogen, sulphur, etc. 
 
 These substances are usually designated by their chemical symbols, that 
 is by the initial letter of the name. They combine in certain fixed quan- 
 tities, called their chemical equivalents. Thus — 
 
 Carbon, symbol C, chemical equivalent 12. 
 Hydrogen, ,, H, „ „ 1. 
 
 Oxygen, „ O, „ „ 16. 
 
 Nitrogen, ,, N, ,, „ 14. 
 
 Sulphur, ,, S, ,, ,, 32. 
 
 A pound of carbon is capable of developing, during combustion, a certain 
 quantity of heat, called the total heat of combustion, and is measured by units 
 of heat. 
 
 The British standard * unit of heat is defined as that quantity of heat which 
 will raise one pound of pure water one degree Fahrenheit in temperature. 
 
 * Joules found the thermal unit to be equal to 772 foot-pounds ; recently, however, 
 with better appliances and possibly greater care in manipulation, Profs. Rowald, Griffiths, 
 and Schuster find it to be 778. 
 
 35 
 
546 MANUAL OF MARINE ENGINEERING. 
 
 The total heat of combustion of one pound of hydrogen is measured in 
 the same way* and it is also found that the total heat of combustion of any 
 compound of carbon and hydrogen, is the sum of the quantities of heat which 
 the hydrogen and carbon contained in it would produce if burnt separately. 
 
 If a fuel contains oxygen as well as hydrogen, it is known that eight 
 parts by weight of the former unite with one of the latter to form water,, 
 which exists as such in the fuel, and does not add to the total heat of combustion. 
 If there is, however, an excess of hydrogen beyond what is required to form 
 with the oxygen the water, the remaining hydrogen does ado to the total heat 
 of combustion, and may be reckoned in estimating its value. 
 
 Hydrogen gas requires 8 pounds of oxygen, and consequently 36 pounds 
 of air to consume it ; its total heat of combustion is 62,032 units of heat, 
 and, therefore, it can evaporate 64 pounds of water from and at 212°. 
 
 Carbon, when fully burnt, requires only 2*7 pounds of oxygen, or 12 
 pounds of air to consume it — that is, to convert it into carbonic acid : its 
 total heat of combustion is 14,500 units of heat, and can, therefore, evaporate 
 15 pounds of water from and at 212°. If, however, it is only partially burnt 
 or turned merely into carbonic oxide, half the quantity of air is consumed, 
 and the total heat of combustion is only 4,400 units of heat. 
 
 Sulphur exists only in small quantities in good coal, and the total heat 
 of combustion is only about 4,000 units. 
 
 From this information, the following rule is deduced for the total heat 
 of combustion for substances containing carbon, hydrogen, and oxygen. 
 
 Total heat of combustion of 1 pound of fuel 
 
 = 14,500 j C + 4-28 (H - ^U, . . . (1 
 
 and 
 
 Theoretical evaporative power of 1 pound of fuel 
 
 = 15Jc+4-28(H-^)|. 
 
 C being the weight of carbon, H that of hydrogen, and O that of oxygen, 
 all expressed in fractions of a pound. 
 
 Example. — To find the evaporative power of a pound of Welsh coal, 
 whose constituents are, carbon 85, hydrogen 4, oxygen 2"5, others 8*5. 
 
 E = 15 j 0-85 + 4-28 (004 - ^)\ 
 = 15 - 125 pounds. 
 
 The quantity of air required to burn a pound of fuel may also be estimated 
 in a somewhat similar way. 
 
 Weight of air required to burn 1 pound of fuel 
 = 12 C + 36 (H -- ?\ 
 
 To burn ordinary coal or coke, 12 pounds of air are required on the 
 average. 
 
 To provide for.' the dilution of the gaseous products so that free access is 
 
ARTIFrCIAL DRAUGHT. 547 
 
 given to the air to reach the fuel, much more is required ; so that to consume 
 one pound of fuel in practice, 24 pounds of air are often used. 
 
 At the temperature of 62°, the volume of 1 pound of air is 13-14 cubic 
 feet ; therefore, to consume 1 pound of coal or coke 315 cubic feet of air 
 are necessary. If, however, the draught is very good, such as found with 
 artificial means, 250 feet are sufficient.* 
 
 Sir Alex. Kennedy found that the best results were obtained when 
 18-14 pounds of air were used, and noted the following : — 
 
 Air per pound of coal, 
 
 . lbs., 
 
 14-5 
 
 17-4 
 
 17-8 
 
 18-14 
 
 Loss per cent., 
 
 . 
 
 92 
 
 5-0 
 
 3-6 
 
 0-5 
 
 Tables li. and lia. give the composition, total heat of combustion, and 
 evaporative power of the various fuels. 
 
 Rate of Combustion. — The quantity of coal burnt on a square foot of 
 grate depends partly on its nature, but mostly on the draught. Some hard 
 coals, like anthracite, require a very strong draught to burn at all, and some 
 qualities of anthracite burn slowly even then. Bituminous coal burns much 
 more freely than anthracite, and some of the softer kinds consume very 
 rapidly. AH coal burns very much more rapidly with a strong draught, 
 as might be supposed, and for that reason, when only a comparatively small 
 boiler can be fitted to supply steam for the small grate possible with it, 
 artificial draught is a necessity. It is also claimed that when the draught 
 is very strong, a smaller iveirjht of air suffices to complete combustion. 
 
 Artificial Draught. — There are four general methods of causing an arti- 
 ficial draught: (1) The steam blast, by which the products of combustion 
 are ejected from the funnel in the same way that the exhaust steam from a 
 locomotive produces a draught ; (2) by an air-blast delivered under the 
 fires with a closed ashpit ; (3) by making the boiler room air-tight, and 
 forcing air into it by a fan, until the pressure is above that of the atmo- 
 sphere, the only vent being through the furnaces to the funnel ; and, (4), 
 by a fan placed near the base of the funnel which draws from the smoke-box 
 and delivers into the funnel ; this latter is known as the induced draught system. 
 
 The first of these methods is the older and costlier form ; it is not very 
 effective, and by no means economical, as there is loss of both coal and 
 fresh water ; it also quickly wears out the funnel ; but it has the merit of 
 being cheap in first cost, and does well enough if it is only required to be 
 used occasionally to quicken the fires in getting up steam, or during a hot 
 day when steaming with a fair wind, or no wind at all. The second plan 
 has the merit of simplicity, and made by the late Mr. James Howden and 
 others very effective by their special systems, it is now very generally used 
 with most satisfactory results in the mercantile marine. There is no danger 
 with it, and it may be applied to any kind of ship. The third plan is one 
 almost universally adopted in warships, and also in many large and small 
 express merchant ships. It is a costly plan, and necessitates a complete 
 change in the ship arrangements in the boiler compartments. It is undoubt- 
 edly a most efficient plan of artificial draught, but the cost must enter into 
 
 * The temperature oi ignition of Scotch coal is 760° F., of Tyneside 770°, Durham 790°, Yorkshire 
 810°, Welsh 875°, and anthracite 925°. 
 
548 
 
 MANUAL OF MARINE ENGINEERING. 
 
 the calculation of practical efficiency, as must also the risk run. The firemen 
 are practically imprisoned in the stokehole, as access to, and egress from it 
 are through an air-tight lock, and when a panic arises the result under such 
 
 SC- 
 SI 
 
 58 
 
 Q 
 1 
 
 fc*. 
 
 r. 
 
 a 
 
 -a 
 
 o 
 
 S 
 
 bio 
 
QUANTITY OF FUEL BURNT ON THE GRATK. 549 
 
 circumstances might be serious ; on the other hand, the fireman is always 
 working in fresh, cool air, and, therefore, more comfortable and efficient 
 than in the ordinary stokehole. The fourth plan is simple and less costly, 
 and was claimed to be less trying to the tubes, etc., but why is not obvious. 
 The Ellis-Eaves patent system, which includes such an eduction fan, has 
 been tried with success ; but the fan must be most carefully designed and 
 fitted to enable it to deal with gases at 600° F. successfully. 
 
 Howden's System of Forced Draught.— The second method of artificial 
 draught above mentioned was elaborated and perfected by the late Mr. James 
 Howden, who was one of the first to take up the question of forced draught, 
 and to persevere both in experimenting and in educating the public mind, 
 so that to-day his system is very extensively adopted in the Mercantile Marine 
 of all nations with more or less beneficial results. Fig. 203 shows his arrange- 
 ment whereby the air is supplied to the furnaces by conduits at the sides of 
 the smoke-boxes after having passed through a chamber containing a system 
 of vertical tubes through which the hot air and gases from the smoke-box 
 flow to the funnel. The supply of air is kept up and regulated by means of 
 a fan, and is usually delivered at a pressure equal to half an inch to one inch 
 of water. This system is, therefore, on the regenerative principle, whereby 
 the waste heat in the chimney is utilised to heat the in-coming air to the 
 furnaces, and although the direct saving by using this waste heat is not very 
 large, owing chiefly to the low specific heat of air, the indirect one is con- 
 siderable from the fact that the combustion of fuel is much more complete 
 and efficient the nearer the air supply approaches the temperature of com- 
 bustion. Mr. Howden claimed a saving of coal by this system amounting to 
 a considerable percentage, and no doubt this was so when applied to a boiler 
 whose design and proportions were not of the best, but in every case there 
 must be a material gain by supplying the air raised by waste heat at a high 
 temperature * at a rate to suit the conditions prevailing at the time. 
 
 Ellis & Eaves' System differs from the Howden system chiefly from the 
 fact that the products of combustion are drawn at the funnel base and delivered 
 up to the funnel by a fan after passing through the tubes of a heating chamber 
 similar to Howden's. In this case the inflow of air is through the heating 
 chamber and downwards to the furnaces, induced by the suction of the 
 fan, and hence called the " induced draught " system. This system also 
 differs from Howden's in some details. Among others, it was usually fitted 
 with boilers having Serve tubes of considerable diameter, as against the tubes 
 of small diameter fitted with retarders in Howden's case. This system was 
 fitted by Messrs. John Brown & Co., of Sheffield, to several ships, as well as 
 to boiler installations on shore, with beneficial results in all cases. It need 
 hardly be said, however, that the suction fan had to be very carefully designed 
 and fitted to avoid damage from the heat of the gases that pass through it. 
 
 Quantity of Fuel Burnt on the Grate. — With good stoking and the ordinary 
 funnel draught, as much as 20 pounds of coal can be burnt per square foot 
 of grate per hour, and under most favourable circumstances with short 
 grates 22 to 25 pounds. In the mercantile marine, 15 pounds is the 
 average amount burnt when working economically, and all calculations 
 for a merchant ship should be based on this ; for although 20 pounds 
 
 * 300° F is often attaineii when v.-orkihg at full power with a pressure in the ashpits from J to 11 iucb-cs 
 of water. 
 
550 MANUAL OF MARINE ENGINEERING. 
 
 can bo consumed when the fires are forced, or on a very windy day 
 when the draught is good, no grate of ordinary length should be supplied 
 with more than 15 pounds to obtain complete combustion, and this is all 
 that will be consumed on a sultry day. On a short grate, say 13 to 1'5 the 
 diameter of the furnace in length, 20 pounds of coal may. however, be burnt 
 economically. With moderate forced draught, 20 to 30 pounds can be 
 readily burnt on a square foot of grate, and 30 to 50 pounds with good air 
 pressure either forced or induced, and with short grates and good stoking 
 even as much as 60 pounds. A locomotive with a draught produced by the 
 exhaust steam can burn, as a rule, 65 pounds of coal per square foot of grate ; 
 and the modern express engine burns as much as 80 pounds. 
 
 The coal consumed in the old torpedo boats with the closed stokehole, 
 and a pressure of air in the stokehole above that of the atmosphere equal to 
 6 inches of water, was as much as 96 pounds per square foot of grate, and 
 was even 62 pounds at a pressure of 3 inches only. Mr. Thornycroft eclipsed 
 even this in the yacht " Gitana," where 138 pounds of coal were burnt per 
 square foot of grate ; this is exceedingly high, and is only such as can be 
 obtained under the most favourable conditions with every care taken in the 
 ventilating arrangements ; the evaporative power under such conditions is. 
 of course, by no means high, although, perhaps, higher than could have been 
 anticipated. In the Navy an air pressure of \ inch is allowed to count as 
 natural draught, and 2 inches is the limit allowed for forced draught in large 
 ships with tank boilers. With those pressures the combustion is very good 
 and evaporative power very fair. 
 
 Size of Funnel. — The natural draught, or that obtained by means of a 
 funnel, is very much influenced by the area of its transverse section, and by 
 the height or distance of its top from the level of the fire-bars. The draught 
 in a funnel is due to the difference in density of the column of gas in it from 
 that of the surrounding atmosphere ; the contents of the funnel rise upwards, 
 on the same principle that a bubble of air rises through water. The density 
 depends on the temperature of the gases at the funnel base, and for this 
 reason a good draught cannot be obtained without .a comparatively high 
 temperature. The good draught observable on a windy day is due to the 
 tendency to form a vacuum at the mouth of the funnel, by the wind blowing 
 across it. Professor Eankine has given the following formula? for chimney 
 draught : — 
 
 Let iv be the weight of fuel burned in a given furnace per second in pounds 
 V the volume at 32° of the air supplied per pound of fuel. 
 r the absolute temperature at 32° Fahr., which is 461° -f- 32°. 
 r L the absolute temperature of the gas discharged by the chimney, whose 
 sectional area is A ; then 
 
 Velocity of the current in the chimney in feet per second is 
 
 _ w xY X r x 
 A X r 
 
 The density of that current in pounds to the cubic foot is very nearly 
 
 = lo (o-0807 + ^-) ; 
 that is to say, from 0*084 to 0-087 X {r Q -rr,!. 
 
SIZE OF FUNNEL. 551 
 
 Let I denote the whole length of the chimney, and of the flue leading to 
 it in feet ; 
 
 m its " hydraulic mean depth ; " that is, its area divided by its perimeter; 
 which, for a square or round flue and chimney, is one quarter of the 
 diameter ; 
 
 /, a coefficient of friction, whose value for currents of gas moving over 
 sooty surfaces is estimated by Peclet at 0-012 ; 
 
 Gr, a factor of resistance for a passage of air through the grate, and the 
 layer of fuel above it ; whose value, according to the experiments of Peclet 
 on furnaces burning from 20 to 24 pounds of coal per square foot of grate, 
 is 12. 
 
 Then according to Peclet's formula, 
 
 The head required to produce the draught in question is 
 
 -£(i + o + £). 
 
 2g\ mj 
 
 which, with the values assigned by Peclet to the constants, becomes 
 
 IJ? ( 0012 x l \ 
 = TgV 6 + m J' 
 
 When the head is given the value of /& may be calculated, and then, 
 Weight of fuel which the furnace is capable of burning completely per 
 hour 
 
 ft x A x r 
 
 V o x r i 
 It is usual to reckon the head by taking 1 inch of water as the unit ; 
 then, 
 
 Head in inches of water = 0-192 x h x ^ (0-0807 + ^-\ 
 
 Mr. Thornycroft found, by careful experiment with steam launches and 
 torpedo boats having " loco " boilers and working with a plenum (that is, 
 with a closed stokehole into which air is forced), " that of the initial 
 pressure, the resistance of the tubes accounts for about seven-tenths of the 
 whole, the resistance of the fires and fire-bars being only about one-tenth ; " 
 and that " the pressure in the funnel, as measured, was sensibly equal to 
 atmospheric pressure." 
 
 Professor Rankine also stated that if H be the height of the funnel, r 2 
 the absolute temperature of the external air, then — 
 
 Head produced by chimney draught 
 
 = H (o-96 !± - lY 
 
 or, taking h as the head, 
 
 Height of chimney required to produce a given draught 
 
 = h + (o-96^ - lY 
 
 The velocity of the gas in the chimney is proportional to Jh, and 
 
 therefore to v0 - 96 r 1 - r 2 . 
 
 The density of that gas is proportional to — . 
 
552 
 
 MANUAL OF MARINE ENGINEERING. 
 
 The weight discharged per second is proportional to velocity X density, 
 
 n/O-96 r, 
 
 and, therefore, to - — ; which expression becomes a maximum, 
 
 25 Tl 
 
 when r x = — t 2 . Therefore the best chimney draught takes place when 
 
 the absolute temperature of the gas in the chimney is to that of the external 
 air as 25 to 12. 
 
 When this condition is fulfilled h — H. 
 
 That is, the height of the chimney for the best draught is equal to the 
 head expressed in hot gas, and the density of the hot gas is half that of the 
 air. 
 
 Then, if T x be the temperature in degrees Fahr. of the air, the best 
 
 ... .--. 5,993 + 25 T x 
 
 temperature of chimney base is = — 
 
 12 
 
 = 500 + 2-08 T x . 
 
 Then the following holds good when H = 
 
 h 
 
 
 
 
 
 i 
 
 Temperature of external air, 
 
 40° F. 
 
 50° F. 
 
 60° F. 
 
 70° F. 
 
 80° F. 
 
 90° F. j 
 
 i 
 
 Best temperature at funnel base, - 
 
 583° 604° 
 
 625° 
 
 646° 
 
 666° 
 
 687° 
 
 The Size and Height of Funnel are governed in practice rather by cir- 
 cumstances than by scientific investigation ; the diameter is fixed by arbitrary 
 rules based on successful practice, and the height such as suits the appearance 
 or service of the ship. Of late years tall funnels have been the fashion.* 
 
 Rule.— Height of funnel H = -007 (-) ; 
 
 A = 
 
 C x -084 
 x/H 
 
 H is height in feet ; A, the area of section in square feet ; and C, the 
 consumption in pounds per hour of fuel on the grates connected to the chimney. 
 
 Evaporation. — The heat of the gases from the furnace should be absorbed 
 by the surfaces with which they come in contact on their passage to the 
 chimney, and the efficiency of this part of the boiler depends on the capa- 
 bility of those surfaces readily to take up the heat, on the material to transmit 
 it by conduction to the inner surface or that with which the water is in 
 contact, and on that inner surface being in such a condition as to give up 
 the heat to the water. The internal efficiency of the boiler depends on the 
 convection or circulation of the water in the boiler, whereby fresh portions 
 are successively brought in contact with the hot surlaces. The importance 
 of this latter factor was seldom sufficiently appreciated in estimating the 
 efficiency of a boiler, although now engineers give it first consideration. 
 
 When the furnace is internal — that is, when it forms a part of the boiler 
 proper and is surrounded with water — a large proportion of the total heat 
 of combustion is absorbed by it, partly by direct contact with the hot fuel 
 at the sides, and partly by radiation from the glowing surface of the incan- 
 descent fuel when coked. The furnace also absorbs heat from the hot gases 
 passing along its surface. 
 
 * The funnels in Atlantic express steamers are now 100 to ISO feet high above the grate levels, oval 
 in cross-section and as large as !*4 feet x 17 feet. 
 
HEATING SURFACE. 553 
 
 The furnace or fire-box of a locomotive boiler is usually made of copper, 
 and it has been proved by experiment that a very large proportion of the 
 whole heat generated in it is absorbed by it, consequently its evaporative 
 efficiency is very high. The furnace of the ordinary marine boiler is ol 
 steel, which is not so good a conductor of heat as copper, and, therefore, 
 does not transmit such a large proportion of the total heat of combustion, 
 although its evaporative power is still high. (For conductivity of metals 
 see Chap, xxx.) 
 
 Combustion. — The combustion of the fuel is not always comjAeted in the 
 furnace of the marine boiler, as the gases distilled from it during the process 
 of coking escape to the chamber beyond the furnace before sufficient air 
 has been supplied to complete combustion ; and also it sometimes happens 
 just after firing that the temperature above the fuel is not sufficiently high 
 to cause ignition. Moreover, the carbon often unites in the furnace with 
 only sufficient oxygen to form carbon monoxide, and this flows unchanged 
 into the combustion chamber : if, then, a further supply of air is found 
 there, another portion of oxygen is taken up, and carbon dioxide or carbonic 
 acid gas is formed ; that is what usually takes place with bituminous coal, 
 and unless this second supply of warm air is provided, the combustion of a 
 large portion of the fuel is not completed in the boiler. If the fire is com- 
 pletely covered with green — that is, fresh fuel — the part next the hot surfaces 
 gives up its volatile elements, which rise and become cooled in expanding, so 
 that when they come in contact with the oxygen of the air, the temperature 
 is not high enough to cause them to chemically combine — that is, to ignite, 
 and consequently they merely mix mechanically and flow on until they 
 pass out at the mouth of the funnel and appear as smoke. 
 
 If the smoke-boxes leak air, as they only too often do, ignition will some- 
 times take place in them, and flame will ascend the funnel and appear at 
 the top, where there is a further supply of air sufficient to complete com- 
 bustion. 
 
 When, however, by careful stoking and regulating the supply of air. sc 
 that combustion is completed in the combustion chamber, the evaporative 
 power of that part of the boiler is high, and with the furnaces is the most 
 valuable part of the boiler for transmitting heat to the water. 
 
 Heating Surface. — As has been stated, the efficiency of the heating surface 
 of the boiler depends on the material, its thickness, and on the state of the 
 surfaces in contact with the hot gases and water. The furnaces and com- 
 bustion chambers are of steel, whose conductivity is inferior to that of copper 
 (16 to 91), but is still good. 
 
 The superior evaporative power of the furnace is due in great measure 
 to the cleanness of the surface exposed to heat ; there is no deposit of soot 
 or ash on it, and the smallest possible amount of oxide ; the combustion 
 chamber also is generally in the same condition. The roughness of the 
 surface exposed probably increases its power of receiving heat, not so much 
 from any abstract virtue in that state, as from the actual surface being greater 
 than if smooth. Very much depends on the condition of the inside surface 
 exposed to the water ; if it is quite clean and absolutely smooth, it is not so 
 efficient as slightly dirty and rough ; the best condition being roughness 
 with freedom from coating of bad conductors. If a metallic surface is 
 smooth and clean, evaporation from it is slow and intermittent, because on 
 
554 MANUAL OF MARINE ENGINEERING. 
 
 it is formed a film of steam, which is a bad conductor, and this will only 
 disperse when its buoyancy overcomes the attraction ; when the attraction 
 is overcome, the steam rises suddenly en masse, the surrounding water then 
 flows into its place, when a film is again formed, and a bubble accumulates 
 as before. If, on the other hand, the surface be rough, the film is broken 
 up by numerous points on the surface, which serve as accumulators and 
 starting points for myriads of small bubbles ; these are formed quickly, rise 
 freely and continuously, giving a rapid and steady supply of steam. It is 
 for this reason that many boilers prime — that is, work with violent ebullition 
 — when quite new 
 
 Tube Surface. — The tubes of a marine boiler represent the greater portion 
 of the total heating surface, and it is to this part that great attention is 
 required, both in designing and working a boiler. They are usually made 
 of iron in the mercantile marine, and of steel in naval boilers. The con- 
 ductivity of the brass, which was formerly used in the Navy, is theoretically 
 double that of the iron (30 to 16) ; but in practice with foul surfaces, there is 
 not really so great a difference, their relative value being then about 3 to 2 ; 
 to obtain, therefore, results commensurate with their extra cost, brass tubes 
 had to be kept very clean always. 
 
 Much stress is sometimes laid on certain experiments which show the 
 relative evaporative power of certain portions of the tubes, to prove thereby 
 that the final foot, or two feet, of tube is practically useless in every boiler. 
 It is not very surprising to find that the first foot of tube has a high evapo- 
 rative power, and that the power decreases rapidly the farther removed the 
 portion is from the furnace or combustion chamber. Considering that the 
 temperature of the gas on entering the tube is 2,000° to 2,400°, and on leaving 
 it should be not more than 600°, while the temperature of the water sur- 
 rounding the tube is very nearly the same (386° F.) at every part, the trans- 
 mission of heat should be very much greater at the entering end, where the 
 difference is 1,700° to 2,000°, than at the other, where it is only 220° ; more- 
 over, the end next the funnel is more liable to get dirty from deposit of soot, 
 etc., and so to rapidly fall off in efficiency. But in continuous practice it 
 was not found that there was such a large difference in the value of the two 
 ends of the tubes, because the end at which there is the rapid evaporation 
 soon .became covered with scale, and its efficiency reduced when salt feed 
 water was used ; the gas temperature was then not so much reduced after 
 passing through the first portion, and for this reason the last portion was found 
 to have a higher efficiency than before. The size of tube has also some 
 influence on its efficiency, for whereas the surface increases as the diameter, 
 the contents increase as the square of the diameter. If a tube, then, of 
 4 inches diameter be substituted for two of 2 inches diameter, the surface 
 is the same, while the cubic capacity is doubled ; if the rate of flow be the 
 same in both cases, the 4-inch tube will pass twice the quantity of hot gas 
 through it that flows through the two 2-inch ones, and have only the same 
 surface to absorb and transmit the heat. If the quantity flowing through 
 the tubes be the same in both cases, the 4-inch tube is still at a disadvantage, 
 inasmuch as the mean distance of the gas from its surface is greater than 
 that in the 2-inch ones. The velocity through the small tubes will, in the lattei 
 case, be double that in the 4-inch one, and will therefore cause a brisker circu- 
 ation of the hot gas, and the liability of soot and ash deposit is considerably 
 
EFFICIENCY OF A BOILER. 555 
 
 reduced. Hence tubes of smaller diameter are used with advantage with 
 forced draught, for the same evaporation is effected with smaller amounts 
 of surface. The " Serve " tube made by J. Brown & Co., having 6 or 8 ribs 
 projecting inside, has the great advantage of large absorbent surface in pro- 
 portion to its cubic capacity. The twisted strips of iron placed inside the 
 tubes by Mr. Howden, and called " retarders," have the effect of reducing 
 the cubic capacity as well as circulating the hot gases ; the retarders, however, 
 can transmit very little of the heat they absorb, in the way the ribs of 
 the Serve tubes do. 
 
 Evaporative Power. — The probable evaporative power of a boiler may be 
 found approximately by the following formula : — 
 
 Let e x be the theoretical evaporative power of the fuel, F the weight of 
 coal burned on the grate in pounds per hour, and K the total heating surface 
 in square feet ; then 
 
 Pounds of water evaporated per pound of fuel burnt 
 
 = 1 ' 833 (nrFr) *■: 
 
 Example. — To find the evaporative power of a boiler which burns a fuel 
 whose theoretical evaporative power is 15, the number of pounds burnt per 
 hour on the grate is 800, and the total heating surface is 1,000. 
 
 The efficiency of the boiler is, by this rule, 
 
 1,833 (2KTf)' OT ' 655 - 
 
 This rule is only approximate, however, because the heating surface is 
 not always wholly real heating surface, some being only nominal ; many 
 boilers as made formerly have given better results after the removal of tubes, 
 which very materially decreased the nominal heating surface. If a boiler 
 has just enough surface to absorb heat from the gases, so that the temperature 
 at the funnel base is only such as is sufficient to produce the required draught, 
 then the heating surface is effective ; any surface added to this is superfluous, 
 and in very many cases does positive harm. 
 
 The Efficiency of a Boiler under various conditions of working can be 
 ascertained as follows : — 
 
 W is the weight of water actually evaporated per pound of fuel. 
 C ,, calorific value of the fuel used in British thermal units. 
 H ,, total heat absorbed in raising a pound of steam from a pound of 
 
 water at 32° F., and h that of the feed-water. 
 T ,, temperature of the steam produced in degrees Fahr. 
 
 If superheated steam is produced at a temperature T s , so that the amount 
 of superheat is T g — T, the additional H s will be found by multiplying 
 (T s — T) by the specific heat of steam. 
 
556 MANUAL OF MARINE ENGINEERING. 
 
 W (H — h) 
 
 (1) Efficiency of the boiler = — —^ * 100 per cent. 
 
 (2) Efficiency of superheating = — ^— -' x 100 per cent. 
 
 The factor of evaporation = 
 
 (H+H«-fi) 
 
 966 
 
 The specific heat of steam at, say, 200 lbs. absolute and superheated to 
 480° F. is 0-59, and if to 600° F. it is 0*545. This is the mean specific heat 
 between 381° and 480° and 600° F., and is the ratio of it to the heat required 
 to raise a pound of water by the same amount. Taking the British thermal 
 unit as 778 foot-lbs., it requires on the average 459 foot-lbs. per degree to 
 superheat a pound of steam of 200 lbs. pressure to 480° F. ; or, in all, 
 (480 - 381) X 0-59, or 58'3 B.T.U. 
 
 The temperature at uptake depends on the temperature of the water in 
 the boiler ; for instance, at 300 lbs. pressure, it will be 422° F. ; so it will 
 be no use to try to bring the temperature of the gases below 400° or even 
 450°, and with such a temperature there the efficiency of the tubes will be 
 very low at 200 lbs., inasmuch as the water then is 387° F. It is, therefore, 
 manifest that it is better to have a higher temperature at the smoke-box, and 
 reduce at funnel base by air-heaters, feed-heaters, or super-heaters to dry the 
 steam. 
 
BOILERS. 557 
 
 CHAPTER XXII. 
 
 BOILERS — TANK BOILER DESIGN AND DETAILS. 
 
 There are iu use on shipboard to-day two distinct types of boiler, viz. : — 
 
 (1) The Tank Boiler, consisting of an external shell, in which is contained 
 the water and the means whereby that water may be converted into steam ; 
 it also has within it a space for the temporary storage of the steam when 
 formed ; and 
 
 (2) The Water-Tube Boiler, consisting of a structure of tubes joined 
 together, intercommunicating, and connected to a receiver at the top, and 
 generally to water chamber or chambers at the bottom. The water is in this 
 class always within the tubes, and circulates when heat is applied to their 
 surfaces. The fire is generally on a grate immediately below the nests of 
 tubes, so that the hot gases ascend and pass through them. 
 
 The Tank Boiler to-day exists in various forms, but the shell is always 
 cylindrical or nearly so. The furnaces within it are always cylindrical in 
 general form, but they are often stiffened to resist collapse by corrugations 
 or equivalent means. The furnaces deliver the hot gases into a chamber, 
 within which combustion may be completed, and the contents distributed 
 over a large number of small tubes, through which they are conveyed to the 
 smoke-box, and by which their heat is absorbed on the way there, so that 
 they play the part of a surface condenser to them. They pass the heat 
 into the water, whereby it is heated and some of it converted into steam. 
 In the past there have been many forms of tank boiler other than the cylin- 
 drical one, of which 
 
 (a) The Oval Boiler, as shown in fig. 204, which, when side room is limited, 
 permits of a larger and more powerful boiler than if cylindrical ; and although 
 not so light per unit of power as a cylindrical, it preserves most of the good 
 features of that form, together with a simple system of staying whereby 
 strength is obtained. 
 
 (6) The Vertical or Haystack Boiler, having a cylindrical shell with a 
 hemispherical top, as in fig.' 213. Or it may have a flattened or oblate'spherical 
 top with the funnel passing through it, as in fig. 212. 
 
 (c) The Locomotive Boiler in a modified form was used very generally 
 for small craft, and even for destroyers of 3,500 I.H.P., but they were not 
 altogether successful, especially under severely forced draught. The form 
 of locomotive boiler, as shown in fig. 214, was, however, quite satisfactory 
 in torpedo gunboats of considerable power with a fairly high air pressure. 
 
 The Boiler used in the Mercantile Marine is to-day almost exclusively 
 of the tank type, and simply cylindrical in form, made of steel, and fitted 
 with iron or steel lap-welded tubes. The furnaces are often, in small ships, 
 plain cylinders welded and flanged to meet the combustion chamber in such 
 
•558 
 
 MANUAL OP MARINE ENGINEERING. 
 
 a way that they can be passed through the hole, provided for them in the 
 front plate of the boiler (v. fig. 206). In larger and more important ships, 
 Avhere supervision and care is better observed, the furnaces are of the corru- 
 gated or ribbed variety. 
 
 Cylindrical Boilers with Two Furnaces, as in fig. 205, are used in small 
 
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 shipe where one such is sufficient, and in other shallow ships having more 
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 yachts and some passenger steamers ; more than one boiler is often decided 
 oil as a means to ensure immunity from disablement in case of derangement 
 of one ol them. Two furnace boilers may be made of considerable diameter 
 
CYLINDRICAL BOILERS. 
 
 559 
 
 if required, as the furnaces as now made may be large— as much as 54 inches 
 in diameter for moderate pressure, and 50 inches for 200 lbs. per square 
 inch. 
 
 
 
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5*0 
 
 MANUAL OF MARINE ENGINEERING. 
 
FOUR-FURNACE CYLINDRICAL BOILER. 
 
 561 
 
 fig. 206, end view ; but it has been made often in the past with one chaml er 
 common to the three. 
 
 Cylindrical Boilers with Four Furnaces may be as shown in fig. 207, with 
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 MANUAL OF MARINE ENGINEERING. 
 
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THE DOUBLE-ENDED ROILER. 
 
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564 
 
 MANUAL OF MARINE ENGINEERING 
 
 
 
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THE WATER-TUBE BOILER. 565 
 
 fig. 209 ; or there may be, and often was, to save weight and space, one 
 chamber common to four furnaces, as shown in fig. 208. Double-ended 
 boilers are now made of enormous size, as much as 18 feet diameter and 
 as large as 24 feet long. 
 
 The double-ended boiler is lighter and cheaper to make per unit of power 
 than a single-ended, and occupies less space. Being of smaller diameter 
 than a single-ended one of equal power or heating surface, the shell can be 
 made of thinner plates, which for high pressures used to be a matter of serious 
 consideration ; but to-day, when steel makers can and do roll plates of any 
 thickness up to 2 inches of large area at moderate costs, and most boiler- 
 shops are equipped with tools capable of handling and dealing effectively 
 with such plates, this arrangement has little or no weight. But the straining 
 effect of high temperature, differences of temperature, and changes of tem- 
 perature are as severe and active now as they ever were ; and although the 
 care and methods of the modern boiler-maker are beyond all praise, and 
 produce work far superior to that of a few years ago, the effects of racking 
 are liable to appear, although they are not so evident in the form of leaks 
 as were only too often the case formerly ; it still remains, therefore, that 
 the single-ended boiler can be trusted to withstand rougher usage than the 
 double, and is, therefore, more generally used in the mercantile marine than 
 formerly. Moreover, since quite large units of the single-ended type are 
 made now, there is not the same necessity for the double-ended in full- 
 powered express steamers as obtained formerly ; nevertheless, in these 
 latter the double-ended boiler is still frequently used to economise space 
 and weight. In the Navy the double-ended boiler was never a favourite 
 with the engineers responsible for them, and although largely used in cruisers 
 on account of their lighter weight from 1887 to 1897, they have since been 
 seldom used, and latterly the tank boiler has been abandoned altogether. 
 There was, however, a special form of cylindrical boiler introduced into the 
 Navy so far back as 1854, which has survived, and has been often adopted 
 for use in express steamers, and known as 
 
 The Gunboat Boiler (fig. 211), with the furnaces at one end and the tubes 
 at the other, with the combustion chamber between them, is a very 
 efficient boiler, and the lightest cylindrical one per unit of power. It is a 
 very convenient design for shallow ships having plenty of fore and aft room 
 to stow the boiler in. 
 
 The Naval Boilers of to-day are. however, almost exclusively single-ended 
 cylindrical and water-tube boilers of sorts ; in small ships exclusively the 
 latter. In larger ships the rule was to have a portion of water tube with a 
 portion of cylindrical, the latter being used for cruising at slow speeds, and 
 the former added to them when high power or full speed is required ; now 
 water-tube boilers only are used. 
 
 In the Mercantile Marine the Water-tube Boiler has practically found 
 no favour in this country, and not very much elsewhere ; it has undoubtedly 
 some good claims for serious and favourable consideration, especially for 
 certain services, and it is very remarkable that on shore the very large instal- 
 lations at electric generating stations water-tube boilers are almost exclusively 
 used, in spite of the fact that electrical engineers are very keen observers 
 of fuel consumption, and that they have had experience with quite good 
 forms of the cylindrical boiler at some of these stations. It is, of course, 
 
566 
 
 MANUAL OF MARINE ENGINEERING. 
 
 admitted that in the ordinary merchant steamer on long voyages there is 
 for days no variation in the demand for steam, and from noon to noon hardly 
 a half revolution difference in speed, whereas at an electric light station 
 the demand for steam is always changing, and a sudden fog may make a 
 sudden demand for an enormous increase in it, which necessitates the forcing 
 
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 of such boilers as are on service, as well as the addition to their number as 
 rapidly as they can be possibly got into action ; in both cases the boilers are 
 severely tried, and only those which can withstand the racking due to sudden 
 changes of temperature can be employed. 
 
THE DOUBLE-ENDED BOILER. 567 
 
 Small Boilers have usually two furnaces, and with this number are more 
 efficient than with three even when of moderate size. 
 
 The number and size of the furnaces must, however, depend on the size 
 of the boiler and the heating surface it is to contain. It is found in practice 
 that large furnaces are more efficient as coal consumers than small ones, and 
 the reason is not far to seek. The grate area with the same length of fire- 
 bar increases as the diameter, while the section through which the air passes, 
 both above and below the bars, increases as the square of the diameter ; it 
 is also possible to give a good inclination or rake to the bars with a large 
 furnace, which very materially assists combustion. In practice the grates 
 are not of course always of the same length, but they do not increase in 
 length as the furnace does in diameter, and consequently the air passages 
 increase more rapidly than does the grate area when the diameter of furnace 
 is increased. Furnaces should be not less than 36 inches, nor more than 
 48 inches in diameter, except under special circumstances. Taking this as 
 a rule for guidance, boilers may be made up to 9 feet diameter with one 
 furnace, up to 13 feet 6 inches diameter with two furnaces, up to 15 feet 
 with three furnaces, and beyond that diameter four furnaces are necessary 
 to avoid too long length of grate. 
 
 A single-furnace boiler has of course one combustion chamber, a two- 
 furnace boiler may have one chamber common to the two furnaces (fig. 205), 
 or a separate one to each. When there is only one boiler in the ship the 
 latter plan is preferable, as then the bursting of a tube cannot wholly dis- 
 able the boiler ; when there are two or more boilers one chamber common 
 to the two furnaces is preferable, as by stoking the fires alternately an even 
 supply of steam is kept up and the smoke consumed. A three-furnace boiler 
 has usually three separate combustion chambers, and the same remark applies 
 to it as to the two-furnace boiler. The four-furnace boiler has generally only 
 two combustion chambers, one wing and one middle furnace having a com 
 mon chamber ; but some engineers prefer three chambers, the two middle 
 furnaces having one in common, and each wing furnace a separate one. 
 
 The chief objection to two large furnaces instead of three smaller, and to 
 three larger ones instead of four smaller, is the longer grate required to get 
 the requisite area, and to the large amount of dead water between the furnaces 
 at the bottom. There is also to be considered the limit placed by the Board 
 of Trade and Registry rules to avoid risk, of collapsing by direct crushing 
 of the metal, which prevents the adoption of the larger furnace with very 
 high pressures. 
 
 It is unusual and certainly most difficult to use plates above If inches 
 thick in the construction of a boiler shell, and it is this consideration which 
 often fixes the limit of diameter. Boiler-shop appliances are now made to 
 deal with even thicker plates when necessary. Boiler shell plates up to 
 2 inches thick can be obtained 12 feet wide and of 250 square feet area at quite 
 moderate prices in mild steel from 27 to 36 tons ultimate tensile strength. 
 
 The Double-ended Boiler (fig. 209) has furnaces at both ends with return 
 tubes over them, and is generally tantamount to two single-ended boilers 
 back to back, but with the backs removed. It is made up to 18 feet diameter 
 and as much as 24 feet long ; but such very large boilers are unusual, partly 
 owing to the want of facilities for moving such great weight, and partly 
 because the conditions under which such large boilers are possible are limited 
 
568 MANUAL OF MARINE ENGINEERING! 
 
 % 
 
 to very large steamers. For the lower pressure of steam for turbines such 
 large boilers are now frequently supplied. 
 
 The double-ended boiler is lighter (vide Table Hi.) and cheaper in pro- 
 portion to the total heating surface than the single-ended boiler, and its 
 evaporative efficiency in practice is generally higher. On the other hand, 
 greater care is necessary in designing and in working it. That it is lighter 
 is obvious, and that it is cheaper may be inferred from the fact that there 
 is less material, and less labour consequent on the reduced quantity of 
 material. 
 
 The simplest form of this kind of boiler is one in w 7 hich all the furnaces 
 open into one common combustion chamber (fig. 208) ; this form, although 
 at one time common enough, is now seldom adopted. The objections to 
 it are, that the bursting of one tube may disable the whole, that the cleaning 
 of one fire causes the efficiency to sink very low on account of the whole 
 being effected by the inrush of cold air, and that unless special means be 
 provided to promote proper circulation there is a strong tendency to prime. 
 Some of these objections are got over by building a thin brick wall with 
 special bricks across the middle. Mr. Howden recommended this arrange- 
 ment with his system of forced draught, such being necessary whenever 
 there is a chamber common to opposite furnaces. 
 
 The next simplest form is one in which opposite furnaces have a com- 
 bustion chamber in common (fig. 209) — that is, it differs from the first by 
 having the combustion chamber divided longitudinally by water spaces. 
 This avoids the chief objections raised against the first form while retaining 
 its chief advantages, which are, simplicity of construction, by discarding the 
 flat back of the combustion chambers, with the necessary stays, etc., and the 
 greatest heating surface within the smallest limits of length. It is often 
 urged against this form of boiler that the tubes are very liable to leakage at 
 their back ends, arising from the rush of cold air against the tube plates 
 when the door of the furnace opposite it is open, causing it to buckle. It 
 sometimes happens that the tubes in this kind of boiler do show a tendency 
 to leak, but it is then generally due to the want of expansion on the part 
 of the first row of stays above the combustion chamber when they are placed 
 too close to the tubes. If these stays are at least 12 inches above the tubes, 
 so as not to hold the front tube plates too rigidly, then when steam is being 
 got up the expansion of the tubes simply causes the plates to spring very 
 slightly instead of to start their ends and cause them to leak. The leakage 
 from springing of the tube plate from exposure to cold air can only take 
 place when the combustion chamber is unduly short, and when there is an 
 insufficient number of stays to the tube plates. 
 
 This particular form of boiler is very generally used for large power and 
 high speed ; the evaporative results obtained from it are most satisfactory. 
 Care, however, is required in raising steam, and the opening of fire-doors 
 to check evaporation is a reprehensible practice at all times for them, as 
 it is indeed for all boilers. A brick semi-partition in the middle of the com- 
 bustion chamber will prevent the cold air* rushing on to the opposite tube 
 plate, and it acts also as an equaliser of temperature in the combustion 
 chamber at all times. If, however, the combustion chamber is too small, 
 this will only magnify the defect by causing intense local heat, and thus 
 tend to crack the plates. 
 
DRY COMBUSTION CHAMBER BOILER. 569 
 
 Another common form of double-ended boiler has the furnaces at one end 
 with one chamber common to them, and those at the other end with another 
 chamber in common. The boiler is then longer than either of the other 
 forms, and more expensive ; the combustion chambers have large flat backs, 
 requiring a very large number of stays, which prevent their being properly 
 cleaned from scale. 
 
 The last form, which is by far the most expensive and heaviest, but now 
 often adopted, is one in which each furnace has an independent combustion 
 chamber. There is little need of description, as it is to all intents and pur- 
 poses two single-ended boilers, except that the water and steam are common 
 to the two parts.* 
 
 Oval Boilers are included under the generic term of cylindrical, as they 
 partake of the principal features of that class. The transverse section is, 
 however, not an ellipse, but is really formed by two semicircles with a rect- 
 angle intervening. The flat sides thus left between the semi-cylinders require 
 staying, the first rows being at the commencement of the flat. There are 
 both single- and double-ended oval boilers, which for pressures under 120 lbs. 
 may be made both simply and economically to very large sizes, as the thick- 
 ness of shell plate depends on the diameter of the cylindrical part. Two 
 very large furnaces may be thus fitted into a cylindrical part of comparatively 
 small diameter, sufficient heating surface being obtained by giving the requisite 
 height. This form is most convenient when the boilers have to be stowed 
 fore and aft, and the diameter is limited by the breadth of the ship between 
 the stringers ; also when forced draught is employed they permit of the large 
 amount of heating surface in proportion to the grate area which is necessary 
 for economic results. 
 
 Holt's Boiler. — An ingenious form of double-ended boiler, used formerly 
 by the late Alfred Holt, consists of two oval end parts united by a cylindrical 
 part whose axis passes through the upper focus of the oval ; in the bottom of 
 the end parts are the furnaces, which are each connected by a large tube to 
 a combustion chamber in the middle of the cylindrical part, from which the 
 tubes extend to the front above the furnaces, and are consequently much 
 longer than usually found in a marine boiler. These boilers are made of 
 great length, and practice proved them to be very good and efficient steam 
 generators, lasting very much longer than the ordinary double-ended boiler, 
 being more elastic and the middle portion raised far above the influence of 
 bilge-water. 
 
 Dry Combustion Chamber Boiler. — The boiler in this case is a simple 
 cylindrical or oval shell, having the furnaces extending from end to end, 
 and the tubes over them likewise from end to end. The combustion chamber 
 is external, and does not form an integral part of the boiler ; but is built 
 of brickwork so as to enclose one end, and form a connection between the 
 furnaces and tubes. If two such boilers are placed back to back and some 
 distance apart, and the space between them enclosed by brickwork, a double- 
 ended arrangement is formed similar to the first described, except that the 
 water and steam are not in common to the two parts, except by means of 
 special connections. Such a boiler is very cheap to manufacture, as the 
 wet combustion chamber with its flat sides and stays is avoided ; the ad- 
 vantage of two boilers without their cost is obtained, and the evaporative 
 
 * The naval practice is to have a separate chamber to each furnace in all boiler?, 
 
570 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Oo**n Cast 
 
 Dottn Cast' 
 
 Down Cos? 
 
 Fig. 2 i 2. — Scotch Haystack Boiler. 
 
VERTICAL CYLINDRICAL JiOILER. 
 
 571 
 
 efficiency is very high. The intense heat from the brickwork, however, is 
 liable to crack the tube plates, unless care is taken ; but if ample space is 
 provided in the combustion chamber, there is not so much danger of this, or 
 of damaging the tube ends. 
 
 Such boilers are now, however, not so much in use ; although, among 
 other things, they accomplish a satisfactory consumption of smoke, and 
 a very even rate of evaporation. Their reputation, however, has been on 
 some occasions somewhat damaged by a too intemperate advocacy, and by 
 improper construction and design of the combustion chamber when attempting 
 too much with them. 
 
 For pressures above 100 pounds in ships liable to rough usage, such 
 boilers may be adopted with advantage on account of the absence of the 
 internal combustion chamber. 
 
 Gunboat Boilers (fig. 211). — This type of boiler, first adopted by the 
 Admiralty for gunboats, and afterwards chosen for corvettes, is something 
 between the locomotive and cylindrical boiler. The shell is cylindrical, and 
 contains the furnaces at one end and the tubes at the other, the combustion 
 chamber being in the middle between them ; the top of the furnaces is. 
 therefore, level with the top of the tubes, and no part of the heating surface 
 is far removed from the water level. The flame and hot gases flow from 
 the furnaces into the combustion chamber, are there sometimes slightly 
 diverted and spread by means of a hanging bridge, and flow onward with 
 only this slight interruption into and through the tubes to the smoke-box. 
 It is not surprising, then, that this boiler burns its coal freely, and evaporates 
 very quickly and efficiently. Two such 
 boilers, having a total grate area of 68 
 square feet, and a total heating surface 
 of only 2.200 square feet, supplied 
 steam to triple - compound engines de- 
 veloping over 1,340 I.H.P. The coal 
 consumed was about 36 pounds per square 
 foot of grate per hour ; the weight of water 
 evaporated by one pound of coal, at a 
 somewhat reduced speed, was about 9J 
 pounds. 
 
 The chief objection to this class of boiler, 
 which prevents its more general adoption in 
 the mercantile marine, is the great length 
 required ; a space is necessary at the back 
 end to get at the tubes, and to admit of 
 the smoke-box, etc. ; and the total heating 
 surface is also very small for the space occu- 
 pied by it. It is, however, very convenient 
 for shallow ships, when it is necessary to 
 have clear decks, and has consequently been 
 used for modern express river steamers, as 
 well as gunboats. 
 
 Vertical Cylindrical Boiler.— This kind of 
 boiler, with many variations of internal arrangement, is used for auxiliary 
 purposes, and then generally called the donkey boiler. On a large scale it 
 
 Fig. 213.— Cochran's Boiler. 
 
572 MANUAL OP MARINE ENGINEERING. 
 
 was much used by Scotch engineers for river steamers, where rapid evapora- 
 tion is of more importance than economy of fuel. It is light and inexpensive 
 in proportion to the grate and heating surface, and occupies a small amount 
 of floor space, capacity being obtained by the height. Fig. 212 is a modern 
 example of such a boiler as has been made for large power. Fig. 213 is a 
 design of vertical cylindrical boiler as invented by Mr. Cochran, now of Annan, 
 N.B., and made by him in large numbers for the auxiliary service of large 
 ships and the main boiler of small ones ; "it is a very convenient form for 
 these purposes. 
 
 Locomotive Boiler. — This boiler has a fire-box of rectangular section, 
 both horizontally and vertically, enclosed in a shell of somewhat similar 
 shape, except that the top is often semi-cylindrical. The tubes are contained 
 in a cylindrical barrel, extending horizontally from the fire-box shell, and 
 at the end of this is the smoke-box. As the name implies, it is similar to 
 the boiler of the ordinary locomotive. 
 
 This form of boiler is only employed in the steam launches of the mer- 
 cantile marine, but was used in the Navy very extensively for torpedo boats 
 as well as launches, and was even adopted on a large scale in torpedo gun- 
 boats, as well as in some of the early destroyers. It is a very convenient 
 form for the naval service, as it is, with the exception of water-tube boilers, 
 the lightest kind of boiler for the heating surface contained ; and as it is 
 invariably used now with an artificial draught, the smallness of the grate 
 area is no detriment to it. It is also especially well adapted for high pressures, 
 as the flat surface can be stayed without affecting the accessibility, and 
 the cylindrical barrel is of such small diameter, as to be made of very light 
 plates for even very high pressures. Much difficulty is, however, found in 
 keeping the tube ends tight, especially when the draught is forced much ; 
 and when pressed to their utmost capability with engines having large 
 cylinders, they are very liable to prime excessively. The only part which 
 presents any difficulty of construction is the furnace crown, which, being 
 flat, requires an extensive amount of stays, etc., and as the evaporation 
 is very rapid from it. there is great liability of a heavy scale being formed 
 if any sea or "hard" water is used, which prevents the heat from passing 
 to the water, and causes it to destroy the plates. 
 
 It has the disadvantages of the gunboat boiler, and on that account is 
 not likely to be used in the merchant service, except in very light-draught 
 river steamers in the colonies and foreign countries, where it has for many 
 years done good service. 
 
 Wet-Bottom Locomotive Boilers. — The success of the locomotive boiler 
 for naval small craft caused the authorities to adopt them for a bigger class 
 of ship, as well as for a large number of torpedo gunboats. A modification 
 of the design, however, was made, which converted it into a wet-bottom 
 boiler, the air being admitted beneath the bars through an aperture in the 
 front, as in the furnaces of the ordinary boiler. The change, however, was 
 a most unfortunate one, as, with the exception of the design as fitted by 
 Laird Brothers in H.M.S. " Rattlesnake," this form of locomotive boiler 
 proved troublesome and inefficient. The circulation is bad, so that the tube- 
 ends at the fire-box are so seriously affected as to cause considerable leaks 
 and general trouble. The tubes are much nearer to the fire than in the 
 ordinary locomotive boiler, and, consequently, become more readily choked ; 
 
DOUBLE- ENDED LOCOMOTIVE BOILERS. 
 
 573 
 
 in iact, after some hours' 
 steaming, owing to the sooty 
 deposit in the iorm of a bird 
 nest at the end of each tube, 
 the draught is so seriously 
 impeded as to render tbo 
 boiler practically useless. 
 
 The wet - bottom boiler, 
 however, had certain advan- 
 tages over the dry-bottom, 
 and, when carefully designed, 
 may be fairly satisfactory. 
 Fig. 214 shows the boiler as 
 fitted in the "Rattlesnake.'' 
 whence it will be seen that 
 the flat tubes connecting the 
 top to the bottom of the fire- 
 box provide a splendid means 
 for circulation, and form a 
 good support to the roof. The 
 fire-bars must be lowered, at 
 the back ends at least, so as tc 
 be in the same relative poei- 
 tion as in the dry -bottom' 
 boiler. The fire-box itself may 
 be divided into two indepen- 
 dent parts by a longitudinal 
 water-space, instead of the 
 row of vertical tubes connect- 
 ing the top and bottom, as in 
 fig^ 214. By either method an 
 increase of heating surface is 
 made in the fire-box, and a 
 better means of circulation 
 provided, as also a good 
 supply of water to the top of 
 the fire- box ensured. 
 
 Double-ended Locomotive 
 Boilers.— This form of boiler 
 consists of a fire-box with a 
 tube-barrel at each end, which 
 is to the ordinary locomotive 
 boiler what the double-ended 
 cvlindrical boiler is to the 
 single-ended. The firing-holes 
 are on either side, and in the 
 case of a big boiler they will 
 be two in number, so • that 
 through the four the grate can 
 be well worked. Moreover. 
 
 mmggggj bt 
 
574 MANUAL OP MARINE ENGINEERING. 
 
 from their position and size, the opening of them does not so seriously 
 affect the tube-plates as with the ordinary locomotive boiler. This form 
 of boiler is much lighter than the ordinary one, and is convenient in many 
 ways, especially for vessels of limited beam. 
 
 Seaton's Locomotive Boiler. — By making the tube barrel of the locomotive 
 boiler a truncated cone instead of* a cylinder, and arranging the tubes so that 
 the top rows are nearly horizontal, while the others slant more and more 
 as they approach the bottom, a better means for circulation is provided by 
 the wider spacing of the tubes at the fire-box and the declination of the boiler 
 barrel at the bottom, while, by the inclination of the tubes upwards, a better 
 draught is ensured. 
 
 Dimensions of a Boiler. — The amount of grate area is the consideration 
 which chiefly affects the choice of dimensions of boiler, and to a very large 
 extent the number and form of the boilers also are governed by it. 
 
 The cylindrical boiler is not elastic in the hands of the designer : to 
 increase the number of furnaces in it its diameter must be increased, which 
 means that both breadth and height are affected. If two furnaces of 40 
 inches diameter be the limit for a boiler of 10 feet in diameter, that there 
 may be adequate heating surface, and 14 feet is the suitable diameter for 
 three furnaces of 40 inches diameter, the grate bars being of the same length 
 in both cases, the increase in boiler capacity is 96 per cent, for an increase 
 pf 50 per cent, of grate. The smallest diameter of shell into which three 
 40-inch furnaces can be fitted so as to give adequate heating surface, is 
 13 feet 6 inches, which is an increase in capacity of 82 per cent, over the 
 boiler 10 feet in diameter. Four 40-inch furnaces require a shell of at least 
 16 feet diameter, which means an increase of 156 per cent, capacity to obtain 
 100 per cent, increase of grate. To arrange four 40-inch furnaces so as to be 
 convenient for stoking, a shell of 17 feet diameter is required, which means an 
 increase of 189 per cent, over the shell of 10 feet diameter ; if. instead of 
 increasing the number of furnaces by increasing the diameter of shell, the 
 number of shells be increased, the space occupied is considerably in excess 
 of the direct ratio of grate areas. 
 
 It is true that to some extent increase of grate area may be obtained 
 by increasing the length of furnace, but the efficiency of a grate in practice 
 is nearly inversely as its length ; for a long grate cannot be nearly so well 
 attended to as a short one, nor is the air supply either under or over the 
 bars so good with a long furnace, since the area of section at the mouth, 
 with the same diameter of furnace, is the same whether the bars be short 
 or long. It is no doubt for this reason that Macfarlane Gray's rule, that 
 the consumption of coal is very nearly proportional to the diameter of furnace, 
 is found to be so nearly correct in every-day practice. 
 
 Area of Fire Grate. — The area of fire grate required for the evaporation 
 of a certain weight of steam depends on the quantity and quality of the 
 fuel burned on it ; the quantity of coal is generally dependent to a large 
 extent on the quality, as may be seen by reference to Table li., 
 and by the draught. It may be assumed that 1 pound of good steam 
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 in a locomotive boiler, as fitted to torpedo boats, when not being forced, and 
 6 pounds when forced to the utmost ; also that in the mercantile marine, 
 where the coal is only of moderate quality, 8 to 9 pounds is a fair result, and 
 
DIFFERENT TYPES OF MARINE BOILERS. 
 
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576 MANUAL OF MARINE ENGINEERING. 
 
 6 to 8 pounds only can be obtained with the coal supplied at some foreign 
 ports. The quantity of coal burnt on a square foot of grate per hour with 
 natural draught is about 20 pounds, under favourable circumstances ; with 
 good stoking on short special bars and very good draught as much as 25 pounds 
 may be consumed ; but under ordinary circumstances and natural draught, 
 only 15 pounds should be supplied to obtain complete combustion and 
 economical results. With Howden's system of forced draught and warm 
 air 50 pounds per square foot can be burnt, if the coal is fairly good and 
 clean ; with closed stokeholes, and an air pressure of h inch of water, 30 
 pounds of good Welsh and 35 of North country coal can be burnt. 
 
 From this it will be seen, (1) that the greatest weight of steam evaporated 
 per square foot of grate per hour, under the most favourable circumstances 
 and natural draught, is 10 X 25, or 250 pounds ; (2) that with bad fuel and 
 economical stoking it may be onty 6 x 15, or 90 pounds ; (3) that with 
 fairly good fuel and favourable circumstances it may be 9 x 20, or 180 pounds, 
 and (4) that with fairly good coal and careful stoking about 150 pounds may 
 be expected. In practice, therefore, for short trial trips with picked Welsh 
 coal and picked stokers, calculations may be based on an evaporation of 
 250 pounds ; for mail steamships using good English coal, calculations 
 should be based on an evaporation of 150 pounds for natural draught, and 
 for forced draught from 270 to 400, depending on the air pressure ; and if 
 a ship is going to trade in the East or localities where inferior coal only can 
 be obtained, the boilers should be designed on the assumption of an evapora- 
 tion of only 100 pounds of water per square foot of grate natural, and 200 to 
 280 forced draught. 
 
 If the weight of steam required per hour for a given engine be calculated 
 and divided by one of these numbers, the result will be the number of square 
 feet required. 
 
 If the draught be increased by artificial means to a still higher intensity, 
 as is the case with torpedo boats, etc., the quantity of fuel consumed per 
 square foot of grate may be as high as 100 pounds per hour, with an air 
 pressure of 6 inches in the stokehole ; and 50 pounds with only 2 inches, 
 the corresponding evaporations being 570 pounds and 350 pounds per square 
 foot of grate. 
 
 Consumption of Fuel. — The consumption of fuel per I.H.P. per hour for 
 engines working at full power was 4 pounds, with surface-condensing expansive 
 engines, using steam of 30 pounds pressure above the atmosphere ; 3^ to 
 3| pounds with similar engines of best make and large size ; 2f pounds 
 with compound engines when forced, and 2 J to 2| pounds when of moderate 
 size and working at two-third power ; 2J pounds with compound engines of 
 moderate size and as generally fitted in the mercantile marine when working 
 at full speed ; 2 pounds with the best compound engines well designed 
 and carefully worked at sea full speed ; If pounds with large compound 
 engines when working at sea full speed with best fresh Welsh coal : 
 1| pounds with good triple-expansion engines using English and Welsh coal 
 ©f really good quality, and If pounds when ordinary good steam coal is used ; 
 1^ to 1-h pounds with triple- and quadruple-expansion engines using steam 
 at 200 to 180 pounds pressure ; the consumption of water with these latter 
 engines being about 14 to 15 pounds ; the consumption in torpedo boats 
 with compound engines was 3i to 4 pounds when working nearly full speed, 
 
HEATING SURFACE. 577 
 
 and 2h pounds with triples in Destroyers. In H.M. Navy using Welsh coai 
 and an air pressure of \ inch of water the consumption is 1-75 pounds including 
 auxiliaries ; with Howden's system in the mercantile marine 1*3 to 1-4 pounds 
 of good Welsh coal with large engines and 1-35 to 1-5 with smaller is general. 
 With superheaters for the steam in addition to the hot air of the Howden 
 and Ellis-Eaves systems the consumption is still lower, and generally does 
 not exceed 1-3 pounds per I.H.P. per hour for main engines only.* 
 
 Assuming the consumption of coal to be \\ pounds per I.H.P. per hour, 
 and the grate to burn 15 pounds per square foot, there should be 0-1 square 
 foot of grate per I.H.P. If the sea full speed I.H.P. of a merchant ship be 
 multiplied by 0*1 it will give sufficient grate area for that power. 
 
 That is, the grate area required for that power = I.H.P. -~ 10. 
 
 On trial trips with good coal, natural draught, and the engines working 
 at full speed, the triple-compound engine will develop 14 I.H.P. per square 
 foot of grate ; hence, one-fourteenth of a square foot per I.H.P. may be taken 
 as the proper allowance in designing furnaces for engines to develop a certain 
 power on favourable occasion, and one-thirtieth with Howden fittings 
 
 As the sea full speed power is usually, on long voyages, about three- 
 fourths that developed on a trial trip, the proportion of|X , T V, or - 095 of 
 a square foot per I.H.P. developed at sea, corresponds with that given above. 
 But with the higher funnels of the larger steamships even better results 
 are obtained, and it is found now that even 0*08 square foot per I.H.P. as 
 developed at sea is sufficient for the main engines only ; for the auxiliaries 
 and domestic purposes a special addition must be made in proportion to 
 their needs. 
 
 Heating Surface. — Strictly speaking, all surfaces exposed to heat which 
 are capable of absorbing, and their bodies of transmitting, that heat to the 
 water or steam are heating surfaces ; but technically only certain parts are 
 reckoned as effective heating surface, and the aggregate of such surfaces is 
 called the total heating surface. The surface of the upper half of the furnace, 
 or the part above the level of the fire-bars, that of the combustion chamber 
 above the level of the bridges and back slabs, including the actual surface 
 of the back-tube plates, are reckoned as the effective heating surface of 
 furnaces and chambers, and are stated separately, chiefly on account of the 
 metal forming them being three to four times the thickness of the tubes. 
 The surface of the tubes measured externally — that is, the area obtained by 
 multiplying the external circumference by the length between the tube plates, 
 is called the tube surface. 
 
 The Admiralty reckon tube surface by taking the area obtained by multi- 
 plying the external circumference by the length over the tube plates ; and 
 in reckoning the total heating surface, the surface of the back tube plates 
 is omitted. The calculation is in this way simplified, and the total heating 
 surface is practically the same as if calculated strictly, for the surface of the 
 parts of the tubes covered by the plates is very nearly the same as that of 
 the back tube plates. 
 
 The front tube plates should be, and usually are, omitted in calculating 
 the total heating surface, as they cannot be considered as effective. 
 
 The amount of total heating surface must depend on the quantity and 
 quality of the fuel burnt on the grates in a fixed time — that is, on the 
 quantity of heat generated in a unit of time, and also on the quality of the 
 
 * Experiments on ships of large size fitted with Robinson superheaters showed the consumption of 
 good coal with triples to be 1-4, and with Howden's fittings 1-18. QuadruDles 1-15 lbs. per hour. 
 
 37 
 
578 
 
 MANUAL OF MARINE ENGINEERING. 
 
 surface, etc. But since a grate may at some time have to burn the best of 
 fuel, and the engine to have the advantage, the total heating surface should 
 be adequate for such an occasion. If, however, no increase in engine power 
 is required, the damper must be partly closed or a portion of the grate bricked 
 up ; the latter is by far the better plan on economic grounds, as the brighter 
 the fire the more effective is the combustion, and the consumption corre- 
 spondingly decreased. 
 
 Tube Surface. — When possible, there should be 1*33 square feet of iron or 
 steel tube for each pound of coal burnt per hour — that is, in the ordinary 
 marine boiler there should be about 27 square feet of iron tubes per square 
 foot of grate ; with brass and copper tubes it was usual to provide about 
 25 per cent, less surface. 
 
 Since on trial trip with compound engines 10 I.H.P. were usually 
 developed per square foot of grate, there should be 2 square feet of brass, 
 and 2 - 7 square feet of iron tubes per I.H.P. developed on trial trip, or 2'66 
 and 3'6 square feet respectively per I.H.P. developed at sea. And since 
 with triple-expansion engines and natural draught as much as 14 I.H.P. are 
 developed from a square foot of grate, there need only be 1*93 square feet 
 of iron tube surface per I.H.P. in the boilers for them. When weight of 
 boiler is of as much consideration as economy of fuel 1*75 square feet of 
 tube surface per I.H.P. is sufficient with triple-expansion engines. With 
 forced draught by closed stokeholes and 30 to 40 pounds of coal are con- 
 sumed the heating surface per foot of grate must be accordingly increased ; 
 the tube surface should be, therefore, still 1*33 feet per pound of fuel, and, 
 consequently, 40 to 53 times the grate area. 
 
 Total Heating Surface must be governed as to amount by the ability 
 to produce steam, and the weight of steam per horse-power required by 
 the engine, etc. In general practice with modern marine boilers of good 
 design the following holds good for the production of steam per square foot 
 per hour under favourable conditions as to the state of inner and outer sur- 
 faces of tubes, the draught, air temperature and humidity, etc., viz. : — 
 
 (a) Best class of naval water-tube boiler, highly forced draught, . 13 lbs. 
 
 (6) „ „ „ moderately „ . 10 „ 
 
 (c) Ordinary water-tube boiler, . . lightly „ . 75 ,, 
 
 (d) Cylindrical boilers, . . . . moderately „ . 9 ,, 
 
 (e) „ ...... ordinary draught, . . 65 „ 
 
 The steam consumption on various ships, including that required for 
 auxiliaries of every kind as well as the demands for domestic purposes, may 
 be taken as follows : — 
 
 Description of Ship and Service. 
 
 Cargo steamer, general service, quadruple com- 
 pound engine, 
 
 „ „ triple compound 
 
 engine, 
 
 „ ,, geared turbines, 
 
 Passenger express steamers, Atlantic, turbines, 
 
 „ „ Foreign, reciprocators, . 
 
 ,, ,, Foreign, turbines, geared, 
 
 „ ,, Cross-Channel, turbines, 
 
 Naval ships, Battleships and Cruisers, turbines, 
 
 „ Destroyers, Scouts, etc., turbines. 
 
 Main 
 
 Auxiliary, 
 
 Engine. 
 
 <fec. 
 
 14-0 
 
 0-70 
 
 14-7 
 
 0-74 
 
 12-5 
 
 0-85 
 
 12-5 
 
 2-50 
 
 13-5 
 
 1-50 
 
 12-5 
 
 1-50 
 
 130 
 
 165 
 
 13-5 
 
 2-00 
 
 140 
 
 2-00 
 
 Margin Total 
 
 Supply. Consumpt 
 
 1-40 
 
 1-47 
 
 1-25 
 
 1-25 
 
 1-35 
 
 •25 
 
 •30 
 
 •35 
 
 •00 
 
 16-1 
 
 16-9 
 14-6 
 16-25 
 16-35 
 
 15-25 
 15-95 
 16-85 
 17 00 
 
TOTAL HEATING SURFACE. 
 
 579 
 
 If these figures be taken in conjunction with the amounts evaporated 
 as above as a basis from which it may be deduced that the ordinary cylindrical 
 
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580 MANUAL OF MARINE ENGINEERING. 
 
 boiler on service should supply 8 lbs. on short service with forced draught, 
 and 7 lbs. on long service with good draught as with Howden's apparatus, 
 or without but with high chimneys, and 6 lbs. for ordinary cargo ships. The 
 express steamer with water-tube boilers and forced draught should have 
 10 lbs. on short service and 9 lbs. on long. The Scout and Destroyer will have 
 12 lbs., and the Cruiser and Battleship 9 lbs. The merchant ship on long 
 service usually proceeds with the engines at three-quarter power ; hence 
 the evaporation then will be 75 per cent, of these allowances and the efficiency 
 should be 0-80. The express cross-channel steamer will usually work with 
 85 per cent, of the power, so that the demand per square foot will then be 
 8-5 lbs. Naval ships proceed at very varying speeds. 
 
 The amount of total heating surface which a cylindrical boiler may have 
 i? given approximately by the following rule : — 
 
 (1) Total heating surface = (diameter of shell in feet) 2 
 
 X (length of tubes in feet -j- 3). 
 
 Example. — What amount of heating surface should be found in a boiler 
 12 feet diameter, the tubes being 7 feet long ? 
 
 Total heating surface = 12 2 X (7 + 3), or 1,440 square feet. 
 
 Example. — What heating surface should a double-ended boiler contain 
 whose diameter is 10 feet, and the length of tubes 5*5 feet, the combustion 
 chambers being common to opposite furnaces ? 
 
 Here total heating surface = 10 2 X 2 X (5 - 5 + 3), or 1,700 square feet. 
 
 If the combustion chambers are the same as in a single-ended boiler, 
 the quantity will be the same as in two single-ended boilers of the same 
 diameter and length of tubes. 
 
 Efficiency of Heating Surface. — The rate of transmission of heat from the 
 fuel and hot gases in a boiler to the water on the other side of the metal 
 separating them depends on the differences of temperature, the thickness of 
 the plate and its material, but chiefly on the state of the receiving and 
 transmitting surfaces. Some few years ago elaborate and careful experi- 
 ments were made by Sir John Durston and the late Mr. Blechynden inde- 
 pendently, which demonstrated that, when the plate is clean on both sides, 
 there is not a great difference between the temperature on the sides of the plate 
 itself, the dry side being, in practice, only 67° to 86° F. hotter. If, however, 
 the transmitting side is dirty with scale, etc., it may be very much greater ; 
 and, with greasy mud, it was over 500°. That a slight layer of soot or 
 tarry matter prevented the dry side from getting very hot was demonstrated, 
 as also that, until the plate is raised to a certain temperature, the actual or 
 burning gas flame does not come into real contact. Further, that the 
 thickness of plate when clean is not of much consequence, and is less so 
 when dirty ; it is, however, of some importance as being a regulator of the 
 transmission of heat. Neglecting thickness altogether, it was found that 
 the efficiency varies as the square of the differences of temperature, and the 
 following formula may be taken as correct : — 
 
 (T - fl« 
 Q = — — , 
 
 Q being the units of heat per hour per square foot of heating surface. 
 T the temperature of the hot gas. 
 t „ „ water. 
 
 a is a factor which Eankine valued at 1G0 to 200 ; Blechynden at 40 for 
 tubes when cleaned, and 50 for plates f inch thick when clean. 
 
CAPACITY OF BOILER SHELL. 58] 
 
 There is reason to think that Bleehynden is nearer the truth than Rankine 
 for modern boilers worked on modern conditions. 
 
 There is every reason to suppose that the efficiency is seriously affected 
 by such inequalities of surface as permit of disturbing the flow of hot gases 
 over it and thereby bringing really more heat in contact. 
 
 Efficiency Of Boilers.— Generally, it may be taken that the efficiency of a 
 boiler is best gauged by taking the water evaporated by a pound of coal 
 from and at 212° F. and divide it by the amount such fuel could evaporate 
 if all its heat were used for the purpose. In practice it is, of course, impossible 
 to utilise the whole, or even nearly the whole, heat of combustion. It is, 
 moreover, not to be expected of certain boilers that the efficiency so measured 
 can be very high, but it is necessary to compare them with others of their kind 
 as well as with boilers generally. 
 
 A boiler for express service, such as that of a torpedo boat or a short- 
 distance passenger ship, must be light, and the measure of its efficiency for 
 this purpose will be gauged better by knowing how many pounds .of steam 
 it can produce per square foot of heating surface per hour. Now, weight is 
 always a factor of some value on shipboard. If, therefore, the amount be 
 multiplied by the number of pounds of water it evaporates (from and at 
 212" in each case), or the number of pounds of steam it produces per pound 
 of standard fuel, the result must express the general value of the boiler for 
 steamship purposes. The weight of steam supplied per hour for each ton 
 weight of boiler and appurtenances, including water, is really a practical 
 criterion of the efficiency of a boiler for marine purposes. 
 
 Area through Tubes. — The sectional area through the tubes, or that area 
 through which the hot gases and smoke pass from the combustion chamber 
 to the funnel, should not be less than one-seventh the area of grate with 
 the natural draught, and is usually about one-fifth. Too large an area pro- 
 duces a low velocity, which permits a deposit of soot and ash to form, with 
 the consequent reduction of evaporative efficiency. Too small an area 
 checks the draught, especially when the surfaces have become dirty. With 
 forced draught the area through tubes can be smaller if necessary : and 
 when it is more than one-seventh the grate (for natural draught) it is ad- 
 vantageous to have retarders in the tubes. These spiral strips not only 
 prevent a too rapid flow of gas but force the hot current over every portion 
 of the tube's circumference. 
 
 Capacity of Boiler Shell. — To contain the requisite heating surface, and 
 to leave sufficient steam Space, the boiler shell should contain 3 cubic feet 
 per I.H.P. for the mercantile marine, and 2-5 cubic feet for the Navy and 
 other service which has equally quick-running engines, when made for com- 
 pound screw engines, and 2 and 1-6 respectively for triple engines; when 
 for paddle engines it should be larger. In the mercantile marine a slightly 
 larger allowance is sometimes made, especially when for steamers making 
 long voyages. Since the volume of steam produced at a pressure of 210 lbs. 
 is only about a third that at 70 lbs. when the same weight is used, the steam 
 space for boilers working at the high pressures now obtaining may be con- 
 siderably less than was usual. Good results can be got now with an allowance 
 of If to 2 cubic feet of boiler per I.H.P. for natural draught, and 1£ to 1J for 
 forced draught. In fact, the steam space must bear a relation to the high- 
 pressure cylinder capacity in every case ; the intermittent demand of the 
 
582 MANUAL OF MARINE ENGINEERING. 
 
 single high-pressure cylinder of a paddle engine affects the equilibrium of 
 the boiler very differently from that of the two high-pressure cylinders of 
 a twin screw ship running at four times the revolutions, but using the same 
 quantity of steam. 
 
 * Steam Space. — The top row of tubes in a cylindrical boiler should be 
 not less than 0"28 of the diameter of the shell from, the top. The tubes are 
 sometimes placed higher, but there is then risk of priming owing to contrac- 
 tion of water-surface area as well as to the conformation of the sides. Priming 
 is often due rather to contracted area of water surface than to small steam 
 space, although the latter is generally set down as the cause when the tubes 
 are high. If the water surface is so contracted that little or no part of the 
 boiler where there may be down currents lies immediately under it, priming 
 is sure to ensue. 
 
 The capacity of the steam space depends on the quantity of steam used 
 in a fixed time, and on the number of periods of supply to the engines in 
 that time. The effect of admission to the cylinders is to reduce the pressure 
 in the steam space ; if that reduction in pressure be sensible there will be 
 an augmentation of ebullition at that period ; if the reduction in pressure 
 is serious there will be excessive ebullition, resulting in priming. For this 
 reason it is that a slow-moving paddle engine, which takes its steam in a 
 series of gulps, requires to have boilers with much larger steam space than 
 is sufficient for fast-running screw engines using the same weight of steam. 
 
 There should be - 8 of a cubic foot of steam space per I.H.P. for a slow- 
 running paddle engine, and 0*65 of a cubic foot per I.H.P. for compound 
 screw engines, and as low as 0*55 of a cubic foot for fast-running mercantile 
 and naval screw compound engines. Boilers for triple- and quadruple- 
 expansion engines may have 25 per cent, less steam space than this with 
 natural draught, and 50 per cent, less with quick-running ones and forced 
 draught ; while with turbines the space is even less. The amount of steam 
 Apace in a boiler is not dependent on the weight of steam used per stroke, 
 but on the volume, and for that reason a boiler constructed for, say, 200 lbs. 
 pressure does not require so much steam space as a similar boiler constructed 
 for only 75 lbs. pressure ; for if both boilers have the same grate area and 
 heating surface, they will evaporate the same iveight of steam, but the volumes 
 will be as 3 to 8 nearly. If these two boilers supply steam at the same rate 
 to engines running at the same number of revolutions, their steam spaces 
 may be as 3 to 8 nearly. 
 
 It is no doubt on account of the high pressure and large number of 
 revolutions of engines, together with the shaking when running, that a 
 locomotive boiler works so well without priming, in spite of the small steam 
 space and contracted water surface. 
 
 Area of Uptake and Funnel Sections. — Although in practice the funnel is 
 often designed to suit the general appearance of the ship, and it is also found 
 that, whereas some engineers prefer a small high funnel, there are others 
 who, for some other reasons, resort to the practice of making the funnel short 
 and of large diameter; still there is undoubtedly a certain diameter and a 
 certain height that will give the best result, and that cannot be determined 
 from external considerations.! 
 
 * When turbines take the steam tho space docs not require to be so large as obtains with reciprocatore. 
 t Consult Index for other information. 
 
AREA OF FUNNEL SECTION. 
 
 583 
 
 In the Navy with natural draught the funnel is usually made with a 
 sectional area equal to one-eighth the area of the grate (natural draught size). 
 In the mercantile marine a somewhat larger funnel usually obtains, the area 
 being from one-fourth to one-sixth that of the grate ; in general practice a 
 funnel whose sectional area is one-fifth to one-sixth that of the grate, and 
 whose top is at least 40 feet from the level of the grate, will give a very good 
 result. The objections to a large funnel, beyond that of space occupied 
 and cost, are resistance to the wind and large surface exposed to the cooling 
 action of both wind and water, whereby the hot column within is partially 
 cooled, and the draught thereby checked. On the other hand, a small funnel 
 is liable to become excessively hot, and when the fires are freshly charged to 
 become choked with smoke, and at all times it tends to check the draught. 
 The funnel of a warship may be small, because it is so seldom that the boilers 
 are urged to the utmost, and it must be as small as possible for obvious reasons. 
 When the draught is forced either by a blast or by other artificial means, the 
 funnel may be short, and of comparatively small diameter. The area at 
 the base of a locomotive is seldom more than one-tenth the area of fire-grate, 
 and often as small as one-twelfth. The size, for appearance sake, is now 
 got by making the outer casing of the required form. 
 
 Area of Funnel Section should be such that (1) for natural draught 
 Merchant service condition there are 1*25 square inches for each pound of 
 fuel consumed per hour, or, say, 1*88 square inches per I.H.P. of trial trip. 
 
 (2) For Naval service and short Express service conditions there should 
 be 17 square inches with assisted draught, and 1*25 with forced draught 
 per I.H.P. of trial trip, or 0*9 square inch per pound of coal burned per hour. 
 
 Diameter funnel in inches = Vl.H.P. X F. 
 
 For ordinary Merchant steamer F = 2 - 38 natural draught. 
 ,, Naval or Express „ F = 2 - 16 assisted ,, 
 „ ,, F = 1 "66 forced ,, 
 
 TAP>LE LIHa. — Capacity of Funnels for Coal Burnt in Lbs. per Hour. 
 
 
 
 
 
 
 Height of Funnel above Dea 
 
 cl Plates 
 
 in Feet. 
 
 
 
 
 
 Diameter 
 
 of 
 Funnel. 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 20 
 
 25 
 
 30 
 
 35 
 224 
 
 40 
 
 50 
 
 60 
 
 70 
 
 80 
 
 90 
 
 100 
 
 110 
 
 120 
 
 140 
 
 2-0 
 
 167 
 
 187 
 
 207 
 
 239 
 
 264 
 
 288 
 
 311 
 
 333 
 
 
 
 
 
 
 2-5 
 
 261 
 
 292 
 
 320 
 
 346 
 
 370 
 
 413 
 
 453 
 
 489 
 
 522 
 
 
 
 
 
 
 
 3-0 
 
 380 
 
 421 
 
 464 
 
 500 
 
 535 
 
 595 
 
 653 
 
 705 
 
 753 
 
 ■ 
 
 
 
 .. 
 
 
 
 3-5 | 
 
 515 
 
 Ol .7 
 
 630 
 
 681 
 
 728 
 
 813 
 
 891 
 
 963 
 
 1,028 
 
 
 
 
 
 
 
 4-0 
 
 
 750 
 
 822 
 
 888 
 
 948 
 
 1,061 
 
 1,161 
 
 1,254 
 
 1,341 
 
 1,422 
 
 
 . . 
 
 
 
 . 
 
 5-0 
 
 . . 
 
 , , 
 
 1,282 
 
 1,385 
 
 1,480 
 
 1,654 
 
 1812 
 
 1,956 
 
 2,088 
 
 o 221 
 
 
 
 . , 
 
 
 
 6-0 
 
 , . 
 
 , . 
 
 . , 
 
 
 2,140 
 
 2,380 
 
 2,609 
 
 2,820 
 
 3,012 
 
 3,196 
 
 3 366 
 
 
 , , 
 
 
 
 7-0 
 
 , . 
 
 
 
 
 2,916 
 
 3,252 
 
 3.564 
 
 3 852 
 
 4,112 
 
 4.364 
 
 4,580 
 
 
 . , 
 
 
 
 8-0 
 
 , . 
 
 
 , , 
 
 
 
 4,228 
 
 4.644 
 
 5,016 
 
 5,448 
 
 5,688 
 
 5,980 
 
 6.283 
 
 .. 
 
 
 
 9-0 
 
 , , 
 
 
 . , 
 
 
 
 
 5,884 
 
 6,356 
 
 6,788 
 
 7,204 
 
 7.592 
 
 7,961 
 
 8,310 
 
 8,9S( 
 
 10-0 
 
 
 
 , , 
 
 . , 
 
 
 
 7,248 
 
 7,824 
 
 8,352 
 
 8.885 
 
 9,360 
 
 9,820 
 
 10.250 
 
 11.070 
 
 11-0 
 
 
 
 , . 
 
 . , 
 
 
 . . 
 
 , , 
 
 9,466 
 
 10,110 
 
 10,730 
 
 11.310 
 
 11,840 
 
 12.380 
 
 13,380 
 
 12-0 
 
 
 
 
 •• 
 
 
 
 " 
 
 11,280 
 
 12,050 
 
 12,700 
 
 13,460 
 
 14.120 
 
 14.750 
 
 15.930 
 
584 
 
 MANUAL OP MARINE ENGINEERING. 
 
 Sq. Ft. Total H.S. per 
 Ton of Boiler. 
 
 ■^ [^ - l-' (S Clip oo -• O M !C O ■* « ^ ih t^. M l> C) O ■* O '- ?) 
 
 ■<f n o; to oo - a o a s to ■* o ic 6 •* i m in i t'^ c ffi « c i ?) 
 
 c to c to « to i.o a ■* to iq o io io io Tt ^m o ■* o o i"o c 15 
 
 m -tNtOtONCMtOW-iO-teCtOiOlON-H (MtC0OC0lO?l»O-<* 
 
 g^^ipMiprflffl-fwacci^oao^ooitswtooiHOOO 
 
 H CM CM <M CM CM CI CM (M 00 Cl CM CI CO CM CO CO CO CM CI CO M CO CO CI CO CO 
 
 «i5'*00t0t000'*h>O'*tM-H^t0C»f-Oi0C>l<MeCCSClOi5 — 
 GGOGpGpGOC»CJ0Olp:oppqip-<pcMpp — '-h OICI- o- — 
 
 »iooaoooi»ccMr-t^«ts tO'*i--too^"t 01 ci © co 
 
 g iC tptpiS r- -J<tprf'Ot01^Xi05C/3C5 CI © 00 00 -— 1^. Op Cl S3 © 
 
 H'^'H'HiHfHiHrt'-ijlrtrtFHr-I^FHtMCM'HHN'HCMU-Hr-rt 
 
 j;„iOO >0 U5 -# 15 CO 
 
 gci co t< p p r- ci r- ip cm p ip t;~ -rf ip ci p -h to tS r- cm ci to "i — 
 
 Po oot--n^toco-Hic(MMcMto — Gotbotbtbiraotb"'oc»-Hoo 
 
 fClCOCOt-l— ©©l5i5©©l5lSlS"*-*-<#C0C0CCCOCiOl — I CM i— < 
 
 • >ra »o ■<* co 
 
 2p©©©©L^01CC01»p©ipLOCMCMip©©pCMI>>ppiS©© 
 
 Pt>ib6toi^xo^tsMt»i6^iMMTiit>iooojoii 
 
 t H CCM«NNN(Mt>lcUt»t>l-H-HCTrtrtrtFHrt-.H-H 
 
 • >s O 15 
 
 gttr-HOp ^CCt^p©CJCMcpTt<p.^'©CMpCMCl©CliS 
 
 PoOI^I^i5-^(SOOrt^iCftOini>-'HOt^KSl«OtDOIMrt'- 
 
 Hoisio-tfis-^cccccoco-'tcocococococicicioioi— i — — — — 
 
 *=OOu0OOOOO'0OOOOOOOOOOif5OOOi.0 01O 
 
 wiotM^oo- itooootoio- i^ectstoo-tiocot- o 15 c. © 
 
 O.I— ©CO©©00Cl-H©COOl©a0r^lSCOCCClCl©>^a0CO©©© 
 CO CO MCCC0C5CIHCM-it»CNrt- l -Hrtrtrt«MrtW 
 
 ^©■<*i©©©i>-iO©©co©r-^-'*<©©-^<io©©©co-*oi©'— ico 
 «sh.^o»- ir^-t<-Heociccoo-}'i-~cM-* , ©©co— ■ — -*©•>* t- 
 crM ©oo-^co^t— t>-co©oo«:*#cocM©©©cixai«D©»5iSTf< 
 
 KnWNNtMt>lHHi-ltN«HHHHHP-IH 
 
 ^©oociTf-^coicccr-ccococo©©©^©©©^©©^-^ 
 g©©©©©to«i©t-»©oo©i>.t-c-i>'©©oo©©©©©©L5 
 
 i-icocococococococococococococoeococococoeocococococo'co 
 
 t0©tO00t0'*<'#t0a0't0)l0t0CM©O)0)0J30(MC0C0O©-t-+> 
 Ct:03tOtOMtOtOO!0*t«»(MCSl^iOaoOIM'*tO-i — OOO 
 
 -HCoeocoeocMoitM-HrHcoeoeococMco— 1— 1 -h m ci ci — i ci — ci 
 
 C01GO©-1<©eOCl©0000 01COCMCOGC©01Cl©ei©COr-~01--© 
 M'*C0-TC0-^l'tCO'<*COCOTj<C0-^lCO-*C0'*C0-*COTtC0C0C':C0C0 
 
 to © © --O © ■>$< ■<*< M" •* ■<* CO CO CO CO CM CO CM Ol Ol 01 OJ CM 01 01 01 Ol 
 
 iOOOOiO^OOOOOOiScMOOOOOOOOCOiCS 
 J3 CO CO 3 15 tO Ifl 15 15 LO CO tO tO 15 ■* O 15 O 15 19 lO •* O 15 15 15 LQ 
 
 G H» T-"-HiTra'>"7»Hc<> "fr-K^'-u^o^^p^Rd^ "loSr-y-HtOHXsin-^-Hoo 
 
 ctOOOWiOOMOMCStSMtOtOOtClO-OtOtOKnfflMO 
 
 ^2^2rt t0 2 t0i0 i2 <x>c0OC5OO ~' c>0505 ' - ' 35050505CC0500 
 
 e©co©eo©©©©cMcc©co©©©©co©©~. ©©©©co© 
 
 I— I 
 
 ■^■^•^■"tcocococM— i — 01"+tj<-^<cococooioi-h-h-h©©c;©go 
 
 tu to ^TotT'of 
 
 s -.-. ^-^. p o r 3 r d 
 
 5 8 ~ - • >>>>>> 
 
 • H * ' H ,, r S , r 2' r 2 
 
 -S ~ * ^ * 
 
 •> 8 •> ■ 'P _ cS <S etj 
 
 2^ s 2 g ■ • ' a 
 
 H c Eh g r* H H 
 
 O eo O k " o":r;:" ^"o £ "o 
 
 |1 s s 5 : :|% s s : =| fe'fife s = | 
 
 tc 
 
 a 
 3 
 
 a 
 •J 
 
 o 
 
 a 
 a 
 
 p 
 
 a 
 o 
 
 G 
 
 pq 
 a 
 
 S 
 
 <: 
 
 o 
 
 a 
 o 
 
 p 
 c 
 
 H 
 
 cr 
 
 < 
 
 PC 
 
 « 8 
 
 a g 
 
 Total. 
 
 Water. 
 
 Boiler. 
 
 Total. 
 
 Water. 
 
 Boiler. 
 
 Total Heating Surface. 
 
 Surface. 
 
 Length. 
 
 Diameter. 
 
 Number. 
 
 Number of 
 Combustion Chambers. 
 
 Thickness. 
 
 Diameter. 
 
 Number. 
 
 Working Pressure. 
 
 Thickness. 
 
 Length. 
 
 Diameter. 
 
 o 
 - 
 
LEADING PARTICULARS OF BOILERF. 
 
 585 
 
 
 
 e© m 
 
 t>»lO 
 
 00 
 
 -H SO 
 
 ■* 
 
 1/5 CO CO 
 
 00 
 
 Tt9« 
 
 o^ i' 
 
 | t~- CO ■** o 00 
 
 
 per Ton of Boiler. 
 
 -H t~- 
 
 cm m oo 
 
 CO 
 
 CO so 
 
 O 
 
 SO CM CO 
 
 rt< 
 
 = CO CO t— O ^< 
 
 
 CO lO 
 
 ffltsuj 
 
 CO 
 
 m co 
 
 CO 
 
 CO lO U0 
 
 "0 
 
 m 
 
 — ^~ 
 
 t+i m •* m ■-T 
 
 
 4a 
 
 
 ui ■>* SO 
 
 o o to 
 
 OO 
 
 CO CO 
 
 C55 
 
 CO OJ CO 
 
 <N 
 
 CO 
 
 in 
 
 co co co r- r- 
 
 
 » 8 
 
 Total. 
 
 § "P e- 
 
 K5 -H< CO 
 
 so 
 
 r- O 
 
 l-H 
 
 •* l>. UO 
 
 t- 
 
 
 OJ 
 
 so r- t— co -^ 
 
 
 ^5 3 
 to <s 
 
 
 HCM CM 
 
 NOIN 
 
 OJ 
 
 OJ <M 
 
 <M 
 
 CM CM 01 
 
 CM 
 
 CO 
 
 CM 
 
 SO CM SO SO CO 
 
 
 
 «CM o> 
 
 -hoow 
 
 CO 
 
 O CO 
 
 >o 
 
 -HttOO 
 
 p 
 
 >* 
 
 ■<# 
 
 oj cm t^ t- •r 
 
 
 Water. 
 
 OJ 00 OJ 
 
 00 
 
 O 00 
 
 t-- 
 
 OJ M CO 
 
 OJ 
 
 CM 
 
 o 
 
 ~h O CO O CM 
 
 
 
 HO O 
 
 o o o 
 
 © 
 
 — o 
 
 O 
 
 o o o 
 
 o 
 
 -H 
 
 -H 
 
 i-H i-H f-H i-H i-H 
 
 GO 
 
 
 acq -H< 
 
 OJ OJ — t 
 
 o 
 
 00 o 
 
 Ttl 
 
 r^o o 
 
 <M 
 
 (M 
 
 F-H 
 
 ■*OhOM 
 
 P 
 
 o £ 
 
 Boiler. 
 
 §=p »> 
 
 in in r- 
 
 55 
 
 r-cM 
 
 rf 
 
 ipOJI^ 
 
 CO 
 
 OJ 
 
 OJ 
 
 -h r---H pCM 
 
 CM hH CM CM CM 
 
 f-H 
 
 
 
 
 lO 
 
 
 
 
 
 
 
 
 to in in in 
 
 
 Total. 
 
 oO 1(5 
 
 t-m CM 
 
 o 
 
 o-< 
 
 t- 
 
 OCO t*i 
 
 o 
 
 o 
 
 i-H 
 
 SO CO O -h CO 
 
 > 
 
 
 oj co r- 
 
 OS 
 
 SO i—i 
 
 LO) 
 
 i— i oj r- 
 
 CM 
 
 CO 
 
 ^_l 
 
 oj r- co in o 
 
 -^- 
 
 
 
 H W OJ 
 
 CM "H — I 
 
 o 
 
 p— 1 t^* 
 
 CO 
 
 r- io uo 
 
 >n 
 
 M< 
 
 T(< 
 
 CM OJ CM CM CM 
 
 — 
 
 o 
 
 
 i-H -H 
 
 i-H i— • t -H 
 
 — - 
 
 — 
 
 
 
 
 
 
 
 fc 
 
 
 . 
 
 in o o 
 
 
 
 
 o 
 
 
 
 
 in m m 
 
 
 Water. 
 
 g m 
 
 r-r-r- 
 
 >o 
 
 1ft r^ 
 
 »n 
 
 O O u-5 
 
 CM 
 
 >n 
 
 CO 
 
 lO C) l> h- Tji 
 
 a 
 
 to 
 
 
 CO OJ -h 
 
 o 
 
 O O 
 
 0-1 
 
 O OJ OJ 
 
 t^ 
 
 CO 
 
 so 
 
 o O CO 00 tr- 
 
 
 *"■ 
 
 
 ^T ^T ^* 
 
 ■* 
 
 T* CO 
 
 CM 
 
 71HH 
 
 i-H 
 
 i-H 
 
 '- , 
 
 -— i-H 
 
 rH 
 
 
 
 o 
 
 
 
 
 id 
 
 
 
 
 1ft 
 
 **» 
 
 
 Boiler. 
 
 go U» 
 
 ONW 
 
 lO 
 
 10) o 
 
 C>» 
 
 O CO CO 
 
 oo 
 
 «3 
 
 CO 
 
 00 CO so •»* •* 
 
 
 
 
 £so os 
 
 co co in 
 
 00 
 
 CM — i 
 
 CO 
 
 ioi^ 
 
 ■^ 
 
 CO 
 
 (-^ 
 
 CO l^ t^ CO so 
 
 
 
 H CO t» 
 
 00 !>• t- 
 
 CO 
 
 !>• "* 
 
 T* 
 
 •* TJ< so 
 
 co 
 
 Ol 
 
 CM 
 
 -— i i-H -— r^ rH 
 
 Total 
 
 V<N CM 
 
 O O UO 
 
 IO 
 
 i-H iO 
 
 o 
 
 O CO Ol 
 
 r- 
 
 m 
 
 r- 
 
 O O O o o 
 
 p 
 
 ft-n r, 
 
 — CO — 
 
 r- 
 
 I-- i— 1 
 
 o 
 
 CO CO CO 
 
 o 
 
 CO 
 
 Ol 
 
 CO O 01 OJ o 
 
 Heating Surface. 
 
 
 Ol 00 ^ 
 
 ■c -># ■* 
 
 
 O Tf 
 
 ■* SO 
 
 o 
 
 CO 
 
 CO — CM 
 CM CM CM 
 
 OJ 
 
 --H 
 
 CO 
 
 CO 
 
 -H 
 
 CO O GO 00 o 
 
 /~\ 
 
 
 
 «co o 
 
 O o o 
 
 CO 
 
 O CO 
 
 ■* 
 
 tn t-~ co 
 
 (M 
 
 I-- 
 
 CO 
 
 CO — i O O -h 
 
 
 
 Surface. 
 
 ft oj r- 
 
 ■* CO CO 
 
 I-- 
 
 CO 01 
 
 OS 
 
 OOW 
 
 OJ 
 
 CO 
 
 CO 
 
 oo co in in co 
 
 
 
 c?oi — ' 
 
 m -i i» 
 
 lO 
 
 ■* CO 
 
 ■* 
 
 CO CO t~» 
 
 "* 
 
 o 
 
 o 
 
 co i-- co co m 
 
 ■«*; 
 
 
 
 j/2 "H^ Tfl 
 
 •* Tf CO 
 
 so 
 
 so oi 
 
 CM 
 
 CM — i — 
 
 I-H 
 
 l^-t 
 
 I-H 
 
 
 
 
 r+M 
 
 ttiao 
 
 -<M 
 
 
 ■Ooo 
 
 -«i 
 
 CO 
 
 
 
 HM.HIN 
 
 
 
 
 = 01 o 
 
 O CO — i 
 
 r^- 
 
 S CO 
 
 P-H 
 
 O OJ — i 
 
 
 a* 
 
 to 
 
 co co m o oj 
 
 a 
 
 CO 
 
 01 
 
 Length. 
 
 HH 
 
 — 1 i-H 
 
 
 
 
 
 
 
 
 
 
 a 
 
 H 
 
 
 fie- ■>• 
 
 r~t*» t^ 
 
 tr- 
 
 CO CO 
 
 t- 
 
 CO l-«ao 
 
 t-- 
 
 CO 
 
 c- 
 
 CO CO CO CO CO 
 
 
 Diameter. 
 
 eW oW 
 
 m« cm 
 
 •eWxeWHn 
 
 rH?l 
 
 KHiMr* 
 
 a** 
 
 r.hc-Ki-<t;) 
 
 r+* 
 
 HIM 
 
 H_ 
 
 1 l*J ' I'J' ' I'J' ' vj' ' 1y 
 
 
 OJ CN OJ 
 
 CM 
 
 (M <0J 
 
 CM 
 
 CM CO CO 
 
 CO 
 
 co 
 
 CO 
 
 so so so so so 
 
 H 
 
 
 CO o 
 
 <M CO O 
 
 o 
 
 (M CO 
 
 CO 
 
 OJ 00 o 
 
 "* 
 
 CO 
 
 CM 
 
 Ol CO CO OO CM 
 
 
 
 Number. 
 
 ■* •* 
 
 «C C-l Tj< 
 
 CO 
 
 r— ( r^ 
 
 OJ 
 
 O CO ■># 
 
 so 
 
 r- 
 
 r- 
 
 CO Tj< ~H — < 00 
 
 o 
 
 
 
 CO so 
 
 OOODt- 
 
 r~- 
 
 CO CO 
 
 "<* 
 
 •* CM(M 
 
 CM 
 
 |-H 
 
 r-H 
 
 ~ ' rt "^ ^ 
 
 X 
 
 Number of Com- 
 
 
 . . 
 
 so 
 
 •' CO 
 
 "* 
 
 
 co 
 
 co 
 
 CM 
 
 CM CM CM CM — " 
 
 ft 
 O 
 
 -Ji 
 
 bustion Chambers. 
 
 
 
 
 
 
 
 
 
 
 
 00 
 
 Thickness. 
 
 
 "!^ u*c "!=* 
 
 C*X1 
 
 l-'^t-Itl 
 
 «^1 
 
 3t*3! 
 
 * 
 
 
 r.ci 
 
 -m 
 
 •£-«•«»<£:£ 
 
 o 
 
 
 
 p4S 
 
 
 -fC 
 
 
 
 
 
 
 
 < 
 
 s 
 
 Diameter. 
 
 u o o 
 
 ■* SO o 
 
 CI 
 
 O CO 
 
 (M 
 
 «5 ■* •* 
 
 CO 
 
 o 
 
 o 
 
 0»* 
 
 i>. co >o in oi 
 
 s> 
 
 a 
 ft 
 
 
 w ■* ■* 
 
 ■*-*■■# 
 
 ■* 
 
 ■* SO 
 
 -* 
 
 ^ H-JI *3* 
 
 CO 
 
 -* 
 
 ^ 
 
 SO CO CO so so 
 
 
 Number. 
 
 CO 00 
 
 O CO CO 
 
 CO 
 
 CO CO 
 
 •* 
 
 CO CO CO 
 
 CO 
 
 SO 
 
 CJ 
 
 CM OJ Ol CM Ol 
 
 
 
 .op 
 
 «^ieffi r» 
 
 «^l 
 
 » to -p 
 
 J" 
 
 '*' u, te n' 3 
 
 ,|o 
 
 ^w 
 
 
 
 < 
 
 Ah 
 
 
 Thickness. 
 
 a«« -4C 
 
 r^|?5 ^Pl [rH 
 
 -m 
 
 r*"p* 
 
 in 
 
 -mW 
 
 H 
 
 JT 
 
 -*0 
 
 --h ^^-h m^ci;m 
 
 
 (jo co 
 
 CO CO O 
 
 CO 
 
 CO O 
 
 OJ 
 
 OJ O CO 
 
 o 
 
 CO 
 
 CO 
 
 CO CO CO CO OJ 
 
 ^ 
 
 • 
 
 Length. 
 
 M — ' 
 
 r^ 
 
 
 
 
 
 
 
 
 
 z 
 
 
 
 *3t>- t- 
 
 ffiCOO 
 
 o 
 
 a ca 
 
 CO 
 
 — H F«H .— 1 
 
 --H 
 
 o 
 
 -H 
 
 OJ OJ OJ OJ OJ 
 
 S 
 < 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 rJiO ■* 
 
 CO SO lO 
 
 CO 
 
 CM G5 
 
 o 
 
 O CO CO 
 
 o 
 
 o 
 
 CO 
 
 o o o o oj 
 
 1 
 
 
 Diameter. 
 
 «!S co 
 
 UJOi 1 
 
 >* 
 
 •* CM 
 
 - 
 
 m ■* ■* 
 
 CO 
 
 CO 
 
 - 
 
 hhOOOO 
 
 1 
 
 
 
 O 0) ■* 
 
 o 
 
 CO O 
 
 o 
 
 oo o 
 
 o 
 
 o 
 
 o 
 
 o o o o o 
 
 
 Working Pressure. 
 
 Jffl CO 
 
 — 05 —i 
 C>1 - 71 
 
 o 
 
 CN 
 
 —i aD 
 
 0-1 — i 
 
 Ci 
 
 CO o CO 
 — i CM — i 
 
 o 
 
 CM 
 
 CO 
 
 o 
 
 (M 
 
 CO O CO 00 
 
 01 r-H 0J — 1 -H 
 
 
 I -»_-_»■ 
 
 — i « i 
 
 
 - . 
 
 
 , 
 
 
 - I 
 
 -^v^ 
 
 --^ 1 1 1 I 1 
 
 W 
 
 0<M 
 
 
 
 -3 
 
 
 -3 
 
 
 
 
 "g o3 
 
 
 J-1 
 
 a m 
 
 T3 a) 
 t* — • 
 
 o s 
 8-h 
 
 S-l 
 
 
 ci ' 
 
 
 3 
 ® 
 
 
 eS 
 
 
 ci o 
 05 2- 
 
 i • • i aT 
 
 =H 
 
 <; cj 
 
 00 
 
 03 
 
 ~ 03 
 
 on" 
 
 -H 
 
 . os" 3 
 
 
 — ^3 
 
 l*-l ^ *»H 
 
 o flo 
 
 
 o "B o 
 
 riH •. 
 tf 03 
 
 o "2 
 
 
 »5 
 
 73 O-P 
 
 o o 
 
 
 3 J = : § 
 
 CB 
 --J o 
 
 -w - - 
 
 5M 
 
 O S- a) 
 
 So • - - = 
 
 
 
 PQ 
 
 pa 
 
 
 25-; 
 
 =5 
 
 ^v^ 
 
 
 h-2 
 
 25 
 
 i-3 25 
 
 
 M 5 
 
 Q ^~ 
 
 'oczh 
 
 ^ 
 
 'qQl 
 
 UJ 
 
 ooz 
 
 o" 
 
 "< 
 
 15" 
 
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586 MANUAL OF MARINE ENGINEERING, 
 
 CHAPTER XXIII. 
 
 WATER-TUBE BOILERS. 
 
 Such boilers as are composed wholly, or nearly wholly, of small tubes with 
 the water within them are called "water-tube" boilers. Singular to say, 
 the same causes that brought this type of boiler to the front in our time 
 operated over eighty years ago on the minds of the engineers of that day, 
 to produce many inventions and numerous patents, some of which are the 
 prototypes of the most successful boilers of their kind in use at the present 
 time. Others, while displaying a very considerable knowledge of the laws 
 governing the circulation, so necessary for the success of every boiler, and 
 indicating a mechanical knowledge of the highest order, were almost im- 
 possible of manufacture then, and have not survived, owing to theii 
 inferiority to other designs. In the third decade of the nineteenth century, 
 when the railway locomotive had been tried and found to be a success for 
 traction, many minds were engaged on designing and perfecting steam- 
 driven road carriages ; in other words, they were attempting to substitute 
 for the stage coach and the four-horsed travelling carriage a similar con- 
 veyance to go on the highways with a steam engine as the motive power. 
 Now, on an iron track the weight of the engine was not of very serious moment ; 
 but to run successfully on a turnpike road, without doing such damage as 
 would prohibit its use, the machinery was necessarily as light as possible, 
 because it was carried on the same four wheels as had hitherto supported 
 only the passengers and luggage. Naturally, therefore, attention was drawn 
 to the boiler as being by far the heaviest part of the machinery, and of that 
 part the water was no inconsiderable item ; hence, engineers availed them- 
 selves of their knowledge of the properties of the cylinder to devise boilers 
 on the tubulous system. Of them, by far the most ingenious were those 
 patented by the well-known physician-engineer Goldsworthy-Gurney in 1827 ; 
 the best of which consisted of a horizontal cylindrical drum on the top, and 
 a somewhat smaller drum at the bottom, these two being connected by 
 down-cast pipes at the ends, and by up-cast pipes at the sides, the latter 
 being in the shape of the letter S, and each pair together forming the figure 
 8 (see fig. 215). This boiler is interesting, as being the fore-runner and 
 type of the Thornycroft (fig. 230) and other similar boilers of to-day. There 
 were, besides this, other forms almost equally interesting, especially one 
 in which the up-cast tubes connecting the upper and the lower drums being 
 like the letter U, projecting horizontally, as in the case of the Du Temple 
 boiler, with the fire underneath them. In another ingenious design by the 
 same gentleman there was a similar boiler, but with one nest of larger tubes, 
 the bottom parts of the nest forming the fire-grate ; while in yet another 
 instance another effective arrangement was attained by having two sets 
 of tubes (right and left-handed) on this system, and interlaced into one another 
 
WATER-TUBE BOILERS. 
 
 587 
 
 
 9 
 
 a 
 
 o 
 
 C 
 
 a 
 
 I 
 
 id 
 
 £ 
 
 so as to form a central furnace, the fire-bars, of course, being the lower part 
 of the tubes. Gurney's boiler was intended for a working pressure of 130 lbs. 
 per square inch ; but even in those days it was anticipated that steam of 
 
588 MANUAL OF MARINE ENGINEERING. 
 
 much higher pressure might be used, and it was probably ouly the want of 
 means of manufacture and the absence of a suitable design of engines for 
 such pressures, as well as the necessity for obtaining chemically pure water, 
 that prevented the general use of the tubulous boiler at a very early date. 
 
 From the beginning of the nineteenth century engineers have, both in 
 this country and America, tried the water-tube type of boiler for high-pressure 
 steam. That made by Col. Stevens in 1805 consisted of copper tubes 1 inch 
 diameter and about 4 feet long. From his time to the present the number 
 of inventions of a tubulous form of boiler is almost legion, as every engineer 
 at some time of his life has experienced the necessity to satisfy one or other 
 of the conditions that go to make the tubulous boiler superior to every other 
 form. These conditions may be laid down as — 
 
 (1) Simplicity of form of element. 
 
 (2) Simplicity of construction due to smallness and lightness of element. 
 
 (3) Great strength of element in proportion to the working pressure and 
 consequent large margin of safety. 
 
 (4) Small quantity of water contained, especially in proportion to the 
 water evaporated per hour. 
 
 (5) General lightness of structure compared with the ordinary marine 
 boiler. 
 
 (6) Immunity from injury, and eventual destruction by (a) rapidity of 
 raising steam, (b) forcing when at work, and (c) sudden cooling by drawing 
 or putting out fires. 
 
 Every water-tube boiler worthy of notice must fulfil the above require- 
 ments ; but, to be successful for every-day use, it should also satisfy the 
 following conditions : — 
 
 (1) The elements of the boiler must be so disposed and distributed as 
 to be capable of taking up the heat generated by the fire without allowing 
 to pass to the chimney more than is sufficient for draught purposes. 
 
 (2) The circulation of the water must be positive, continuous, and uniform. 
 
 (3) The internal and external parts must be capable of easy examina- 
 tion and cleaning. 
 
 (4) Every part liable to deterioration or derangement must be capable 
 of removal with a minimum disturbance of other parts, or be put out of 
 action without loss of general efficiency. 
 
 (5) The steam in the receiver must be properly separated so as to be 
 fairly dry. 
 
 The generic form of a large class of water-tube boilers is pretty much 
 that shown on Mr. Gurney's design (fig. 215) — namely, an upper chamber, 
 into which the steam is delivered by a series of tubes surrounding the fire, 
 and a lower chamber or chambers connected to this upper chamber by them 
 and by downcast tubes, the latter serving to keep up the circulation through 
 the other tubes. 
 
 Another favourite type, however, consists of horizontal, or nearly hori- 
 zontal, tubes above and around the fire connecting two chambers, on the 
 top of one of which is the steam drum, from which a connection is made 
 to the other, to the bottom by preference, for proper circulation. 
 
 The great attraction for this latter type of boiler is the ease with which 
 
WATER-TUBE BOILERS. 589- 
 
 eke tubes may be examined, cleaned, or removed, and the freedom liom 
 priming when the parts are properly proportioned and arranged. It has, 
 however, the objectionable feature of flat surfaces which, in most of the 
 forms, require staying. 
 
 A third type consists wholly of small pipes, the steam drum (if existing 
 at all) being in a rudimentary form. It need hardly be said that this class 
 of boiler is the lightest, but it is likewise the most dangerous of the water- 
 tube type, and its employment is dependent on a regular supply of perfectly 
 pure water. It is generally known as the flash boiler, from the fact that the 
 water flashes into steam when forced into it. 
 
 The importance of the tubulous form of boiler has increased very much 
 of late years, in consequence of the demand for extremely light machinery 
 in the very high-speed vessels now required for war purposes, and the possi- 
 bility of employing them successfully, is the result of the introduction of the 
 evaporator, by which pure water can be distilled from sea water. For, 
 although previously the surface condenser and a very small supply of fresh 
 water carried in tanks permitted of the use of water-tube boilers in the mer- 
 cantile marine, they were in almost every case a failure, owing to salt water 
 eventually being used to make up the inevitable losses experienced in every 
 ship ; but the failure in these cases was mostly due to imperfect circulation. 
 It is a little singular that the evaporator suggested and fitted by Hall to 
 ships first 85 years ago having his surface condenser should have been allowed 
 to go out of use and be forgotten. 
 
 The general adoption of the triple-compound engine was undoubtedly 
 delayed for many years in consequence of the failure of the water-tube boiler 
 in the s.s. " Propontis." But it must not be forgotten that, in past days, 
 here and there water-tube boilers have worked successfuly under the ordinary 
 every-day circumstances of a merchant ship including the use of salt water, 
 and likewise that the boilers now fitted to ships have better means for exami- 
 nation and cleaning than existed in former times ; besides which the supply 
 of fresh water is practically ensured. 
 
 It is not claimed for these boilers that they are any more efficient as 
 steam-generators than the ordinary marine tank boiler ; but, on the other 
 hand, it is not admitted that they are of necessity worse generators. On 
 the contrary, it is very possible that, under centain conditions, a water-tube 
 boiler may be more economical in steam production than the ordinary 
 cylindrical boiler, inasmuch as it is capable of obtaining a better circulation 
 with less expenditure of energy, and is not so liable to losses by radiation. 
 Its efficiency, however, depends more on clean surfaces and tight casings 
 than does the tank boiler, and generally the latter is easier cleaned, etc., than 
 the former. 
 
 In marine practice, the heating surface in proportion to the grate has 
 been generally small, and the makers and users of those boilers have been 
 content with the other and real advantages derived from them rather than 
 to effect the utmost economy by arresting the waste of heat which can, un- 
 doubtedly, take place in some forms. That the whole surface of the tubes 
 is not efficient heating-surface goes without saying, just as no one counts 
 as heating-surface the lower part of a furnace or flue, or the parts of a 
 Lancashire boiler not exposed to heat. Probably only 0-7 of the circum- 
 ference of the horizontal tube of these boilers acts as real heating-surface ; 
 
590 MANUAL OF MARINE ENGINEERING. 
 
 but, for convenience of expression, the whole tube hitherto has been counted, 
 to serve frequently as an argument to their detriment by hostile critics. 
 
 It must also not be overlooked that, in some designs of water-tube boilers, 
 a considerable number of the tubes, whose surface nominally counts as heating- 
 surface, are practically not exposed to heat at all, and consequently really 
 act as downcasts. Reckoning, however, the whole surface of all the tubes 
 which are or may be exposed to heat as heating-surface, the allowance should 
 be 40 per cent, more per T.H.P. than given usually for cylindrical boilers 
 when weight is not of first consideration. In express boilers, where weight 
 is of first consideration, as low as 1-5 square feet of nominal heating-surface 
 per S.H.P. has given very fair results with a coal consumption of 75 lbs. per 
 square foot of grate. It is, however, better to allow 1-8 square feet, and limit 
 the consumption to 65 or 70 lbs., then 26 to 29 S.H.P. per foot of grate is 
 obtained, and 1-15 lbs. of Welsh coal* is burned per hour for each square 
 foot of heating-surface. That is to say, at or about 200 lbs. pressure an 
 efficient water-tube boiler can evaporate 10 to 12 lbs. of water per square foot 
 of nominal heating-surface per hour. Superheaters can be, and often are, 
 fitted to water-tube boilers with advantage, and are a means of considerable 
 improvement in efficiency ; Mr. Harold Yarrow's experiments proving that, 
 with a superheat of 100° F., his boilers gained in economy 8 to 10 per cent., 
 and, with an increase to 150° F., the gain was even 11 to 13 per cent. Under 
 these circumstances, when using 1 -237 lbs. of oil fuel per square foot of heating- 
 surface, as much as 14-6 lbs. of water were evaporated per lb. of fuel, and with 
 a consumption of 0-542 lb. it was 15-9, or 8-6 lbs. per square foot of heating- 
 surface. 
 
 Grate Area. — The area of the grate must in this type of boiler, as in every 
 other, depend on the draught available, and consequently the size of grate 
 may, to some extent, follow the ordinary rules given elsewhere ; but in most 
 forms of water-tube boilers it is not possible to work economically with 
 so high a pressure in the ashpit as is used with the locomotive or ordinary 
 marine boiler. In some few forms, the obstruction to the passage of hot 
 gases is sufficiently great to permit of a higher pressure, but in almost every 
 case a large grate is necessary for the water-tube boiler, in order to distribute 
 the heat evenly over the whole of the boiler. Moreover, it is not a good 
 thing to have the fire too intense when any portion of the tubes is touching 
 it or near it ; in fact, the, tubes which are contiguous to the fire, and receive 
 the direct radiation from the glowing coals, should be made of somewhat 
 thicker material as well as larger in diameter. With express boilers, and 
 the tubes arranged nearly vertical for rapid and certain circulation, the 
 foicing of the fires may be made by an air pressure as high as 4 inches, and 
 Sh to 4 inches is now the practice for full speed when the grates are getting 
 dirty. The consumption of coal in such cases is 65 to 70 lbs. per square 
 foot of grate, and 2-4 to 2-6 lbs. per I.H.P. (including that required for steering 
 gear and all auxiliaries). 26 to 29 I.H.P. per square foot of grate is obtained. 
 For turbines the consumption is not so large in normal conditions, being 
 1-65 to 1-8 per S.H.P. of coal, or 1-25 to 1-4 of oil. 
 
 Tubes. — It is well to arrange the tubes so that access to the more remote 
 ones permits of their receiving a fair amount of heat, and prevents those 
 nearer the fire from getting an undue amount. The tubes should be made 
 * If oil is burned the rate is 9 lb. per square foot of S.H.P. 
 
STAYS. 591 
 
 of solid drawn steel, of best quality. Copper has been tried, but the diffi- 
 culty of detection of a fault in solid drawn copper detracts very much from 
 its usefulness, and is more than a set-off for the superior conductivity and 
 the low corrosive character of that metal irrespective of the cost. The 
 Admiralty insist on steel tubes being galvanised on the outside and not on 
 the inside, the decomposition of the zinc when inside the tubes tending to 
 produce hydrogen gas, by which an explosion might be caused on the incautious 
 introduction of a lamp when empty. No doubt, for ships liable to be laying 
 up out of use for long periods, the galvanising is an advantage ; but for a 
 ship that is generally in use, the necessity for it does not exist. There is, 
 however, a further advantage in galvanising, inasmuch as in the process 
 of pickling, defects can be discovered in the tubes which might otherwise 
 escape detection ; scouring with a fast-revolving brush will also answer this 
 purpose. The small light boilers, employed in torpedo boats and torpedo 
 boat destroyers, have up-cast tubes generally at least an inch external dia- 
 meter and from 14 to 16 L.S.G. thick. They are expanded into the tube 
 plates in the usual way, and when stay tubes are not used, the ends are 
 slightly drifted so as to prevent drawing ; if the tubes are galvanised, it is 
 better to remove the zinc from the part fitting into the tube plate. For 
 general purposes an inch tube is too small, except in small craft; in 
 ordinary ships, as in the larger vessels for naval service, the tubes should be 
 1J to l| inches diameter in certain forms of boilers, such as the Thornycroft, 
 Yarrow, or Normand, and from 1£ to 3 inches in those boilers in which the 
 tubes are horizontal, or nearly so. Where weight is not of such paramount 
 importance, the tubes of the horizontal boiler may be even somewhat larger 
 still, especially those in the immediate vicinity of the fire. 
 
 Circulating Tubes. — The down-cast pipes are better to be of compara- 
 tively large size and few in number, so situated as not to be exposed to heat 
 sufficient to convert them into upcasts,* and to be in the way of the natural 
 horizontal flow of the water current in the steam receiver. These down-cast 
 pipes are, as a rule, 3 to 6 inches diameter, but when a water wall is 
 formed by tubes placed touching one another in the main part of their length, 
 but their ends worked zig-zag fashion at the upper and lower receivers, they 
 may be of the same size as the ordinary tubes, these, of course, acting as 
 down-casts. 
 
 Stays. — In designing a water-tube boiler the same care must be exercised, 
 as with the ordinary boiler, in arranging the stays so as to resist deformation 
 and destruction. Many engineers, following the example of their locomotive 
 brethren, depend on the ordinary tubes to effect this purpose ; and, no doubt, 
 so long as they are in normal condition they may be relied upon ; but when 
 the boiler is such that the tube ends might at any time become abnormally 
 hot and be suddenly cooled, it would not be safe to depend on ordinary 
 tubes for holding the parts together. When straight tubes are employed, 
 stay tubes may, of course, be fitted, as in the ordinary boiler. General stiff- 
 ness, as well as local stiffness, is usually obtained by making the tube plates 
 abnormally thick, and, when the tube plate forms the part of a cylindrical 
 drum, this thickness is necessary for strength to resist tangential strains. 
 
 * Mr. Yarrow's experiments show that the down-cast pipes may be heated to advan- 
 tage under certain conditions, but it is certain that the continuous flow must not be 
 jeopardised. 
 
592 MANUAL OF MARINE ENGINEERING. 
 
 Steam Drums. — The size of the steam receiver is dependent, as usual, on 
 the volume of steam produced, and in order to keep the size as small as 
 possible, as well as to prevent priming, it is customary to work these boilers 
 at a very high pressure, there being little difference in weight between such 
 a boiler and one made for a much lower pressure. If, however, the boiler 
 is to be worked at the same pressure as the engines, and that pressure is 
 only 180 lbs., the steam dome must be of considerable size ; in fact, it must- 
 follow the rules laid down for ordinary boilers ; it is better, however, to use 
 water-tube boilers with higher pressures, 200 lbs. being about the lowest 
 advisable. 
 
 The Two Types. — The modern marine water-tube boilers may be roughly 
 divided into two classes, the one having large or comparatively large tubes 
 and suitable for large ships and long runs, the other having small tubes and 
 suitable for small ships or for short runs at very high speed, and conse- 
 quently called " Express Boilers." Each of these can be subdivided into 
 two divisions, the one having drowned tubes (that is, tubes whose upper 
 ends are submerged in water), and the other having the upper ends exposed 
 to steam only when the circulation is maintained by priming or raising of 
 the water with the steam in the pipes, and the discharging of it into the 
 bottom of the steam receiver. It is claimed by the makers of the latter 
 class that the steam obtained is quite as dry as that produced by the sub- 
 merged tube boiler. Of course, dash plates and other means are provided 
 for separating the steam from the water besides the usual internal steam 
 pipe : even with submerged tubes it is sometimes found advantageous to fit 
 baffle plates so as to ensure steady circulation and prevent priming. The 
 design of the priming boiler is one that admits of a greater length of tube 
 and a longer exposure to heat than is usual with the other forms in which 
 the tubes are straight or nearly straight, and their advocates claim for them 
 an elasticity and power of resisting sudden changes of temperature superior 
 to that of the straight tube boiler. On the other hand, they have the obvious 
 disadvantage ot cost of manufacture as well as impossibility of internal 
 examination, besides the difficulty and costliness of removing and replacing 
 an injured tube. 
 
 Herreshoff Boiler. — In the very early days of the last century others, 
 besides Mr. Jacob Perkins, attempted to produce high-pressure steam with 
 the lightest possible apparatus, by making a boiler entirely of small tubes, 
 and another inventor went so far as to construct one consisting of a single 
 tube coiled roughly into the shape of a bell or beehive, and placed this over 
 a fire. Water was pumped into the lower end of this coil, and steam emitted 
 from the upper end. As might be anticipated, the generation of steam was 
 most rapid, and attended with considerable risk to the onlookers. So long 
 as the feed supply was regular and the water perfectly pure, this boiler 
 could be used with impunity ; but even now, as in those early days, no feed 
 pump could be relied upon to work with the necessary uniformity, and as 
 in practice, all water contains more or less solid matter, such boilers were 
 found impracticable, and the more so as, from the curved form of the tube, 
 it was impossible to examine or clean them mechanically. The Herreshoff 
 boiler is a modern form of this type, which has always been a favourite one 
 for certain purposes in America ; its inventor subsequently perfected a 
 design of it which worked with a considerable degree of success, and it was 
 
WARD S SECTIONAL COIL BOILER. 
 
 593 
 
 introduced by him into this country, at the time that Mr. (now Sir J. I.) 
 Thornycroft introduced a similar boiler as regards coils, and fitted one in 
 the s.s. "Peace" (1883). Eventually, however, it met with the fate of its 
 prototype, so that it is now seldom or never used. His boiler consisted of 
 
 Fig. 216,— Ward's Sectional Coil Boiler—Heating Surface, 2,4S0 square feet; 
 
 Grate, 77 square feet. 
 
 two coils of pipe ; the outer, which is in the form of a closed cylinder, is 
 made of tubes of uniform section, connected at their ends by welding so as 
 to be continuous : the inner, which is in the shape of a bell, is likewise made 
 
 38 
 
591 
 
 MANUAL O* MARINE ENGINEERING. 
 
 of continuous tubes, but those forming the sides of the bell are of larger 
 section than that forming the crown. The top of the outer coil is connected 
 to the top of the inner, the bottom of the outer is connected to the feed 
 
BABC'OCK AXD WILCOX BOILER. 
 
 595 
 
 pump, and the bottom of the inner to a steam chum or receiver, from which 
 the engine takes steam. The fire is in the inside of the bell, which is sup- 
 ported on brickwork, and is similar to that of a vertical donkey boiler. The 
 outer coil is cased with sheet iron, and the inner is partly so covered. The 
 water, in traversing the outer coil, is gently heated by the waste heat ; on 
 traversing the crown of the bell it is rapidly heated, so that when it enters 
 the larger pipe it is in a state of ebullition, part steam and part water ; when 
 it has traversed the whole length of the coil it should be all steam. Another 
 coil boiler, well-known in America, is Ward's ; it has been used in the U.S. 
 Navy, and of its kind is a good one {v. fig. 216). It, as well as Thornycroft's 
 and HerreshofFs, has, however, the fatal fault of inaccessibility for cleaning, 
 but with anthracitic coal giving off little or no smoke or tarry vapour, and 
 using pure water, this fault is not of serious consequence. 
 
 Babcock & Wilcox Boiler (fig. 217).— The two gentlemen whose names 
 were joined in the patent and gave the title to the Company that manufactures 
 these boilers, devoted many years to the perfecting of their system. In 
 this case the flat boxes of the Watt and D'Allest boilers are replaced with 
 narrow sinuous chambers (called " headers," fig. 218), placed nearly vertically 
 side by side. With this arrangement staying is unnecessary, but opposite 
 the tube ends are the usual openings, as in the above-named boilers, fitted 
 
 Fig. 218. — Headers for Small Tubes. 
 
 with doors or plugs. The top of each front header is connected by a pipe 
 to a drum of considerable size, placed immediately over the back headers, 
 and to which it is connected also by pipes. The front and back headers 
 are connected by a large number of inclined tubes ; but in the case of this 
 boiler they are of much smaller diameter than in others of the same type. 
 The circulation of the water is obvious and good, and of such a nature that 
 this boiler may be classed as a priming one. The furnace is, of course, 
 beneath the tubes, and extends nearly to the full width of the boiler. Some- 
 times on each side of the boiler, forming a water wall, is a system of small 
 vertical tubes, fitted into a horizontal box at the bottom, and into a similar 
 horizontal box of square section at the top. The top boxes are connected 
 also to the steam drum, and at the opposite ends the top and bottom boxes 
 are connected by external circulating tubes. These boxes and the " headers " 
 are of square section, and made of wrought iron or steel, welded up and 
 neatly finished. The tubes throughout are simply expanded in the tube 
 holes, and although it appears to be a most difficult operation to thus secure 
 them, and a still more difficult one to. tighten them if they leak, in practice 
 all such difficulties have been overcome, and these boilers found to work 
 most satisfactorily. This, however, is now seldom supplied, as brickwork 
 to the height of the lower tubes is better, and a good casing above in wake 
 
596 
 
 MANUAL OF MARINE ENGINEERING. 
 
 TABLE LVL- 
 
 -Basin and Sea-going Trials of H.M.S. 
 in 1898-99 (Babcock & Wilcox). 
 
 Sheldrake 
 
 Nature of Trial. 
 
 oq 
 
 E- 
 0) 
 
 I" 
 
 ■5.S 
 d 
 
 2 
 
 DO 
 - • 
 
 ^ > 
 
 %- <1> 
 
 C u 
 
 6 a 
 
 2 
 
 Duration in 
 hours. 
 
 as 
 
 ca to 
 168 
 
 a 
 
 4) 0> 
 
 *- "X 
 3 5 
 to .-- 
 to O) 
 
 Q.° 
 
 Si 
 
 .2*0. 
 
 
 
 Ctf 
 
 o* . 
 
 ■ 
 
 
 
 M 3 
 
 Ashes per 
 Pound Coal. 
 
 Grate 
 Surface. 
 
 Kvaporative, . 
 
 1116 
 
 
 
 169 
 
 150 
 
 10 5 
 
 
 126 
 
 ,, .» 
 
 2 
 
 2 
 
 8 
 
 179 
 
 1292 
 
 
 
 1-46 
 
 15 
 
 105 
 
 
 126 
 
 
 2 
 
 2 
 
 8 
 
 169 
 
 1761 
 
 
 1-78 
 
 25 
 
 9 05 
 
 
 126 
 
 
 2 
 
 2 
 
 8 
 
 165 
 
 1873 
 
 
 1-67 
 
 25 
 
 9 34 
 
 • • • 
 
 126 
 
 S hours at 2500 I. H. P., . 
 
 4 
 
 
 8 
 
 152 
 
 2642 
 
 
 
 1-43 
 
 150 
 
 
 ... 
 
 252 
 
 3 hours at 3000 I. H. P., . 
 
 4 
 
 
 3 
 
 151 
 
 4050 
 
 •43 
 
 1-57 
 
 25-6 
 
 
 
 .•■ 
 
 252 
 
 3 hours commissioning, . 
 
 4 
 
 
 3 
 
 119 
 
 2735 
 
 •14 
 
 1-64 
 
 17-8 
 
 
 
 
 252 
 
 1000 miles at 1500 1. H. P., 
 
 3 
 
 1 
 
 69 
 
 120 
 
 1303 
 
 
 
 1-61 
 
 12-8 
 
 
 
 •152 
 
 189 
 
 „ 1500 „ 
 
 3 
 
 1 
 
 68 
 
 120 
 
 1506 
 
 
 
 1-6 
 
 12 67 
 
 
 
 150 
 
 189 
 
 
 „ 1500 „ 
 
 3 
 
 1 
 
 70 
 
 135 
 
 1534 
 
 •o 
 
 1-75 
 
 142 
 
 
 
 •200 
 
 189 
 
 
 „ 1500 „ 
 
 3 
 
 1 
 
 68* 
 
 130 
 
 1539 
 
 
 
 159 
 
 131 
 
 
 
 •150 
 
 189 
 
 
 1S00 „ 
 
 3 
 
 1 
 
 67 
 
 135 
 
 1S29 
 
 
 
 1-6 
 
 15 4 
 
 
 
 •216 
 
 189 
 
 
 1800 ,, 
 
 3 
 
 1 
 
 66* 
 
 140 
 
 1838 
 
 •o 
 
 1-68 
 
 164 
 
 
 
 •220 
 
 189 
 
 
 „ 2000 „ 
 
 3 
 
 1 
 
 59 
 
 145 
 
 2033 
 
 
 
 1-57 
 
 170 
 
 
 
 •220 
 
 189 
 
 
 2000 „ 
 
 3 
 
 1 
 
 6U 
 
 140 
 
 2042 
 
 
 
 1-56 
 
 108 
 
 
 
 •230 
 
 189 
 
 
 ,, 2250 „ 
 
 3 
 
 1 
 
 56$ 
 
 150 
 
 2245 
 
 
 
 163 
 
 19 4 
 
 
 
 •250 
 
 189 
 
 Jc 
 
 tal, • 
 
 632J 
 
 ... 
 
 ... 
 
 ... 
 
 ... 
 
 • •• 
 
 ... 
 
 ... 
 
 Mean, .... 
 
 ... 
 
 ... 
 
 146 
 
 1974 
 
 ... 
 
 1-63 
 
 ... 
 
 9-85 
 
 •198 
 
 ... 
 
 of the tubes answers every purpose. It need hardly be said that the use 
 of clean, soft water is essential. In the later marine type the tubes adjacent 
 and near to the fires have been enlarged very considerably, so that, instead 
 of being about 1\ inches diameter, as were the other tubes, they are now 
 at least 2| inches diameter (v. fig. 218«). These boilers are lighter in. pro- 
 portion to their output of steam than the ordinary cylindrical boiler ; but 
 they are not so light as some of the other forms of water-tube boilers ; in 
 
 Fig. 218a. — Babcock Tube and Headers. 
 
 this case, however, there is not the same necessity as exists in others for 
 working at a high pressure to be successful, and in this respect it is much 
 superior to one or two of the other forms of water-tube boiler when required 
 to be of large size. In the American Navy, as in H.M. service, these boilers 
 have given great satisfaction, and proved themselves to be the most for- 
 midable competitors for general adoption, both against the cylindrical and 
 the other types of large water-tube boiler. The simplicity of design and 
 
B'ABCOC'K AND WTLCOX BOILER t 
 
 51)7 
 
 .cheapness of manufacture will also recommend them for use generally. In 
 H.M.S. " Sheldrake" these boilers, with best Welsh coal and a consumption 
 
 Fig. 219.— Babcock & Wilcox Boiler (Small Tube), Naval Design. 
 
 ■of 407 lbs. per square foot of grate, evaporated 9*15 lbs. water per lb. of coal, 
 from and at -212' F., and at a half that rate the evaporation was 11 lbs., the 
 
D!>8 
 
 MANUAL OF MARIN3 EKGINEERIKO 
 
 
 Fig. 220.— Babcocfc & Wilcox Narnl "Boiler (Mixed Tubes). 
 
BABCOCK AND WILCOX BOILER. 
 
 599 
 
 evaporation per square foot of heating surface being 8"3 and 5*0 lbs. respec- 
 tively. At sea, and when running full speed, the consumption per I.H.P. 
 
 Fig. 221.— Babcock & Wilcox Naval Boiler (Mixed Tubes). 
 
 was only 157 lbs., and at two-thuds power 1*429 lbs. Equally good results 
 have been obtained in U.S. Naval ships with Pocahontas coal. 
 
60U 
 
 MANUAL OF MARINE ENGINEERING. 
 
 The Babcock & Wilcox Company sometimes fit into the uptake what is 
 virtually an extension of the boiler, and call it a feed, heater ; this, of course, 
 materially adds to the economy of the system, as it also adds considerably 
 to the weight. 
 
 The Special Boiler Committee appointed by Parliament reported favour- 
 ably on this boiler, as one of those suitable for use in the Navy ; consequently 
 a large number of ships — mostly battleships and large cruisers — are fitted 
 with them. The Admiralty require such boilers to evaporate 12 lbs. of water 
 from and at 212° when the coal consumption is at the rate of 18 lbs. per 
 square foot of grate, 11 J lbs. when the rate is 24 lbs., and 11 lbs. when 30. 
 
 ITORMANB 
 
 BOILEK 
 
 Fig. 222.— Normand Boiler. 
 
 Fig. 217 shows the naval boilers, which are made entirely of large tubes, 
 and the baffling arrangements, which have been found beneficial on shore ; 
 this type is also the one best suited for the mercantile marine. Fig. 219 
 is a design of boiler made entirely of small tubes, suitable for cruisers and 
 express steamers generally. Fig. 220 shows the boilers of the cruiser type, 
 which are of a lighter design than fig. 217, having large tubes at the bottom 
 only, and smaller higher up ; no baffles are deemed necessary. 
 
 Normand Boiler. — Fig. 222 shows the form of boiler of ibhe late M. Nor- 
 mand. of Havre. It consists essentially of one horizontal upper drum of 
 considerable diameter, and two lower drums of much smaller diameter, the 
 
seaton's boiler. 601 
 
 lower ones being placed on each side of the fire, and connected to the upper 
 one by tubes of small diameter, bent to such form as to obtrude the greater 
 part of their length into the space above the fire, and so receive heat from 
 it, and at each end so as to be normal to the surface into which they fit. The 
 upper and lower drums are likewise connected at their ends by larger tubes, 
 which act as downcasts, so that the water circulates up through the smaller 
 tubes into the upper chamber, delivers the steam with which it is charged, 
 and flows away to the end of it, and down through the outer tubes to the 
 water drums below. This boiler exposes a very large amount of surface, 
 both to the direct action of the fire and to hot gases. The circulation is 
 perfect ; there are no flat surfaces requiring to be stayed ; the drums them- 
 selves are of the cylindrical form, with spherical ends, so as to be as light 
 as possible consistent with strength : and by the disposition of the tubes 
 at the front and back there is a minimum amount of brickwork required for 
 a boiler of this kind, while the tubes have a sufficient amount of curvature to 
 permit of considerable elasticity in the whole structure, so as to enable it 
 to withstand the rough usage of rapidly raising steam. On the other hand, 
 the tubes cannot be cleaned mechanically nor examined, either internally 
 or externally, and. in the event of a tube splitting, or becoming otherwise 
 so damaged as to necessitate removal, unless it be an outside one, great 
 expense is entailed by having to remove the other tubes that surround it to 
 get at it. M. Normand arranged the inner and outer rows of tubes to form 
 water walls on each side of the grate, and on each outside of the boiler ; he 
 left, however, some open work on each side of the grate near the front, through 
 which the hot gases can pass among the tubes on each side to the back end of 
 the boiler, where there are additional tubes massed behind the brickwork of 
 the furnace. The hot gases then pass upward into an uptake leading to 
 the funnel base. By this method the hot gases are made to flow uniformly 
 over practically the whole of the tube surface. The Normand boilers in 
 H.M.S. "Ferret" had 8,112 square feet of surface and 154 square feet of 
 grate, and gave steam for 4,774 I.H.P. Similar boilers in the French ship 
 " Forban " proved very economical, the consumption of fuel at full speed 
 being only 1-36 lbs. per I.H.P. with triple reciprocators. 
 
 The Normand Siguady boiler consists of two Normand boilers placed 
 back to back, and connected by the steam collectors and water drums. 
 
 Ferguson & Fleming's Boiler (fig. 223).— Messrs. Ferguson & Fleming 
 make a boiler similar in principle to the original Normand, and get over 
 the difficulty of drawing the tubes by limiting their length from end to end 
 to the diameter of the upper drum, so that each and every one of them can 
 be drawn into this drum without affecting the surrounding tubes. 
 
 Seaton's Boiler (fig. 224). — In this boiler there are two water drums at 
 the bottom, as in the Normand, and two steam drums at the top ; there is 
 an intermediate water drum between them of sufficient size to admit of the 
 tubes connecting it to the upper and lower drums being drawn into it and 
 removed. In this case the tubes are straight, so that from the middle drum 
 everv tube can be examined and cleaned, as well as removed ; the transverse 
 section of the boiler is the shape of the letter X. In addition to the advan- 
 tages of being able to examine, clean, or remove every tube at the middle 
 chamber, there is in this boiler the additional one of the hot gases being 
 forced to flow twice through the tube before emerging into the funnel. The 
 
80'2 
 
 MANUAL OF MARLNE ENGINEERING 
 
 upper and lower drums in this boiler are connected by a water wall of tubes, 
 which act as downcasts. 
 
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 Fig. 224a.— Seaton's Boiler. 
 
€04 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Yarrow Boiler (fig. 226). — This boiler, the invention of Mr A. F. Yarrow, 
 has proved to be one of the most successful in practice, both in the production 
 of steam, and its capability of being examined, repaired, or cleaned, and 
 it can, besides, withstand the same rough usage as the Normand. It differs 
 from the Normand boiler, however, in having straight tubes only, so that 
 
 Fig. 225.— Yarrow Boiler (Large Tube Tyjje), 
 
 from the upper drum they can all be quickly examined and cleaned. In the 
 earlier forms of it, this drum was made in halves, the top of it being con- 
 nected to the lower by a flange, with bolts and nuts ; it was, therefore, capable 
 of easy removal, and when this was so the bottom of the boiler consisted, 
 as in the Normand. of two cylindrical drums ; but as in the larger boilers 
 
YARROW BOtLER. 
 
 605 
 
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 it was found almost impossible, as well as inconvenient, to have a bolted 
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606 
 
 MANUAL OF MARINE ENGINEERING. 
 
 with a tube plate and pocket bolted to it, as shown in the illustration, and 
 to remove a tube necessitates the taking off of one or other of the pockets. 
 
 The Yarrow boiler was one of the few recommended by the Boiler Com- 
 mittee to the Admiralty for further trials and use in large ships, but with 
 larger tubes than usual in Express boilers. This kind (fig. 225) has been 
 fitted in many British and foreign warships to be used in connection with 
 cylindrical boilers. The lower or water drums are in this case made approxi- 
 mately cylindrical, like those of M. Normand, and the tubes are removed 
 and replaced by removing those in direct line with the ones requiring it, and 
 dodging from hole to hole as far as the slackness of the new tube in the old 
 holes will permit. The Yarrow boiler is now used very extensively in cruisers 
 and battleships, as well as in the smaller craft ; its straight tubes in a nearly 
 vertical position render it a very trustworthy and efficient one. It has also 
 
 Fig. 227. — Japanese Navy Boiler (Yarrow Type) for Destroyers and Torpedo Boats. 
 
 been adopted in foreign Navies for cruisers and small craft. Fig. 227 is an 
 example, being the particular form of it developed by the Japanese. 
 
 Blechynden's Boiler (fig. 228) is like the Yarrow as now made, but the 
 tubes are very slightly curved, and the upper end enlarged so as to permit of 
 them being withdrawn through a few small holes on either side of- the top 
 middle line of the steam drum : in recent boilers the tubes have been arranged 
 to draw through one set of holes on this centre line. A double row of tubes 
 forming water walls on each side, with the usual openings at the upper part, 
 was often added to act as down-comers as well as walls. 
 
 White Forster's Boiler is somewhat like Blechynden's, but it has the 
 tubes curved in a longitudinal plane so as to be drawn into the drum and 
 through the manhole at its end. 
 
BLECHYNDEN S BOILER. 
 
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 MANUAL OF MARINE ENGINEERING. 
 
 
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DU TEMPLE BOlLEtt. fi09 
 
 Thornycroft's Boiler. — The boiler known as the 'Speedy" type was the 
 direct descendant of Goldsworthy-Gurney's invention of 1827 (see fig. 215), 
 and consisted essentially of the top and bottom drums, as in the Normand 
 and Yarrow type?, but the tubes in this case were not only bent in so as to 
 obtrude themselves on the space over the fire, but they were carried up and 
 around so as to enter the top of the drum instead of the bottom ; it is, there- 
 fore, distinctively a priming boiler, inasmuch as there can be no circulation 
 without the flow of water through the tubes into the steam space. There 
 were the usual downcast pipes at each end, which were of considerable size. 
 In this boiler there is. of course, a greater length of tube exposed to the hot 
 gases, and consequently, if necessary, a large amount of surface per foot 
 of grate can be provided. It had considerably more elasticity than the 
 Normaud type of boiler, and likewise possessed the defects of the latter, as 
 it was absolutely impossible to examine the tubes, and, in case of damage to 
 one of them, the difficulty of removal and replacement considerable. < wing 
 to the extreme curvature of the tubes, there is the greater liability of the 
 tubes to choke with foreign matter that may be in the boiler, or get into it 
 with the water. For these reasons it was objected to by engineers, and 
 soon discarded by the Admiralty, although experience proved it to be a very 
 efficient evaporator, a rapid generator of steam, and capable of very rough 
 usage. Fig. 229 is a further invention of Sir J. I. Thornycroft, and known 
 as the " Daring " type. With the same facility for getting sufficient heating 
 surface as with the Speedy, this boiler permits of a larger grate, and allows 
 of a better circulation of hot gases as well as better uptake and funnel ; it 
 has. however, the same objections due to the great curvature of its tubes. 
 
 Mumford's Boiler (fig. 230) is a modification of the Fleming & Ferguson 
 (fig. 223). made to suit the service conditions in cruisers and small craft ; it is 
 very simple in design and construction, and although the curvature is too great 
 to permit of a visual examination of the tubes, it enables them to resist the 
 tendency to contortion from the severe heating some of them get when working 
 under forced draught. Moreover, each one of them can be withdrawn into 
 the steam drum, and from it passed out for careful examination in daylight, 
 and likewise replaced in an equally easy and expeditious way. These boilers 
 are found to be safe and economic in working of launches, pinnaces, and 
 torpedo boats, and equally efficient when of the larger sizes now made by 
 Mr. Mumford, where the total heating surface in a single-ended unit is 1,180 
 square feet, the tubes in that case being If inches diameter, 10 S.W.G. thick, 
 and varying from 4'7 to 6"0 feet long ; its steam drum has a diameter of 
 5 feet. 
 
 Du Temple Boiler. — This consists of the usual upper drum, which is of 
 considerable size, and two lower chambers, of square section, placed above 
 the fire level on each side of the grate. The top drum is connected to them 
 by a nest of tubes bent zig-zag, so as to cross the flow of gases five times. 
 The portion of tube exposed to direct action from the fire is generally made of 
 smaller diameter than the rest of its length. There are the usual down-cast 
 pipes at the end. This boiler possesses many of the characteristics of the 
 Thornycroft boiler, including some of the evils in a magnified form. The 
 contraction of diameter of the tubes in the neighbourhood of the fire has 
 certain advantages ; but, on the other hand, modern practice has shown 
 that there should be larger tubes exposed to direct radiation from the fire, 
 
 39 
 
610 
 
 MANUAL OF MARINE ENGINEERING. 
 
 iilK-iii 
 
 
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HEED'S BOILER. 
 
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 and that those tubes in the immediate vicinity, although not exposed to 
 direct radiation, are the better for enlargement. 
 
 Reed's Boiler. — Fig. 231 is a view of the boiler designed and patented by 
 
 Mr. J. W. Reed, of Jarrow. It will be seen to consist of the usual steam 
 receiver at top and the two wing water receivers at the bottom, both of 
 which are in all essentials the same as found in the Normand. Thovnycroft, 
 
612 
 
 MANUAL OF MARINE ENGINEERING. 
 
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BELLEVTLLE ROILER. 613 
 
 or Yarrow boilers, but in this case they are connected differently — viz., by 
 tubes zig-zagged somewhat like those in the Du Temple boiler, but not to the 
 same extent, nor is there any reduction in the diameter of these small tubes 
 as in the Du Temple. Mr. Reed has, besides, methods of his own for securing 
 the small tube ends in the tube holes, whereby a tube can be easily fitted and 
 withdrawn ; and. when in place, it is capable of a certain amount of self- 
 adjustment without any tendency to leak. This is effected by screwing the 
 ends of the tube and fitting them with nuts formed to suit the chamfering or 
 spherical recessing at the mouth of the tube holes. 
 
 Of the " Express " boilers none have done better than Mr. Reed's, either 
 in respect to the quantity of steam generated per square foot of heating 
 surface or to the number of pounds of water evaporated per pound of fuel 
 When severely forced, it seems to suffer little or no harm from the process, 
 and at cruising speeds it is very economical. It, however, has the objection 
 common to the Normand, Thornvcroft, and certain other boilers, of having 
 the tubes so considerably bent as to prevent examination internally. 
 
 Belleville Boiler (fig. 232). — This boiler had been for very many years 
 a favourite in France, and fitted into many warships, as also into a large 
 number of their merchant ships. It had also been used extensively in ; 
 America, both for marine and land purposes, when it was at last more favour- 
 ably looked upon in this country, the lead in the matter being taken by 
 the Admiralty, who replaced the wet-bottom locomotive boilers of H.M.S. 
 " Sharpshooter " with boilers of this kind : the first-class cruisers " Powerful " 
 and ;i Terrible." each of 25,000 I.H.P., and many other ships, cruisers, battle- 
 ships, and gunboats were afterwards fitted with them. For large ships this 
 boiler seemed suitable, and can compete successfully with the other forms 
 of water-tube boiler in the question of weight : for smaller sizes, however, 
 several of the before-mentioned boilers have advantages over the Belleville. 
 It was for some time the cause of considerable trouble and anxiety, and 
 although, now when they are better understood and more carefully worked, they 
 do good service but they are no longer fitted to new ships. It consists of the 
 usual top drum and bottom feed collector, connected by what is virtually a 
 set of flattened spirals ; in other words, each element is of the form that a spiral 
 spring would be if heated and flattened. There are, of course, the usual down- 
 cast pipes for circulating, besides special apparatus for preventing priming, 
 for feeding, and for reducing the pressure. This boiler, like most of its con- 
 geners, must, to be successful, be worked with steam at a high pressure, 
 or, to speak more correctly, with steam of low volume per lb. ; consequently, 
 the manufacturers of it usually design it for a much higher working pressure 
 than is required, and fit it with an ingenious, and fairly reliable, reducing 
 valve. Each element is made up of a series of tubes of considerable diameter, 
 extending the whole length of the boiler, connected at the ends by junction 
 boxes of malleable cast iron (fig. 232a), to which they are screwed, and they 
 are zig-zagged so as to form the flattened spiral. These front junction boxes 
 have, opposite each tube end, a hole of sufficient size to admit of the thorough 
 examination and cleaning of the tubes, and closed with an oval door clamped 
 in the same way as a manhole door. The top of each element enters the 
 upper drum with a short internal pipe, and inside the drum is a series of dash 
 plates, so arranged as to thoroughly separate the water and the steam. There 
 is further, however, external to the boiler, a separator, having an internal 
 
614 
 
 MANUAL OF MARINE ENGINEERING 
 
 F:=9 
 
 Fig. 232. — Modern iielleville Boiler, with Econoniwei 
 
BELLEVILLE BOILER. 
 
 615 
 
 spiral dash plate, and, at the bottom, an automatic drain. This boiler is 
 a most costly one to manufacture, and even when made with an accuracy- 
 exceeding that usually followed in engine building it loses so much water by 
 the very large number of small leaks as to seriously reduce its efficiency, 
 as well as to require a large reserve of feed water and a constant use of the 
 distillers. After a most lengthy and carefully made series of experiments 
 with it and the cylindrical boiler, the Committee appointed by Parliament 
 condemned its use in H.M. Navy on these and other grounds ; at the same 
 time, it must be said that, while recommending that all large ships should 
 have a few cylindrical boilers for cruising and use in port, certain other 
 forms of water-tube boilers should be further tried and worked tentatively 
 with the object of using them in such ships solely when high speed is necessary. 
 Now water-tube boilers are exclusively used. 
 
 The boilers so recommended, besides the Babcock & Wilcox, and the 
 Yarrow with large tubes already described, were the Diirr boiler and Niclausse 
 boiler, both well known and used on the Continent ; the latter in the French 
 
 lug. 232a. — Tube and Headers of a Belleville Boiler. 
 
 and the former in the German Navy are favourites. Each of them has its 
 origin in the patents of Jacob Perkins of 1831, and Howard of 1869. Perkins 
 adopted an internal tube to cause a circulation in the large one exposed to 
 fire ; this idea was enlarged and improved by Field, by whose name such 
 tubes are known instead of by Perkins. Howard placed his large tubes at 
 a small angle to the horizontal and connected their front ends to a vertical 
 chamber, and fitted them with concentric circulating tubes as far as the 
 water level. In 1870 Miller improved on this by fitting a vertical division 
 plate in the chamber, into which plate he fitted the circulating tubes, and 
 placed on top of the chamber a horizontal cylindrical receiver, with which 
 both divisions of the vertical chamber had easy communication, so that 
 the water flowed down into the front or outer part of it and through the cir- 
 culating tubes into the main tubes, and from them into the back or inner 
 part of the chamber and thence to the receiver, where it gave up the steam, 
 and flowed back again over the same course. 
 
C!6 
 
 MANUAL OF MARINE ENGINEERING. 
 
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JJJCLAUS6E BOILER. 
 
 617 
 
618 
 
 MANUAL OF MARINE ENGINEERING. 
 
 The Durr Boiler is substantially the above, with improvements in detail. 
 
 The Niclausse Boiler (fig. 233) differs from both Howard's, Miller's, and 
 Diirr's in having a series of vertical boxes or headers fitted with a division 
 plate, and having properly formed mouthpieces to allow of the tubes having 
 a considerable slant while they themselves are vertical. These headers are 
 made of steel or malleable cast iron, and connected to the under side of the 
 steam receiver by neck pieces. The tubes in this case pass through from 
 front to back of the headers, and fit quite steam-tight in a slightly-taper 
 hole in the division plate and back. The portion of the tube between these 
 is pierced with large oval holes to permit the water passing freely from it 
 to the inner division of the header. The inner or circulating tube is fitted 
 through a hole in the plug fitted in the end of the large tube (fig. 233a). and 
 held in place by a neck piece which fits through and steam-tight in a hole 
 in the stopper of the hole of the larger tube. The back ends of the big tubes 
 are reduced in diameter, screwed, and fitted with cap nuts, which are taken 
 
 Tig. 233a. — Niclausse Tubes as fitted in place. 
 
 off when the boiler is to be drained dry ; therein lies one of the chief objections 
 to this and the Durr boiler, as the time taken to remove and replace so many 
 cap nuts is serious in all cases, and especially so in ordinary mercantile ships 
 with a limited staff. So long as distilled water is used to make up waste 
 there will be little need to remove them for cleaning, and it would only be in 
 ' naval ships where so much harbour service would require frequent drying out. 
 The Niclausse boiler, as shown in fig. 233, is a favourite one in France, 
 and used in the French Navy ; it has also been used extensively in the Navies 
 of the United States of America and of Japan. It possesses several notice- 
 able features, but perhaps the strongest one to recommend its use is that there 
 is a positive and constant circulation of water in each and every tube, and 
 a steady transmission of the steam generated in them to the steam chest 
 without in any way interfering with it. In the English Navy it has proved 
 to be an economic and reliable generator of steam, and in that respect com- 
 pares favourably with the other types of water-tube boilers, and when the 
 
STEAM TRIALS WITH NICLAISSE BOILERS. 
 
 619 
 
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620 
 
 MANUAL OF MARINE ENGINEERING. 
 
 fleet ran from Gibraltar to Quebec as a trial of endurance, etc., the following 
 facts are worth recording : — 
 
 Tons. 
 
 Coal consumed by H.M.S. " Bedford " with Belleville boilers was 1,378 
 
 „ "Essex" „ „ „ 1,334 
 
 ,, ,, "Cornwall" ,. Babcock & Wilcox 
 
 boilers was 1,245 
 
 „ "Cumberland" ., Belleville boilers was 1,223 
 
 „ „ "Berwick" ,, Niclausse „ 1,161 
 
 Fig. 234. — Stirling Boiler ^ Marine Design). 
 
 Table lix. gives some results of trials with these boilers in various 
 ahips of the British Navy, which are altogether quite satisfactory considering 
 
I'HK HOHENSTEIN I30ILER. 
 
 621 
 
 that the evaporation is equal to about 10 lbs. of water per lb. of coal at full 
 power The water consumption of county class of cruiser per I.H.P. is 
 very high at full power. 
 
 The Stirling Boiler (fig. 234), which bears a strong family likeness to 
 Rowan's boiler without its defects is now coming into use for marine work 
 after having given satisfaction to its users on shore. It has the merit of 
 simplicity, and practically straight tubes of good size placed nearly vertically. 
 
 The Hohenstein Boiler. — This boiler (fig. 235) is a favourite one in the 
 
 Iff. o 
 
 Inlnluhil 
 
 Fig. 235. — Hohcuatein Boiler. 
 
 United States, for general use in the Navy of that country. It has been 
 experimented upon very considerably by United States oinoials, and with 
 good results. It consists essentially of two steam receivers at top, parallel 
 with one another, and two water receivers at bottom, immediately below 
 the steam receivers and connected to them by downcast pipes: or there 
 
022 
 
 MANUAL OF MARINE ENGINEERING. 
 
 may be only one at the back, as shown in fig. 235. Extending above these 
 water receivers are rectangular section chambers or headers, and similar 
 chambers extend from and below the steam receivers. These chambers 
 are connected by a pair of pipes placed diagonally across the fire and 
 connected at their remote ends by a junction box. The water, therefore, 
 circulates by rising from the water receivers into the chambers, passing 
 upwards through the tubes connected to it through the junction boxes, 
 and so through the second set of tubes to the chambers underneath the steam 
 receivers, and from them into the steam receivers themselves, and. finally, 
 back again by the downcasts to the water receivers. It will be seen that the 
 circulation is positive and regular. The tubes thoroughly intercept the heat 
 in its passage from the grate to the chimney. They are straight, and easily 
 examined, cleaned, or removed. 
 
 SCALE Of rttt 
 I Z 
 
 Fig. 236. — Miyabara Boiler of I. Japanese Navy. 
 
 The Miyabara Boiler (fig. 236), the invention of Admiral Miyabara of the 
 I. Japanese Navy, is ingenious and somewhat like the Hohenstein, but con- 
 sists of two sets of cylindrical drums placed parallel, and some 6 to 8 feet 
 apart ; each set consists of 3, one above the other. The middle drum of 
 each is connected to the top and bottom drums of the opposite set by 
 tubes 2 inches to 3 inches diameter. The top drum in each set is con- 
 nected to the bottom one by downtakes. It is very simple, and easy to 
 make, having nearly straight tubes with a good inclination ; the hot gases 
 will be well broken up and distributed, and limited spaces are provided 
 for them to come in contact with the outer casing. It is now used in 
 
THE THORNY CROFT-MARSHALL BOILER. 
 
 623 
 
 the cruisers and battleships of Japan, and the following is given as the 
 results of experiments: — 
 
 TABLE LX. — Trials of Miyabara Boiler. 
 
 
 Xo. 1. 
 
 No. 2. 
 
 No. 3. 
 
 No. 4. 
 
 Type of boiler, .... 
 
 Miyabara 
 
 Miyabara 
 
 Miyabara 
 
 Miyabara 
 
 Nature of trial, . . . 
 
 Coal only 
 
 Coal only 
 
 Coal only 
 
 Coal only 
 
 Duration of trial, . . in hours 
 
 3 
 
 3 
 
 8 
 
 3 
 
 Grate area (G.S.), . . in sq. ft. 
 
 53-906 
 
 53-906 
 
 53-906 
 
 53-906 
 
 Heating Surface (H.S.), . „ 
 
 1911-77 
 
 1911-77 
 
 1911-77 
 
 1911-77 
 
 Ratio of — ', . . • • 
 
 35-466 
 
 35-466 
 
 35-466 
 
 35-466 
 
 Barometer, . . .in inches 
 
 29-990 
 
 29-818 
 
 30-123 
 
 29-990 
 
 Boiler pressure, . . in lbs. 
 
 233-84 
 
 233-95 
 
 235-88 
 
 233-00 
 
 Total fuel burnt per square foot of 
 
 
 
 
 
 grate per hour, . ' in lbs. 
 
 17-721 
 
 29-535 
 
 37-411 
 
 44-302 
 
 Total fuel burnt per square foot of 
 
 
 
 
 
 heating surface per hour, in lbs. 
 
 0-4997 
 
 0-8328 
 
 1-0548 
 
 1-2492 
 
 Quantity of ash and clinker during 
 
 
 
 
 
 trial, ... in lbs. 
 
 210-729 
 
 373-959 
 
 767-966 
 
 285-714 
 
 Kind of fuel, .... 
 
 Welsh coal 
 
 Welsh coal 
 
 Welsh coal 
 
 Welsh coal 
 
 Temperature of feed water, at F. ° 
 
 75-05 
 
 75-00 
 
 73-33 
 
 74-63 
 
 rr, /'Outside of boiler room, 
 Temperature t F ° 
 
 ., , 1 Inside of boiler room, 
 Atmosphere | at F 
 
 
 
 
 
 79-68 
 
 85-37 
 
 83-16 
 
 80-95 
 
 
 
 
 
 91-47 
 
 91-16 
 
 89-84 
 
 92-37 
 
 Water evaporated from and at 212° 
 
 
 
 
 
 F. per lb. of fuel, . in lbs. 
 
 10-439 
 
 9-2646 
 
 9-1603 
 
 8-2293 
 
 Water evaporated from and at 212° 
 
 
 
 
 
 F. per square foot of heating 
 
 
 
 
 
 surface, . . in lbs. 
 
 5-2162 
 
 7-7150 
 
 9-6627 
 
 10-2800 
 
 The Thornycroft-Marshall Boiler (fig. 237) is also one with large tubes, 
 that seems a good design for marine or land. Its sectional form, which 
 is a good one for ship use, consists of a steam drum, with headers 
 standing vertically downwards, capable of taking twp vertical rows of tubes, 
 and spaced so that similar headers standing upwards from the feed collectors 
 or water drums may fit between them. The bottom hole in the lower header 
 is fitted with a tube, whose back end fits into a junction box like that of 
 a Belleville boiler ; another tube is fitted into this box, and the bottom 
 hole of the header from the steam drum, and so makes the connection from 
 the feed to the st«am receiver ; and so on with the other elements, as shown 
 in fig. 237, like those in the Hohenstein. Here, too, the tubes are straight, 
 and at a good angle ; they can be readily examined and replaced or cleaned, 
 and the -circulation is certain and good. This boiler can also be readily 
 drained and cleaned. 
 
 There are some other boilers worthy of notice in a special treatise, but 
 not, however, of the same interest to the marine engineer as the foregoing. 
 
 The water in the water- tube boiler is, as a rule, very considerably less 
 than that in the ordinarv boiler, and to this fact is in no small measure due 
 the exceedingly light weight of certain boilers of this type. To the designer 
 
624 
 
 MANUAL. OF MARINE EXGINfEKRTXO. 
 
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THE THORNYCROFT MARSHALL BOILER. 625 
 
 this is of course a great advantage, and in ease oi accident, cither from within 
 or from without — from a shot, for example — the possible amount of horin 
 is limited, especially if there be no obstruction with the draught : but to the 
 firemen it is otherwise, being a source of trouble, as it permits of such great 
 variations in steam pressure from temporary causes. The automatic reducing 
 valve, as fitted to the Belleville and other boilers, tends to prevent any ill 
 effect from this being experienced in the engine-room, for, so long as the 
 pressure in the boiler does not fall below that at which the apparatus is set 
 to deliver, there will be little or no variation at the high-pressure valve-box. 
 The total weight in the boiler-room includes, besides the boiler proper, 
 the water, mountings, and fittings ; the furnace fittings, brickwork, etc. ; 
 the casings and uptakes to the funnel base ; in fact, everything is included 
 that is necessary to make a fair comparison. Taking naval boilers of the 
 old double-ended type, the average weight of water is about 29 per cent, 
 of the total weight ; while with single-ended naval boilers, it is about 26 per 
 cent. The steam pressure for which these boilers were designed is, however, 
 only 155 lbs. With double-ended ■ boilers made in accordance with the 
 Board of Trade Rules for pressures of 160 lbs., the water was about 33 \ per 
 cent, of the total weight, and about 32 per cent, with the single-ended boilers. 
 With the gunboat type of boiler, the water is 33 per cent, when designed 
 in accordance with the Admiralty Rules. With the Belleville boilers the 
 water was only about 8 per cent, of the total weight ; with the Babcock & 
 Wilcox boilers it is about 14i per cent. ; with the Yarrow boiler it is about 
 15| per cent. ; with the Thornycroft boiler 15 per cent. ; and with the 
 Normand boiler about 24 per cent. The dry-bottom locomotive boiler, 
 which was the rival of the smaller class of water-tube boiler, has water to 
 the extent of 30 per cent. 
 
 The total weight per I.H.P. at forced draught was on the average with 
 the ordinary old naval double-ended boilers about 88 lbs., and with the 
 single-ended ones 108 lbs. The lightest of all was 80-6 lbs. The Belleville 
 and the Babcock & Wilcox are both under 80 lbs., and with an air pressure 
 of 1 inch of water the weight is as low as 70 lbs. with the Belleville. Taking 
 the naval cylindrical boilers with natural draught the average weight per 
 I.H.P. was 121 lbs. with double-ended boilers, and 133 lbs. with single-ended. 
 It is true that the latter figure was for the larger sizes for battleships and 
 cruisers. With the Belleville boilers the weight per I.H.P. is 82 lbs., and 
 the Babcock & Wilcox 88*7 lbs. ; and with the large allowance provided 
 in the cruisers it was 110 lbs., or 11 ibs. less than the old double-ended boilers 
 for 155 lbs. pressure, and 23 lbs. less than the single-ended boilers for 155 lbs. 
 pressure as supplied to ';he Navy. It is manifest by these figures that there 
 is a distinct saving in weight with the water-tube boilers in large ships. 
 
 It is, however, in the smaller class of vessels that the great advantage in 
 using these boilers is experienced. Here types of boilers can be employed that 
 would hardly be admissible in the larger ships *and the full advantage of the 
 system can be utilised. By again referring to the table, it will be seen that 
 with the locomotive boiler the weight per I.H.P. is 33 lbs., while with the 
 Yarrow boiler the weight is only 25 lbs., considerably under the locomotive, 
 as are all the other forms now fitted into torpedo boat destroyers. The great 
 saving in each case is, of course, due in large measure to the small auantity 
 of the water. Referring to Table Ixv. it will be seen that, when taking all 
 
 * The Yarrow boiler, with somewhat larger tubes than formerly obtained, is now fitted in all elasseg 
 of war ship and in large and small express passenger ships. . „ 
 
626 MANUAL OF MARINE ENGINEERING. 
 
 the weights which come into and are included in the boiler-room weights 
 according to Admiralty Rule, for each ton with single-ended boiler, etc., 
 1753 I. H. P. was developed by the engines; with double-ended boilers, 
 201 to 23-8; with Belleville boilers, 21-7 to 22*5; with Thornycroft, etc., 
 in 3rd class cruisers. 431 ; with locomotive in torpedo gunboats, 29*5 ; with 
 Babcock & Wilcox, in same ships, 325 ; with Niclausse, 26*3. With the 
 express boilers in torpedo boat destroyers as much as 95 - 9 I.H.P. has been 
 developed with Thornycroft's, 84 - 6 with Normand's, 79'1 with Yarrow's, 
 and 706 with Reed's, as against 60 with locomotives in the " Havock." 
 
 When the cylindrical boiler is designed in accordance with the Board 
 of Trade Rules, the difference is more marked, for here the double-ended 
 boiler is such that the total weight per I.H.P. is 149 lbs. when the working 
 pressure is 160 lbs., and as much as 196 lbs. when it is 210 lbs. pressure. 
 Single-ended boilers are, of course, heavier still. The Belleville boiler is 
 only 107 lbs. per I.H.P. with a working pressure of 250 lbs., and the Babcock 
 & Wilcox 115 lbs. for a pressure of 200 lbs. 
 
 The space occupied by these boilers does not differ materially from that 
 required for cylindrical or locomotive boilers, and the stokeholes, as a rule, 
 must be as long ; but with some water-tube boilers a comparatively short 
 stokehole is sufficient. The water-tube boiler is capable of subdivision to 
 an extent not possible with the cylindrical, and in this respect it is in the 
 hands of the designer much as the old box boiler was. 
 
 Feed Arrangements. — It is, of course, essential to the good working of 
 these boilers, as of all boilers containing such comparatively small quantities 
 of water, that the supply of feed-water shall be regular and delivered in such 
 a way as to improve rather than to retard the circulation. To satisfy the 
 first condition, it is necessary to have good feed-pumps and a good store of 
 water, from which they may draw whenever required. The ordinary engine 
 pump drawing from the hot-well is always more or less intermittent in its 
 action, and while no great inconvenience is felt when it is supplying the 
 ordinary tank boiler or the larger kinds of water-tube boiler, it would not 
 suit the more delicate ones. It may, therefore, be taken, as a rule, that 
 this class of boiler should have an automatic apparatus that effectively 
 controls the feed. The Belleville boiler was always fitted with an arrange- 
 ment which consists essentially of a chamber in which there is no disturbance 
 by circulation or other cause, but so connected to the boiler that the water 
 level in it is the normal level of the water in the boiler when " standing." 
 It contains a " float " arrangement which operates in an ingenious, if some- 
 what complicated way, a controlling valve on the feed-pipe, so that, as the 
 water falls in the boiler, the feed-water is allowed to pass, and as the water 
 rises above the working level it is gradually shut off. Sir J. Thornycroft. Mr. 
 Mumford, and others have somewhat similar methods for regulating the supply. 
 Other engineers, following Mr. Yarrow, have, however, preferred to fit to each 
 boiler an independent donkey pump, and to set it working with steam from 
 its own boiler at such a speed as to just keep up the supply. This method, 
 while convenient, cannot be said to be automatic ; it might, however, be so 
 if to the feed-valve, or even to the steam-valve, of the donkey is fitted a 
 means of controlling by a similar apparatus to the above. As the variation 
 of level in most water-tube boilers is somewhat limited, the " float " apparatus 
 cannot have a great range, as may be seen in fig. 279. 
 
 • To-day with turbines and more modern designs of these boilers 25 per cent, better output of S.H P. 
 if obtained. 
 
WATER CONSUMPTION TRIALS. 
 
 627 
 
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632 MANUAL OF MARINE ENGINEERING. 
 
 Mr. Yarrow, however, adopted an ingenious arrangement for controlling 
 the feed donkeys which is at the same time very simple and effective. There 
 is an internal steam pipe in the steam receiver specially fitted for the feed 
 donkey ; its open end is surrounded by a small trough, and is at the highest 
 working level of the water. (The trough is shaped like an inverted cone, 
 and so keeps the water fairly smooth within it.) If the water rises above 
 this point, water enters the steam pipe, gags the pump, and is carried through 
 its exhaust pipe to the condenser or hot-well tank. 
 
 It is a disputed point whether to admit the feed-water into the upper or 
 lower chamber ; but if the principle of aiding and not obstructing circula- 
 tion is observed, there need be little or no controversy on this point. Each 
 case must be -dealt with on its own merits ; but, as a general rule, unless the 
 feed-water is well heated, it is better admitted to the bottom chamber or at 
 the downcast part of the upper chamber in a direction of the downward 
 Mowing currents ; by preference the latter, as it is more likely to aid circulation. 
 In this case, the feed-pipe may with advantage be made small, so that the 
 water is injected at a high velocity. When the water is well heated before 
 admission to the boiler proper it is not of much importance where it is taken, 
 but an internal pipe with small outlets to insure distribution should be fitted. 
 Mr. Yarrow caused the incoming feed to be heated by passing through the 
 outer rows of tubes. 
 
 It is usual and advisable also to fit water-tube boilers with a settling 
 chamber into which any solid matter admitted with the feed-water can be 
 deposited and blown out. The principle on which these act is that solid 
 matter of higher specific gravity than the, water will deposit in an enlarge- 
 ment of a channel where there is a reduction in the velocity of flow ; a change 
 in direction of flow also helps to deposit. A certain amount of lime solution 
 is used with advantage in these boilers, and a means for admitting it is 
 necessary. As a rule, it may be introduced with the feed-water by placing 
 it in the tanks. 
 
 In arranging casings and uptakes great care should be taken to insure 
 a good distribution of the hot gases over the whole of the tubes. The ten- 
 dency to shorten circuit is, of course, strong, and to prevent it considerable 
 judgment is necessary in the placing of baffle plates and other devices. It 
 is a matter of experiment, however, to so place them as to get the best results. 
 
 The fire-places in water-tube boilers are generally of such a nature as to 
 require a brickwork erection ; in fact, in the case of the Belleville boiler the 
 brickwork was very extensive, but in the case of most of the others, and 
 especially of the lighter kinds, the brickwork is limited to the immediate 
 neighbourhood of the fire, the rest of the boiler up to the base of the funnel 
 being enclosed with light ironwork, either lined with asbestos, tiles, or in some 
 other way protected from the direct action of the heat of the gases. As far 
 as possible it is better to use water walls than brickwork, both of which are 
 only absolutely necessary to resist contact of the flame or direct radiation 
 from the glowing fuel. Casing plates outside the tubes, not subject to the 
 direct action of the fire, may be of very thin sheet steel, especially if stiffened 
 with light angles or by corrugations ; it is generally made portable, so that 
 it can be easily removed for the cleaning of the outside of the tubes or the 
 repairing of the boiler when necessary. Sheet asbestos £ inch thick forms 
 a very good protection to these thin casings ; it requires to be carefully 
 
FIRE-PLACES IN WATER-TUBE BOILERS. 633 
 
 secured and protected from damage, and for this purpose it is often com- 
 pletely covered by thin sheet steel, but this covering should be avoided if 
 possible. Two layers of asbestos millboard with a space between them 
 practically sealed, so as to prevent air circulation, makes a still better shield, 
 and with care in fixing it will withstand a considerable amount of wear and 
 tear. The Admiralty do not now require the fitting of dampers to the up- 
 takes of water-tube boilers as found in the torpedo boat destroyers, and, 
 indeed, the neoessity for them does not exist, while their presence may be a 
 source of danger. The evaporation with these boilers is quickly checked by 
 leaving the fire door open, or, much better still, by shutting off altogether 
 access of air to the fire. Without the damper the bursting of a tube or other 
 serious leakage is not dangerous, as the steam escapes immediately up the 
 funnel. The Admiralty also require to be fitted an apparatus for drenching 
 the fires with water, which can be operated both from the stokehole and 
 on deck. The tubes do not get covered with soot and dust so badly as might 
 be expected, and when Welsh coal is used in the boilers the deposit is easily 
 removed. Oil fuel, however, is now almost exclusively used. 
 
 The life of these water-tube boilers is rather too short to be satisfactory, 
 and the trouble experienced with the Belleville, due, however, in no small 
 measure to the want of care and judgment by those in charge of them, caused 
 them to be condemned. Others less in first cost, and not having such delicate 
 fittings and casings, etc., not Tequiring so much care and attention, might 
 be more generally used for express steamers as they are in naval ships. It 
 is, however, imperative for success in either service that the tubes last at 
 least as long as those of a cylindrical boiler ; that they be straight, or nearly 
 so, and inclined at a good angle with the horizon — say, at least, 10 degrees ; 
 that the circulation be certain, uniform, and rapid ; that the surfaces exposed 
 to heat are to be easily cleaned and kept clean ; that the fittings be as few 
 in number as possible and absolutely tight ; that the casings be so designed 
 as to remain air-tight and to keep good as long as the boiler itself, and that 
 the grates be of such a size and so placed as to get good combustion as well 
 as uniform distribution of the heat produced either by coal or oil. 
 
634 MANUAL OF MARINE ENGINEERING. 
 
 CHAPTER XXIV. 
 
 HOILERS — CONSTRUCTION AND DETAIL. 
 
 Boilers for the mercantile marine are nearly always constructed in accord- 
 ance with the rules prescribed for the scantlings by the Board of Trade, 
 Lloyd's or other Registry. If a ship requires a passenger certificate, the 
 boiler must be built in accordance with the Board of Trade rules and reeu- 
 lations ; and if to class at Lloyd's, those of Lloyd's Registry must be followed. 
 If both a Board of Trade certificate and a Lloyd's machinery certificate 
 are requisite, the boiler must be so designed as to accord with the rules of 
 both. It is much to be regretted that these two public institutions do not 
 agree to one set of rules which shall be acceptable to both, and in accordance 
 with the best engineering knowledge and experience of the day. Both sets 
 of rules are based on scientific principles and practical experiments on a 
 large scale, and differ only slightly and that in form from one another in 
 consequence. They chiefly differ on the question of the factor of safety, 
 and the differences on this point arise from the allowance included covertly 
 for wear and tear ; the survey in the one case being yearly, and in the other 
 often at longer intervals. 
 
 Ships built for mercantile purposes, but neither classed at a Registry nor 
 requiring a passenger certificate, may have boilers constructed in accordance 
 with the rules of one or the other institution. 
 
 The Testing of Boilers by Hydraulic Pressure was primarily intended to 
 prove their structural strength and local stiffness to withstand the steam 
 pressure at which they were intended to work ; incidentally the workmanship 
 was also proved, but as the working pressure in early days was very low, 
 this latter was quite a secondary consideration. As there were practically 
 no recognised rules such as now obtain to guide both the designer and boiler- 
 maker, such a test was absolutely necessary to avoid catastrophe. The extent 
 of test pressure beyond the working limit was a matter of a few pounds in early 
 days, and so long as the box boiler was in use the margin never exceeded 
 35 lbs. per square inch. The old custom of double the working pressure 
 as the measure of the test was not unreasonable, therefore, as the margin 
 above working pressure by such a rule gave all that was necessary and no 
 more. The cylindrical boiler as made to-day in accordance with well-proved 
 rules of the best and tested materials, and its workmanship quite beyond 
 reproach, would command confidence if not tested by water at all, and any 
 leaks when steam was raised would be quite small and insignificant. Never- 
 theless, it is good for all concerned that there should be a water test of some 
 kind at a pressure above that at which it has to work ; but what that margin 
 should be still excites much controversy and not a little heart-burning. It 
 must be always remembered that the cold-water test conditions are quite 
 
TESTS OF BOILER MATERIAL. 
 
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636 MANUAL OF MARINE ENGINEERING. 
 
 different from those prevailing on service, for then, instead of being at an 
 uniform temperature throughout, and that, too, at or about that at which 
 the structure was rivetted together, the variation between the parts may 
 be considerable, and all much hotter than when constructed, thereby changing 
 the conditions of stress and strain considerably. The furnaces, chambers, 
 and tubes will all be hotter than the shell and its long stays, and therefore 
 should expand more ; on the other hand, the considerable length of the stays 
 in the steam space permits of greater actual extension under load, so that 
 the equilibrium is probably maintained at the high as at the low temperature. 
 High temperature affects the structure of the boiler, but the local stresses 
 and strains only are severe, and it is really on getting up steam that faulty 
 or imperfect workmanship is made manifest; it is for that reason many 
 engineers require that steam shall be raised in one at least of a set of boilers 
 before being placed in the ship after being tested by water. A box boiler, 
 as made years ago for working pressures of 25 to 35 lbs. with flat sides and 
 ends, flattened furnaces, internal uptakes, the stays complicated and riveted 
 to the boiler, and, therefore, with no initial tension, made of untested and 
 somewhat uncertain material by means also uncertain, and designed through- 
 out more or less by rule of thumb, required some drastic tests to give it a 
 character, and ample proof of its reliability was absolutely necessary. More- 
 over, such a boiler, with its deadweighted safety-valves and huge grates, 
 was liable to have an increase of steam pressure on the sudden and unexpected 
 stoppage of the engines far in excess of the working pressure, and might 
 amount to the 25 to 35 lbs., especially when ships had no auxiliary steam- 
 worked feed pump. Altogether, then, the rule for testing boilers to a pressure 
 equal to twice the working pressure was quite a reasonable one, and was 
 never questioned. 
 
 To-day, however, everything is different ; there is no uncertainty in design 
 or manufacture, the construction is carried out under constant and careful 
 inspection, and the whole of the material is tested and its character deter- 
 mined ; moreover there is no possibility of steam accumulating to an extent 
 as to produce a dangerous pressure ; the Admiralty, therefore, were fully 
 justified in changing the rule for water tests years ago on the use of Siemens' 
 steel becoming universal, and the working pressure raised to something 
 over 100 lbs. per square inch. The Bureau Veritas and German Govern- 
 ment have also acquiesced in the reasonableness of these views, and for 
 a very long time all locomotive engineers have been content with quite a 
 moderate margin over the working pressure as the test of these boilers. The 
 Board of Trade, Lloyd's Register, and the British Corporation, however, still 
 require all new boilers to be tested by water in the presence of their Surveyor 
 to a pressure double that at which they license them to work.* 
 
 The Admiralty Rule for all tank boilers is that they be tested to a pressure 
 equal to 90 lbs. in excess of the working pressure, so that if it is 180 lbs. the 
 test will be 270 lbs. 
 
 The Bureau Veritas require a test of double the working pressure, so long 
 as that pressure does not exceed 142 lbs. per square inch ; where the working 
 pressure is over 142 lbs., the margin need not exceed 142 lbs. ; so that if 
 the working pressure is intended to be 180 lbs., the French test is 180 + 142, 
 or 322 lbs., instead of 360. as required by English rules. 
 
 The German Government require the hydraulic test to be twice the working 
 
 •The Rule now established by the British Marine Engineering Design and Construction Committee 
 is that, foi boilers whose working pressure does not exceed loo lbs., the hydraulic test is 2 x W P ; when 
 over luo lbs. the test is 1-5 x W P-f 50 H>s. 
 
CYLINDRICAL BOILER SHELL. 637 
 
 pressure when it does not exceed 5 atmospheres ; when the working pressure 
 intended is above this (that is, 75 lbs.) the margin need not exceed 5 atmo- 
 spheres, so that a German boiler for 180 lbs. W.P. will be tested to (180 -f- 75), 
 or 255 lbs. only. If a boiler requires testing after it has been at work in the 
 ship the pressure need not be more than 50 per cent, over the working pressure, 
 and not exceed by 85 lbs. the W.P. 
 
 It will be seen, then, that for a modern boiler made for a working 
 pressure of 180 lbs. per square inch, the Admiralty would test to 270 lbs. ; 
 the German Government to 255 ; the French officials to 322 ; and the 
 English to 360 lbs. As a consequence of the latter, a boiler can never 
 be made with such light scantlings as the circumstances would warrant, 
 as the stress would be at test so close to the elastic limit of the material as to 
 be dangerous, whereas under German, French, and Admiralty conditions 
 a factor of safety of 4, which is ample, could be indulged in generally. 
 
 The Admiralty rules are not so stringent nor so extended as those of the 
 Board of Trade, as the circumstances of naval construction require a little 
 more elasticity. 
 
 Boiler Shell, Cylindrical. — This is the simplest and strongest form to 
 withstand internal pressure ; because, since a circle is the figure of least 
 perimeter for a given area, there is no tendency to change of form, the metal 
 is strained in one way only — viz., tangentially and in tension.* 
 
 The total pressure tending to rupture a cylinder is the part of all the 
 pressures acting at the various points in direction normal to the surface 
 resolved in one direction. This is equivalent to the pressure on the plane 
 through the axis of the cylinder. 
 
 Rupture is resisted by the two sections of metal at the sides. 
 
 Let D be the diameter of the thin cylinder, and L its length in inches ; 
 f the effective pressure per square inch, and t the thickness of metal in fractions 
 of an inch. 
 
 Then the pressure tending to burst it = p X D X L. 
 
 The stress per square inch on the metal resisting this is 
 
 _ pxDxL __ pxD 
 L x 2* 2f 
 
 Let T be the ultimate strength of the material in pounds per square 
 inch, and F be the factor of safety deemed advisable, and let T -f F =/ 
 Then the safe working pressure for a boiler shell, or other cylindrical 
 
 part subject to internal pressure = — =—-- 
 
 This holds good only when there is no joint or other cause of reduction 
 of effective area of plate section. 
 
 Since a boiler shell is made of one or more plates connected by riveted 
 joints, the effective area of plate is that part remaining between the rivet 
 holes (neglecting the effect of the friction between the plates). 
 
 If the plates are connected by means of a single row of rivets, the average 
 value of the part remaining between the rivet holes is generally 56 per cent, 
 of the whole plate ; so that in this case 
 
 a* i- 2*X/ 56 tXf 1#10 
 
 bale working pressure = — =p^ X jw~ = -yp X * **• 
 
 * For this and the other parts of the boilers, ride new Rules established by the British Marine 
 Engineering Design and Construction Committee as set out in the Appendix. 
 
638 
 
 MANUAL OP MARINE ENGINEERING 
 
 When the plates are counected by two rows of rivets, so that the joint is 
 eaid to be a double-riveted one, there is on the average 70 per cent, of the 
 plate remaining between the holes ; in this case 
 
 Safe working pressure = 
 
 D 
 
 100 
 
 D 
 
 X 140. 
 
 It will be seen that the double-riveted joint is more than 25 per cent, 
 stronger than the single-riveted. This difference is due to the wider spacing 
 of the rivets which can be allowed, in consequence of the increase of rivet 
 area obtained with the two rows. 
 
 The value of F depends on the kind of material, and on its quality. 
 
 The Board of Trade requires that the steel used in the shells of boilers 
 shall have a minimum strength of 27 tons and a maximum of 32 tons* per 
 square inch, with an elongation of not less than 18 per cent, in 10 inches, 
 and allows 28 tons to be used for calculations. If the plates stand a test of 
 over 28 tons, that higher figure may be taken for calculation purposes. Steel 
 up to 35 tons has been used for boiler shell, and is likely to be so generally. 
 The Board, however, requires a minimum factor of safety of 4*5, while the 
 Admiralty was content with 4"0, which is quite sufficient, considering that 
 the stress on the material is gently applied and remains quite a steady 
 continuous one during work. 
 
 Table lxvii. gives the tests of the materials required by the Admiralty 
 Board of Trade, Lloyd's Register, etc. 
 
 Riveting. — The longitudinal seams or joints of a boiler shell are made in 
 one of the following ways : — 
 
 (1) Lap joint and single-riveted. — The plates in this case are lapped one over 
 another, and united by a single row or rivets. This method is only adopted 
 when the plate is thicker than required for mere strength, so that 56 per 
 cent, of it is sufficient to safely withstand the pressure. This generally 
 arises in the case of small boilers and steam receivers, where the diameter 
 is comparatively small, or with low pressures, when the thickness of plate 
 which is sufficient to withstand the pressure would not permit of caulking, 
 nor give the allowance for deterioration. 
 
 TABLE LXVIII. — Strength of Various Riveted Jotnt.-. 
 
 Description of Joint. 
 
 1. Lap joint, ordinary single-riveting, 1 rivet to each pitch, 
 
 2. .. ,, double-riveting, 2 rivets to each pitch, 
 
 3. ,. ,, treble-riveting, 3 rivets to each pitch, 
 
 4. ,, special treble-riveting, 4 rivets to each pitch, 
 
 5. Butt joint, double straps, ordinary double, 2 rivets to each pitch, 
 
 6. 
 
 7. 
 
 8. 
 
 9. 
 10. 
 11. 
 
 special double, 3 
 
 ordinary treble, 3 
 
 special treble, 5 
 
 ordinary quadruple, 4 rivets to each pitch, 
 
 special quadruple, 9 „ „ 
 
 9 9 ,' II ,, «, 
 
 Per cent, of 
 Soliil Plate. 
 
 56 
 
 66 
 
 72 
 
 80 
 
 75 
 81 to 84 
 80 to 82 
 84 to 88 
 
 84 
 
 92 
 93 to 94-5 
 
 * The standard steel plates for boiler shells have a tensile strength of 28 to 32 tone. 
 
X 
 
 RIVETING. 
 
 G39 
 
 Rivet metal* is always softer than boiler plate, and its resistance to shearing 
 
 is less than that of the latter to tension, for this reason the area of the rivets 
 
 should be greater than that of the plate remaining between the rivet holes ; but, 
 
 on the other hand, whereas the plate is subject to reduction by wear, the rivet 
 
 section is not so affected, and in consequence it might suffice to allow the same 
 
 27 
 area of rivet. The Board of Trade requires the rivet area of section to be ^-, or 
 
 1*174 times the area of section of plate between the holes for steel whose 
 tensile is 27 tons. 
 
 Taking 56 per cent, as the proportion of plate between the holes, p the 
 pitch, and d the diameter of the rivets, 
 
 Pitch of rivets 100 
 
 diameter 100 — 56' 
 
 or Pitch of rivets = 2*273 x diameter. 
 
 Since the area of rivet = 1174, the portion of the plate between holes, 
 
 n iz 27 / jw 56 , 27 
 
 4^ 2 = 23(P-<*)*=l00i>X<X23' 
 
 Substituting the value of p as found above, 
 
 Diameter of rivet = 1*9 x thickness of plate. 
 
 Example. — To find the pitch and diameter of rivets for a single-riveted 
 lap joint, with plates J inch thick, strength of joint being 56 per cent. 
 
 Diameter of rivet = 19 X \, or 0*95 inch. 
 
 Pitch of rivets = 2*273 x 0-95, or 2*16 inches. 
 
 The lap of the plates is three times the diameter of the rivet ; if it is 
 made more than this, the plate will spring in caulking ; if made less, there 
 is danger of bulging or cracking the edges, and also there is no margin for 
 recaulking if required. 
 
 The following table gives the pitch, etc., as found in general practice : — 
 
 TABLE LXIX. — Lap Joints, Single-Riveted, Tight Work. 
 
 Thickness 
 
 of 
 
 Plate. 
 
 Diameter 
 
 of 
 
 Rivet. 
 
 Pitch 
 
 of 
 Rivet. 
 
 fireadth 
 of 
 
 Lap. 
 
 Percentage of 
 
 Board of 
 
 Trade 
 
 P. C. Joint. 
 
 Plate. 
 
 Rivet 
 
 ft 
 
 s 
 
 § 
 
 i 
 A 
 
 1 
 
 i 
 i 
 
 1 
 
 i 
 
 1A 
 
 l! 
 
 2i 
 
 2* 
 
 11 
 2J 
 2f 
 21 
 3i 
 
 58 
 58-3 
 57 1 
 
 58-8 
 55 6 
 57o 
 
 66 1 
 65 4 
 
 67 3 
 65 2 
 70 
 63 3 
 
 56-2 
 
 56 
 
 57 3 
 559 
 59 6 
 540 
 
 (2) Lap joint and double-riveted. — Here there are two rows of rivets. 
 Sometimes, as in shipbuilding, the rivets of one row are exactly in line with 
 those of the other, and thus called " chain " riveting ; but more frequently 
 those of one row are in line with the middle of the spaces of the other, and 
 it is then called "zigzag" riveting. The latter plan requires less lap, and 
 
 • Steel parts for rivets and screwed stays have an ultimate tensile of 26 to 30 tons and its resistance 
 to shear is returned as 23 tons. 
 
610 MANUAL OF MARINE ENGINEERING. 
 
 makes tighter work than the former, but does not leave the plate so strong, 
 
 especially when the holes are punched. Here again the rivet area may be 
 
 made the same as the area of the plates between the holes, but the Board of 
 
 27 
 Trade requires the area of rivet section to De ~-„- or 1*174 that of the plate 
 
 between the rivet holes Taking the proportion of the plate between the 
 holes at 70 per cent. 
 
 Pitch of rivets 100 
 
 ~~ dTameteF" ' TOO — 70' 
 
 oi, Pitch of rivets = 3*333 x diameter. 
 
 « Trrf 2 70 27 no 
 
 Also. 2 X -j- - jqq p X t X ^ = 0-8 p X t. 
 
 whence, substituting the above value of p, 
 
 Diameter of rivet — 1*7 X thickness of plate. 
 
 Example. — To find the pitch and diameter of the rivets for a lap joint, 
 double-riveted : the plates being f-inch thick, and the strength of joint 
 70 per cent. 
 
 Diameter of rivet = 1*7 x f, or 1*275 inches. 
 
 Pitch of rivet = 3*333 x 1*275, or 4*35 inches. 
 
 Also, if treble riveted, and the strength of joint 72 per cent. 
 Pitch of rivets = 3*57 X diameter. 
 
 a*d 2 72 27 ' JB 
 
 T~ = 100 V X l X 23 = °' 84 ° P X *' 
 whence, substituting the above value of p, 
 
 Diameter of rivet = 1*28 X thickness of plate. 
 
 Example. — To find the pitch and diameter of the rivets for a treble- 
 riveted lap joint, the plates being 1 inch thick. 
 
 Diameter of rivets = 1 X 1*28, or 1*28 inches. 
 Pitch of rivets = 3*57 X 1*28. or 4i inches. 
 
 A Special Arrangement of Treble Riveting is sometimes resorted to whereby 
 the strength of joint is 80 per cent, of the solid plate, and has double the 
 number of rivets in the middle row than in either of the two outer. In this 
 case it is obvious that the pitch in the outer rows must be such that inner 
 row rivets must be far enough apart to give at least 56 per cent, of plate 
 between them. Table lxxii. gives the diameter, pitch, etc., for such a type 
 of joint, and it will be seen that the ratio of d to t is high — viz., 1-50. 
 
 General Rules for Riveted Joints are based on the allowances made by 
 the Board of Trade for mild steel — viz., that the shell plates may vary in 
 tensile strength from 27 to 36 tons, and that the ultimate shearing resistance 
 of rivet steel is 23 tons per square inch. The minimum tensile for the plates 
 of 27 tons is assumed in the rules and calculations, and by referring to Table 
 lxix. the adjustment can be made when steel of higher minimum tensile 
 strength is intended to be used. 
 
 For lap joints gene/ally, where p is the pitch of outer rows, and d the 
 diameter of rivet — 
 
RIVETING. 641 
 
 Area of rivet X number of rivets per pitch X 23 = (p — 3) X I X 27. 
 
 (a) Then thickness of plate = ■— -^-r X 0-669. 
 
 p — a 
 
 For a single -riveted joint, p — d is usually 56 per cent, of p. 
 
 ., double- ,., p — d ,, 66 „ 
 
 „ treble- „ p — d „ 72 „ 
 
 „ special „ p — d „ 80 „ 
 
 Substituting these values in equation (a), the following rules hold good : — 
 
 Thickness of plates when lap-jointed =fxd 2 s-p. 
 
 For ordinary single-riveting =/ — 1*195. 
 
 „ double -riveting = /= 2-027. 
 
 „ treble-riveting (3 rivets per pitch) =f= 2-788. 
 
 For special treble-riveting (4 rivets per pitch) =f— 3*345. 
 
 Also it follows that : — 
 
 For ordinary single-riveting d = 0-44 p ; and p = 2-273 d. 
 
 „ double- ,, d = 0-34 p ; and p = 2-94 d. 
 
 ,, treble- ,, d — 0-28 p ; and p = 3-57 d. 
 
 For special treble-riveting d = 0-20 p ; and p = 5-00 d, 
 
 Example. — A plate | inch thick if connected to another by a lap joint 
 will have : — 
 
 For single-riveting, 1^ inches diameter rivets, 2-98 inches pitch. 
 „ double- „ 1| ,, „ 3-30 „ 
 
 „ treble- „ 1 „ „ 3-57 „ 
 
 ,, special „ 1| „ „ 5-625 „ 
 
 For Butt Joints and Double Straps the Board of Trade permits only 
 1-875 times the single shear instead of the double. Here area of rivet section 
 X number per pitch of outer rows X 23 X 1-875 = (p — d) X t X 27. 
 
 (b) That is, the thickness of plate = _ X 1-2543. 
 
 Single -riveting is seldom or never adopted with butt joints. 
 
 For ordinary double -riveting, (p — d) is usually = 0-75 X p. 
 
 treble- „ (p - d) „ = 0-80 X p. 
 
 For special 3-rivet double-riveting (p — d) „ — 0-81 X p. 
 
 „ 5-rivet treble-riveting {p — d) „ = 0-84 X p. 
 
 For ordinary 4-rivet quadruple-riveting (p — d) „ — 0-84 X p. 
 
 For special 9-rivet „ (p — d) „ = 0-92 X p. 
 
 11-rivet „ (p - d) „ = 0-94 X p. 
 
 Substituting these values in equation (b), then : — 
 
 Thickness of plates when double-strapped butt-jointed = d 2 X K -r- p. 
 
 For ordinary double-riveted joint, K = 3-333. 
 
 special ,, ,, K = 4-652. 
 
 ordinary treble-riveted joint, K = 4-704. 
 
 special treble-riveted joint, K — 7-466. 
 
 ordinary quadruple-riveted joint, K = 5-973 
 
 special 9-rivet' „ .. K =12-26. 
 
 „ 11-rivet . „ K =14-68. 
 
 41 
 
642 MANUAL OF MARINE ENGINEERING. 
 
 In case of the ordinary double-riveted, d = 0-25 p, or p = 1-00 d. 
 
 special double-riveted, d = 0-19 p, or p = 5-26 d. 
 
 ordinary treble-riveted, d — 0-20 p, or p = 5-00 d. 
 
 special treble-riveted, d = 0-16 p, or 7) = 6-25 d. 
 
 ordinary quadruple -riveted, d = 0-16 p, or p = 6-25 cZ. 
 
 special 9-rivet quadruple-riveted, d = 0-08 p, or p = 12-50 d. 
 
 special 11-rivet quadruple-riveted, d = 0-06 p. or p = 16-67 d. 
 
 This Special Form of Treble-riveted Butt Joint now in common use for the 
 longitudinal joints of boiler shell plates, there are three rows of rivets on each 
 side of the joint, and the number of rivets in the outer row, or that next the 
 edge of the butt straps of each, has half the number of rivets that is in each 
 inner row. That means that, instead of six rivets acting to each (outer) 
 pitch, there are only five ; or, to make the comparison, there are only 2-i for 
 each inner or general pitch ; therefore, if p is taken as the pitch in outer 
 row, n is 5, and by substituting the value of p from p — d = 0-84 p, 
 
 p = 6-25 d. 
 
 d 2 V ^ d 2 
 
 (c) That is, thickness of plate = ^ X 1-1707 ; or — X 6-97. 
 
 ' r p— a p 
 
 Substituting the value of p = 6-25 d, then 
 
 Diameter of rivet = 0-837 X thickness of plate. 
 
 Hence for an inch plate with this style of joint the rivets would be f , 
 and the pitch 5| inches. 
 
 The advantage of this kind of joint is perceptible for thick plates, as 
 may be seen by calculating the size and pitch of rivets for a plate, say, 
 \\ inches thick. 
 
 Then the diameter is 1*5 and the pitch 9'4 inches, whereas if with the 
 ordinary treble riveting the diameter would be as much as If inches, and 
 the pitch for all the rows 8| inches, and the strength of the joint 5 per cent, 
 weaker, and not as capable of remaining tight under steam as the other 
 joint with its inner rows of rivets pitched no more than 4-£ inches. 
 
 The thickness of the butt straps should vary with the pitch of outer 
 rows to resist forcing when being caulked. In practice, the thickness of 
 inner strap should be not less than one-eighth the pitch of the outer rows, 
 so that for the r5-inch plate the inner strap should be at least 1-fg-. 
 
 The Thickness of Butt Straps required by the Board of Trade is five-eighths 
 the thickness of plate for each of two straps ; or nine-eighths of one only 
 for ordinary styles of riveting ; with the special one, as above described — 
 
 Thickness of butt straps = — t X — — ^— r- 
 
 8 p — 2d 
 
 For the 1'5-inch plate with the pitch as above 
 
 The thickness required by B. of T. = | x 1*5 X ~^~-^- — M58 inches. 
 
RIVETING. 643 
 
 Breadth of Butt Straps for double riveting is 2"5 x pitch ; for treble riveting, 
 3 X pitch; and for treble riveting with the alternate rivets in outer rows 
 omitted, 2 45 x the pitch. 
 
 "Breadth of lap joints single-riveted is 3 X diameter of rivet. 
 
 ,, „ double- „ 3 x diameter -f 0*55 X pitch. 
 
 „ „ treble- ,, 3 X diameter -j-1-10 X pitch. 
 
 Treble-riveted Lap Joints with the number of rivets in the middle row 
 double that in the two outer is a good form of joint, and has similar advantages 
 to those possessed by the double butt straps treated in a similar way. The 
 strength of that joint may be calculated on the basis that there are four 
 rivets in single shear for each unit of pitch. 
 
 d 2 x 4 
 That is, the thickness of plate = — ^ x 0-669. 
 
 p — d 
 
 The highest strength of joint with reasonable size of rivet is 80 per cent. — 
 
 that is, p — d = 0-80 X p, or p = 5*0 x d. 
 
 d 2 d 
 
 Then thickness of plate = — X 3*345 = = X 3*345. 
 
 p 5 
 
 Then diameter of rivet = 1*49 x t. 
 
 Example. — If an inch plate is jointed with this type of riveting, the 
 diameter will be 1| inches, and the pitch 7*5 in the outer rows and 3| inches 
 in the middle one. The percentage of joint through that row will be 
 
 3 ' 75 8 7 5 1 ' 5 X 100, or 60. 
 
 This is rather too close for good practice ; the middle row, therefore, 
 should always have a pitch not less than 2*8 X diameter. 
 
 With Double Butt Straps and Double Riveting, with the outer rows having 
 half the number of rivets as in the inner, the following holds good. Here 
 there are three rivets in double shear for each unit of pitch, and assuming 
 that the pitch of the inner row is 2*6, and the outer 5*2 times the diameter, 
 
 5. 2 1 
 
 Then strength of joint = — =-^ — , or 0*808. 
 
 _,, . ., ,., . ,. 3x 1*875 x 23 x 0-785 rf 2 d . 700 
 The strength of the riveting = - -0 , ~ , or - 0*722. 
 
 If it is 81 per cent., then 
 
 d - X 0*722 = 0*81, or d = 1115 x t 
 
 For example, if an inch plate is connected to another with double butt 
 straps and the outer rows with half the number of rivets in the inner, the 
 rivets will be diameter = ^1*115 X 1, or 1^; the pitch of the middle row 
 will be 2|, and the outer rows 5J. By the ordinary methods the rivets 
 would be \\ diameter, and the pitch 4^ inches with 75 per cent, of plate 
 and 72*8 of rivets (B. of T. rules), as against the 80*8 per cent, of each member 
 by the above system. It should be noted, however, that the butt straps in 
 
644 
 
 MANUAL OF MARINE ENGINEERING. 
 
 this case must, be of such a thickness that together they give at least 81 per 
 cent, of solid plate with only 60 per cent, of the pitch between the rivet holes. 
 
 81 
 The sum of the thickness of butt straps = ~ x t, or 1 -35 X t. 
 
 Each strap, therefore, ought to be - 7 X t ; and if the thickness be taken 
 at pitch -T- 8 for the inner strap, then for the above example its thickuess 
 would be 51 ~ 8, or - 688 inch. The inner strap for such joints should be 
 0'75 X t, and the outer 0-675 X t 
 
 The Board of Trade Rule for the maximum pitch of rivets is as follows, 
 but the limit for any thickness of plate is 10-J inches : — 
 
 Maximum pitch = cXl+ 1*625 inches. 
 
 c = 2 "62 for lap joints with two rivets in the pitch space. 
 
 V — -±1 „ 
 
 c = 4-14 „ 
 
 four 
 
 55 55 
 55 55 
 
 c = 3-50 for butt 
 
 two 
 
 55 55 
 
 c = 4*63 ,, , 
 
 three 
 
 55 55 
 
 c = 5*52 ,, , 
 
 four 
 
 '5 55 
 
 c = 6-00 
 
 five 
 
 55 55 
 
 TABLE LXX. — Lap Joint, Double Riveting Zigzag. Strength of Joint 
 66 per cent, of Solid Plate. Breadth of Lap = 4'62 d. 
 
 
 Rivets. 
 
 
 
 Rivets. 
 
 
 Thickness 
 of Plate. 
 
 
 
 Breadth 
 of Lap. 
 
 Thickness 
 of Plate. 
 
 
 
 Breadth 
 of Lap. 
 
 
 
 
 
 
 Diameter. 
 
 Pitch. 
 
 
 
 Diameter. 
 
 Pitch. 
 
 
 1 
 
 3 
 4 
 
 2i 
 
 3| 
 
 7 
 
 8" 
 
 1-Ar 
 
 x IS 
 
 •>8 
 
 6 
 
 » 
 
 TU 
 
 1 a 
 iff 
 
 21 
 
 91 3 
 
 °T8 
 
 1 5 
 
 if 
 
 4 T ' C 
 
 6i 
 
 f 
 
 15 
 T5 
 
 211 
 
 4J 
 
 1 
 
 iA 
 
 4 
 
 6f 
 
 1 1 
 
 16 
 
 1 
 
 2 T § 
 
 4f 
 
 1A 
 
 1A 
 
 4.5. 
 
 74 
 
 i 
 
 It 1 * 
 
 31 
 
 4y§ 
 
 n 
 
 ll 
 
 4.1 3 
 
 7.V 
 
 13 
 
 1& 
 
 3i 
 
 H 
 
 1 a 
 
 "•TS 
 
 if 
 
 •A 
 
 8iV 
 
 TABLE LXXI. 
 
 -Lap Joints, Treble Riveting. Strength of Joint, 
 72 per cent. of solid plate. 
 
 Breadth of Lap = 3 d + 1-35 Jd (2 p + d) = 6-852 d, 
 
 Thicknes 
 
 of Plate. 
 
 Rivets. 
 
 Breadth 
 of Lap. 
 
 Thickness 
 of Plate. 
 
 Rivets' 
 
 Breadth 
 of Lap. 
 
 Diameter. 
 
 Pitch. 
 
 Diameter. 
 
 Pitch. 
 
 1 
 
 M 
 
 7 
 
 B 
 1 5 
 It 
 1 
 
 1 h 
 
 1 
 
 l'A 
 ll 
 if 
 lft 
 
 if 
 
 3ft 
 
 4 
 4J 
 4ft 
 4*1 
 
 6| 
 
 7f 
 
 8 
 
 8* 
 9" 
 9j% 
 
 11 
 
 lf% 
 
 11 
 
 1& 
 It 
 
 1A 
 
 If 
 
 lxl 
 
 If 
 11 
 
 5& 
 
 t>fs 
 
 6rf 
 
 Cft 
 
 6J 
 
 0i 
 
 10 
 
 ion 
 ill 
 
 12-0 
 123 
 
RIVETING. 
 
 645 
 
 TABLE LXXII. — Lap Joint, Treble Riveting with Double the Number 
 of Rivets in Middle Row that of Outer Ones. Strength of 
 Joint, 80 per cent, of Solid Plate. Breadth of Lap = 7'id. 
 
 
 Rivets. 
 
 
 
 Rivets. 
 
 
 ■' 
 
 Thickness 
 
 
 
 Breadth 
 
 Thickness 
 
 
 
 Breadth 
 
 of Plate. 
 
 
 
 of Lap. 
 
 of Plate. 
 
 
 
 of Lap. 
 
 
 Diameter. 
 
 Pitch. 
 
 
 
 Diameter. 
 
 
 Pitch. 
 
 
 3 
 
 4 
 
 1| 
 
 5| 
 
 8| 
 
 1^ 
 
 119. 
 
 
 8 
 
 llf 
 
 in 
 IB 
 
 1/s 
 
 H 
 
 9 
 
 H 
 
 lli 
 AlB 
 
 
 8£ 
 
 12\ 
 
 8 
 
 1A 
 
 6& 
 
 y 4 
 
 1A 
 
 MS 
 
 
 9 
 
 13 
 
 Hi 
 16 
 
 1 1 3 
 
 7 
 
 ina 
 
 iu 8 
 
 li 
 
 11 
 
 
 03 
 
 J 8 
 
 13| 
 
 1 
 
 n 
 
 7* 
 
 n 
 
 1t% 
 
 1H 
 
 
 Oil 
 
 1-ti 
 
 N.B. — The pitches are in all cases in this joint in excess of the Board of Trade Rule ; 
 for example, by the rule an inch plate may have 5| pitch only, which means the middle 
 row is only 2J inches. 
 
 The chief difficulty with lap joints is in the working of the corners of 
 the plates where the next strake of plating covers the lap ; these corners 
 are hammered or machined to a taper, so as to nearly conform to a circle, 
 and the covering plate is slightly joggled, so as to lie evenly on the deformation 
 caused by the lapping. Most boilermakers now. to ensure a good fit, go 
 to the expense of planing or milling the corners fair, and even to extend the 
 taper part beyond the butt end of the plate itself, so as to cause as little 
 deformation as possible. 
 
 Butt Joints with Double Straps and Single-riveted. — This form of joint 
 is not often resorted to, as there are two rows of rivets, and only a shearing 
 area of twice that of the one row of rivets, besides all the expense of double 
 straps, which entail the caulking of four seams ; the sole advantage it pos- 
 sesses over the double-riveted lap joint is the absence of smithed corners, 
 and that the plates lie wholly in the circle without deformation ; this, how- 
 ever, does not compensate for the extra expense and the liability of leakage 
 from the two extra seams. 
 
 The strength of this, joint is seldom more than 65 per cent, of the solid 
 plate, as more cannot be obtained without placing the rivets so far apart as 
 to prevent the strap from being caulked tight. If the straps are made of 
 the same thickness as the plate itself, 70 per cent, may be obtained. Taking 
 70 per cent, as the strength of joint, the diameter and pitch of the rivets are 
 the same as given for the double-riveted joint, and the breadth of the strap is 
 six times the diameter of the rivets. For example, if the plate is f-inch 
 thick, the rivets should be 1^ inches diameter, and 4| inches pitch, the 
 breadth of strap being 6 X 1^-, or 7| inches. 
 
 Butt Joints with Double Straps and Double-riveted. — This is a very general 
 and deservedly favourite form of joint for thick plates, and when well made 
 gives every satisfaction. There is no necessity for smithing or machining 
 the plates, nor of joggling the covering plate of the next strake, although 
 some boilermakers, to avoid the caulking of the ends of the strap where it 
 butts against the next strake, thin down the end and notch out the covering 
 plate, so as to lap over the strap. This makes a very good joint, but is 
 somewhat expensive, and if the plates are properly fitted, there should be 
 no need of such an elaboration. 
 
646 MANUAL OP MARINE ENGINEERING. 
 
 For each portion of plate between the holes, there is a rivet area equal 
 to four times the area of section of one rivet ; but the Board of Trade used 
 to allow only 3J times the area of one rivet, but now 3| times. 
 
 Under these circumstances the following rules hold good for a joint equal 
 to 75 per cent, of solid plate : — 
 
 Pi tch of rivets 100 
 
 diameter " 100 — 75' 
 ^ or, Pitch of rivets = 4 X diameter. 
 
 * . Q „ xd 2 75 27 
 
 Also, since 3| X -r- = Vqq PXIX^ 
 
 Diameter of rivets = 1*196 x thickness of plate. 
 
 The thickness of each butt-strap must be at least § that of the plate, and 
 when the strap between contiguous strakes is simply butted against them, 
 it is better to be of the same thickness as the plate. 
 
 Example — To find the diameter, pitch of rivets, and thickness of straps 
 for a butt-joint double-riveted, and equal to 75 per cent, of solid plate, whose 
 thickness is 1 inch, 
 
 Diameter of rivets = 1 X 1*196, or 1*196 inches. 
 Pitch of rivets = 4 X 1*196, or 4^f inches. 
 
 Thickness of long strap § inch, and of short strap 1 inch if butted, and 
 | inch if fitted under the covering plates. 
 
 Butt Joints with Double Straps Treble-riveted. — This form has become a 
 necessity since very thick plates have been used for boiler-shells, in order 
 to get adequate sectional area of rivet ; it also admits for the same reason 
 of thinner plates being used when desired, so that when there are only a 
 few butt joints it is really more economical to adopt this form of joint. In 
 this case, for each portion of plate between the holes, there is a rivet area 
 equal to six times the area of section of one rivet ; but, as before, the Board 
 of Trade only allow this as 5f. The percentage of plate between the 
 rivets is generally 80 per cent, with this type of joint. 
 
 Here the pitch of the rivets is 5 x diameter, and, taking the Board of 
 Trade allowances of 23 tons for shear and 27 for tension, then — 
 
 5f X^x23 = ^x*x27. 
 
 and diameter of rivet = 1 *06 X thickness of plate. 
 
 Butt Joints, treble-riveted, with half the number of rivets in the outer 
 than in the inner two rows, are now generally adopted as the best for the longi- 
 tudinal joints of modern boilers, inasmuch as with rivets of quite compara- 
 tively small diameter a stronger joint is obtained ; in practice it is generally 
 84 per cent, of the solid plate, and may often be even higher. Taking this 
 percentage and the ratio of tension to shear as before, the following holds 
 good. There are here five rivets for each space in the outer rows, and with 
 the allowance of If for double shear, the acting rivet area =5 x If, or 
 9g that of one rivet. Then pitch of rivets is 6*25 diameters, and 
 
 9f X ~ x23 = ^ V Xtx27. 
 4 100 r 
 
 Therefore, diameter oi rivet = 0*806 x thickness of plate. 
 
RIVETING. 
 
 647 
 
 In actual practise the rivets are somewhat larger than given by this rule 
 and often equal to the thickness. 
 
 TABLE LXXIII. — Butt Joints, Double Straps, Double Riveting 
 Zigzag. Strength of Joint, 75 per cent, of Solid Plate. 
 Breadth of Strap = 6 d + 1-3 Jd (2 p + d) = 9-9 d. 
 
 Thickness 
 
 Rivets. 
 
 Inner Strap. 
 
 Thickness 
 
 of 
 
 Plate. 
 
 Rivets. 
 
 Inner Strap. 
 
 of 
 
 Plate. 
 
 Dia- 
 meter. 
 
 Pitch. 
 
 Least 
 
 Thickness. 
 
 Breadth. 
 
 Dia- 
 meter. 
 
 Pitch. 
 
 Least 
 Thickness. 
 
 Breadth. 
 
 8 
 
 a 
 
 4 
 
 rg 
 
 1 1 
 
 TB" 
 7 
 8" 
 29 
 "if 5 
 
 I 
 
 3 
 H 
 
 H 
 
 4 
 
 12 
 
 "37 
 1 4 
 75 
 1 6 
 37 
 16 
 37 
 
 n 
 
 8| 
 
 9 
 10 
 
 7 
 8" 
 
 k 
 
 i 
 
 ift 
 
 1ft 
 
 ift 
 i 9 
 
 *77 
 
 4£ 
 
 H 
 
 17 
 37 
 
 II 
 
 19 
 77 
 2 1 
 
 37 
 
 10* 
 
 Hi 
 
 12| 
 
 TABLE LXXIV. — Butt Joint with Double Strap, Double Riveted 
 Zigzag, with Alternate Rivets in Outer Rows omitted. 
 Strength of Joint, 81 per cent, of Solid Plate. Breadth of 
 Strap = 106 d. 
 
 
 Rivets. 
 
 Inner Strap. 
 
 
 Rivets. 
 
 Inner Strap. 
 
 Thickness 
 of 
 
 
 
 
 Thickness 
 of 
 
 
 
 
 
 
 
 
 
 
 
 
 Plate. 
 
 Dia- 
 
 Outer 
 
 Least 
 
 Breadth. 
 
 Plate. 
 
 Dia- 
 
 Outer 
 
 Least 
 
 Breadth. 
 
 
 meter. 
 
 Pitch. 
 
 Thickness. 
 
 
 meter. 
 
 Pitch. 
 
 Thickness. 
 
 » 
 
 23 
 37 
 
 4 
 
 i« 
 
 37 
 
 71 
 
 14 
 
 16 
 
 1 2 
 
 x 77 
 
 5£ 
 
 2 2 
 
 "3 7 
 
 Hi 
 
 n 
 
 16 
 
 25 
 77 
 
 4ft 
 
 1 7 
 
 77 
 
 »i 
 
 1 
 
 1 4 
 
 5| 
 
 24 
 37 
 
 11-B 
 
 ¥ 
 
 2 7 
 75 
 
 4*f 
 
 is 
 
 U 2" 
 
 "16 
 
 1ft 
 
 1 6 
 
 hi 
 
 6ft 
 
 2 5 
 
 77 
 
 12ft 
 
 13 
 
 29 
 75 
 
 4. ii 
 
 1 9 
 32 
 
 Q5 
 
 a 8 
 
 H 
 
 1 8 
 
 1 77 
 
 6£ 
 
 26 
 157 
 
 13ft 
 
 8 
 
 32 
 77 
 
 t>16 
 
 21 
 77 
 
 10i 
 
 ift 
 
 1 1 1 
 
 x 77 
 
 7 
 
 28 
 37 
 
 14ft 
 
 A T .5. — By Board of Trade Rules for the above pitches the butt straps should be 
 33 per cent thickes. That is, what is ^ should be f f foe the Board of Trade require- 
 ments. 
 
 TABLE LXXV. — Butt Joints, Double Straps, Treble Riveting. 
 Strength of Joint, 80 per cent, of Solid Plate. Breadth 01 
 
 Strap = 6 d + 2-7 Jd (2 p + d) = 14*9 d. 
 
 Thickness 
 
 of 
 
 Plate. 
 
 Rivets. 
 
 Inner Strap. 
 
 Thickness 
 
 of 
 
 Plate. 
 
 Rivets. 
 
 Inner Strap. 
 
 Dia- 
 meter. 
 
 Pitch. 
 
 Least 
 Thickness. 
 
 Breadth. 
 
 Dia- 
 meter. 
 
 Pitch. 
 
 Least 
 Thickness. 
 
 Breadth. 
 
 3 
 
 4 
 
 13 
 
 16" 
 
 8 
 
 .15 
 
 16 
 
 1 
 1ft 
 
 13 
 
 16 
 
 7 
 
 8" 
 
 15 
 
 16 
 
 1 
 
 1ft 
 
 li 
 
 4ft 
 
 m 
 5 
 
 5ft 
 
 H 
 
 16 
 
 77 
 
 18 
 77 
 19 
 "3"5 
 20 
 77 
 21 
 77 
 23 
 77 
 
 121 
 
 13| 
 14 
 
 14| 
 
 15H 
 
 16| 
 
 H 
 
 ift 
 
 ii 
 
 ift 
 
 if 
 
 ift 
 
 1 7 
 
 l 77 
 
 ill 
 
 lT7 
 
 I 1 * 
 
 115 
 
 x 37 
 ] 1 7 
 
 6i 
 6f 
 
 7ft 
 7ft 
 
 25 
 
 77 
 
 25 
 37 
 27 
 TS 
 29 
 "57 
 30 
 157 
 31 
 77 
 
 17* 
 
 18| 
 
 19| 
 
 m 
 
 22 
 
 23 
 
648 
 
 MANUAL OF MARINE ENGINEERING. 
 
 TABLE LXXVI. — Butt Joints, Double Straps, Treble Riveting 
 with Alternate Rivets in Outer Rows omitted. Strength 
 of Joint, 84 per cent, op Solid Plate. Breadth of 
 Strap = 6 d + 2'5 Jd (2 p + d) = 15-2 d. 
 
 Thickness 
 
 Of 
 
 ; /ate. 
 
 Rivets. 
 
 Inner Strap. 
 
 Thickness 
 
 of 
 
 Plate. 
 
 Rivets. 
 
 Inner Strap. 
 
 Dia- 
 meter. 
 
 Pitch. 
 
 Least 
 Thickness. 
 
 Breadth. 
 
 Dia- 
 meter. 
 
 Pitch. 
 
 Least 
 Thickness. 
 
 Breadth. 
 
 1 
 
 H 
 
 i& 
 
 H 
 
 1A 
 
 if 
 
 1A 
 
 7 
 » 
 3 l 
 5" 5 
 
 H 
 iA 
 
 5i 
 
 7 
 8J 
 
 TTJ 
 26 
 TS3 
 26 
 ^2~ 
 2T 
 TT5 
 2 S 
 
 30 
 
 32 
 ST 
 33 
 3"? 
 
 13& 
 14| 
 
 15| 
 16| 
 17* 
 
 19 
 20 
 
 H 
 l& 
 
 if 
 
 m 
 
 113 
 A 16 
 
 Al6 
 
 in 
 
 1 6 
 1A 
 
 i ,? 
 if 
 
 121 
 Its 
 
 8re 
 
 ol3 
 
 8^ 
 
 9f 
 
 9A 
 
 9M 
 
 *10| 
 
 3 4 
 3 2" 
 3 S 
 TT3 
 36 
 7T5 
 38 
 If 2 
 
 3 9 
 "52" 
 
 4 
 
 7T2" 
 
 41 
 7T5 
 42 
 3~2" 
 
 20| 
 
 21 
 
 St 
 
 90 5 
 
 23? 
 24| 
 25J 
 
 Treble-riveted (zigzag) Butt Joint, having half the rivets omitted in the 
 outer rows, and the outer butt strap only covering all three rows ; the inner 
 strap covering the inner rows and permitting of heavy caulking. The pitch 
 of the outer row is p, the diameter of rivets d, and the thickness of plate t. 
 
 In this case, for each pitch of outer rows there are four rivets exposed to 
 double shear and one to single shear, hence 
 
 Effective rivet area = U 2 + 4 X ^d 2 ) 0-7854 = 6-676 d 2 . 
 
 Taking 23 tons as the resistance to shear, and 27 tons to tensile — 
 
 6-676 d 2 x 23 = (p — d)t x 27. 
 
 The strength of this joint should be 85 per cent, of the solid plate, so that 
 p = 6-67(7. 
 
 m , , 27 5-67 
 Then d = 23 X 6^696 
 
 X t = t. 
 
 The pitch of inner rows equal 3-333 d, and the strength of joint at them 
 
 2-333 ,n 
 = 3^33, or 70 per cent. 
 
 With such an arrangement of straps and riveting the limit of pitch of 
 10| inches will apply only to the inner rows. 
 
 Quadruple - riveted (zigzag) Butt Joint, with straps covering all the rows 
 of rivets : p is the pitch of the outer row, that of the two inner rows is one- 
 third p, while in the second row every third rivet is omitted compared with 
 the inner row — that is, for each pitch of the outer row there are two rivets 
 in the next and three in each of the others — hence for each pitch there are 
 nine rivets in double shear. The thickness of the plate is t. Then, 
 
 Effective rivet area = 9 X ^ X 07854 d 2 = 13-257 d 2 . 
 
 o 
 
 * N.B. — The Board of Trade Rules do not allow of a greater pitch than 10 J inches 
 for any thickness of plate. 
 
RIVETING. 
 
 649 
 
 Taking 23 tons as resistance to shear and 27 tons tensile, 
 
 13-257 d 2 X 23 = (p - d) t x 27. 
 
 It is generally convenient to make trie rivets of the same diameter as the 
 thickness of plate ; then as d = t — 
 
 13-257 d X 23 = (p - d) 27. Or, p == 12-29 d. 
 
 The strength of the joint is then 11-29 ~ 12-29, or 91-8 per cent. ; that of 
 
 the inner part is 
 O 
 
 41 
 
 41 
 
 or 75-6 per cent. 
 
 • o 
 
 Fig. 237a. — Special 9-Rivet Quadruple Joint. 
 
 Another Quadruple-riveted Joint has half the number of rivets in the 
 second row than in the two inner, and a quarter the number in the outer 
 row. That is, for each pitch of outer rows there are two rivets in the second 
 and four in each of the two inner rows — that is, there are 11 rivets in double 
 shear per pitch. 
 
 The effective rivet area is then = 11 X ~ x 0-7854 d 2 , or 16-2 d 2 . 
 
 Then, 16-2 d 2 X 23 = (p — d) t X 27. 
 
 As before, assuming d = I, then 16-2 d X 23 = (p — d) 27. Or, p = 14-8 d 
 
 O 
 
 o o o o o o 
 
 oooooooooo 
 ooooooooooo 
 
 ooooooooooo 
 
 oooooooooo 
 o o o o o o 
 
 o 
 
 Fig. 237?;. — Special 11-Ri.vet Quadruple Joint 
 
 14-8-1 
 
 o 
 
 o 
 
 o 
 
 o 
 
 o 
 
 The strength of this joint is then 
 
 14-8 
 
 , or 93-2 per cent. That through 
 
 3-93 \ 
 
 the inner row is -— ^t.- — } or 76-4 per cent., and through the second row 
 
 7-4 
 
 393 
 
 — :— , or 86-5 per cent. 
 
650 
 
 MANUAL OF MARINE ENGINEERING. 
 
 It is necessary with these quadruple-riveted joints that the butt straps 
 be cut away, following the contour of the riveting into what are called 
 Vandykes, in order that the caulking may be effective, and also to save weight. 
 
 The following table gives the thickness of shell plates for minimum tensile 
 test strengths from 28 to 38 tons to equal those of 27 tons, for which the 
 tables of pitch and diameter of rivets, etc., are calculated. For example, 
 if a boiler is to be of steel whose miniumm tensile test shall be 36 tons, the 
 riveting suitable is got by finding the thickness of 27-ton steel for such a boiler 
 and working pressures ; if it works out to f inch it will be seen that if of 
 27-ton stuff the thickness would be 1 inch ; the riveting given in the table 
 for it will be that to employ. 
 
 TABLE LXXVII. — Relative Thickness of Boiler Shell Plates 
 for Different Tensile Strengths. 
 
 
 
 Minimum Tensile Test 
 
 Strength in Tons pef 
 
 Square Inch. 
 
 • 
 
 
 27 
 
 28 
 
 29 
 
 30 
 
 31 
 
 32 
 
 33 
 
 34 
 
 35 
 
 36 
 
 37 
 
 38 
 
 0-50 
 
 0-483 
 
 0-466 
 
 0-450 
 
 0-486 
 
 0-422 
 
 0-409 
 
 0-397 
 
 0-386 
 
 0-375 
 
 0-365 t 0-355 
 
 0-60 
 
 0-579 
 
 0-559 
 
 0-540 
 
 0-524 
 
 0-507 
 
 0-491 
 
 0-476 
 
 0-463 
 
 0-450 
 
 0-438 
 
 0-427 
 
 0-70 
 
 0-675 
 
 0-652 
 
 0-630 
 
 0-611 
 
 0-591 
 
 0-573 
 
 0-551 
 
 0-540 
 
 0-525 
 
 0-511 
 
 0-498 
 
 0-80 
 
 0-772 
 
 0-745 
 
 0-720 
 
 0-699 
 
 0-675 
 
 0-655 
 
 0-635 
 
 0-617 
 
 0-600 
 
 0-584 
 
 0-569 
 
 0-90 
 
 0-868 
 
 0-838 
 
 0-810 
 
 0-784 
 
 0-760 
 
 0-736 
 
 0-714 
 
 0-694 
 
 0-675 
 
 0-657 
 
 0-640 
 
 1-00 
 
 0-965 
 
 0-931 
 
 0-900 
 
 0-873 
 
 J-844 
 
 0-818 
 
 0-793 
 
 0-771 
 
 0-750 
 
 0-730 
 
 0-711 • 
 
 1-10 
 
 1001 
 
 1-.024 
 
 0-990 
 
 0-961 
 
 0-929 
 
 0-900 
 
 0-872 
 
 0-848 
 
 0-825 
 
 0-803 
 
 0-782 
 
 1-20 
 
 1-157 
 
 1117 
 
 1-080 
 
 1-048 
 
 1-013 
 
 0-982 
 
 0-925 
 
 0-925 
 
 0-900 
 
 0-876 
 
 0-853 
 
 1-30 
 
 1-254 
 
 1-211 
 
 1-170 
 
 1135 
 
 1-097 
 
 1-064 
 
 1031 
 
 1-003 
 
 0-975 
 
 0-949 
 
 0-924 
 
 1-40 
 
 1-350 
 
 1-304 
 
 1-200 
 
 1-223 
 
 1-182 
 
 1-145 
 
 1-110 
 
 1-080 
 
 1-050 
 
 1-022 
 
 0-996 
 
 1-50 
 
 1-447 
 
 1-397 
 
 1-350 
 
 1-310 
 
 1-266 
 
 1-227 
 
 1-190 
 
 1-157 
 
 1-125 
 
 1-095 
 
 1-067 
 
 1-60 
 
 1-543 
 
 1-490 
 
 1-440 
 
 1-397 
 
 1-351 
 
 1-309 
 
 1-269 
 
 1-234 
 
 1-200 
 
 1-168 
 
 1-138 
 
 1-70 
 
 1-640 
 
 1-583 
 
 1-530 
 
 1-484 
 
 1-435 
 
 1-391 
 
 1-348 
 
 1-311 
 
 1-275 
 
 1-241 
 
 1-209 
 
 1-80 
 
 1-736 
 
 1-676 
 
 1-620 
 
 1-572 
 
 1-519 
 
 1-473 
 
 1-428 
 
 1-388 
 
 1-350 
 
 1-314 
 
 1-280 
 
 1-90 
 
 1-833 
 
 1-769 
 
 1-710 
 
 1-659 
 
 1-604 
 
 1-554 
 
 1-507 
 
 1-465 
 
 1-425 
 
 1-387 
 
 1-350 
 
 2-00 
 
 1-929 
 
 1-862 
 
 1-800 
 
 1-746 
 
 1-688 
 
 1-636 
 
 1-586 
 
 1-542 
 
 1-500 
 
 1-460 1-421 
 
 I 
 
 Circumferential Seams* are almost always lapped and double-riveted. 
 Boilers of small size and for working pressures under 100 lbs. are single- 
 riveted, and as one row of rivets is quite sufficient to ensure tightness" of 
 joint, which is all that is required in this joint, the single-riveted joint answers 
 the purpose. Boilers made for very high pressures are often treble-riveted 
 in the middle circumferential seams, the Board of Trade rules admitting of 
 a thinner plate when this is done. In the case of large double-ended boilers, 
 the circumferential seams should be always treble-riveted except the end 
 ones, which seldom need be done in this way, and then only when the plates 
 are very thick. 
 
 Methods of Work.— Formerly most of the holes in boiler plates were 
 punched and drifted fair. Since the Board of Trade placed a premium on 
 drilled holes, machines have been made which compete successfully, both 
 in the cost and speed of output, with the punching machine, so that now 
 all the holes in a boiler are drilled, and generally done in place 
 
 ^S8ESitiES£gtt :K£„(^l't;, ; " kd and 64 for *-*i— * thick 
 
ALLOWANCE FOR WEAR. 651 
 
 Material. — The shell of a cylindrical boiler is now made of steel plates with 
 a strength up to 32 tons, while the Admiralty limit it to 30 tons ; it is expected 
 to stretch 20 per cent, per fracture (see Table lxvii.). Steel of a higher 
 strength is sometimes used, and that with 35 tons as the limit must have 18 
 to 20 per cent, extension ; sometimes steel as high as 40 tons per square inch 
 has been used for shells ; it should have the same extension. 
 
 Siemens steel has quite taken the place of iron in boiler-making, and, 
 as it possesses a much higher tensile strength, with greater toughness, it 
 is a more suitable material for the purpose. A boiler made wholly of 
 steel was cheaper than when wholly of iron, which no doubt was one 
 cause of the increased demand for steel boilers. Steel plates can, at a trifling 
 extra expense, be supplied oi very large sizes, exceeding largely those made 
 by the old Yorkshire ironmasters. As a matter of fact, the overhead 
 price of a heavy specification of steel plates was not seriously greater than 
 that of a light specification, while there was a very considerable difference 
 for iron if large and heavy plates were included. The size of plates to be 
 used in the construction of the shell depends now on the appliances of the 
 boiler -maker ; the breadth of plate is limited by the depth of gap in the 
 riveting machine, and the length of plate by the capabilities of the planing 
 machine and squeezer or rolls ; and the weight of the plate is limited to the 
 strength of the various small cranes, etc. Steelmakers can make plates 
 up to 12*5 feet broad and 50 feet long, and If inches thick (Chap. xxx.). 
 
 Nowadays rolls or squeezers for curving plates, planing machines for 
 truing the edges, drilling machines for dealing with all the holes, and riveting 
 machines for closing the joints are made in such large sizes that a double- 
 ended boiler of the largest kind can be made in two drums, and each drum 
 in two plates ; single-ended boilers in one drum, and each drum in two 
 plates ; some people prefer to make each drum in one plate when the diameter 
 permits of it. As a rule, it may <be assumed that a single-ended boiler has 
 one drum or strake of plating, the strake consisting of two plates ; a double- 
 ended boiler should have three strakes of plating, and the longitudinal seams 
 should be so arranged that no two of them come in line nor interfere with the 
 seams at the ends, and they should be well above the furnace line. 
 
 Allowance for Wear.* — All boilers are designed so as to last as long as 
 possible, and, since wear takes place by corrosion, some additional thickness 
 must be provided at first to meet this condition. The Board of Trade tacitly 
 make this allowance by using a high factor of safety ; but since the factor 
 of safety causes the additional thickness to be proportional to the total thickness* 
 while the wear takes place independently of thickness, it does not properly 
 meet the case. A boiler with plates | inch thick will waste the same quantity 
 of steel per square foot as one with plates 1 inch thick if worked under similar 
 conditions. Suppose such waste to be £ inch in a certain time, the loss 
 in one case is 25 per cent., but only 12| per cent, in the other. To meet 
 the case properly, the factor of safety should be reduced, and a constant 
 quantity added as is done by the Bureau Veritas ; for example, in the case 
 mentioned above the plates should be ^ inch and 1 inch, so that at the end 
 of the time the former should be 12-| per cent, under \ inch in thickness. If 
 any further proof be needed, it is only necessary to calculate the thickness 
 
 * Now that salt water is seldom admitted to a boiler, and greater care is taken, 
 the waste of material is only slight. 
 
€52 
 
 MANUAL OF MARINE ENGINEERING. 
 
 of plates for boilers of small diameter and low pressure to find it such as 
 would be impossible to rivet and caulk tight. 
 
 The following rules make due allowance for such contingencies : — Let D 
 be the diameter of the shell in inches, p the working pressure in pounds 
 per square inch, F a factor ; then 
 
 Thickness of shell plates 
 
 D x p 
 F 
 
 + 2 (in 32nds). 
 
 TABLE LXXVIII— Values of F on Condition. 
 
 
 
 Tensile of Pistes. 
 
 \\ hen the Longitudinal Joints are 
 
 28 Tons 
 
 35 Tons. 
 
 1. Lap, ordinary, single-riveting, .... 
 
 2. ,, ,, double-riveting, .... 
 
 3. ,, „ treble-riveting, .... 
 
 4. ,, special, treble-riveting (4 rivets per pitch), 
 
 5. Butt, double straps, ordinary, double-riveting, 
 
 6. ,. ,, special, 2 rows, 3 rivets per pitch, 
 
 7. ,, „ ordinary, treble-riveting, . 
 
 8. „ ,, special, 3 rows, 5 rivets per pitch, 
 
 9. ,, ,, quadruple, 4 rows, 9 rivets per pitch, 
 10. ,, „ quadruple, 4 rows, 11 rivets per pitch, 
 
 
 562 
 662 
 719 
 762 
 750 
 812 
 812 
 850 
 920 
 941 
 
 702 
 
 827 
 
 9C0 
 
 952 
 
 938 
 
 1.015 
 
 1,015 
 
 1,062 
 
 1,150 
 
 1,176 
 
 D is the internal diameter in, inches. 
 ■p is the working pressure. 
 
 Boiler Ends. — There are several methods of connecting the end-plates to 
 the shell. 
 
 (1) Flanging the end-plates. — This is now the universal and the best plan, 
 for there are only one set of rivets and two caulking edges, and the room 
 occupied is less than with angles. Flanging is done by special appliances 
 at a trifling cost in various designs of hydraulic press. The flanges (fig. 238) 
 are usually inside the boiler, but when the shell is of so small a size that the 
 rivets cannot be held up inside, or it is desired to rivet the joint by machine, 
 the front end is placed (fig. 239) with the flange outside. In either way there 
 is no tendency to force the joint by pressure on the ends, as there is with 
 
 (2) Flanging the shell plates, which some few engineers, in order that a 
 thicker plate may be provided to withstand the wear that takes place at the 
 bottom corners, or rather edges, of the cylindrical boiler, and to avoid the 
 fitting of the flanged ends into the shell with fair accuracy, resort to turning 
 it inwards (fig. 240), so as to form a connection for the ends. This has some 
 bad features, among which may be cited the difficulty and cost of flanging 
 thick plates across the grain, especially after being bent ; the necessity for 
 such shell plates to be of low tenacity to have so much ductility ; and lastly, 
 the pressure on the end always tending to open the joints. It has, however, 
 
THICKNESS OP THE END-PLATES. 
 
 653 
 
 some few advantages, but they are dearly bought, for, to avoid the chance 
 of leakage at the corners of the end-plates, there is far greater risk of leakage 
 from the horizontal shell joints, which are of necessity flanged over too, unless 
 welded. It may be added that such an arrangement quite prevents the 
 furnaces from being near the shell. 
 
 Riveting. — The cross seams of the backs are usually double-riveted, 
 although for pressure under 100 lbs. single-riveting does quite well. The 
 riveting of the ends to the shell is usually of the same design as that of the 
 other circumferential seams, but, in double-ended boilers, they need not be 
 treble-riveted because the middle seams are. 
 
 The Quality of Plate depends largely on the amount of flanging to be 
 done. To stand flanging around the edges of the circular plates, they may 
 
 Fig. 238. 
 
 Fie. 239. 
 
 Fig. 240. 
 Figs. 238 to 240.— Methods of Connecting Shell and End Platee. 
 
 be of the same quality as the shell ; but, when the furnace holes are flanged 
 to meet the furnaces, a softer quality is desirable, and it is better to use the 
 mild kind as ordered for internal parts, having an extension of 25 per cent. 
 at least to stand such severe treatment. 
 
 The Thickness of the End-plates and the pitch of the stays are inter- 
 dependent to a certain extent ; but, since the stays in the upper part of the 
 boiler must be wide enough apart to admit of a man passing between them, 
 the plates at the upper part of the ends must be made thick enough to suit 
 this pitch of the stays, which are usually, in consequence, somewhat thicker 
 than the other part of the ends ; some makers, however, prefer to make the 
 ends of one uniform thickness, and stiffen the top plates to stand the wide 
 pitch of the stays by riveting on thick and large washers in wake of these 
 
<554 MANUAL OF MARINE ENGINEERING. 
 
 stays as provided for in the Board of Trade rules. The end-plates are gene- 
 rally from f to | inch thick, although a few makers prefer to have thicker 
 plates, so as to avoid the necessity for doubling plates about the man- and 
 mud-holes, and for fitting nuts and washers to the screwed stays. Taking 
 15 inches as the smallest convenient pitch for the stays in the steam space, 
 and suppose them to have riveted washers as well as double nuts, Lloyd's 
 requires the plate to be ^ inch thick for 150 lbs. working pressure, and jfr inch 
 thick for 200 lbs. The Board of Trade permit plates ^ inch less for these 
 pressures. 
 
 Furnaces. — Those fitted in the cylindrical boiler are invariably of circular 
 section, that being the best form to resist a uniform external pressure. The 
 strength of such a furnace varies as the square of the thickness of plate, and 
 inversely as the diameter and length. From experiments made on a large 
 scale, there is, however, reason to doubt the supposition that the strength 
 varies exactly inversely as the length. 
 
 Let D be the external diameter, in inches, of the furnace whose length is L feet, 
 t the thickness of the plates in parts of an inch. 
 
 Safe working pressure = . j r (Board of Trade). 
 
 o f ,- 89,600 X t* /T1 „ _, 
 Or, Safe working pressure = — _ _— (Lloyd s Registry). 
 
 When the plain part of a furnace exceeds 120 times the thickness of the plate, 
 
 51*5 
 Safe working pressure = -=— (18-75 t — 12-36 L) (British Corporation). 
 
 If the furnaces are constructed by welding the plates together, or connecting them 
 
 by a butt joint with double straps single-riveted, or by a single butt strap double-riveted, 
 
 the Board of Trade allow the above factor. To avoid making by the above rules a furnace 
 
 which might give way by the crushing of the material, the Board of Trade insist that, in 
 
 9 900 X t 
 no case shall the working pressure, in pounds per square inch, exceed — - — — , and 
 
 ((300 t 12 L)\ 
 sr 1, when the length* 
 
 of the plain part is less than 120 times the thickness of plate. 
 
 Rule for Thickness of a Plain Furnace.*— The following gives a suitable 
 thickness for the furnace plates in 32nds of an inch, and makes due allowance 
 for uniform wear of the surfaces : — 
 
 -^- 
 
 ?+2. 
 
 F 
 
 Here L is the length and D the diameter, both in inches ; and F a factor 
 which is for ordinary mild steel, 1,100. 
 
 Furnace plates are made from § inch to f inch thick, the most general 
 sizes being ^ to ^ inch. Some engineers, however, in spite of this, to 
 avoid using corrugated furnaces or fitting stiffening rings, made the furnace 
 of plates as thick as f inch ; if the boiler is kept quite clean there is no 
 objection. 
 
 If the furnace is so long that -^-inch plates are insufficient by the rules 
 for the working pressure, some means of stiffening it should be resorted to ; 
 such stiffening was generally effected by means of rings, as the effective length 
 for purposes of calculation is thus virtually the longest distance between 
 
 • For new Rules of the British Marine Engineering Design and Construction Committee, v. Appendix. 
 
METHODS OF STIFFENING FURNACES. 
 
 655 
 
 such rings, or between such rings and the ends, it may be taken as the value 
 of L in the foregoing formulae. 
 
 The Methods of Stiffening Furnaces were— (1) By making the furnace in 
 two or more drums and connecting them by means of a {J-shaped hoop, called 
 the " Bowling hoop " (fig. 241), because first made by the Bowling Company. 
 These hoops are weldless, and possess a considerable amount of elasticity ; 
 so that, in addition to stiffening, they allow expansion longitudinally on the 
 part of the furnace. This was a very convenient plan, as it admitted of the 
 furnace being partially withdrawn in case of damage, etc., and notwithstanding 
 that there are two thicknesses of plates, and two laps at each joint of the 
 furnace, it gave every satisfaction when tried. No one, however, makes 
 these hoops now,, so that this type of joint has gone out of use. 
 
 (2) By making the furnace in two or more drums, and connecting them by 
 means of flanges, formed by turning the plate end outwards (fig. 242). To 
 allow of a caulking edge on both sides of the lap, a thin ring is introduced 
 between the flanges. This was a favourite method, because no joint or 
 riveting is exposed to the fire. Such joints generally give trouble from the 
 
 Fig. 241. 
 
 Fig. 242. 
 
 Fig. 243. 
 Figa. 241 to 243. — Methods of Stiffening Furnaces. 
 
 strain on the boiler end tending to open the joint, and by the wearing away 
 of the metal at the root of the flanges, due to mechanical action. Furnaces 
 made on this plan also required more room in the boiler, and the flanges 
 blocked up the space between them and the boiler shell. 
 
 (3) By making the furnace in two or more drums of different diameters ; 
 those of small diameter are flanged out so as to fit into, and be connected to, 
 the larger ones with lap joints single -riveted (fig. 243). It possesses one 
 or two very useful features, among which may be reckoned the capability of 
 the furnace, being small at the mouth, to leave good space on the boiler 
 fronts for manholes, etc., and while small at the combustion-chamber ends, 
 it is of good diameter in the middle. It has, however, the objection of 
 presenting joints to the direct action of the fire. 
 
 (4) By making the furnace with a series of corrugations or ridges. — There 
 are now several ways of accomplishing this, the best known being that of 
 the late Mr. Fox (fig. 244), and that of Mr. Purves (fig. 245), made by 
 John Brown & Co. of Sheffield, the more recent patent being that of Mr. 
 
656 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Morison (fig. 246) ; this form possesses the good features of the Purves, and 
 avoids the fault of the Fox. Fig. 247 is the modified form of corrugated 
 furnace made by C. D. Holmes. Fig. 248 is the form in which Mr. Deighton, 
 of Leeds, makes furnaces which have proved quite satisfactory. 
 
 The corrugated furnace was an extension of the Bowling-hoop principle. 
 
 ■^~<^j 
 
 Fig. 245. — Purves' Furnace in Section. 
 
 Fig. 244. — Fox's Furnace. 
 
 Fig. 246. — Morison's Furnace in Section. 
 
 Fig. 247. — Holmes' Furnace. 
 
 and its genesis is best illustrated by reference to fig. 247, which shows the 
 plan followed by Mr. Holmes. Here there are comparatively few corruga- 
 tions, but still sufficient to give the necessary stiffness to the furnace. For 
 increased pressure, such a furnace as this must either be made of thicker 
 
 Fig. 248. — Deighton's Furnace. 
 
 plate or have the corrugations closer together ; consequently, for the same 
 pressure and dimensions, the Fox furnace will be thinner than that of the 
 Holmes. On the other hand, the Holmes furnace is more rigid longitudinally 
 than the Fox furnace. The corrugated furnace made by the Farnley Company 
 has the corrugations formed spirally around the furnace, and are said thereby 
 
METHODS OF STIFFENING FURNACES. 657 
 
 to give greater longitudinal rigidity, without sacrificing too much of its 
 transverse stiffness ; it must, however, tend to set up twisting strains when 
 end-pressure is applied, which would bring shear on the rivets, and the trans- 
 verse stiffness can be only little more than that of the Holmes furnace. 
 ■ These special furnaces soon became a necessity for large diameters and 
 high pressures ; but, although immensely strong so long as the metal is 
 cold, corrugated ones will probably collapse longitudinally when red hot 
 quicker than an ordinary furnace, from the fact of there being superabundance 
 of plate between the extreme points of support to supply the extra length of 
 the arc over that of the chord ; a common furnace cannot come down in this 
 way without stretching the metal ; in the Fox design the corrugations are 
 simply drawn out of shape. 
 
 The Purves furnace was practically an extension of the Adamson-joint 
 principle, and is shown in fig. 245. This furnace possesses quite as much 
 transverse stiffness as the Fox, while being superior in longitudinal rigidity ; 
 it is. moreover, easier to clean and to repair. 
 
 For corrugated furnaces, the following rules hold good : — 
 
 „. . . 14,000 X thickness in inches ,,. , , _ , . 
 (a* Working pressure = — =r (Board of Trade), 
 
 D being taken as the smallest outside diameter in inches. 
 
 C x (T 2) 
 
 (b) Working pressure = .=r (Lloyd's Register). 
 
 T is the thickness in sixteenths of an inch. 
 
 D is the smallest diameter outside. 
 
 C = 1,259 for Fox's, Morison's, Deighton's, Beardmore's, or Leeds Forge bulb 
 
 furnaces where made of steel, 26 to 30 tons tensile strength. 
 C = 1,160 for Purves', when the ribs are 9 inches apart ; or Brown's, ribs 8 or 
 
 9 inches apart. 
 G = 945 for Holmes', and C = 912 for Farnley's. 
 
 q v (T 2) 
 
 (e) Working pressure = ^r (British Corporation^, 
 
 C = 1,250 for Leeds Forge bulb furnaces. 
 
 C = 1,160 for Fox's, Purves', Deighton's, Morison's, and Brown's. 
 
 C = 950 for Holmes' and Farnley's. 
 
 (>p 2) 
 
 (d) Working pressure = C - — „ (Bureau Veritas). 
 
 C = 1,260 for corrugated and bulb furnaces. 
 C = 1,100 for ribbed furnaces. 
 
 Fig. 249 shows a Morison furnace with its back end so formed as to permit 
 of the furnace being withdrawn through the hole in front of the boiler, without 
 sacrificing any of the good features of the method of jointing of the com- 
 bustion chamber. 
 
 Messrs. John Brown & Co. have also a furnace with a back end formed 
 so as to permit of the furnace being withdrawn in the same way. In this 
 case, however, the furnace is reduced somewhat in diameter and flanged 
 outward all round {v. fig. 251a). This flange is not concentric with the body 
 of the furnace, its centre being higher so as to permit of the connection of the 
 tube plate in the usual way, the back tube plate being flanged and riveted 
 into the combustion chamber all round, and formed with a round hole corre- 
 sponding to that of the furnace neck. 
 
 * For new Rules of the British Marine Engineering Design and Construction Committee, v Appendix. 
 
 42 
 
658 
 
 MANUAL OP MARINE ENGINEERING 
 
 Methods of Connecting Furnaces to End-plates. — There are two distinct 
 ways of accomplishing this : — 
 
 (1) By flanging the furnace to meet the front plate. — This was formerly 
 a common one, because the iron of a furnace was always of a soft high quality, 
 capable of being easily flanged, but the method is objected to as the pressure 
 on the ends tends to open the joint. The flanging is invariably outwards, 
 and when done the root should have good curvature, the radius of the outer 
 surface being at least 1| inches. Since steel has been used for boiler ends 
 this method has been practically dropped, as it does not admit of the furnace 
 being withdrawn in case of need. 
 
 1 
 1 
 1 
 1 
 
 1 T 
 
 —i r 
 
 -i r 
 
 
 ' '1 t 
 
 i ' i 
 i ; | 
 
 ilUi| 
 
 F 
 
 — i — i 
 — i r 
 
 — , — T 
 
 / 
 
 J/ 1 
 
 1 
 
 Fig. 249. — Removable Furnace by the Leeds Forge (Ashlin's Patent). 
 
 (2) By flanging the front plate inwards (fig. 250) or outwards (fig. 251) to 
 meet the furnace. The former is the most general method, the latter being 
 resorted to only when the boiler is so small as to prevent the rivets from 
 being properly " held up," and when it is necessary to have the furnaces 
 very n ar the shell. In naval ships the boilers were often made in this way. 
 This flanging in or out makes the best finish to the boiler front, and leaves 
 more room for man- and mud-holes, etc. , and permits of much neater furnace 
 fronts. The joint (fig. 250) is capable of being caulked at both edges, and 
 
COMBUSTION CHAMBERS. 
 
 659 
 
 the strain on it is across the rivets, and does not, in consequence, tend to 
 start the caulking. 
 
 Combustion Chambers. — The length of a combustion chamber, measured 
 in line with the furnace, should be such that its capacity above the level of 
 the fire-bars is equal to the total capacity of the furnace, when the boiler is 
 single-ended ; when double-ended and one combustion chamber is common 
 to opposite furnaces, the capacity of the combustion chamber should be 
 equal to three-fourths of the combined total capacity of the two furnaces. 
 
 To obtain such a capacity of combustion chamber when the boiler is 
 single-ended, or double-ended and divided transversely, the length must be 
 about two-thirds the diameter of the furnace, and when common to two 
 opposite furnaces, it must nearly equal the diameter of furnaces. 
 
 Combustion chambers are generally formed with flat tops, but made 
 sometimes by curving the back plate over the top to meet the flange of the 
 tube plate. The latter plan avoided the necessity of the girder stays 
 to support the flat top, and reduced the number of joints of plating, 
 
 Fig. 250. 
 
 Fig. 250«. 
 
 Fig. 251. 
 
 Fig. 251a. 
 
 Figs. 250 to 251a. — Methods of Connecting Furnaces to Boiler Ends and Tube 
 
 Plates. 
 
 but the capacity of the combustion chamber is less, and the space for tubing, 
 etc., contracted. It used to be claimed for this form that staying is avoided, 
 but this is not a substantial gain, as, in a single-ended boiler, the stays which 
 are necessary for the back end-plates form the stays of the chamber, and 
 in a double-ended boiler, if stays are omitted between the chambers, the 
 Board of Trade surveyors require additional staying in the steam space 
 to tie the ends of the boiler together. Although the plan was a favourite one 
 with engineers a few years ago, when the pressure was under 100 lbs., it is 
 now seldom seen. 
 
 The thickness of plates and pitch of stays are, of course, interdependent, 
 but, as a rule, the chambers of large boilers, whose working pressure is 150 lbs. 
 per square inch and upwards, are made of ^-inch to f-inch plates, and those 
 oi smaller boilers, or those working at lower pressures, are made of ^-inch 
 to |-inch plates. 
 
660 MANUAL OF MARINE ENGINEERING. 
 
 By the Board of Trade rules, a stay-bar If inch in diameter, screwed 
 10 threads per inch, will sustain 73 square inches at a working pressure of 
 150 lbs., and only 55 square inches at 200 lbs., while a If -inch stay, with 
 9 threads per inch, will sustain 77 square inches at a pressure of 200 lbs. 
 A plate \ inch thick requires a stay for 7*74 inches pitch, or 60 square inches 
 at a pressure of 150 lbs., while a f-inch plate requires one for 9-3 inches pitch 
 or 86 square inches for that pressure ; for 200 lbs. pressure, the f-inch plate 
 requires a stay for a pitch of 8 inches or 66 square inches. 
 
 Board of Trade Rule : — Pressure = ^ — ^ — — 
 
 D — O 
 
 T is the thickness of flat plate in sixteenths of an inch ; S, area of surface 
 supported in square inches ; C, a constant which, for screwed stays with 
 nuts and plates exposed to flame, but in contact with water, is 100. 
 
 For stays in steam space fitted with riveted washers, two-thirds the 
 pitch in diameter, nuts, etc., 210 if washers are of the same thickness as 
 plate, but only 165 if with plain washers two-thirds the thickness of plate 
 and three times the diameter of stay. 
 
 The bottom of the chambers should be -j^ to £ inch thicker than the 
 sides, as from various causes there is often rapid wear in that part ; also to 
 avoid excessive staying and to provide for burning, which sometimes takes 
 place there, the plates at the top should be ^ inch thicker. 
 
 Some steel makers now supply plates having a varying thickness so that 
 one single plate can be wrapped around the combustion chamber so as to form 
 the top, sides, and bottom, with the latter of the necessary extra thickness. 
 
 The back tube plates vary in thickness from -^ in small boilers for low 
 pressures, to § inch, and even f inch, in large ones for high pressures. Gene- 
 rally, in modern boilers of ordinary sizes and pressures, the back tube plate 
 is f to f inch thick, the former being the best size when possible, as with it 
 the tubes can be made quite tight, and there is less liability of cracking the 
 plates or burning the tube ends than with the thicker plates. 
 
 The back plate and tube plate of the combustion chamber are almost 
 invariably flanged inwards to take the side plates and those on the top and 
 bottom ; some makers have tried to make the chambers by flanging the 
 sides top and bottom to meet the back and tube plates (v. fig. 203) ; but, as 
 this is very troublesome to effect, and prevents the tubes from being extended 
 to the sides and top, it is seldom followed. The flange of the top plate of the 
 furnace should be inside the chamber, and connected to the tube plate with 
 counter-sunk rivets ; the landing edge is then turned away from the " wash " 
 of the flame, and no rivet heads are exposed to it. The landing edges of all 
 joints of plating exposed to water should be downwards, so that deposit 
 cannot lodge on them ; when they are upwards on the water side the deposit 
 on them is very considerable, and it is found that rapid corrosion then takes 
 place in the angle beneath it. The back tube plate is by some engineers 
 welded to the furnace, so that no joint is exposed to the severe action of the 
 flame issuing from the fuel. It is an expensive thing to do, and utterly 
 prevents the withdrawal of a damaged furnace when done. 
 
 Tubes. — The tubes in the ordinary marine boiler are from 2£ inches to 
 4 inches external diameter, the usual sizes being from 2| inches to 3| inches 
 
TUBES. 
 
 661 
 
 in the mercantile marine, and 2| inches to 2| inches in H.M. Navy. With 
 the ordinary natural draught the tubes should not be more than 24 diameters 
 long ; with the forced draught they may be as much as 60 diameters long, 
 as in a locomotive boiler, but the practice in torpedo boats and steam launches 
 was about 35 diameters long, and generally with Howden's forced draught 
 is 30 to 36. The length, however, does not matter so much with forced 
 •draught. 
 
 The spacing of the tubes often depends on circumstances, but in the 
 mercantile marine, where space and weight of machinery are not of such 
 moment as in the Navy, the pitch of the tubes is usually 1-4 x diameter. 
 There is less liability to prime when the tubes are widely spaced, and they 
 are more easily cleaned from scale. For the latter purpose they are arranged 
 in rows, both horizontally and vertically, and not zigzag, as often seen in 
 locomotive boilers. 
 
 Tubes are manufactured of a certain minimum thickness, and said to be 
 "" according to list " when so made. If the pressure they are intended to 
 withstand does not exceed 60 lbs., they may be " according to list " ; if it 
 does not exceed 100 lbs., they should be " 1 gauge thicker than the list " ; 
 if the pressure does not exceed 150 lbs., the tubes should be " 2 to 3 gauges 
 thicker than the list." By rule, L.S.C. = 300 -^ cl x Vp. 
 
 The following table gives the thickness under the various circumstances 
 in the numbers of the Legal Standard Gauge : — 
 
 TABLE LXXIX. 
 
 — Boiler Tubes. 
 
 
 
 
 
 External diameter of tubes, Ins., 
 
 2 
 
 2i 
 
 2i 
 
 2| 
 
 3 
 
 31 
 
 10 
 
 3J 
 
 10 
 
 3| 
 
 4 
 
 Thickness for 60 lbs., - L.S.G., 
 
 12 
 
 12 
 
 11 
 
 11 
 
 11 
 
 10 
 
 9 
 
 Thickness to 100 lbs., - L.S.G., 
 
 11 
 
 11 
 
 10 
 
 10 
 
 10 
 
 9 
 
 9 
 
 9 
 
 8 
 
 Thickness to 150 lbs., - L.S.G., 
 
 10 
 
 10 
 
 9 
 
 9 
 
 8 
 
 8 
 
 7 
 
 7 
 
 6 
 
 Thickness to 200 lbs., - L.S.G., 
 
 9 
 
 9 
 
 9 
 
 8 
 
 8 
 
 7 
 
 7 
 
 6 
 
 6 
 
 It is usual to make the tubes of slightly larger (^ to | inch) diameter at 
 their front end, so as to draw out easily when once started from the plates. 
 Tube manufacturers will swell the ends to -^ inch larger diameter without 
 extra charge. 
 
 Boiler tubes were formerly of iron, hard copper, or brass ; now they are 
 exclusively of iron or steel ; when of the former they are made from strips 
 of best Staffordshire, or other good iron having a tensile strength of 20 tons 
 with an elongation of 12 per cent. When of steel, they are made either of 
 strips having a tensile not exceeding 27 tons, or solid drawn from mild steel 
 billets. Those for water-tube boilers are always solid drawn, cold finished, 
 and carefully made to gauge thickness and character. 
 
 Iron tubes are generally used in the mercantile marine ; brass tubes were 
 used in the Navy, partly because of their superior conducting power, but 
 chiefly on account of their endurance and reliability, as the iron tubes then 
 
662 
 
 MANUAL OF MARINE ENGINEERING. 
 
 made seldom lasted more than four years, and after only as many days a few- 
 would sometimes prove defective from small holes being formed, which 
 rapidly enlarged with the rush of water and steam through them. 
 
 Brass tubes were made of a composition of 68 per cent, of B.S. copper and 
 32 per cent, of spelter, which, when drawn out into the tube, has a very 
 high tensile strength and is very tough. Such tubes lasted ten or twelve 
 years under ordinary circumstances, but, if used with coal containing much 
 sulphur, they perished more rapidly and lost the toughness. They cost 
 about four times the price of iron tubes, but, when condemned, were worth 
 about half the original cost, and, since they lasted at least double the time 
 and their superior efficiency was a sufficient set-off for interest on capital, 
 the brass tubes were more economical than the iron ones. 
 
 Steel tubes are now being used on a large scale ; the early experience 
 with this metal for tubes was not always happy ; there is, however, steel 
 and steel, and because tubes of that material failed egregiously many years 
 ago, it is no reason why tubes made of modern steel should not be used now, 
 especially in steel boilers. The Admiralty use welded steel tubes in cylind- 
 rical boilers, and the tube-makers are confident that they can make as perfect 
 a weld with the mild steel strips they use as with iron. For internal pressure, 
 however, a solid-drawn tube is preferred, and always used. 
 
 To preserve the ends of the tubes in the back tube plate from being 
 wasted away and the severe action of flame on the tube plates with forced 
 draught, the Admiralty used iron ferrules fitted to them. These ferrules 
 (fig. 252) were made of malleable cast iron ; they have a slight taper and a 
 
 mushroom-shaped flange, so that when 
 driven home they are very tight and pro- 
 tect the end of the tube and surrounding 
 plate completely. 
 
 Stay Tubes. — The tube plates are 
 usually held and stiffened by some of the 
 tubes of greater thickness being arranged 
 as stays ; for that purpose they are screwed 
 at the ends with a fine thread (9 to 10 
 threads per inch), and either tapped into 
 both plates or tapped into the back plate 
 and screwed by nuts to the front plate. 
 The better plan is the former, the back- 
 end thread is minus and the front end plus, and the tube screwed into both 
 plates at the same time. The section at thread in this plan is in excess 
 of that of the tube owing to its large diameter, and the tube can be with- 
 drawn at any time without disturbing the others. If fitted with nuts, there 
 is great difficulty in getting the tube out, generally necessitating the with- 
 drawal of a whole row, while there is really no difficulty in tapping the holes, 
 and no necessity for nuts. The stay tubes in the Navy have a plus thread 
 at each end, the diameter of the front end being larger than that at the back ; 
 in this case, the thickness at bottom of thread is the same as that of the 
 body of the tube. 
 
 Stay tubes are \ to | inch thick in the body, and as the thread is 
 <& inch deep, they are -& to {« inch thick at the bottom of thread. Sonic 
 makers prefer to fix thick tubes, and space them farther apart. For 100 lbs. 
 
 Cup ferrule for ffot!cr_JTubes_ . 
 
 Fig. 252. 
 
STAYS. 663 
 
 working pressure and upwards, the stay tubes are J to | inch thick, each 
 alternate tube should be a stay tube— that is, in a nest of, say, 64 tubes, 
 there will be 16 stay tubes. 
 
 Stay tubes generally outlast two sets of the ordinary tubes. 
 
 Serve tubes, manufactured by John Brown & Co., Sheffield, have a series 
 (6 to 8) of longitudinal ribs running the whole length of the tube inside 
 and projecting inwards about f inch, thus forming an additional absorbing 
 surface for the heat. Besides the additional surface, this tube causes a 
 better circulation of the hot gases within it. In small boilers especially is 
 this tube a useful one, as, with a large grate area, a sufficient heating 
 surface can be obtained. 
 
 Retarders. — By introducing into the inside of a plain tube a twisted strip 
 of metal of the same breadth as the inside diameter, Howden found a con- 
 siderable gain in efficiency, owing to the hot gases getting stirred and circu- 
 lated in their passage through it. 
 
 Stays. — Flat surfaces have to be stiffened and tied together by bars called 
 stays. When the surfaces are close together and the plates comparatively thin, 
 so that the stays are short and numerous, they are screwed into both plates, 
 and the ends either riveted over or fitted with lock nuts ; such stays are 
 usually called " screwed stays." As has been said, the thickness of plates 
 and pitch of stays are interdependent. The size and number of the stays 
 depend on the pressure they have to withstand. The stays in the steam 
 space must be so spaced that a man can pass between them, and for this 
 purpose they should never be nearer than 14 inches, centre to centre, and 
 are usually 15 to 17 inches centres, which gives a clear space of 12 to 14 inches 
 between them. These stays are seldom more than 3 inches effective diameter, 
 and as the exact spacing of them depends on the form and size of the boiler, 
 they are generally arranged to suit the particular case, and the diameter 
 varied to give a section adequate to the load each has to bear. To admit 
 of easy access, these stays are arranged in horizontal and vertical rows as 
 nearly as possible. 
 
 The Admiralty allow a stress of 18,000 lbs. per square inch of effective 
 area of stay when at the test pressure, which is usually 90 lbs. above the 
 working pressure. If the stays are below 1| inches diameter, only 16,000 lbs. 
 
 The Board of Trade allow 9,000 lbs. per square inch at working pressure.* 
 
 Lloyd's Registry allow, on stays not exceeding 1| inches smallest diameter 
 at working pressure, 8,000 lbs. on " screwed " stays and 9,000 lbs. on others ; 
 on stays above 1| inches smallest diameter, 9,000 lbs. on " screwed " stays 
 and 10,000 lbs. on others. 
 
 British Corp oration Rule is D = */ — ~ (- J. 
 
 S = the surface in square inches supported by the stay. 
 D = the effective diameter of stay. 
 \Y -- the working pressure. 
 
 C = 8.000 for steel screwed stays ; 8,900 for steel longitudinal stays. 
 
 * Tested wrought iron of 21"5 tons tensile may be used for screwed stays instead of 
 steel, if preferred. Longitudinal stays may be of same quality as shell plates and 
 stressed to one-sixth the ultimate strength at working pressure. 
 
664 MANUAL OF MARINE ENGINEERING. 
 
 Bureau Veritas has the same rule, but C = 300 x tensile strength of 
 metal in tons, which, for iron, is 22 tons, and for steel according to test. 
 
 The large stays are sometimes made with a plus thread ; this necessitates 
 " upsetting " the ends so that the body is of somewhat smaller diameter 
 than at the bottom of thread of the ends. This is a somewhat expensive 
 process, and is not so rehable as simply screwing a rolled bar with a minus 
 thread at each end. The latter plan, especially since the making of steel 
 boilers has become general, is now fast taking the place of the former; it 
 has, too, the advantage of excess of section in the body where most corrosion 
 takes place. These stays are secured to the plate with a nut and washer 
 outside, and a nut inside to lock, whose length is two-thirds that of the out- 
 side, which is one diameter long. When weight is of consequence, steel 
 bars may be swelled and have a plus thread by " upsetting " the ends, as the 
 Board of Trade now permit this. 
 
 The screwed stays are usually from 1 J to 2 inches diameter, with a standard 
 thread 9 per inch. The most useful sizes are 1^, 1J, and If inch, suitable for 
 jg to f inch plates, and pressures from 60 to 220 lbs. per square inch. When 
 screwed through a plate whose thickness is less than half the diameter of the 
 stay in length there should always be a lock-nut with a thin washer, the nut 
 being two-thirds the diameter of the stay in length. The practice of 
 " nobbling," or riveting over the ends of these stays, is very objectionable 
 when they pass through thin plates, as extreme pressure is very apt to cause 
 the stay to draw completely through the plate, and this is especially so when 
 the plate is ductile and soft like mild steel. Screwed stays had fine threads, 
 which were in accordance with Whitworth's rule as to the number per inch — 
 viz., II diameter, 11 threads ; 1| diameter. 10 threads ; If diameter, 9 threads ; 
 and 2 inches diameter, 8 threads. It is, however, more convenient to have 
 only one number for all sizes, so now the Engineering Standards Committee 
 has decided on 9 per inch for all stays from 1J to 2 inches; ordinary stays 
 above 2 inches, 6 per inch, with nuts on each side of the plate. 
 
 When several stays are fitted with nuts and washers at their ends, the 
 following rule holds good : — 
 
 tv , , , . , n n /(Thickness of plates in sixteenths) 2 
 
 ritch of stavs in inches = 11 x \ ™ — rr — • 
 
 \ Working pressure 
 
 Example. — What pitch of stays is suitable for a plate f inch thick for a 
 working pressure of 160 lbs. ? 
 
 = 11 X y] ] 
 
 1Q2 
 
 Xa/ttx, or 8-73 inches. 
 
 C x (T -J- l) 2 
 Board of Trade Rule* — Working pressure = — -^ — ~^— '-, 
 
 S — 6 
 
 T being thickness of plate in sixteenths of an inch. 
 S the surface in square inches. 
 
 C being 100 for screw stays with nuts, 165 for longitudinal with nuts 
 and washers two-thirds the thickness of plate. 
 
 * For the new Rules of the British Marine Engineering Design and Construction 
 Committee for stays and flat surfaces, v. Appendix. 
 
CONSTRUCTION OF MARINE BOILERS. 665 
 
 C X T 2 
 
 Lloyd's Rule : — Working pressure = 
 
 P 2 
 
 P 2 being the mean of squares of pitches in rows and between rows, 
 and C and T as before. 
 
 C, for screwed stays with nuts, 110 with plates under ^ inch thick ; 
 over -& and under -fe inch, 120 ; over -& inch, 135 ; longi- 
 tudinal stays, with double nuts and outside riveted washers, 
 two-fifths the pitch in diameter, and half the thickness of 
 plate, 200. 
 
 f 1 v T2 
 
 British Corporation Rule : — Working pressure = =^5 . 
 
 pa _|_ ^2 
 
 Here P is the greatest and p the least pitch. 
 
 C is 265 for screwed stays with nuts, and 370 for stays with double 
 nuts and outside riveted washers, the latter being two-thirds 
 the thickness of plate and one-third the pitch in diameter. 
 
 Bureau Veritas Rule is similar to that of the British Corporation, but 
 the value of C changes with the tensile strength of the plate, and for screwed 
 stays with nuts 415-3 for 27-ton steel, and 461*5 for 30-ton. For longitudinal 
 stays with inside and outside nuts and washers, the outside washers being 
 0-4 of the pitch in diameter, and two-thirds the thickness of plate, C is 435*4 
 and 483*8. 
 
 Flat plates may be stiffened to allow of wider spacing of the stays than 
 given by this rule, by fitting thick washers of large diameter to each stay, or 
 by connecting the stays to the plates by means of angle-irons or T-bars ; 
 the latter plan possesses the advantage of distributing the strain over a 
 large area, and that without a doubtful joint, as is the case with nuts. The 
 old plan of riveting a doubling plate of common iron in wake of the large 
 stays has almost disappeared. 
 
 Continental Practice differs from British in two respects in the design 
 and construction of marine boilers. Almost invariably the tops of the com- 
 bustion chambers are in a horizontal line in whatever position they are placed 
 in a ship ; it is true that when a yacht or other vessel having sail power, 
 and requiring to use it whenever possible, as is the case with cruisers in the 
 Pacific and other extensive oceans where coal is scarce and dear, has the 
 boilers placed axially fore and aft, it is desirable that the chamber tops shall 
 be sloped, that when the ship is listed over under canvas the chamber top 
 on the weather side shall not be in danger of emersion and exposure to 
 steam only with the liability of overheating. In other ships with boilers 
 placed in this way there is, of course, some liability to list, and when rolling 
 the chamber tops may get their ends bare. In the latter circumstances 
 it is of no consequence, as the exposure is only temporary, and the emersion 
 only for very short periods, but with a list more or less permanent there may 
 be some risk of damage. It is, however, very slight, and British engineers 
 design the boilers under the belief that it is the duty of those in charge to 
 keep a sufficient amount of water in the boiler under any circumstances 
 to submerge the parts exposed to heat. If 1 oilers are made with the com- 
 bustion chamber tops sloped away, as is common in German practice (v. Fig. 
 210). there is considerable sacrifice of total heating surface as well as tube 
 
666 MANUAL OF MARINE ENGINEERING. 
 
 surface, and an increase in the weight of water in the boiler. On the other 
 hand, with boilers made in this way there may be at all times less water over 
 the chamber tops at middle than would be prudent with horizontal top ones. 
 
 Continental boiler-makers also adopt chain riveting frequently for shell- 
 plate joints instead of zigzag, and when employing the special arrangement 
 of treble riveting, whereby alternate rivets are omitted in the outer rows, 
 the inner butt strap extends only to the double rows of closer-pitched rivets 
 and the outer strap only is taken by the whole six rows of rivets. In this 
 way the edges of the inner strap can be caulked more severely, and the effective 
 rivet area per pitch is 8 '50 times the section of one rivet, as against 9 "37 
 when both straps are taken by the rivets, and a considerable weight of butt 
 strap is saved. In the old days of small plates, the latter consideration had 
 much more weight than it has to-day, when the largest boilers have six 
 longitudinal joints, whereas in those days a small single-ended boiler would 
 have that, and the largest double-ended ones 12 to 16 joints. 
 
 When weight is of very great importance, the butt straps can be vandyked 
 instead of being straight at the edges, and when this is so the caulking can be 
 heavier, as the distance apart of the rivets parallel to the edge is less than 
 the pitch of the outer rows. (v. figs. on. p. 649). 
 
 This vandyking with the outer and omission of cover with the inner 
 straps is followed likewise when the double-riveted joint is adopted with 
 alternate rivets in its ' outer rows omitted. In this case the effective rivet 
 area is 4*75, against 5-63, as commonly done. A few boiler-makers- will even 
 employ two sizes of rivet in these special joints with three rows, so that 
 those of wide pitch in the outer row are of larger diameter than those m the 
 inner rows. For example, with 10 inches pitch and lj-inch rivet the plate 
 left is 85 per cent, of the solid plate ; the inner row of rivets may be 1 J inches 
 diameter with the plate 75 per cent., instead of 70 per cent, with 1^ inches. 
 
 It is, however, in practice very inconvenient to work with two sizes of 
 rivet, and it is only under certain very special circumstances that it is worth 
 while departing from the common well-understood practice. 
 
 Water Spaces. — The spaces between the furnaces themselves, between 
 the furnaces and shell, and between the combustion chambers, although 
 sometimes diminished, should not be less than 5 inches ; that between the 
 backs of combustion chambers and shell should taper from 6 inches at the 
 bottom to 9 and even 12 inches at the top, to allow of the free current upward 
 of the steam generated on the surfaces. If the spaces are less than 6 inches, 
 it is very difficult to hold up rivets, to clean the surfaces from scale, or to 
 get a good circulation. 
 
 The space between the nests of tubes should not be less than 10 inches, 
 and, when possible, should be 12 inches. This permits a man to go down 
 to clean across, and ensures good circulation. 
 
 Man-holes. — The chief one in the shell should be oval, 16 inches by 12 
 inches; those in the ends, etc., may be 15 inches by 11 inches; and the 
 smallest through which a boy can pass is 14 inches by 10 inches. Mud- 
 holes are generally 9 inches by 6 inches, and peep-holes 6 inches by 4 inches. 
 The hole cut in the shell plating should be surrounded by a doubling plate 
 or angle bar ring, to compensate for the metal cut away, and the holes through 
 thin plates should have such rings to stiffen the edges. A better plan is to 
 flange the plate inwards so as to stiffen the edge and give a good face for the 
 
MODERN CYLINDRICAL BOILERS — SCANTLINGS. 
 
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66S MANUAL OF MARINE ENGINEERING. 
 
 door joints ; this is now very general. The doors are usually placed inside 
 the boilers, and held to their faces by studs screwed into them, which pass 
 through strong cross bars or " dogs," held by square nuts ; the main door 
 in the shell is, however, sometimes fitted externally, and connected to a 
 flanged ring with bolts, in the same way as a steam chest door on the engines. 
 The doubling ring is, in this latter case, formed of a very thick angle-iron, 
 whose deep tveb is flanged to fit to the boiler, and whose flange forms the 
 face for the- door. 
 
 The doubling plate at a shell man-hole is now as a rule flanged inwards, 
 .and the edge of the flange machined for the door to fit on. 
 
 The weight of boiler is approximately = ^ — -^- tons. 
 
 For single-ended boilers with a chamber to each furnace, F is 740. 
 For double-ended boilers with two furnaces to each chamber, F is 810. 
 For double-ended boilers with a chamber to each furnace, F is 780. 
 
 The toted heating surface is approximately = D 2 x L X K square feet. 
 
 For single-ended boilers, K is 0-97 to 1-15. 
 For double-ended boilers, K is 1-0 to 1-10. 
 
 The higher value of K in each case is with small tubes well packed ; the 
 lower value is for the boilers of ordinary merchant steamers with at least 
 3-inch tubes. In the above both L and D are in feet. 
 
SMOKE-BOX. (J69 
 
 CHAPTEE XXV, 
 
 BOILER MOUNTINGS AND FITTINGS. 
 
 Smoke-Box. — This appendage to the boiler is for the purpose of receiving 
 the products of combustion as they emerge from the tubes, and conducting 
 them through the " uptakes " into the funnel. In the old box form of boiler, 
 it was built inside, and formed an integral part of the boiler ; but, with 
 the modern cylindrical boiler, it is a separate structure, secured to the boiler 
 front by studs. It is constructed of iron or steel plates and angles of " ship " 
 quality, and made " smoke tight " only ; it should, however, be caulked 
 if necessary, to prevent air passing to the inside. The advent of cheap mild 
 steel permits of flanging instead of angles. In front of the tubes are a number 
 of doors hung on hinges, and so arranged that every tube may be swept or 
 removed in case of necessity. The doors should be rigid, and of such a size 
 as to be easily handled, and when the nests of tubes are so large as to cause 
 the door to be too ponderous if made in one, two doors may be fitted to 
 close on a portable stanchion. The doors are sometimes arranged to open 
 on a horizontal and sometimes on a vertical axis ; the latter is preferable 
 when possible, as then they are more easily handled. 
 
 The bottom of the smoke-box should be at least 12 inches broad, measured 
 in direction of the length of the boiler, and, when possible, as much as 15 
 inches. If too narrow, it is soon filled with soot and ashes, so as to cover 
 the ends of, and render useless, the bottom rows of tubes ; the baffle plates 
 on the doors are also soon burned away. The bottom plate should be at 
 least 2 inches below the bottom row of tubes, and the side plate the same 
 distance from the side tubes, so that the tubes may be drawn clear of the 
 2-inch angle-iron rim around the doorways. The front of the smoke -box is 
 sloped outwards, so as to be about twice the breadth of the bottom from the 
 boiler front, above the level of the top row of tubes. Above this, the smoke- 
 box contracts towards the funnel base, and its configuration must depend on 
 the position of this, and on the consideration that the section transverse to 
 the flow of gases must have an area at least equal to the area through the 
 tubes. The part between the smoke-box and funnel is called the " uptake," 
 or " take-up " ; it should have easy bends, and lead as directly as possible to 
 the funnel, and be without recesses and obstacles where eddies are formed 
 and the draught may be baffled. 
 
 The bottom and sides of the smoke-box and sides of the uptake should be 
 of £-inch plates for large boilers, -^ for smaller ones, and where weight is of 
 more consideration than endurance | is sufficient. The smoke-box doors 
 should be of the same thickness, and have baffle plates ^ inch thick on the 
 inside, and air or screen plates of the same thickness on the outside ; these 
 screen plates prevent radiation of the heat to the stoke-holes, and, for the 
 
670 
 
 MANUAL OF MARINE ENGINEERING. 
 
 same purpose, the sides of the smoke-box and uptake should be fitted iu the 
 same way. When steel is used, there may be a reduction of 20 per cent, 
 in these thicknesses, and, in naval ships, the outer casings may be jg inch 
 only. Steel sheets are now cheaper than iron, and will stand flanging, bending, 
 and working in a way that is impossible with iron ; consequently iron is now 
 not used for such purposes as these. 
 
 To protect the boiler front, which has only steam on its inner surface, 
 the uptake should have a back commencing from just above the level oi the 
 
 -Naval and Yacht Funnel Top. 
 
 Fig. 254. — Ordinary Funnel, with 
 Hood, etc. 
 
 top row of tubes. When this cannot be done, a good and well-fitting bailie 
 plate should be fixed to the boiler front. 
 
 A casing is also fitted round the funnel from its base to the level of the 
 deck casing, to prevent radiation. A corresponding casing is fitted round 
 the funnel, above the level of the deck coamings, to a convenient height, 
 and over this is fitted a hood secured to the funnel, so as to prevent water 
 
TUNNEL. 
 
 (571 
 
 passing down, while allowing the hot air to come out. This hood is called 
 by various names, as " cravat," " bonnet," etc. (fig. 254). It is now often at 
 the top of the funnel, so the latter is completely protected from the weather 
 (fig. 253). 
 
 Where there are several boilers discharging smoke to one funnel, each 
 smoke-box should have a separate uptake, so that the smoke from one does 
 not enter the box of another ; and, when there are no good ashpit doors, 
 there should be a damper in each of these uptakes, so as to regulate the 
 draught and get a uniform evaporation from all the boilers, and, in case of 
 necessity, to isolate a particular boiler. 
 
 TABLE LXXXI. — Pitch, etc., op Riveting for Funnels, Casings, etc. 
 
 (Admiralty Work). 
 
 Description. 
 
 Thickness of 
 Plate. 
 
 Uptakes, 
 
 Funnels, 
 
 Funnel casing, 
 
 Deck casings, 
 
 Inches. 
 
 ft 
 
 1 
 4 
 
 ft 
 3 
 
 k 
 3 
 Iff 
 
 I 
 
 ft 
 
 3 
 * 
 
 ft 
 
 Cowls and trunks (exposed to 
 weather), ... - 
 
 { 
 
 Screens and ventilating trunks, - -J 
 
 Other work, 
 
 ft 
 
 JL 
 
 8 
 
 ft 
 
 h 
 
 ft 
 
 ft 
 h 
 
 ft 
 
 Diameter of 
 Rivet. 
 
 Inches. 
 3 
 8 
 
 h 
 
 & 
 
 8 
 
 1 
 
 s 
 
 a 
 
 8 
 
 3 
 
 i 
 i 
 
 i 
 
 a 
 
 8 
 
 1 
 4 
 
 1 
 4 
 
 Pitch cf 
 Rivet. 
 
 Inches. 
 
 2 
 
 2 
 
 -2 
 
 H 
 
 o 
 O 
 
 2 
 9 
 
 3 
 2 
 
 2 
 
 2 
 2 
 
 2 
 
 2 
 2 
 
 4 
 
 3 
 
 2 to 3 
 
 2 to 3 
 
 2 
 
 Funnel. — This is usually of circular section, but sometimes, to minimise 
 the transverse size of the boiler-hatch, it is made of oval section. The funnels 
 of men-of-war are often made of oval section for the same reason, but, instead 
 of the section being an ellipse, as is sometimes the case in the mercantile 
 marine, it is like that of an oval boiler (fig. 204). The best height to look 
 well is 5 to 7 diameters above the taffrail, the latter when there are high 
 bridges or boats in wake of the funnel. For the same reason, the ring for 
 the shrouds should be -^ the diameter from the top. Funnels are made of 
 
672 MANUAL OF MARINE ENGINEERING. 
 
 ship quality plate, lap -jointed, or butt-jointed with single straps inside ; the 
 latter costs more, but when so made is more durable. Another method 
 much in fashion at one time, and which presents a good appearance, is to 
 make the longitudinal joints with inside butt-straps, and the circumferential 
 with a band of iron of a flattened U section (fig. 254). 
 
 The funnel plates should, for strength, be thicker at the base than at the 
 top, but the top plates wear out faster than those at the bottom. The 
 following may be taken as the approximate thickness of these plates : — 
 
 Top plates = 01 inch + 0*025 for each foot of diameter. 
 Middle „ = 0-125 „ +0-026 
 Bottom „ = 0-15 „ + 0-027 
 
 If the funnel is stiffened with angle or T bars (fig. 253), it may be made 
 of somewhat thinner plates. The funnels of naval ships are, of course, 
 made as light as possible, and the plates composing them are seldom more 
 than 5 ^y inch thick, and of steel. 
 
 Furnace Fronts and Doors. — Although often made of cast iron, they are 
 better when of wrought iron to withstand the rough use to which they are 
 exposed. It is from this cause that all the improved doors which have been 
 tried have been finally rejected ; and because of this rough usage all attempts 
 at refinement in the fittings, etc., meet with want of success on board ship. 
 The smaller the door the better, as, when open, an excess of air passes into 
 the furnace and lowers its efficiency ; on the other hand, it must be large 
 enough to stoke, work, and clean the fires properly. A long grate requires 
 a larger door than does a short one. Furnaces of large diameter — that is, 
 above 42 inches — should have a pair of doors to be used alternately. The 
 amount of opening when stoking or cleaning fires is thereby reduced, and 
 the sides of the grate are better attended to. 
 
 The doors should be so arranged as to remain open in a seaway when 
 required ; this may be effected by making a projection and corresponding 
 recess in the hinge. A star damper should be fitted to the fire door, by which, 
 when open, a supply of air is admitted to the fires. The baffle plate inside 
 the door should have a number of small perforations in this case, to finely 
 divide and to distribute the extra supply of air. 
 
 Fire-bars. — The length of grate should never exceed twice the diameter 
 of the furnace, as the fires cannot be properly worked when this is the case. 
 To get the highest efficiency of grate, it should not be more than H times 
 the diameter. The slope of the grate should be 1 inch to the foot, which 
 may be increased, in furnaces over 42 inches diameter, to 1| inch with ad- 
 vantage. If the grate is over 5 feet long, there is generally some difficulty 
 in properly stoking the back end. and it is only a good fireman" who can 
 properly work the fire on a long grate. The increased slope materially 
 helps to overcome this difficulty, and at the same time the fire is better 
 supplied with air at the back, and not choked by the products of combustion 
 from the front. 
 
 The bridge or brick barrier at the eud of the grate should be built to 
 such a height that the area of passage over it is, not less than | nor more 
 than I the grate area ; and, when possible, the distance from the top of the 
 bridge to the top of the furnace should be sufficient for a man to pass into 
 the combustion chamber. 
 
hendkrson's patent door and bars. G7;3 
 
 When South Wales or other similar coal is to be burnt on the grate. 
 there should be no space between the bars and the side of the furnace ; when 
 the furnaces are corrugated, these side bars should be made to fit into th- 
 corrugations. 
 
 The fire-bars are usually in two lengths ; but the grate is more efficient 
 when they are in one, as the bearer is avoided, which baffles the free flow 
 of air to the fire above it, and prevents the fireman from " pricking " effec- 
 tively. Cast-iron fire-bars to burn bituminous coal may be 5 feet 6 inches 
 long, but if the coal contains much sulphur they are safer in two lengths. 
 The Admiralty do not allow the bars to be longer than 27 inches, even when 
 made of wrought iron, as they generally burn Welsh coal, and may have 
 to use, on foreign stations, coal containing sulphur. 
 
 Fire-bars are made from | to 1J inch broad on the face ; the former 
 is better when the bars are not very long, and when of wrought iron may 
 with advantage be even f inch. In the . mercantile marine 1J inch is the 
 usual breadth when the bars are long. 
 
 The depth of the bar at the middle depends on the length, and should 
 be — 
 
 = 0-6 >/length, when of cast iron, 
 
 and =0*5 \Aength, when of wrought iron. 
 
 The thickness at the bottom should be one-third the breadth at the face, 
 and should taper to two-thirds beneath the flange. 
 
 For burning bituminous coal there should be a space of \ inch at least 
 between the bars, and, when it cakes quickly, there may be as much as 
 | inch ; but, if Welsh coal or American anthracite is to be burnt, there 
 should never be more than \ inch spaces, and with narrow bars the space 
 may with advantage be less. 
 
 Martin's Patent Bars consist of wrought-iron bars of square section placed 
 with the angles upward, and so arranged that the bars may be slightly turned 
 so as to clean the fires. 
 
 Henderson's Patent Door and Bars. — This is one of the most successful of 
 the improved grates. The bars are of cast-iron and ordinary section, hung 
 in a frame, which can be easily moved by means of levers, so as to slightly 
 move alternate bars longitudinally ; this movement, while carrying the fuel 
 gradually to the back of the furnace, breaks up the clinker, and obviates 
 the necessity of cleaning the fire. The door of the furnace is hinged hori- 
 zontally at the bottom, and so arranged as to drop into a recess in the dead 
 plate left for it, which recess is filled by a back piece attached to the door, 
 and turns down as the door opens ; if the fire needs cleaning, the door is 
 turned still farther, so as to leave the gap in the dead plate, and form a slope 
 below it to shoot the clinker and cinders into the ashpit instead of raking 
 them on to the stokehole floor. The furnace front, too, is carefully designed, 
 so as to pass a current of air completely about it between the front and back, 
 thus serving the double purpose of keeping the front comparatively cool, 
 and of heating the air before entering over the fire. This arrangement of 
 grate when properly worked permits of a rapid and complete combustion 
 of coal equivalent to a moderate forced draught, as may be seen by referring 
 to Tables lxxxvii. and lxxxviii. where grates are marked " H." 
 
 43 
 
674 MANUAL OF MARINE F.NG1NEF.RINC. 
 
 To Burn Crude Petroleum Residuals and other similar liquids, a special 
 burner is necessary, and on its efficiency depends the economy of using 
 such fuels. 
 
 The functions of such an apparatus are to direct the fluid in a steady 
 stream to the place where ignition commences ; to pulverise it, or " spray " 
 the stream so as to permit of quick and perfect ignition ; to project the 
 flame in the required direction, and to keep it constantly going with 
 the least possible expenditure of steam, air, or anything else which costs 
 money. 
 
 Mr. James Holden, formerly the locomotive-superintendent engineer 
 of the Great Eastern Railway, after experimenting for a considerable time 
 with the burners used on the Caspian Sea and in America, devised an instru- 
 ment of his own, which has been found to give the greatest satisfaction, both 
 on locomotives and on shipboard, by performing its duties efficiently and 
 without fail over considerable periods. Steam is employed to drive the 
 liquid through the nozzle (fig. 201), but, before doing so, it has induced a 
 flow of air in with it from the central tube, which air mingles with the pulver- 
 ised liquid and helps to burn it completely, as well as to furnish the oxygen 
 required for ignition. The ring around the nozzle is charged with steam, 
 and this forms a heater for the air current, which flows past it to the nozzle. 
 There are, however, other burners which do not require steam to work them 
 (vide Sir F. Flannery's paper in Trans. 1. Naval Architects, vol. xliv.). Fig. 
 201a is a good example of a simple burner worked by compressed air, and 
 found to be very efficient on shipboard. In this case the nozzle can be 
 cleared from obstruction by opening the central valve and admitting steam 
 or air. , 
 
 Stop-valve. — The main stop-valve should be of sufficient size to pass out 
 all the steam the boiler is capable of making with little resistance, but no 
 more than all. The chief loss of pressure at the cylinders is often due to 
 the stop-valves being only partially open ; on the other hand, when the 
 cylinders are large and the strokes of the piston comparatively slow, priming 
 may be effectively checked by partially closing the stop-valve, so that there 
 is no sudden withdrawal of steam, in fact, the valve should be set so as not to 
 pass out more steam than the boiler is making. 
 
 The area of opening, sufficient to pass all the steam a boiler can make 
 can be obtained from the following rule : — 
 
 . . .„ total heating surface X R 
 
 Area of orifice = _ ° _ 
 
 8 x P 
 
 R is the rate of evaporation per square foot of heating surface per hour, 
 P is the pressure absolute. 
 
 The area through a pipe to carry away the steam is two to three times 
 the net area, depending on the length (Tables xxxii. and cxi.). 
 
 The diameter of the stop-valve is often settled from considerations of 
 the size of the main steam pipe at the engines. Let D be the diameter of 
 the main steam pipe, and n the number of boilers (at least two), then 
 
 4 2D D 
 
 Diameter of branch pipe ) -p. /4 2D .,...„ 
 
 , , -, r l >=Da/ - , or —. — , or 1*156 X 
 
 to each boiler j V 3 n ' y^j n ' y/. 
 
 n 
 
STOP- VALVE. 675 
 
 The area of pipe section to suit a boiler may be found by the following 
 rule : — 
 
 (0-25 square inch per square foot of grate -f- 0*01 square \ / 100 
 
 inch per square foot of total heating surface) J X V pressure' 
 
 Exa tuple. — To find the diameter of steam pipe from a boiler whose guate 
 area is 50 square feet, and the total heating surface is 1,500 square feet ; 
 pressure, 80 lbs. 
 
 Area of section = {(0-25 x 50) + (0-01 x 1,500)} x a/~ 
 
 v 80 
 
 = 307 square inches. 
 Therefore the diameter should be 6^ inches. 
 
 Example. — To find the diameter of the main steam pipe of a locomotive 
 boiler whose grate area is 16 square feet, the total heating surface 1,200 
 square feet, and the pressure 150 lbs. 
 
 Area of section = {(0-25 x 16) + (0'01 x 1,200)} x a/-2J 
 
 = 13 - 09 square inches. 
 
 Therefore the diameter should be 4| inches. 
 
 The diameter of the stop-valve should be such that the clear area past it 
 is not less than given by the above rules. 
 
 m1 ,. . . ,, diameter of valve / , . . 
 
 Ihe diameter of spindle = ^ X Vpressure + i inch. 
 
 The valve and seat are of a bronze which should be both hard and strong 
 at the temperature of the steam passing it (Chapter xxx.). The valve should 
 have the boiler pressure always on the side opposite the spindle, so that it 
 helps to open it. The spindle should have a square thread, aftid, when 
 possible, the screwed part should be outside, so that in opening or shutting 
 the valve the spindle does not turja round. As the full pressure is on the 
 valve when shut, the bridge, etc., should be strong enough to withstand it. 
 The seat, when fitted with wings for the guide to the spindle, should be 
 carefully secured and the wings curved, so that, when expanding with the 
 heat, the seat is not distorted. These seats, when fitted into cast iron or 
 steel are very apt to get loose and leak from the permanent set of the metal, 
 induced by the resistance of the cast iron to expansion of the bronze. They 
 should, therefore, be secured by bolts or studs to the casting, and a joint 
 made which will keep tight in spite of expansion, etc., much in the same way 
 as the pipe joints are done. 
 
 The Admiralty require stop-valves and all other boiler mountings to be 
 made of bronze or steel ; for modern high pressures, and especially when the 
 steam is superheated, the material must be such as to retain strength and 
 elasticity with such high temperatures. The immadium bronze made by 
 Parsons M. Bronze Company is found very suitable, as at 550 u F. it has an 
 elastic limit of 8 to 9 tons. 
 
 The Admiralty likewise require all stop-valves to be self-acting — that is, 
 they shall close on the pressure in the boiler being decreased below that in 
 the main steam pipe. This is with the object of localising the danger and 
 
676 
 
 MANUAL OF MARINE ENGINEERING. 
 
 automatically disconnecting the boiler in case of accident to it from shot, 
 etc. Such valves have their spindles turned down at their outer ends, so 
 as to pass through a hole in the screwed shank, fitted to the cross-piece, as 
 shown in fig. 255, and with cross-handles to permit of the valve being turned 
 or closed as required. 
 
 Safety Valve. — As its name implies, this valve is for the purpose of pro- 
 dding a safe and self-acting means of relieving the boiler from excessive 
 pressure. A good safety valve should be (1) large enough in diameter, and 
 nave sufficient lift to allow the steam to escape as fast as it is generated, 
 when the pressure is 5 per cent, above that to which the valve is loaded ; 
 (2) it should be made so that it closes again as soon as the pressure has dropped 
 very slightly below the load ; (3) it should be free to open and shut, so that 
 it may always act efficiently and promptly ; (4) it should be-enclosed so that 
 it cannot be tampered with or accidentally interfered with by pieces of coal, 
 
 tiutwt ad »: ■». 
 
 fOOP^Q PlOf row ^TfrAM FSOM IWLtTStU- 
 , t •-,.«, VAIV£ '() fiON Ht'iiOV »'*»>• '« 
 
 Fig. 255. — Bulkhead Self-Closing Stop Valve (Cockburn's). 
 
 etc., failing into it ; and (5) for marine purposes it must be so constructed 
 as not to be affected by the motion of the ship. 
 
 It is unnecessary to deal with weight-loaded valves, as none are now used, 
 the fifth condition being satisfied by means of steel springs for the load 
 (fig. 256). When weights were used, the amount of lift given to the valve 
 by the steam pressure was very small, and. since the lifting of the valve 
 compresses the spring and increases the load, the spring-loaded valve opens 
 somewhat less. This being so, area of opening can only be obtained by in- 
 creasing the diameter of the valve. Many ingenious methods of increasing 
 the lift have been tried, but all those involving the use of special mechanism 
 have given place to those which do without it ; the most successful and best 
 known of the latter is Richardson's Patent (fig. 257), generally called Adams', 
 after the name of the manufacturer, who purchased, improved, and worked 
 the patent in this couutry, and Cockburn's Valve (fig. "258), designed and 
 
SAJftSTY VALVE. 
 
 677 
 
 made by Messrs. Cock burn, of Glasgow. These consist essentially of an 
 ordinary mushroom valve with a secondary outer rim of U section which 
 overlaps the rim of the seat, so that there 'is a second contracted orifice at 
 the outer edge of this rim. As soon as the valve opens, the steam fills the 
 
 Fig. 267.— Adams' 
 Valve. 
 
 Fig. 258.— Cockburn'a 
 Valve (Navy Type). 
 
 Fig. 256. 
 
 •outer rim, and the valve is then virtually of larger area ; the load on it is 
 so suddenly increased that the valve lifts wide open immediately, and will 
 continue to vibrate with the spring until the pressure falls so as to be insuffi- 
 cient to open the valve when the latter touches the seat in one of its vibrations 
 
678 
 
 MANUAL OF MARINE ENGINEERING. 
 
 and remains closed. The second condition is best fulfilled when the valve 
 can be made to dance over its seat from the vibration of the spring. The 
 third condition can only be fulfilled by making every movable part a very 
 •easy and, in most cases, a very slack fit, and having springs whose resistance 
 to compression is uniform on every side ; this is a quality not always found 
 in the springs supplied commercially, and adjusting a spring when badly 
 made is expensive and troublesome ; to avoid the evils due to this cause, 
 Messrs. Cockburn have devised a joint for the spindle which shall adjust 
 itself to suit the irregularity of pressure when the spring is compressed a3 
 
 Fig 259. 
 
 shown in fig. 259, which is the kind of safety-valve fitted on naval ships 
 where the head room is very limited. Fig. 260 is the design of safety-valve 
 now largely used in naval ships, and made by Cockburn & Co. 
 
 To prevent the valve from being injured by accident or design, it should 
 be enclosed in a case, and the Board of Trade require that such cases shall be 
 locked up, and the key kept by the captain of the 3hip. 
 
 The Size of Safety Valve. — This must suit the volume of steam which 
 can be generated by the boiler in a given time, and that depends on the 
 weight of fuel it can consume, on its efficiency, and on the working pressure. 
 In similar boilers — that is, boilers made on the same general design, and 
 
THK SIZE OF SAFETY VALVE. 
 
 079 
 
 worked with the same draught aud pressure — this volume of steam vaiies 
 with the grate area. The original rule laid down by the Board of Trade 
 for the area of safety valve is based on this, and it was found to work satis- 
 factorily ; since then, however, new rules have been laid down. Strictly 
 speaking, any rule for the safety valve should fix the amount of circumference 
 rather than area, and considerable allowance should be made for the load 
 pressure, as for the same weight of steam the volume varies inversely as the 
 pressure. 
 
 The following are the rules for the size of valve : — 
 
 (1) To satisfy the Board of Trade, the valves must not be in any case less 
 than 2 inches in diameter, and for each boiler with natural draught, the 
 area of valve or valves combined must not be less than given by this rule. 
 
 37 "5 
 
 * Area of safety valve in square inches = —^- X grate area (square feet), 
 
 Fig. 2G0.— Combined Stop and Safety Valves, Admiralty Type (Cock burn's). 
 
 P being the absolute boiler pressure. For forced draught the area found by 
 the above rule must be multiplied by ^, where C is the estimated coal con- 
 sumption in lbs. per square foot of grate per hour. 
 
 (2) Lloyd's retain the old Board of Trade Kule that the valve area must 
 be at the rate of \ square inch for each square foot of grate, but allow special 
 valves of any size, so long as they be satisfactory when tested. 
 
 (3) The French Government Rule is based on the amount of heating surface 
 contained in a boiler, and this, perhaps, is the truest gauge of a boilers 
 capability, as it bears a constant relation to the amount of coal consumed ; 
 allowance is also made for the steam pressure. The diameter is in inches, 
 
 ♦ To suit all conditions of draught and fuel it is better to take a hypothetical grate area of equal to 
 one twenty-eighth of the total heating surface. 
 
680 
 
 MANUAL OF MARINE ENGINEERING. 
 
 the heating surface in square feet, and the pressure in lbs. per square 
 inch. 
 
 Diameter of valve (if only one) = 1 23 » / - — - -. 
 
 J V pressure 4- 9 
 
 (4) The German Government Rule also makes allowance for the steam 
 pressure, and is as follows : — 
 
 To have a clear area of valve or valves, after deducting for the wings or 
 other obstacles, at the rate of so many square millimetres for each square 
 metre of total heating surface, in accordance with the following table : — 
 
 Working pressure in atmo- "1 
 spheres - - - - / 
 
 Number of square millimetres ) 
 per metre of surface - / 
 
 5 
 
 6 
 
 7 
 
 8 
 
 131 
 
 112 
 
 93 
 
 1 
 86 
 
 10 
 
 79 
 
 Working pressure in atmo- \ 
 spheres - - - - f 
 
 11 
 
 12 
 
 13 
 
 14 
 
 i5 
 
 16 
 
 Number of square millimetres \ 
 per metre of surface - / 
 
 66 
 
 60 
 
 56 
 
 54 
 
 52 
 
 51 
 
 (5) An improved rule, which is simple and easily used, is — area of each 
 of two valves = (0-05 square inch per square foot of grate + 0-005 square 
 
 inch per square foot of total heating surface) x */- 
 
 V v 
 
 100 
 
 pressure 
 
 Example. — To find, by the various rules, the size of a pair of safety valves 
 for a boiler whose grate area is 45 square feet, heating surface 1500 square 
 feet, and working pressure 180 lbs. 
 
 (1) By Board of Trade. 
 
 45 37*5 
 Area = — x -^ = 4-33 square inches. 
 
 Therefore the diameter is 2| inches of each valve. 
 
 (2) By French Government rule. 
 
 Diameter = 1*23 
 
 / 1500 
 Vl80 + 9 ~ 
 
 347 inches. 
 
 (3) By German Government rule, for one valve. 
 
 Clear area = 60 x 1500 -r 10-76 = 8365 square millimetres. 
 
 xt 1 8365 
 
 Net area = -^- = 13 square inches. 
 
 Add to this 2 square inches for obstruction of wings, 
 
 gross area =15 square inches, 
 and diameter for onu valve is 4-4 inches, for each of two valves 3-1 inches. 
 
INTERNAL PIPE?. 
 
 681 
 
 '100 
 
 1500) x a/ jg- = 7-3 square inches. 
 
 (4) By the improved rule. 
 
 Area = (0-05 x 45 + 0005 x 
 
 Therefore diameter is 3*1 inches. 
 
 The mitre on a safety valve-seafc should not be more than T \ inch broad 
 except for very large valves, and the bearing area in any case need not 
 exceed that necessary for a pressure of 1200 lbs. per square inch on it when 
 there is no steam pressure on the valve j hence 
 
 Breadth of mitre = diameter of valve x workin « prewnre 
 
 4800 
 
 The Board of Trade rule for the size of steel for the spring 
 
 is 
 
 <-/ 
 
 D 
 
 S is the total load on the valve ; D the diameter of coil,* measured from 
 centre to centre of wire, in inches ; d is 
 the diameter of round wire, and the side 
 of square section wire ; C is 8,000 when 
 the coil is made of round section steel, 
 and 11,000 when of square section (vide 
 Appendix 0). 
 
 Internal Pipes should be fitted from 
 the stop-valves to the highest part of 
 the boiler, and be made with holes or 
 slits, whose collective area is equal to 
 twice the area of section of the pipe. 
 
 The chief object of this pipe is to 
 collect the steam gently from every 
 part of the boiler, so as to avoid setting 
 up a strong current in one particular 
 direction, and thereby induce priming. 
 These pipes are usually made of brass, 
 but some engineers prefer copper, and 
 others made them of cast iron to avoid 
 the risk of galvanic action and reduce 
 the cost. Now that salt water is never 
 used for feed water, these pipes may be of 
 steel galvanised. By fitting an internal 
 pipe, the stop-valve can be placed in 
 a position convenient for examination 
 
 and working, and it should always be situated so as to be easy of access at 
 all timej. Arrangements should also be made for opening and shutting it 
 without going into a position of danger or difficulty, and this can always bo 
 effected by lengthening the spindles or fitting chain gear. The Admiralty and 
 some passenger-ship owners insist on having gear fitted so that the stop-valves 
 can be shut from on deck as well as in the stokeholes. 
 
 In the mercantile marine, the stop- and safety-valve boxes are almost 
 invariably made of cast iron or steel : the valves, seats, and spindles being 
 of bronze. The Admiralty require all boiler mountings to he made of bronze 
 or cast steel, and do not allow sast iron to be used. 
 
 Fig. 261. — Improved Feed Check Valve. 
 
682 MANUAL OF MARINE ENGINEERING. 
 
 Feed-valves. — Each boiler should be fitted with a self-acting, non-return 
 valve, through which the main feed-water is pumped. It should also have a 
 screw spindle, which may be used to regulate the lift, or to shut it down 
 when water is not required. There should also be a similar valve through 
 which the auxiliary pump can discharge water to the boiler. 
 
 The valve is generally of mushroom form, and made similar to the ordinary 
 stop-valve, except that it is detached from the spindle. It is made wholly of 
 bronze, and should be very strong, as at times the pressure on it may be 
 excessive. An inner casing should surround the valve (fig. 261), and be 
 formed or pierced in such a way that the flow is fairly even round the 
 periphery, otherwise the valve tends to cant and wear its seat unevenly, 
 with leakage as a consequence. Locomotive engineers sink the feed check 
 valve seat well below the discharge orifice for this purpose. 
 
 There should be 2 square inches of clear area through the valve and pipe 
 for every 100 lbs. of water evaporated per minute ; or, put in a more con- 
 venient form — 
 
 Area of pipe section = pounds of water evaporated per hour -=- 2.500, 
 
 or diameter feed pipe = ^pounds evaporated per hour -5- 45, 
 
 or area through main feed-valve in square inches 
 
 = total heating surface in square feet -f- 250 ; 
 
 and, area through donkey feed-valve in square inches 
 
 = total heating surface in square feet ■*- 300. 
 
 When feeding with an independent pump the pipes may be 30 per cent, 
 smaller, as the velocity of flow can be varied. 
 
 As the feed-valves cannot always be placed on that part of the boiler best 
 suited to receive the feed-water, and also in order to distribute that water so 
 as to avoid its affecting the boiler plates, an internal pipe should be always 
 fitted. To prevent the necessity of blowing the boiler down in case of accident 
 to the feed-valves, it is a very common practice to fit these valves high up on 
 the boiler, even in many cases above the water-level. This plan also has the 
 advantage of providing a means of warming the feed-water, than which 
 nothing is more essential for the preservation of the boiler ; the heating is 
 effected by the passage of the water through a long internal pipe of brass or 
 copper, which leads it to where there is a down current of water, so that the 
 comparatively cold feed-water may not interfere with the circulation. To 
 permit of the check valves being examined and ground in without emptying 
 the boilers, a stop valve is fitted, as shown in fig. 262, and proved to be very 
 useful, especially in ships making long voyages. 
 
 Some engineers prefer to inject the feed-water in the form of spray r 
 either above or a little way beneath the surface of the water in the boiler ; 
 this avoids all chance of injury to the boiler plates, as any gaseous matter 
 mechanically mixed with the feed-water is at once given up and mixes with 
 the steam. 
 
 Great care should be taken in any case that the internal feed-pipes " run 
 full " — that is, that they are never at any time filled or partly filled with 
 steam, but always with water. 
 
 The dynamic effect of the steam on the feed-water, when mixed inside 
 the pipe, is very startling ; every stroke of the feed-pump delivering cool 
 water produces a concussion bv the sudden condensation of the steam, so 
 
AUTOMATIC FEED VALVES. 
 
 683 
 
 that, in a very short time, both external and internal pipes are damaged 
 6eriously. 
 
 To avoid this, the internal pipe should, when discharging above the 
 water-level, be turned upivard at the end, so as to always remain filled with 
 water ; and when turned downward to discharge under water, the end should 
 be well below the lowest working level. 
 
 An additional means of safety is sometimes afforded by fitting inside the 
 boiler a clack valve, so arranged as to close over the end of the internal pipe 
 or on the spigot of the ordinary check valve ; when this is provided, the 
 latter can be examined when steam is up. A cock also used to be fitted 
 sometimes close to the check valve, so that the supply can be regulated by 
 it, instead of by interfering with the lift of the check valve. 
 
 H'ig. 262. — Feed Check Valve with Shut-off Valve Added. 
 
 Automatic Feed Valves are an advantage to all water-tube boilers, and 
 accessary for their good working when steaming under forced draught, inas- 
 much as the ratio of the water they contain to that evaporated per hour is 
 small ; consequent on this, there might be considerable difficulty in main- 
 taining the water at the right level by hand adjustment, as is effected with 
 tank boilers. It is also necessary for their good working and the production 
 of the full supply of drj steam that the level be as low. as possible consistent 
 with safety. Fig. 262a shows the method for accomplishing this desirable 
 end devised by Mr. A. G. Mumford, and the apparatus supplied by him, 
 and used largely in the British and foreign Navies. The means involved are 
 simple, effective, and free from liability to derangement with ordinary care 
 and attention ; moreover, if out of order, there is no difficulty in restoring 
 
684 
 
 MANUAL OF MA RINK ENG rNEEKINO, 
 
 it. The check valve on the boiler has attached to it a piston, which is a leaky 
 fit in a cylinder below the valve box. and the opening or shutting of the valve 
 depends on the difference in pressure below from that above the piston. 
 So lon.w as the water which leaks past the piston is free to run away, the 
 
 juojj -Jd/iog 
 
 J3/IOQ 03 13/UI - 
 
 pressure of the feed-water is unable to open the check valve. If the flow 
 from below the piston is checked, there is soon equilibrium, and the feed- 
 water forces open the check valve and supplies the boiler ; as soon as the 
 flow is resumed, the piston will be pressed down, and so close the check 
 
SCUM COCK. Gf5 
 
 valve and stop the feed. The water thus leaking away is caused to pass 
 through a small valve, which is operated by the rising and falling of a lloat 
 in a chamber connected with the boiler, so that on the water level falling 
 below normal the leak is stopped, and the check valve permitted to open 
 and remain so until the level is at or slightly above normal when the leak 
 is free and the check valve closed. 
 
 There are other methods of attaining the same end, such as that of Sir 
 John Thornycroft, where the admission valve is operated direct by a float 
 inside the boiler, so that the supply is shut off as soon as the water level 
 rises above normal ; and the ingenious arrangement of Mr. Yarrow, whereby 
 the feed-pump steam cylinder became flooded as soon as the level was too 
 high in consequence of its supply valve being placed at the proper level for 
 good working ; the water was exhausted from the feed donkey to the feed 
 tank, and as soon as the water level fell and steam was again admitted the 
 donkey increased to its regulation speed, and delivered its normal supply of 
 feed-water. 
 
 Blow-off Cock. — A cock used to be fitted at or near the bottom of the 
 boiler, to answer the double purpose of admitting sea-water before getting 
 up steam and to bloiv off brine and some of the water when required. This 
 cock should be a very strong one, as it is liable to rough usage, and, being out 
 of sight and not easily got at, it is very apt to be neglected. For this reason, 
 as well as because a large cock is difficult to open and shut, engineers now 
 prefer a valve to a cock. If, however, a cock is fitted, it should be arranged 
 so that its handle or spanner cannot be removed when it is open. Sea-water 
 is now seldom or never used for filling boilers working at high pressures, so 
 that a connection with the sea is not necessary ; when fitted, the clear area 
 through it was usually 1 square inch -J- (0-2 square inch for each ton of water 
 in the boiler), now it need not exceed 1*5 inches in diameter. 
 
 As it is a very reprehensible practice quickly to blow off a marine boiler 
 when at its normal working temperature, it is now never done; a cock is 
 fitted, however, to the bottom of the boiler, so that, when the pressure of 
 the steam is down, the water may be pumped overboard if required. 
 
 Scum Cock. — A cock, having a clear area through it of one-third that of 
 the blow-off cock, should be fitted to the boiler, near to the level of the 
 water, and to it is connected a perforated pipe, inside the boiler, not lower 
 than the lowest working level. The object of this pipe is to collect all 
 scum and floating impurities from the water and discharge them overboard. 
 Considerable quantities of grease and greasy matter were pumped with 
 the feed-water into the boiler, and had to be got rid of occasionally. 
 A particular kind of grease was sometimes formed in the condensers of 
 engines whose cylinders were lubricated with a certain class of oils which 
 are not pure hydrocarbons ; portions of it were pumped with the feed-water 
 into the boiler in the form of small pellets, which, being of superior specific 
 gravity to pure water, sank to the bottom, and used to remain there until 
 the density of water increased sufficiently to cause it to rise and come in 
 contact with the hot surface. It is, therefore, better to filter the feed-water 
 under all circumstances than to trust to a scum cock or to a blow-off cock, 
 and to use little or no internal lubricant; if a lubricant is used at all, 
 care should be taken that it is specially suited for the purpose and a pure 
 hydrocarbon (vide Chapter xxxi.). 
 
686 MANUAL OF MARINE ENGINEEKiNr?. 
 
 The scum cock was used as a means of reducing the quantity of water in 
 the boiler before adding a fresh supply from the sea ; but, if the surface 
 was clear of dirt, this was better done with the bottom blow-off, especially if 
 it was possible to check evaporation for a few minutes before blowing off. 
 
 Water Gauge. — It is of the first importance that those in charge of a 
 boiler shall know with certainty the position of the water-level within the 
 boiler. The ordinary gauge for this purpose consists essentially of a glass 
 tube, whose ends communicate freely with the inside of the boiler, and so 
 situated on the boiler that the plane of the water surface bisects the tube 
 transversely when at its normal working level. It is, however, found neces- 
 sary in practice to use considerable discretion in the choice of position of 
 this gauge. Since a difference of one-tenth of a pound pressure corresponds 
 to 2 "7 inches of water, it is quite possible so to place the gauge as to give 
 very false readings. The upper end of the gauge should not communicate 
 with the boiler near to any exit for steam, for the rush of steam past the 
 orifice can easily make a reduction in pressure of one-tenth of a pound in the 
 gauge pipe. The lower end should also be clear of any part from which 
 steam is evolved, as steam bubbles might rlow into the pipe, and tend to 
 raise the water-level in the glass. It should also be well away from the 
 currents due to circulation or other causes, and as near to " dead " water 
 as possible. 
 
 It is usual, especially with large boilers, to fit the gauge cocks and test 
 cocks to a brass casting connected by pipes to the bottom and top of the 
 boiler ; this is called a " stand pipe," and is a necessity when the gauge is on 
 the front of the ordinary boiler. The test cocks are placed on the stand pipe 
 at the lowest and highest working levels, for the purpose of checking the 
 glass gauge, and for use when the latter is broken or out of order. In these 
 days of high pressures, however, they are practically of no use, and might 
 well be done away with and their price go towards an additional glass gauge. 
 
 The length of the gauge glass visible should be at the rate of 11 inch for 
 each foot of diameter of the boiler ; the external diameter of the tube is 
 | inch for small boilers and § inch for large ones ; the glass is usually about 
 \ inch thick. The Admiralty use f -inch glasses for all sizes of boiler. 
 
 The gauge is so placed that the water is just disappearing, or, as it is 
 generally said to be just " in sight," when the level is from 2 to 4 inches 
 above the top of the combustion chamber ; the allowance, however, should 
 be 03 inch for each foot of diameter of boiler. 
 
 The pipes connecting the stand pipe to the boiler should be from 1 inch 
 to 1^ inch diameter, and of strong copper, so as to be fitted direct to the 
 boiler. The Board of Trade and some engineers insist on having a cock on 
 the boiler at the top and bottom; but this, like many another intended extra 
 safeguard, is itself a real source of danger, for the cocks are apt to be shut 
 by mistake or carelessness, and thus cause the gauge to show a false level. 
 That this is no mere fanciful danger has been proved on more than one occasion. 
 
 All large boilers., and those in ships which are often under sail, should 
 have two water gauges placed as far apart as possible in an athwartship 
 vertical plane. 
 
 The gauge and test cocks should always be fitted with a small plug in 
 line with the bore, which, on being removed, allows a wire to be introduced 
 to clean it of deposit and scale. 
 
WEfR S HYDROKINETER. 
 
 687 
 
 Some engineers have the stand pipes and connecting pipes so arranged 
 that there is a continuous flow horn top to bottom only via the glass tube, the 
 stand pipe being divided in the middle, or being so formed that the middle 
 is only a connecting bar for top and bottom to prevent the glass tube being 
 drawn out, This is very simple, and safer than the ordinary method. 
 • Steam Gauge.— The steam gauge on Bourdon's principle is now nearly 
 universal and so well known as to need no description. SchcefTer*s original 
 gauge, although less liable to derangement than Bourdon's, is not so accurate, 
 and does not find so much favour. The boiler gauge should have a dial so 
 marked that it may register pressures to at least 25 per cent, higher than the 
 working pressure of the boiler. These gauges should be carefully tested when 
 new, and at frequent intervals after being at work, as it is often found that 
 tkey require some slight adjustment. For this purpose all long-voyage ships 
 should have a standard gauge and test apparatus. 
 
 Sentinel Valve. — The Admiralty used to require each boiler to have a 
 
 Fig. 263.— Weir's Hydrokineter. 
 
 small valve loaded with a weight to a few pounds per square inch above the 
 working pressure, so that in case of the safety valves sticking fast and the 
 gauge being false, an alarm may be given when there is an excess of pressure. 
 Such valves are generally about f inch in diameter, but. sometimes as small 
 as | inch. An arrangement of a small safety valve attached to a whistle 
 has been introduced, but there should be no necessity for such refinements, 
 and it is doubtful if, in time of need, they would be heeded. 
 
 Air Valve. — The old box boilers had a small non-return valve to admit 
 air when the boiler was cooling down to prevent the boiler collapsing from 
 excess external pressure To-day such a valve or cock is desirable to admit 
 air before taking off manhole doors, and avoid accident to the operator. 
 
 Weir's Hydrokineter.— This instrument (fig. 263) is for the purpose of 
 warming the water in the bottom of the boiler when getting up steam. It 
 consists of a series of nozzles, one within the other, each having a grating- 
 
688 
 
 MANUAL OF MARINE ENGINEERING. 
 
 body in rear, through which the water passes on its road to the nozzle, when 
 a current is set up by a jet of steam issuing from the centre one. The steam 
 is obtained from the auxiliary boiler, which has been used to supply the 
 winches. Without this instrument the- bottom of a large boiler remains 
 cold, even after the steam is raised ; with it the temperature of the water at 
 the bottom differs very little from that at the top ; steam can in this way be 
 safely raised in a shorter time than usual, and at no extra cost, and the 
 
 Fig. 264. — Combined Syren and Double 
 Organ Whistle. 
 
 Fig. 264a. — Steven & Strainers' 
 Admiralty Syren. 
 
 endurance of the boiler is very considerably increased. There are many 
 other ways of promoting the circulation when steam is up, but none do this 
 so efficiently during the time of raising steam as the hydrokineter. 
 
 Steam Whistles are of two kinds, known as the bell-whistle and organ- 
 tube whistle ; the latter has now superseded the former, on account of its 
 simplicity of construction and superior tone. An improved form has a 
 division in the tube, or two separate tubes, so as to emit two distinct notes 
 
BOILER CLOTHING, 
 
 689 
 
 which may be in harmony or discord, and when sounded together are heard 
 a long distance. 
 
 It is important that the whistle shall sound as soon as the steam is turned 
 on ; to insure this happening, great care must be taken to keep the whistle- 
 pipe free of water, which is no very easy matter. It may, however, be 
 effected in two ways : first, by leading the pipe from the boiler into the 
 funnel, and keeping it inside as far as the level of the whistle ; second, by 
 taking steam for the steering engine from the top of the whistle-pipe, thereby 
 ensuring a constant flow of steam and no accumulation of water. 
 
 Separator. — This, although not a boiler fitting, is intimately connected 
 with them ; it is almost unknown in the mercantile marine, although it 
 might be used sometimes with advantage there in order to free the steam 
 from water mechanically mixed with it. All men-of-war were formerly fitted 
 with separators, and, from the tendency to prime on the part of their boilers 
 when working at full speed, and the danger to the engines when working at 
 a high velocity of piston from water getting into the cylinders, they were 
 necessary. Now, only ships with certain water-tube boilers have separators. 
 
 The old separator consisted of a vertical cylindrical chamber, having a 
 division-plate extending from the top to about half-way down, and so placed 
 that the steam in going through the separator must pass under this dia- 
 phragm ; the object is to separate out the water mechanically mixed with 
 the steam, by dashing it against the diaphragm, and precipitating it to the 
 bottom of the separator, whence it is blown to the hot-well or sea, whichever 
 is convenient. Most modern separators are constructed on the centrifugal 
 principle, so that the water by its great density follows one course, while the 
 steam follows another. 
 
 Boiler Clothing. — The boiler shell should be well covered with a coating 
 
 of non-conducting material, to prevent loss by radiation from its surface, 
 
 which may amount in some cases to more than 10 per cent. The material 
 
 used should be a non-conductor of heat as well as incombustible and inorganic. 
 
 The following materials* are those in general use for boiler clothing :— 
 
 * Some experiments on the efficiency of different heat insulators are described by 
 Mr. S. H. Davies in a recent issue of the Journal of the Society of Chemical Industry. 
 The following table shows the heat transmitted through 1 foot thickness of the different 
 materials for each 1° F. of temperature between the hot and cold surfaces : — 
 
 
 B.T.U. per 
 
 Weight of 
 
 Cost per 
 Cubic Foot. 
 
 Air-Dried Material. 
 
 Square Foot 
 
 Medium per 
 
 
 per Uour. 
 
 Cubic Foot. 
 
 
 
 
 lbs. 
 
 d. 
 
 1. Slag wool (light), 
 
 054 
 
 8-6 
 
 8-3 
 
 2. Hair felt, .... 
 
 
 
 058 
 
 7 9 
 
 20 6 
 
 3. Light magnesia, . 
 
 
 
 062 
 
 101 
 
 26 
 
 4. Granulated cork, . . . 
 
 
 
 0-069 
 
 61 
 
 98 
 
 5. Slag wool (heavy), . 
 
 
 
 0070 
 
 32 9 
 
 31-8 
 
 6. Kieselguhr, .... 
 
 
 
 073 
 
 150 
 
 96 
 
 7. Flaky charcoal, . 
 
 
 
 082 
 
 14 5 
 
 7 '8 
 
 8. Pumice (J-inch mean diameter), 
 
 
 
 095 
 
 25-0 
 
 241 
 
 9. Sawdust (spruce), . . . 
 
 
 
 096 
 
 131 
 
 05 
 
 10. Asbestos fibre, .... 
 
 
 
 0-136 
 
 14-5 
 
 137 
 
 11. Sawdust (very moist), . 
 
 
 
 293 
 
 ... 
 
 ■ 
 
 Mr. Davies notes that powdered pumice is much inferior to pumice in small lumps. 
 
 44 
 
690 MANUAL OF MARINE ENGINEERING. 
 
 Silicate Cotton or Slag Wool, manufactured from slag and having the 
 appearance of cotton, is eminently fitted for boiler clothing. It is a good 
 non-conductor, incombustible, and imperishable from chemical action ; it is, 
 however, very brittle, and for this reason will not withstand mechanical action ; 
 and, therefore, if loosely packed and subject to vibration, it soon becomes 
 dust, which is most offensive if it gets into the engine bearings. 
 
 Asbestos Fibre has very much the same nature as silicate cotton, it is 
 not so efficient, is more costly, but is more durable. A paste made with 
 this material and magnesia is very efficient. 
 
 Magnesia, either alone as cement, or mixed with asbestos fibre into a 
 paste, is a very excellent covering, and, moreover, is very light in weight, so 
 that an extra thickness of it weighs no more than does an ordinary coat of 
 other compositions of a similar nature. 
 
 Cements of various kinds are used, their efficiency depending generally on 
 the amount of vegetable fibre contained in them. These have not so high an 
 efficiency as the foregoing. 
 
 Fossil Meal, Kieselguhr, or infusorial earth, is a composition consisting ol 
 large quantities of minute shells, and forms a most efficient covering ; 
 besides being inexpensive, it -is also incombustible and durable, and com- 
 paratively light. 
 
 Papier Mache is employed for this purpose, and is a fairly good material, 
 but it is not altogether incombustible and is apt to rot, besides being heavy. 
 
 The last three materials must be put on when the boiler is hot. and be 
 carefully done ; this does not militate in their favour, as it is an objectionable 
 thing to have steam on the boilers when the ship is being finished. 
 
 The silicate cotton and asbestos fibre may be covered with sheet iron. 
 The cements are generally tarred over so as to be waterproof, and the parts 
 exposed to wear covered with sheet iron or lead. 
 
 No wood should be used for clothing when it is possible to avoid it, as it 
 so soon rots, and is at all times liable to take fire. 
 
 Cameron's Patent Lagging is an ingenious arrangement of wrought-iron 
 framework, strapped to the boiler, so as to form a series of segments, which 
 may be filled with any non-conducting substance, and is covered in with 
 squares of corrugated sheet iron, secured in such a way as to be easily and 
 quickly removed for examination or repair. This is the most perfect plan, 
 as it avoids all piercing of the boiler shell with studs and the use of com- 
 bustible materials, and while capable of being well secured, it is such as 
 to be wholly removed and replaced in a very short time. 
 
 A much thicker coating of lagging % is necessary with the high tempera- 
 tures now common, and no boiler, pipe, or cylinder can be considered properly 
 clothed if the temperature at the surface* is more than 30° F. above that of 
 the atmosphere surrounding it. The bottoms of boilers should be as well 
 covered as the tops ; in fact, no heat should be allowed to escape to the air 
 from any part of the boiler that can be kept in. Even the uptakes and 
 smoke boxes may, with advantage, be coated. 
 
ENGINE SEATINGS. 691 
 
 CHAPTER XXVI. 
 
 FITTING IN OF MACHINERY, STARTING AND REVERSING OV ENGINES, ETC. 
 
 Fitting Machinery into the Ship. — As soon as the building of the ship is 
 sufficiently advanced to allow the engineers to commence their work, a line 
 should be stretched in the place intended for the axis of the shafting, and 
 from it reference lines must be scored on the bulkheads, sternposts, and 
 other convenient places for future guidance, and to enable the shipbuilders 
 to set the engine seating and tunnel pedestals with some degree of accuracy. 
 
 For this purpose piano wire answers best, as it can be drawn exceedingly 
 tight without breaking, and the amount of " sag " is very slight, and does 
 not vary. When wire 20 L.S.G. is used, and the tension on it is as much 
 as it will bear with safety (about 200 lbs. is sufficient), the " sag " will not 
 exceed 0-15 inch per 100 feet. At the sternpost and bulkhead holes, cross 
 pieces of wood should be fixed and the centre transferred to them ; with 
 these centres circles should be " scribed " in and marked with a centre punch, 
 when they serve as guides in boring out for the stern tube. It is sometimes 
 found advantageous to verify the centre line markings by the system called 
 "sighting"; this is done by placing battens horizontally at convenient 
 places, whose upper edges touch the centre line of shafting, and, when viewed, 
 should all coincide if they are exactly in line. A similar system of battens 
 should then be placed vertically with the same result, if the line is straight. 
 This plan, however, is a somewhat tedious one, and by no means reliable 
 in most instances owing to the deceptive nature of the light in the hold ol 
 a ship. In long ships a surveyor's telescope may be used to check the 
 centres. 
 
 Boring the Sternpost. — A boring bar with tool head, etc., is fixed 
 accurately in position by means of the reference circles before-mentioned, 
 and the sternpost, which has been roughly bored to within about 5 per cent, 
 of the finished size before being fixed in place, is bored out to the exact size 
 required ; the bulkhead, with its liner, is also bored out, and also any other 
 part into which the stern tube is required to fit accurately. The finishing 
 cut through the sternpost should be commenced from the inside, as the wearing 
 away of the cutting edge of the tool causes the hole to be slightly taper, and 
 this allows the tube to be made a very tight fit in it. 
 
 Engine Seatings. — The superstructure raised on the ship's frames to carry 
 the engines is called by various names, such as engine bed, engine seatings, 
 engine foundation, engine bearers, etc., and is one requiring some skill to 
 design and care to manufacture properly. As the success and efficient 
 working of an engine very materially depend on this structure, too much 
 care cannot be devoted to its construction.* The weight of a marine engine 
 is considerable and concentrated on a comparatively small area ; in a seaway 
 
 * Recent investigations of the source of trouble experienced in sonic: naval ships in America with 
 shafting, etc., have confirmed this view. 
 
692 MANUAL OF MARINE ENGINEERING. 
 
 the inertia forces cause veiy severe stresses on the seating and on the bolts 
 connecting the engine to it ; and, in some cases, the strain of the engine itself 
 when at work is borne largely by the bed on which it rests, owing to the 
 want of rigidity in its own bedplate. 
 
 The ship cannot always be viewed as a rigid structure, for elasticity is 
 observable in all ships, especially when unloaded, and is very marked in 
 those built of considerable length for shallow water navigation. For this 
 reason, the engine seating must be so designed as to add materially to the 
 stiffness of the ship's structure, and be of sufficient strength to distribute 
 any stresses caused by the wf j.ght of the engine over the main framework of 
 the ship. To this end, the vertical portions of the engine seating should be 
 worked in or, better still, be in one with the floor plates and keelsons, and 
 the longitudinals should extend beyond the immediate vicinity of the engine 
 bed. so as to distribute the load over a longer portion of the ship, and not 
 localise it on a few frames. The longitudinal strength thus added to the 
 ship's bottom should not cease abruptly at the bulkhead, as is commonly 
 the practice, but be continued beyond it and decreased gradually. The 
 effect of stopping the engine seating at the aft bulkhead of the engine-room, 
 is to cause a sudden change of flexure in the ship's bottom at that point, 
 when the ship is steaming in a heavy sea ; this change of flexure will produce 
 abnormal stresses on the shafting, especially on the after part of the crank- 
 shaft ; the after bearing of crank-shafts shows this by its tendency to heat, 
 and, in extreme cases, the shaft is broken at that crank-arm, or at its junction 
 with the crank-arm. When the crank-shaft was connected to the thrust 
 shaft by drivers, the working of the ship was manifested by the squeaking 
 sound emitted by them when not oiled. 
 
 The scantlings of the engine seating should never be less than those of 
 the ship's floor plates to which it is fixed, and when the seating is high, they 
 should be in excess of the ship's scantlings. The angle-irons should be 
 carefully fitted, and the riveting more than usually good ; the rivet holes 
 should be fair^ and well filled with the rivet ; and if the holes are not fair, 
 simply drifting them out to allow the rivet to pass through is not sufficient ; 
 these remarks particularly apply to the connection of the top plate with 
 the verticals. The top plate should be at least 50 per cent, thicker than the 
 vertical plates, and well bedded in place. Fifty per cent, of the rivet holes 
 should be drilled through and through, and the remainder rimered fair and 
 true to ensure a thorough trustworthy foundation for a modern engine. 
 
 It was the practice with some engineers at one time to hold the engines 
 down to the ship by a few large bolts which were connected to strong cross- 
 bars under the reverse frames of the ship; this is, however, not a good 
 practice, as the load is localised to an unnecessary degree, and such bolts 
 are very apt to corrode from the action of bilge-water, and, being unseen, 
 to break eventually without being discovered. 
 
 The engine seatings are peculiarly liable to decay from the action of 
 bilge-water and its gases acting on the warm metal ; to prevent this they 
 should be carefully protected by cement where practicable, and well painted 
 where cement cannot be got to stick ; cement wash is. however, better than 
 paint, and if mixed hot. and brushed on, will form a very efficient covering. 
 
 Thrust Block Seating. — This also, from its importance and the nature 
 and the magnitude of the strains on it, must be carefully constructed. There 
 
BOILER SEATINGS OR BEARER. 693 
 
 should be three vertical plates, extending over, at least, four frames in small 
 ships, and six to eight frames in large ones ; the centre plate should be above, 
 and strongly secured to the keelson by angle-irons, its thickness should be 
 50 per cent, more than that of the floor plates ; the side plates should be 25 per 
 cent, thicker than the floor plates, and connected to the reverse frames by 
 strong angle-irons ; all three verticals should be caused, if possible, to abut 
 on the engine seating and tied to it. The top plate should be of the same 
 thickness as that of the engine seating, and when possible in line with it. 
 Stop plates should be riveted to the top plate to serve for the thrust block 
 to abut on. All the rivet holes in this top plate should be drilled quite fair 
 with the holes in the angle-irons and connections, and care should be taken 
 that all the holes in the thrust seating should be rimered fair and all rivets 
 quite fill the holes. In ships of very large power, the base of the thrust 
 block seating should extend over more frames, and the vertical plates should 
 be worked intercostal with the floors, so as to form a direct tie to the ship's 
 bottom plating. 
 
 Pedestals for Tunnel Shafting.— The plunimer blocks for the tunnel shafts 
 rest on the tunnel bottom when the shafting is not high, but when the distance 
 is too great for this, pedestals are built of plates, whose thickness is about 
 the same as that of the floors, connected by angle-iron, so as to form a stiff 
 column ; the top plate should be 50 per cent, thicker. 
 
 Boiler Seatings or Bearers. — The boiler, with its fittings and mountings, 
 together with the water it contains, requires a very strong support and 
 efficient means of keeping it in place when the ship is rolling or pitching. 
 When the boilers are placed athwartships (that is, their axes are athwart- 
 ship), the bearers act as beams, to distribute their weight over a large number 
 of frames, and may be made of H section, so that the lower flange is riveted to 
 the reverse frames, and the top flange carries the chocks, which are wedge- 
 shaped, and shaped to fit under the boilers, and form saddles for them to sit 
 in. Single-ended boilers should have two such saddles, whose breadth of 
 face where the boiler rests should not be less than 9 inches for very large 
 boilers, and 6 inches for small ones. Double-ended boilers of moderate 
 length and size may have three such saddles, but long or large double-ended 
 boilers should have four sets. The chocks are sometimes made of angle-iron 
 and plates, but they are better made of cast iron or cast steel ; when of 
 the latter materials, the patterns can be tried in place after the boilers are 
 in position, and made in such a way that the chocks, when cast, will fit with 
 sufficient accuracy as to require no packings ; when made of wrought iron, 
 much expense is incurred in trying to make them fit, and in the end packing 
 is often necessitated. 
 
 If the boilers are placed " fore and aft " — that is, with their axes longi- 
 tudinally — the bearers are generally laid on the top of individual floors, thus 
 localising the weight on a few frames only. To avoid the straining action 
 proving dangerous, the longitudinals ot the ship in wake of the boilers should 
 be increased, and extra connections* made between them and those frames 
 carrying the bearers. Sometimes the boilers when in this position have 
 been carried by longitudinal bearers inclined so as to be in planes passing 
 through the axis of the boiler. Such bearers distribute the load over a 
 considerable number of frames, but do not so well support the boiler, and 
 moreover prevent access to the boiler bottom for examination and repair. 
 
694 MANUAL OF MARINE ENGINEERING. 
 
 Boilers are now, except in shallow ships, placed well above the bottom 
 plating and bilges, so that there is room for cleaning, examining, and painting, 
 and permits easy access to the double bottom. There is also much less 
 liability to corrosion from bilge water. 
 
 Fitting Machinery on Board the Ship. — The stern tube and the screw 
 shaft are fitted into place, and all sea-cocks and valves fixed to the skin of 
 the ship before it is launched. After it is in the water, the tunnel shafting 
 is placed in position piece by piece, each one being set so that its coupling 
 comes fair and true with that of the preceding one ; the shaft bearings are 
 raised on temporary packings until the whole of the shafting is in place and 
 coupled up ; when this is done, the shafting should be turned around so 
 that the bearings, if not placed in exact position at first, may adjust them- 
 selves ; the bearings should now have the proper packings fitted to them, 
 and when bolted down, the coupling bolts should be withdrawn, and each 
 shaft tried around to see that there is no want of correspondence at the 
 couplings. This may seem a somewhat tedious process, but it is a very safe 
 one, and one which prevents all possibility of shafting being fitted out of 
 line ; the engineers of the ship should occasionally withdraw the bolts, and 
 prove the shafting true, especially after the ship has had cause for straining. 
 Of course, the accuracy of this method depends on the care with which the 
 couplings have been turned ; but, as modern appliances are capable of 
 turning a shaft flange quite true with very ordinary care, there is little 
 cause for fear on this ground. 
 
 The engine bed-plate, or foundation plate, with the crank-shaft in place, 
 is now lowered on board, and placed on temporary packings of iron ; it is, 
 by means of jacks, brought to its exact position, and proved by the shaft 
 couplings as before. Permanent packings of cast iron are now carefully 
 fitted between the temporary ones and the latter withdrawn, and their 
 places filled with hardwood (teak, greenheart, elm, mahogany, etc.) packings 
 formed of pairs of wedge pieces driven from opposite sides. 
 
 The Holding-down Bolts should be carefully fitted so as to distribute the 
 load over the bed, and means provided to prevent the nuts from slacking 
 back ; the Admiralty require check nuts to be fitted to all holding-down 
 bolts, but it is better to prevent movement of the nut by slightly riveting 
 the bolt end, than to trust to check nuts. 
 
 Staying Engines. — Vertical engines, especially those of long stroke, are 
 supposed to require some means of support to prevent undue straining of 
 the columns when the ship is rolling heavily. Such engines, when supported 
 by vertical ivrought-iron columns, do occasionally show signs of such a need ; 
 but it is better to bear this contingency in mind when designing the columns, 
 and make them sufficiently strong and rigid to withstand such strains, than 
 to trust to getting support from the elastic hull of the ship. In case of a 
 collision, such supports, if rigid, might be a positive source of danger, as the 
 force of the blow delivered in their neighbourhood might seriously damage 
 the engines, if not destroy them altogether. 
 
 Such support as is sufficient can be provided by splaying out the front 
 columns, as shown in fig. 32, and making them of strong cast iron or cast 
 steel. 
 
 Ramming Chocks and Stays. — The Admiralty require all vertical engines 
 to ))o fitted with non-rigid stays from the cylinders to some part of the ship 
 
COPPER PIPES. H95 
 
 in such a way as to keep the columns, etc., from being damaged or strained 
 when ramming the enemy. Boilers also must have stays and chocks to keep 
 them from being jerked out of place from ramming, or other inertia effect. 
 As a matter of fact, ramming is no longer looked on as a possible nautical 
 form of attack, but ships do collide accidentally, and sometimes take the 
 ground suddenly. 
 
 If W is the weight in tons of a boiler, its water, fittings, etc., and K is 
 the full speed of the ship in knots. Then — 
 
 Area of section of steel to resist fore and aft movement should not be 
 
 less than — ^ — square inches. 
 
 Boiler Seats. — The boilers require to be carefully fitted in their seats, and 
 secured there so that they cannot be displaced by the rolling or colliding of 
 ■the ship. The boiler should require no loose packings when the chocks are 
 of cast iron ; but when the saddles are of wrought iron, and do not fit the 
 boiler exactly, it is better to interpose iron packings at intervals, and fill in the 
 spaces between with hardwood wedge pieces. To prevent movement longi- 
 tudinally, " toe " plates should be riveted to the frames or other convenient 
 part of the ship's structure, and these should be stiffened by angle-irons. 
 The " toe " plates should stand about 6 to 9 inches above the bottom of the 
 boiler, and be clear of the man-holes and mountings. The boiler is held 
 sometimes in its seating by straps surrounding it, and secured to the bearers 
 when of small size ; but the general practice is to secure it by tie-bars from 
 its upper part to the side of the ship, or by struts formed of plates and angles 
 from the stringers, bearers, etc. Each particular case requires special treat- 
 ment, and it is impossible to lay down any rule, beyond that of providing 
 for every contingency to which a ship is liable. Tie-bars from lugs riveted 
 to the boiler sides to the bearers is a good and perhaps the best way to secure 
 them, and is the Admiralty practice. 
 
 At one time the Admiralty practice was to lay the boiler in a bed of 
 mastic cement, spread on a cradle formed to suit the boiler bottom. This 
 was very necessary for the box form of boiler, especially in wooden ships. 
 This cement also practically insulated the boiler, a practice which may be 
 followed now with advantage to both ship and boiler. A strip of asbestos, 
 | to h inch thick, interposed between the boiler and its saddle, is a good 
 arrangement for preventing thermo-electric action on the ship's frames, 
 tank top, etc., as well as for retaining much heat in the boiler. The boiler 
 bottom in steel ships should be well above the bilges, and room provided for 
 a man to get in to paint or repair ; the boiler bottom and ship's framework 
 can then, and should, be kept well painted. The boiler bottom should 
 also have a good coat of non-conduction material applied to it in such a 
 manner as to prevent radiation, and at the same time such as to permit of 
 easy removal and replacement for examination and repairs to the shaft. 
 
 Copper Pipes. — The whole of the pipes subject to internal pressure should 
 be of steel, wrought iron, or strong copper ; the exhaust pipes may be, and 
 usually are in the mercantile marine, of cast iron. The Admiralty require 
 that all pipes of steel or copper up to 6 inches diameter shall be solid drawn. 
 Wrought-iron and steel pipes are now very generally employed, especially 
 for large sizes and high pressure. If of steel and welded the Board of Trade 
 
696 MANUAL OF MARINE ENGINEERING. 
 
 used to require a strap riveted over the weld. Solid-drawn pipes may be 
 obtained up to 24 inches diameter in steel of any reasonable thickness* 
 
 The thickness of the main steam pipe should when of solid-drawn steel 
 
 = 0-125 + (diameter of bore X pressure -f- 15,000). 
 The thickness of feed pipes 
 
 = 0-125 4- (diameter of bore X pressure -f- 8,000) for copper ; 
 10,000 for steel. 
 The thickness of blow-off and scum pipes 
 
 — 0-125 -r (diameter of bore x pressure -f 9,000). 
 
 When made of steel or iron welded the thickness should be 20 per cent, 
 more than given by the above. 
 
 The thickness of copper or brass main inlet pipes 
 = 0-1 -f (diameter -- 300). 
 
 The thickness of copper or brass main discharge pipes from reciprocating 
 pump = 0-1 -f- (diameter -f- 200). 
 
 If for a centrifugal pump the discharge may be of the same thickness as 
 the inlet pipes. 
 
 The thickness of copper feed suction pipes and bilge-discharge pipes 
 = 0-09 + (diameter -h 200). 
 
 The thickness of copper waste steam pipes 
 = 0-05 + (diameter -=- 500). 
 
 The flanges for copper or brass pipes should always be of tough brass, 
 and of a thickness equal to 4 times that of the pipe ; the breadth of flange 
 should be 2^ times the diameter of the bolts used. For pipes exposed to a 
 pressure of 30 lbs. and upwards, the pitch of the bolts should not exceed 
 5 times their diameter, or 5 times the thickness of flange ; their diameter is 
 usually about the same as the thickness of flange. For pipes not subject to 
 a pressure of more than 30 lbs., the bolts may be 6 diameters apart, or even 
 a little more in some cases. All pipe flanges should be machined ■■true," 
 especially those of the feed and steam pipes. These should fit face to face 
 without forcing, and be jointed with as little putty or paint as possible. 
 Brass flanges on the copper pipes may with advantage have grooves into 
 which is fitted a ring of copper wire soft enough to easily conform to the 
 grooves when pressed by the bolts. This can also be done with iron flanges, 
 but the Admiralty complain that severe corrosion has taken place with 
 them so done. Corrugated rings of soft copper or brass have also proved 
 good jointers when fitted inside the line of bolt holes. The Admiralty require 
 the whole range of piping to be tested by water and then by steam. The 
 jointing that does best for water is not generally so good for resisting steam 
 and heat. The bolting should be close, rather than wider apart with larger bolts. 
 
 Starting and Reversing Engines. — It is most important that an engine 
 shall be capable of having its motion instantly reversed ; in fact, no engine 
 is satisfactory which cannot be stopped and made to move full speed astern 
 in less than 30 seconds from the time of the engineer commencing the evolu- 
 tion. This can scarcely be done by the hand-gear with even comparatively 
 
 • Tlie Chesterfield Tube Company can supply solid-drawn steel tubes 24 inches diameter and up to 
 20 feet loDg. 
 
STEAM CEAR FOR REVERSING* 697 
 
 small engines, and is beyond possibility with large ones. It is essential, 
 therefore, to provide mechanical means for this purpose in all engines of 
 over 100 N.H.P.. if such efficiency is required ; and it should be of such 
 power as to perform the operation without shutting off steam. The simplest 
 method is to fit a steam cylinder with rods, etc., to a lever on the weigh- 
 shaft in addition to the ordinary hand-gear, the piston pushing and pulling 
 as required to help the engineer. There is, however, the objection to this, 
 that the cylinder is then somewhat large, and at times the steam-power 
 masters the hand-power, and overruns its limit, thereby tending to cause 
 damage. This latter objection, however, is easily got over in many ways, 
 the best of which is by adding a second small cylinder containing water or 
 oil, which is forced by its piston from one end to the other, and thereby 
 acts as a brake to the gear. 
 
 Brown's Patent Reversing Gear. — This idea of the brake cylinder has 
 been worked out, and perfected by Messrs. Brown, of Edinburgh, who make 
 a gear which, by the action of a small lever easily moved by one hand, 
 operates on the valve motion instantly, and only to the exact extent intended 
 by the operator, so that if the engineer moves the lever through one-quarter 
 of its angular movement, the link-motion is moved by the gear through 
 exactly one-quarter of its traverse. This is effected by means of a system 
 of compensating levers, so arranged that the gear, by moving, replaces the 
 valves in the exact position from which they were displaced hy the hand- 
 lever. Messrs. Brown have also devised an arrangement whereby the hunting 
 and compensating levers with their rods are on the engine and so the apparatus 
 is self-contained. Figs. 265, 265a, and 2656 show the methods adopted by 
 the firm to make the gear automatic. 
 
 Steam Gear for Reversing. — The simplest and most efficient of the steam 
 gears, and one whose cost is so small that it may be fitted to the cheapest 
 of engines, consists of a small engine, on whose crank-shaft is a worm, which 
 works in a worm-wheel capable of turning freely on a fixed gudgeon on the 
 engine frame or other convenient place ; on this worm-wheel is a stud or 
 crank-pin. to which is fitted a rod connecting it to a lever on the weigh- 
 shaft ; the eccentricity of the crank-pin is equal to half the chord of the arc 
 through which the lever end works. The engine being set in motion, the 
 worm-wheel is caused to revolve, and the motion of the crank-pin causes 
 the weight-shaft to oscillate, and so to reverse the links. The steam 
 cylinder is sometimes fitted with reversing gear, but it is quite un-. 
 necessary, and is really better without it, as the little engine moves so 
 fast that practically no time is lost in making a complete revolution of the 
 worm-wheel. 
 
 The advantage of this gear over others, besides its cheapness, is its simpli- 
 city, safety, and capability of being used to turn the engines when in port, 
 or to work a winch for lifting weights when overhauling the engines ; the hand- 
 wheel is also in this case on the little engine shaft, and acts as a flywheel. 
 The Admiralty engines have a similar starting-gear, but fitted with two 
 cylinders whose cranks are at right angles ; the starting hand-wheel is only 
 connected to the crank-shaft or gear when required to be worked by hand ; 
 this permits of the engine starting from any position without help and stopping 
 at any point without a brake, as is often necessary in the mercantile marine 
 (fig. 32). 
 
698 
 
 MANUAL OF MARINE ENGINEERING. 
 
 oa 
 
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 ee 
 IS 
 
 60 
 
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 c3 
 
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REVERSING OF PROPELLERS DRIVEN BY TURBINES. 
 
 wy 
 
 TABLE LXXXII.— Particulars of All-round Reversing Gears, 
 
 Steam and Hand-moved. 
 
 Ship. 
 
 Cylinders and Steam 
 
 of 
 
 Main Engines. 
 
 H.M.S. M., 
 H.M.S. N., 
 
 s.s. c, . 
 
 H.M.S. E., 
 H.M.S. A., 
 H.M.S. P., 
 H.M.S. B., 
 
 H.M.S. S., 
 H.M.S. P., 
 S.S. 0., 
 S.S. A., 
 S.S. E., 
 S.S. L., 
 S.S. R., 
 
 s.s. v., 
 
 S.S. D., 
 
 Inches. 
 40-58-88 
 
 — tr~ 
 
 36-57-78 
 VZT 
 
 25^-39-60 
 39 
 
 40-59-88 
 48 
 
 33f49-74 
 
 39 
 30f45-68 
 
 33 
 21-31-45 
 
 24 
 19A-28i-43 
 
 18 
 20^-33-54 
 
 27 
 22|-32-45-64 
 
 42 
 
 32-48-80 
 
 48 
 28-43-70 
 
 39 
 24-39-64 
 
 33 
 
 22-33-58 
 
 36 
 23^-85-57 
 
 33 
 
 18-27-48 
 
 27 
 
 Lbs. 
 130 
 
 130 
 
 150 
 
 155 
 
 155 
 
 155 
 
 155 
 
 180 
 
 250 
 
 250 
 
 160 
 
 155 
 
 160 
 
 160 
 
 150 
 
 150 
 
 Reversing Engine 
 Cylinders. 
 
 No. 
 
 1 
 2 
 2 
 2 
 2 
 2 
 2 
 1 
 2 
 
 Dia- 
 meter. 
 
 Ins 
 9 
 
 s 
 
 5 
 
 4A 
 
 4 
 4 
 4 
 
 6 
 
 7 
 
 H 
 
 6 
 
 6* 
 5 
 
 4 
 
 Stroke. 
 
 Ins. 
 8 
 
 5 
 
 H 
 
 5 
 
 5 
 4 
 3 
 5 
 6 
 8 
 8 
 6 
 6 
 6 
 5 
 
 Dia- 
 meter 
 
 of 
 Worm 
 
 Ins. 
 
 7-0 
 
 H 
 
 5i 
 
 °t 
 
 4^ 
 3 
 
 H 
 H 
 
 5} 
 
 5 
 5 
 5 
 4 
 
 Worm-Wheel. 
 
 Dia- 
 meter. 
 
 Ius. 
 23 
 
 23 
 
 23 
 
 26| 
 
 23 
 
 23 
 
 16* 
 
 n 
 
 14$ 
 
 181 
 
 23 
 
 23 
 
 18 
 
 18 
 
 14} 
 
 12 
 
 No. of 
 Teeth. 
 
 41 
 41 
 41 
 42 
 41 
 41 
 41 
 24 
 30 
 38 
 32 
 32 
 38 
 38 
 30 
 30 
 
 Hand-Wheel. 
 
 Dia- 
 meter. 
 
 Ins. 
 54 
 
 36 
 
 39 
 
 54 
 42 
 42 
 36 
 24 
 
 42- 
 
 42 
 
 42 
 
 36 
 
 36 
 
 36 
 
 36 
 
 30 
 
 No. of 
 Turns. 
 
 20-5 
 20-5 
 20-5 
 21-0 
 20-5 
 20-5 
 
 20-5 
 
 H14-0 
 S 20-5 
 H15 
 
 S 60 
 
 19 
 16 
 16 
 19 
 19 
 15 
 15 
 
 Weigh 
 Shalt. 
 
 Ins. 
 61 
 
 6 
 
 4| 
 8 
 6-5 
 
 H 
 
 4| 
 
 hi 
 
 5 
 
 5| 
 
 H 
 
 H 
 
 31 
 
 H 
 
 The Reversing of Propellers driven by Turbines is usually effected by 
 means of a separate, small, additional turbine in rear of, and on the same 
 shaft with, the main or ahead-going one. The astern-going turbines are 
 generally collectively capable of developing S.H.P. equal to 25 to 30 per 
 cent, of that at full-speed ahead-going; in Germany their power is now equal 
 to 50 per cent, of the full-speed power. If the turbine is complete itself, 
 or is the low-pressure member of a compound system, this stern-going 
 instrument may be, and generally is, in the same casing with it, as shown 
 in figs. 58 and 59. The high-pressure member of a compound system 
 has the astern-going turbine in a separate casing in rear of it, and arranged 
 to exhaust to the condenser direct. In this case there is required a pass 
 valve or two valves geared so that steam may be admitted to either one of 
 
'00 
 
 MANUAL OF MARINE ENGINEERING. 
 
 them while shut off from the other, or both may be shut off at the same 
 .time. For this purpose on ordinary slide valve, such as fitted to the 
 
 o 
 J2 
 
 .a. 
 
 H 
 
 _> 
 
 > 
 
 to 
 
 a 
 ••* 
 ■— 
 > 
 
 8 
 a 
 
 22 
 
 (0 
 
 "a.- 
 
 £ 
 
 o 
 
 -0 
 
 ■eylindex ol a reciprocating engine, or, as usual, for alternating in a steering 
 engine, will answer the purpose; or for large sizes two mushroom valves 
 
KEGULATING AND STOP VALVES. 701 
 
 arranged that both may be shut or one lifted without disturbing the 
 other. 
 
 With three propellers the middle one is driven by the H.P. member, 
 and each wing propeller by one of the two L.P. members, and as manoeuvring 
 is usually done by the wing screws only, they have reversing turbines, as 
 in fig. 58. But in this case the main or ahead-going turbine on service obtain 
 their supply of steam from the H.P. member driving the centre shaft, con- 
 sequently in manoeuvring them an auxiliary supply from the boiler is neces- 
 sary for both ahead and astern-going turbines. They may, therefore, have 
 manoeuvring valves for distribution of this steam, as already described, 
 and the middle turbine remains stopped and out of action during the " back- 
 ing and filling " of the others. 
 
 Blenkinsop's Arrangement for Manoeuvring is due to the observation of 
 the late Mr. Blenkinsop, of the Great Eastern Railway Company, that with 
 the centre screw moving in head gear the ship was much more easily handled, 
 inasmuch as the rudder then lent powerful aid in turning under the action 
 of the wing screws. It need hardly be said, however, that it sometimes 
 happens that in executing a turning movement no headway must be made, 
 but. generally speaking, this seldom is rigidly necessary for in approaching 
 and leaving a berth a steamer may with advantage advance, and such advance 
 under any circumstances can be provided for. 
 
 Fig. 266 shows Mr. Blenkinsrm's iiwon+i™ „/u~..~v~ 
 
 The Manoeuvring Valves for Twin-Screw Turbines, 
 shown in Fig. 266, are the invention of Mr. Hamilton 
 Gibson and not Mr. Blenkinsop, as stated on p. 701. 
 
 _ „^,„ u/ uuc cAnausi irom xne li.Jf. turbine is shut 
 
 off, and boiler steam admitted to the L.P. turbine, or the reversing turbine 
 alternately, or to neither as required. Very large valves are necessary 
 for such work, and the gear for working them must be efficient and powerful 
 for rapid handling. In ships of comparatively small power these valves are 
 so large and heavy that steam-worked gearing is necessary for such handling, 
 while in those of large power only such instruments as Brown's push and 
 pull reversing gears are sufficient for the purpose. Fig. 267 shows the 
 arrangement of valves and gear fitted in H.M.S. " Lion " of 80,000 S.H.P., 
 w T hose power is developed in two sets of turbines, which drive four screws in 
 the usual way. 
 
 Regulating and Stop Valves. — As it is necessary that a marine engine 
 when required may run at any number of revolutions from full speed to 
 dead slow, means must be provided for doing this efficiently, especially for 
 warships whose speed when in fleet formation must be regulated to a 
 nicety. In the old days speed was decreased by means of a throttle valve, 
 or by i; notching" up the link motion. The latter method is still adopted 
 with advantage, but it is not sufficient when very low speeds or minute 
 changes in speed are necessary. The old Butterfly throttle valve is no 
 longer used, except for governing purposes, for against the high pressures 
 now obtaining it is never tight enough for practical purposes. 
 
700 
 
 MANUAL OF MARINE ENGINEERING. 
 
 them while shut off from the other, or both may be shut off at the same 
 lime. For this purpose on ordinary slide valve, such as fitted to the 
 
 
 ~S1 
 
 © 
 
 D 
 
 
 ■eylindei oi a reciprocating engine, or, as usual, for alternating in a steering 
 engine, will answer the purpose; or for large sizes two mushroom valves 
 
KEGULATING AND STOP VALVES. 701 
 
 arranged that both may be shut or one lifted without disturbing the 
 other. 
 
 With three propellers the middle one is driven by the H.P. member, 
 and each wing propeller by one of the two L.P. members, and as manoeuvring 
 is usually done by the wing screws only, they have reversing turbines, aa 
 in fig. 58. But in this case the main or ahead-going turbine on service obtain 
 their supply of steam from the H.P. member driving the centre shaft, con- 
 sequently in manoeuvring them an auxiliary supply from the boiler is neces- 
 sary for both ahead and astern-going turbines. They may, therefore, have 
 manoeuvring valves for distribution of this steam, as already described, 
 and the middle turbine remains stopped and out of action during the " back- 
 ing and filling " of the others. 
 
 Blenkinsop's Arrangement for Manoeuvring is due to the observation of 
 the late Mr. Blenkinsop, of the Great Eastern Railway Company, that with 
 the centre screw moving in head gear the ship was much more easily handled, 
 inasmuch as the rudder then lent powerful aid in turning under the action 
 of the wing screws. It need hardly be said, however, that it sometimes 
 happens that in executing a turning movement no headway must be made, 
 but. generally speaking, this seldom is rigidly necessary for in approaching 
 and leaving a berth a steamer may with advantage advance, and such advance 
 under any circumstances can be provided for. 
 
 Fig. 266 shows Mr. Blenkinsop's invention, whereby either wing 
 screw is moved ahead or astern while the other does the reverse — viz., 
 astern or ahead — both turbines taking steam from the centre or high- 
 pressure one. 
 
 Ships with Four Screws driven by Turbines usually have one H.P. and 
 one L.P. machine on each side, the L.P. ones being outermost, and having 
 each a reversing turbine on the same casing, as already described. In this 
 case manoeuvring is done entirely with the outer screws, and all that 
 is essential is means whereby the exhaust from the H.P. turbine is shut 
 off, and boiler steam admitted to the L.P. turbine, or the reversing turbine 
 alternately, or to neither as required. Very large valves are necessary 
 for such work, and the gear for working them must be efficient and powerful 
 for rapid handling. In ships of comparatively small power these valves are 
 so large and heavy that steam-worked gearing is necessary for such handling, 
 while in those of large power only such instruments as Brown's push and 
 pull reversing gears are sufficient for the purpose. Fig. 267 shows the 
 arrangement of valves and gear fitted in H.M.S. " Lion " of 80,000 S.H.P., 
 whose power is developed in two sets of turbines, which drive four screws in 
 the usual way. 
 
 Regulating and Stop Valves. — As it is necessary that a marine engine 
 when required may run at any number of revolutions from full speed to 
 dead slow, means must be provided for doing this efficiently, especially for 
 warships whose speed when in fleet formation must be regulated to a 
 nicety. In the old days speed was decreased by means of a throttle valve, 
 or by ' : notching" up the link motion. The latter method is still adopted 
 with advantage, but it is not sufficient when very low speeds or minute 
 changes in speed are necessary. The old Butterfly throttle valve is no 
 longer used, except for governing purposes, for against the high pressures 
 now obtaining it is never tight enough for practical purposes. 
 
702 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Fig. 268 shows a form of stop regulating valve which has been used 
 successfully in the mercantile marine for many years. The little valve on 
 the spindle is easily opened, and is sufficiently large to supply the necessary 
 steam for manoeuvring purposes if it is about one-third the -diameter of the 
 main valve. Large valves of this kind, however, offer considerable resistance 
 
 Fig. 2(>7. — Manoeuvring Gear, consisting of ahead and astern-manoeuvring valve in 
 front, with the intermediate valve behind, worked by means of a direct-acting steam 
 engine. 
 
 to opening, notwithstanding having a pass valve, and there are other objec- 
 tions, so that many engineers prefer to have some form of balance stop 
 valve, such as is shown in fig. 269, with an independent manoeuvring valve 
 on the side. 
 
703 
 
 ACTUATING 
 VALVE 
 
 Fig. 268. — Engine Stop and Regulating 
 Valve (for small engines). 
 
 Fig. 270. — Cockburn's Double-Beat 
 Type Regulating Valve. 
 
 Pass and 
 Manoeuvring Valve. 
 
 Fig. 269. — Balanced Stop and Regulating Valve (for large engines) 
 
704 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Fig. 270 is a form of equilibrium stop valve, self-adjusting, so that it 
 keeps tight under differing pressures and temperatures. 
 
 The regulating valves, when properly balanced, can be worked by a 
 lever, such as was used in the throttle valve, or with a wheel at the starting 
 platform in the ordinary way. Very large valves, however, are too cumber- 
 some to be easily and quickly worked by hand, and may, therefore, have 
 
 STOK.EH0LB 
 
 an apparatus something like the Brown' steam starting gear attached to 
 them, and operated by a small lever at the starting platform. 
 
 Steam Turning Gear. — Another labour-saving appliance now universal 
 in the Navy and mercantile marine, is steam gear for turning the engines 
 when in port. In all but small ships a separate engine is provided for the 
 purpose, but in small ships the donkey pump engine, or the reversing engine, 
 
GOVERNORS. 705 
 
 when there is one, is employed by using belt or rope gear. The usual plan is to 
 fit a second worm-wheel to the worm-^kaft, and to turn it by a worm on the 
 shaft of the special engine, or on a shaft with a pulley to be worked by an 
 auxiliary engine. 
 
 Steam Ash Hoists. — All large ships require some mechanical means of 
 disposing of the ashes, clinker, etc. ; the simplest form of hoist is only a 
 small winch worked by a steam engine without wheel-gearing ; the barrel is 
 of small diameter, and the flywheel is heavy, and kept running at a constant 
 speed, the bucket of ashes being " whipped " up in the same way that light 
 cargo is got out of the holds with a steam winch. This gear has the merit 
 of cheapness and simplicity, it can be worked by the most ignorant,, and 
 does not easily get out of order. 
 
 Ash Hoist (See's), fig. 271, is an ingenious arrangement whereby the 
 ashes are driven to the top of the tube, tilted over, and shot through the 
 ship's side. With this gear, the fireman is not required to leave the stoke- 
 hole, and no labour is required beyond that of shovelling the ashes into the 
 hopper. Fig. 271 shows the arrangement to consist of a hopper having a 
 hinged water-tight cover ; at its bottom is a special form of nozzle, which 
 discharges water up the delivery pipe from a special duplex pump, when 
 the pressure has been got up in it to 200 lbs. As soon as the gauge shows 
 the pressure, the cock is quickly opened and a rush of water up the pipe 
 takes place, carrying with it the ashes in the hopper ; the pressure drops to 
 150 lbs., but the stream still flows through the pipe carrying air and ashes 
 with it as fast as the latter are shovelled into the hopper. When the whole 
 are disposed of, the cock is closed quickly to prevent back wash ; the hopper 
 is closed, as is also the discharge valve on the ship's side. This apparatus 
 is a great convenience and comfort, especially in bad weather, and saves a 
 huge amount of unpleasant labour ; it also prevents the need of the grimy 
 and unwashed denizens of the stokehole showing themselves on the decks 
 of yachts or passenger vessels. On the other hand, unless care is taken in 
 placing the discharge the detritus is liable to get into the condensers with 
 the cooling water and damage the tubes ; and, further, if the cinders and 
 clinkers are allowed to get into the screw race, injury will be done to the 
 propeller blades, especially if of bronze. 
 
 Governors. — To prevent the engines from racing when the sea is rough, 
 it was usual to fit an instrument which should control the throttle-valve 
 automatically. The governor for shipboard must be designed and arranged 
 so that the pitching and rolling of the ship do not prejudice it, but it should 
 act rather in anticipation of the former. 
 
 There are two distinct classes of marine governor — viz., those whose 
 action is influenced by the variation of motion of the engine itself, and those 
 whose action is caused by the change of pressure at the stern, due to vari- 
 ation of head of water. The former class can only act after a variation of 
 speed has taken place, the latter anticipates and checks such variations. 
 At first sight the latter present the most favourable qualities, inasmuch 
 as they anticipate change of velocity, but they serve only one purpose, that 
 of checking racing of the engine due to the propeller emerging from the water. 
 The other governors are necessarily a little late in action, but they may be 
 made so sensitive as to be almost as quick as the others ; they have, how- 
 ever, one superior merit, and that is, they check racing from any and every 
 
 15 
 
706 MANUAL OF MARINE ENGINEERING. 
 
 cause. II a propeller or shaft break, the pneumatic 01 hydrostatic governor 
 fails to check the engine ; the other governors will check any large increase 
 in velocity and give the engineer time to snut olf steam. 
 
 Both classes of governor are susceptible of subdivision, and may be 
 distinguished as those which act direct on the throttle- valve, and those 
 which act on the valve o! a steam cylinder whose piston operates on the 
 throttle-valve. They are, however, seldom fitted now, as the necessity for 
 them has ceased to be acute, owing to the smaller screws with better sub- 
 mersion, and the multiple-crank compound engines do not race so violently 
 as the old simple condensing. Turbines, however, do require governors for 
 their own safety. 
 
 Dunlop's Governor. — This instrument consists of a vertical cylinder 
 placed close to the stern of the ship, and as low down below the water line 
 as convenient ; there is a communication between the sea and the bottom of 
 this cylinder by a cock or valve of ample size, fitted to the skin of the ship. 
 When the stern of the ship is lifted out of the water, the cylinder is emptied 
 of water, and the pressure in it is that of the atmosphere ; on the stern 
 dipping deep into the water again, the water rushes into the cylinder and 
 compresses the air in it till there is a pressure due to the " head " of water, 
 which may amount to as much as 12 lbs. per square inch in large ships, and 
 5 lbs. even in small ones. The top of the cylinder communicates by means 
 of a pipe with a vessel in the engine-room, which is closed air-tight at its 
 top by a thin corrugated circular diaphragm. The bottom of this vessel is 
 bell-shaped, so that at its top it is of considerable diameter, and the dia- 
 phragm capable of exerting considerable force when even so small a pressure- 
 as 1 lb. per square inch is transmitted from the cylinder. The gear for 
 operating on the throttle- valve is connected to the middle of the diaphragm, 
 and any bulging of it by the pressure causes the throttle-valve to be opened, 
 and when the water cylinder is emptied by the rising of the ship's stern 
 relatively to the sea, the diaphragm assumes its normal position, and the 
 throttle-valve is closed. 
 
 Steam Governors. — The chief cause of failure of the first marine governors 
 ■ was their inability to move the throttle-valve promptly when the spindle- 
 gland was tightly packed ; it was also difficult to set them so as to act 
 effectively when they would move at all. The difficulty was got over by 
 limiting the function of the governor to move the slide-valve of a small 
 steam cylinder whose piston performs the operation of opening and shutting 
 the throttle-valve. The governor is of ample power to move the small 
 slide-valve "with precision, and the steam cylinder can always be made of 
 sufficient size to work the throttle-valve, however tightly its gland is 
 packed. For the purpose of working the small slide-valve, a Silver's or 
 a Meriton's governor may be employed, and many of the old governois 
 of this kind were converted to steam ones by the addition of this steam 
 cylinder. 
 
 Durham's and Churchill's Velometer.— The motion of the engine is com- 
 municated to this governor by a small rope of wire or Manilla on the usual 
 pulleys; it is transmitted from the pulley-shaft to the shaft of a paddle- 
 wheel enclosed in a cylindrical trough, by means of a bevel-wheel, whose 
 axle is free to move about the axis of the shafts in a plane perpendicular to 
 it. and which gears into a similar bevel-wheel on the end of each ah.ift. (This 
 
GAUGES. 707 
 
 is similar to the arrangement provided in traction engines to admit of their 
 going around a curve.) 
 
 The trough is rilled with water or oil, which is carried around with the 
 paddle-wheel, and causes it to resist any sudden changes of motion. The 
 axle of the intermediate bevel-wheel is connected to the small valve of the 
 steam cylinder whose piston operates on the throttle-valve. If the engine 
 races so that the bevel-wheel on the pulley-shaft moves faster than that on 
 the paddle-wheel axle, it will carry the intermediate pinion along with it, 
 until the motion of the paddle-wheel is accelerated by it, and the pinion 
 axle acting on the small slide-valve causes the throttle-valve to be closed. 
 The paddle-wheel is now moving at a higher rate than its normal speed, so 
 that when the engine has slowed down to its normal speed, the bevel-wheel 
 axle is moved in the opposite direction, so as to cause the throttle-valve to 
 be opened, and the engine is thus prevented from further " slowing down." 
 
 This governor, when carefully adjusted, is most sensitive, and will prevent 
 any dangerous racing under the most trying circumstances. 
 
 Coutt's and Adamson's Governor. — This is an extension of Dunlop's 
 principle, and differs from it by the diaphragm being caused to move the 
 slide-valve of a small steam cylinder whose piston operates on the throttle- 
 valve. This has the advantages and disadvantages of Dunlop's governor, 
 except that the work of moving the throttle-valve is done by steam. 
 
 Westinghouse Governor. — A sensitive ball governor, of very small size, 
 is made to operate on a valve which, in opening, allows the steam on one 
 side of the piston of a steam cylinder to escape to the condenser, that on 
 the other side forcing the piston to move quickly and shut the throttle- 
 valve ; the valve closes again as soon as the steam has escaped, and the 
 steam flowing into the cylinder allows the piston to go back to its original 
 position. 
 
 The piston here spoken of is connected to a smaller piston in another 
 cylinder, whose function is to bring it back to its initial position. 
 
 This governor is very sensitive and acts very well, but is liable, from 
 want of attention, to get out of order, and then fail to act at all. Fig. 272 
 is Murdoch's, and very similar and general to the Westinghouse. 
 
 With the advent of triple engines and twin screws the absolute necessity 
 for a governor disappeared ; and to-day, although some racing takes place 
 with all ocean-going marine engines, hardly a single modern ship having 
 reciprocators is furnished with a governor, and when such an instrument has 
 been provided, it generally supplies ample evidence of its non-use if not of its 
 uselessness. 
 
 Gauges. — In every engint-room there should be a steam gauge, which 
 shall show the pressure at the high-pressure cylinder valve-box ; there should 
 be a gauge which shall show the pressure in each receiver (when the engine 
 is compound); that on the low-pressure is called a compound gauge, as it 
 is marked as a pressure gauge when above atmospheric pressure, and as a 
 vacuum gauge when below- -this gauge might, with advantage, be marked 
 eo as to show absolute, pressure ; if this were so an additional stumbling- 
 block would be removed from the path of some young engineers, and permit 
 them to have les<» hazy views on the question of " vacuum " ; and a gauge, 
 commonly e&lled tie vacuum gauge, which shall show the pressure in the 
 •condenser. This gauge i? marked in inches, and so indicates how high tbe 
 
708 
 
 MANUAL OF MARINE ENGINEERING. 
 
 column of mercury would be in a vertical tube whose upper end is connected 
 with the condenser, and the lower open to the atmosphere. The old original 
 vacuum gauge was simply like a barometer, and, when replaced by the 
 Bourdon gauge, to avoid confusion the new ones were marked in this way. 
 To say that there is a vacuum of 12 inches, means that the difference between 
 the pressure in the condenser and that of the atmosphere is equal to the 
 weight of 12 cubic inches of mercury, or due to a " head" of 12 inches of 
 mercury, or 6 lbs. very nearly. The actual pressure in the condenser is, 
 however, not affected in the least by the pressure of the atmosphere, and 
 gauges can now be obtained which give the real actual pressure. 
 
 Fig. 272. — Combination Governor (Murdoch's Patent). 
 
 It would be far simpler, and certainly more scientific, to have the con- 
 denser gauge marked from to 15 lbs. absolute ; the " compound " gauge, 
 or that attached to the valve-box of the low-pressure cylinder, marked from 
 to 50 lbs. absolute ; and that attached to the valve-box of the other cylinders 
 to 250 lbs. absolute, or to such limit as shall be at least 25 per cent, higher 
 than the working pressure. 
 
 The gauges on the boilers might be marked as at present — viz., to indicate 
 the pressure above that of the atmosphere, as it might be more difficult to 
 train firemen to know the meaning of the new markings, but after all it is 
 really immaterial to them how it is graduated, so long as they know to what 
 
CENTRIFUGAL LUBRICATORS. 
 
 709 
 
 mark they must keep the pointer when under steam, and that, when the 
 pointer begins to move, on getting up steam, pressure is forming. 
 
 Care should be taken in setting the gauges in the engine-room that allow- 
 ance is made for the extra pressure due to the " head " of water in the gauge 
 pipe, which will be about 1 lb. for every 2| feet of vertical fall. 
 
 Lubricators and Impermeators. — To obtain perfect lubrication the supply 
 should be steady, uniform, and continuous, or nearly so. This is true for 
 ■every bearing, guide, etc., and equally true for the lubrication of the internal 
 parts. It is usual to rely on capillary attraction to convey the oil from the 
 oil boxes to bearings, guides, etc., by means of worsted syphons ; this is a 
 very simple, well tried, and fairly efficient method, but it has serious draw- 
 backs ; it requires constant attention, as the worsted wick syphons act as 
 filters and become clogged with gluey matter contained in some oils, and 
 there is no definite means of proving that the oil is passing. An equally 
 simple and vei^ satisfactory plan (fig. 273) is to fit an oil box a few inches 
 above each bearing, in such a way that if oil is dropping from it. it can be 
 seen or felt ; there is a small cup to each oil hole leading to the bearing, and 
 over each of these is a small nozzle from the bottom of the oil box. fitted with 
 a small plug so as to regulate the 
 flow of oil, or to stop it altogether ; 
 if preferred, however, syphons may 
 be fitted instead of screw plugs, as 
 in either case the flow of oil can he 
 proved. A perfect lubrication, how- 
 ever, can only be got by forcing the 
 oil by means of a pump through 
 every joint and bearing, collecting 
 the oil as it drops, cooling it, and 
 again pumping it to the joints. This 
 system of forced lubrication intro- 
 duced by Messrs. Belliss & Morcom, 
 and applied to their high-speed elec- 
 tric light and power engines, has 
 proved a great success by permitting 
 of a mechanical efficiency of 93 per 
 oent. with even small engines, and 
 preventing wear of bearings and 
 journals during quite long periods of 
 work. 
 
 Cadman's Patent Lubricators. — 
 Over each moving part required to 
 
 be lubricated is an oil box having, projecting through the bottom, a small 
 plug, held there by a spring, and so set that the oil box on the moving part 
 touches the plug end and opens it so as to let a drop of oil pass. This is 
 especially adapted for the piston and connecting-rod brasses of a vertical 
 engine, and for guides, etc. 
 
 Centrifugal Lubricators. — Whei; engines are running at a high speed, 
 ordinary external appliances, such as telescopic pipes, jointed pipes, pipes 
 fixed to the rod and taking oil from wipers, etc., are not reliable or sufficient 
 for the purposes of thoroughly lubricating the crank-pin brasses. If a system 
 
 Fig. 273. — Centrifugal Lubricator. 
 
710 
 
 MANUAL OF MARINE ENGINEERING. 
 
 of forced lubrication by means of holes in the crank-shaft, its arm and pin, is 
 not adopted, it is usual to have what is known as a centrifugal lubricator, 
 such as is shown in fig. 273, whereby the oil supplied from a fixed lubricator 
 passes through a tube, drops into a circular collar surrounding the shaft, 
 and is whirled into its outer part and from it by means of a tube to a hole 
 through the axis of the crank-pin, from which it flows by smaller holes to the 
 surface and so lubricates the brasses. From the time of leaving the fixed 
 lubricator centrifugal force is operating to compel the oil to flow to the brasses- 
 Lengthened experience with this in the Navy has proved its reliability, and, 
 as it is not an expensive fitting, it might be adopted with advantage generally 
 in the mercantile marine, as less oil is wasted by it than by the ordinary 
 appliances. The hole in the pin also provides a store of oil which would keep 
 the brasses supplied in case of a temporary stoppage of supply from the 
 oil-box. 
 
 Sight Feed Lubricator (fig. 274) is for the purpose of supplying a steady and 
 
 definite supply of oil to the 
 cylinders, and is so designed 
 that the condensed steam in a 
 pipe, etc., leading from the 
 steam pipe, shall carry with it 
 so many drops or globules of 
 oil as it drains back near to 
 the valve -box ; the water con- 
 densed passes through a glass 
 tube, near to the bottom of 
 which the oil enters it 3 and the 
 number of oil globules passing 
 per minute may be counted, 
 and the supply thus regulated 
 to a nicety. This is the most 
 ingenious contrivance, and in 
 the hands of careful men who 
 understand it, effects great 
 economy, while adding to the 
 efficiency of the engine ; but in 
 careless hands it is almost useless. 
 Mechanical Impermeators.— 
 These are most successful in 
 operation, and more reliable 
 than any other form, inasmuch as their action is so simple that every one 
 can understand them. Essentially there is only a force pump, having 
 a very slow motion imparted to its ram or rams by gearing, moved 
 by one of the working parts of the engine. The gearing generally consists, 
 of a ratchet lever worked from the valve-rod of the nearest cylinder, and 
 moving a wheel on whose axle is a worm which gears into a wheel on the 
 rim of a nut ; this nut fits on the thread cut on the ram, and held 
 in position by a bracket ; as the nut is moved round, the arm moves 
 slowly in or out of the chamber. The ram chamber is supplied with oil 
 from a small tank, and is connected by a pipe to the top of the valve-box 
 of the high-pressure cylinder: small non-return valves are fitted, so that 
 
 Fig. 274. — Automatic Drain from Steam Pipes 
 or Receivers. 
 
FEED-HEATERS. 71 X 
 
 the oil cannot flow back to the tank, and the steam cannot force itself or 
 the oil back into the chamber. It is, however, a very common practice now 
 to run engines without internal lubrication, or with only such as enters the 
 cylinders with the piston and valve-rods. It has been found that after an 
 engine has worn its internal moving parts fairly smooth there is no need 
 of other lubricant than the moisture from the steam, and so all risk of damage 
 to boilers from grease is avoided by doing away with internal lubricating 
 apparatus. If, however, superheated steam is used, oil lubrication may 
 become necessary in some engines ; when it does, care should be taken to 
 use only a pure hydrocarbon suitable to the temperature or graphite paste 
 forced in by a similar instrument. 
 
 Drain Pipes from the cylinders and valve-boxes should lead to the con- 
 denser, so that there shall be no loss of fresh water, and no filling of the 
 engine-room with vapour when the cocks are opened. It is customary to 
 connect these pipes to the hot-well, which serves this purpose very well with 
 high-pressure and medium-pressure engines ; but since the pressure in the 
 low-pressure cylinder of a compound engine is only for a very small portion 
 of the stroke above that of the atmosphere, the opening of the cocks will not 
 get rid of the water, but only allow air to force its way back, and so reduce 
 the vacuum. It is sometimes convenient to see if water is flowing, and so 
 prove that the cocks are not choked ; this may be accomplished by fitting a 
 " three-way " cock to the main drain pipe, which permits communication to 
 be made with the condenser or bilge at the will of the engineer. 
 
 Jacket Drains should always lead to the hot-well, and when the engine is 
 working the cocks should be open sufficiently wide to just keep the jackets 
 free of water, or there should be automatic drainers fitted ; the hot water 
 escaping from the jackets then helps to warm the feed- water. 
 
 Feed-heaters. — It is a most essential thing that the feed-water shall enter 
 the boiler as warm as possible, both to obtain evaporative efficiency and avoid 
 wear of the boiler. Economy can only be effected directly by making use of 
 heat that would otherwise be wasted for this purpose; but indirectly, by 
 promoting circulation and reducing the necessity of circulation instead of 
 checking it, considerable economy may be effected, so much so indeed as 
 to warrant the use of heat which is not " waste." Many attempts have been 
 made to heat the feed-water with exhaust steam, hot gases in the uptake, 
 etc. ; but no great measure of success has attended the efforts of those who 
 have paid most attention to this, for the apparatus employed has generally 
 been inefficient and its durability short. In the older expansive engines, 
 where the temperature of the steam at exhaust was often over 230° Fah., 
 great economy was effected by heating the feed-water in a small kind of 
 surface-condenser, placed on top of the condenser so as to intercept the hot 
 current of steam flowing to the latter ; but now with compound engines, 
 sphere the temperature at exhaust is only about 180° Fah. at the most, no 
 such means is efficient for the purpose. No doubt some economy is possible 
 even under these circumstances, especially if the feed-water is permitted to 
 "irculate in the heater for an appreciable time ; and considering that now, 
 at a much lower temperature — viz., 100° F. — a little more heat can be 
 imparted to it from an external heating agent, the temperature being 80 J 
 above it. Some day, perhaps, means will be found to avoid the loss of all 
 the latent heat which takes place, and which is huge compared with the 
 
712 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Atmospheric Va!o»* 
 
 whole of the sensible heat. The exhaust steam being at 180° F., the total 
 possible saving of sensible heat is now 50 ~ J ¥., while the latent heat lost is 
 nearly 1,000° F. 
 
 Weir's Feed-Water Heater and Automatic Regulating Gear are designed 
 to raise the temperature of the feed-water to nearly 212° Fah. by means of a 
 portion of the steam from the L.P. valve-box of the compound engine. Fig. - J70 
 gives a sectional view showing its general construction. The heating steam 
 is taken from the low-pressure casing of the main engine, and the exhaust oi 
 the auxiliary engines, such as feed pumps, electric light, fan engines, etc., 
 is also led into the heater through the non-return valve B on the side oi 
 the apparatus. A circular ring and conical spray piece with perforations 
 are fitted to mix the water and steam uniformly. The feed-water is forced 
 by the main engine feed pumps through the spring-loaded valve D on the 
 cover in a thin sheet, and is instantly heated by contact with the steam. As 
 the pressure in the heater is generally much less than that of the entering 
 water, the effect of this lowering of the pressure, and sudden heating of the 
 water, is to liberate the air in the water, and this is removed to the con- 
 
 . denser, or to the atmosphere, by a 
 small cock K on the air vessel 
 placed on the top of the heater. 
 The feed-water is thus rendered non- 
 corrosive, and falls to the bottom 
 of the heater at the boiling tern- 
 perature due to the pressure. The 
 steam admission valve is of special 
 construction. It can be opened to 
 admit the necessary amount of steam 
 to the heater, but it closes by its 
 own weight in case there is no flow 
 of steam into the heater : this valve 
 is also fitted with a dashpot, which 
 allows the valve to close gradually, 
 and prevents it hammering on its 
 seat in case of fluctuations of 
 pressure. The combination of the 
 automatic regulating gear with the 
 heater has long been a special feature 
 of the Weir apparatus. The float, 
 E. shown in the lower part of the 
 heater is a pan, with water-tight 
 bottom and sides, but open on the 
 top. It is suspended on two levers, 
 so as to move up and down with a 
 parallel motion ; the top lever spindle is carried through the door at one 
 end, and is balanced by a lever and weight. The float is always full of 
 water, and the weight is adjusted to balance when one-half is immersed in 
 water. To the weight lever another lever is attached, which actuates the 
 throttle valve F and controls the supply of steam to the pump drawing 
 from the heater. When the water in the heater rises the float is raised,, and 
 the throttle valve opened, and, when the water level is lowered, the float 
 
 it/O/on tc Unr Ptna 
 
 Fig. 275.— Weir's Feed-Water Heater and 
 Automatic Regulating Gear. 
 
EVAPORATORS OR DISTILLERS. 
 
 13 
 
 follows, and the valve is closed; the level of the water is thus kept constant 
 in the heater, and the pumps are completely filled with water. The regulating 
 valve is a cock with a parallel key : the pressure of the steam keeps it per- 
 fectly steam-tight, although it may have worn slack in the shell ; the pressure 
 also keeps the shoulder of the key against the bottom of the stuffing-box. 
 so that the stuffing gland is always kept slack. A relief valve and the Tieces- 
 sary gauges are also fitted to the heater. 
 
 Evaporators or Distillers. — The advantages of supplying marine boilera 
 
 Fig. 275a. — Live Steam Feed-Water Heater 
 (Caird and Rayner). 
 
 Fig. 27G. — Evaporator. 
 
 with pure water are great, and are so obvious as not to need specifying. 
 Samuel Hall, the successful introducer of the surface condenser, was so 
 thorough in his desire to use only fresh water in the boilers of ships with his 
 condenser that in 1833 he fitted evaporators to several of them made on 
 precisely the same principles as govern the design of our modern ones. He, 
 however, went beyond the modern engineer by placing his apparatus in the 
 steam chest of the boiler itself, so as to lose no heat. The necessity of it 
 was not, however, so severely felt until voyages of considerable length had 
 been made with ships whose boilers work at pressures of 100 lbs. and upwards. 
 The weight of water evaporated in boilers, whose working pressure is 150 lbs., 
 is much greater in proportion to the size than was the case with those working 
 at 75 lbs. ; and the evils arising from the deposit of scale are magnified with 
 
714 
 
 MANUAL OF MARINE ENGINEERING. 
 
 the higher pressure and consequent higher temperature. Again, the liability 
 to put on scale is greater, inasmuch as the losses from leakages are greater 
 with the higher pressures. Hence, the old system of making up loss of water 
 by a supply from the sea, although a very simple and ready one, was not 
 by any means satisfactory, and did not remedy the evil, but rather magnified 
 it. The Admiralty and some private shipowners tried to obviate it bv pro- 
 viding a supply of fresh water in the double bottoms, or in tanks specially 
 fitted for the purpose. This, however, was only half a remedy, inasmuch 
 
 Fig. 277. — Weir Admiralty Evaporators. Fig 277a. — H.P. Evaporator for Destroyers. 
 
 as the fresh water generally obtainable contained large quantities of lime 
 and other salts which gave a hard deposit more difficult to remove. More- 
 over, this fresh water cost money, and was so much extra weight to carry, 
 which also added to its cost. Then recourse was had to the auxiliary or 
 donkey-boiler to obtain distilled water, which meant an expenditure of coal 
 as well as labour in cleaning out these boilers after they had become coated. 
 Besides, these small boilers very soon got so coated that they had to be 
 stopped for a thorough clean-out ; and during the time they were at work 
 
EVAPORATOR!? OR DISTILLERS. 
 
 715 
 
 there was always the risk of damaging them. In spite of these difficulties, 
 however, it was found to be the most satisfactory way of obtaining an extra 
 supply for the main boilers, and, consequently, improvement was made in 
 this direction by designing and supplying a small boiler (fig. 276), whose heat 
 is obtained from either the steam direct from the main boilers or from the 
 exhaust from one or other of the cylinders ; the former plan was found 
 eventually to be the best. 
 
 Fig. 277 may be taken as another example of the type now in general 
 use. It consists of a vertical cylindrical shell fitted with mountings and 
 gear similar to those on a steam launch boiler ; instead of a furnace, com- 
 bustion chamber, tubes, etc., it has a tubulous arrangement ingeniously 
 contrived so that the steam is made to give up its heat to the water within 
 
 Fig. 278. — Fresh-Water Distilling Apparatus for British Battleships and Large Cruisers 
 
 (Caird and Rayner). 
 
 the evaporator as far as possible, and the resultant water to drain away and 
 be returned to the main condenser or hot-well. Steam is in this way raised 
 in the evaporator, and passes from it to the main condenser or, as in naval 
 ships, to the auxiliary condenser, the resultant water being finally pumped 
 into the main boilers in the usual way. Sea water is pumped into the 
 evaporator by a small donkey-pump, and the salt is blown down from the 
 evaporator in the same way as was usual with boilers supplied with sea 
 water. The internal tubulous apparatus is so arranged that it can be easily 
 withdrawn from the shell, as shown in the nature, for a thorough clean-out 
 
716 
 
 MANUAL OF MARINE ENGINEERING. 
 
 when necessary. Instead of the steam from the evaporator being sent direct 
 to the condenser, it can be made to do useful work by admitting it to the 
 valve box of the low-pressure cylinder. 
 
 There are other equally ingenious and efficient evaporators, but they are 
 all worked on the same principle of heating water and converting it into 
 steam with steam made in the main boilers. In all cases, as is only to be 
 expected, tubes are employed in one form or another to effect this purpose. 
 
 Ladders. — The main ladder to an engine-room should not be less than 
 18 ins. wide, and where space permits should be 24 ins., and even more in 
 large ships ; the sides are of flat iron bars usually. 4 X § in. ; the treads 
 or steps are of cast iron, and 10 ins. apart, and from 4J to 7| ins. wide. . The 
 inclination of the ladder to the vertical is usually about 1 in 3 with narrow, 
 and 1 in 1\ with broad steps ; the hand rail is 1 in. diameter when of iron, 
 and from 1 \ to If ins. when of brass ; the former looks better from an engineer's 
 point of view, is more durable, and easier kept clean. Ladders leading to 
 *he various parts of the engine are usually made lighter than the main ladder, 
 
 Fig. 273. — Gravitation Type Feed-Water Grease Extractor (Caird and Rayner). 
 
 and the steps are often formed of three " spills " or bars, § in. diameter, or, 
 better still, square section ; when weight is important they may be of trian- 
 gular section. 
 
 It is very essential that means be provided for the engineers to get easily 
 and safely to every part of the engine requiring attention ; these light 
 ladders are a source of great convenience, and are amply paid for in the 
 better attention given to the w r orking parts. 
 
 Gratings and Platforms. — With the same object in view, good platforms 
 to stand on, and gratings to form roads to the various parts should be pro- 
 vided. The engine-room platform is usually laid with either cast- or 
 wrought-iron chequered plates, the pattern on whose face should be one 
 which will give good foothold, and not prevent dirt from being swept from 
 it. Some engineers prefer to have all bottom platforms made of wood, and 
 laid over with sheet lead ; the lead permits of good foothold, and is easily 
 kept clean ; water runs easily off it. and when in good repair it looks very 
 
STOKEHOLE VENTILATORS. 717 
 
 well ; but it is not. so durable as the iron, and is very liable to damage from 
 weights falling on it, The gangways leading to the upper parts of engines 
 are sometimes made with chequered wrought-iron plates ; but except when 
 they are immediately over the working parts this is not a good practice, as 
 both light and ventilation are obstructed by them, and they require constant 
 cleaning. " Spill " gratings are, therefore, preferable in most cases, as they 
 stop light and ventilation to only a very small extent, and require no clean- 
 ing ; they are made with sides of flat bars 2| X f in., and cross bars f in. 
 diameter, and spaced 2| ins. apart ; those of large size and liable to support 
 heavy weights have sides 3 x-g in., with spills | in. diameter, pitched 
 2-§ ins. 
 
 Feed Filters are often fitted in ships of all sizes with the object of pre- 
 venting any oil or grease getting into the boilers with the feed- water ; they 
 certainly succeed in keeping out grease and solid matter, but fail to do so. 
 with oils. Generally the filtering material is Turkish towelling so placed in 
 one or more layers that all the feed- water must pass through it and leave 
 deposited on it all solids and semi-solids, and to a great extent absorbed by 
 it oils and greases in the liquid state. The apparatus is always arranged so. 
 that the " cartridge " of towelling can be easily withdrawn and a fresh clean 
 one substituted without causing any stoppage of the engines (v. fig. 279). 
 
 Stokehole Ventilators snould have an aggregate transverse area of 0-45 
 square inch for each pound of fuel burned per hour ; or, say, 0*675 square 
 inch per I.H.R of trial trip in the Mercantile Marine, or 0*75 square inch 
 per I.H.P. in Express steamers on short service and naval ships; for turbines 
 0-62 per S.H.P. is sufficient. 
 
 The area of the mouths should not be less than — 
 
 1-35 square inch per pound of iuel in 10 knot ship. 
 
 1-24 
 
 5) 
 
 11 
 
 1-13 
 
 11 
 
 )) 
 
 103 
 
 11 
 
 »! 
 
 0-93 
 
 11 
 
 >> 
 
 0-85 
 
 11 
 
 11 
 
 0-78 
 
 'J 
 
 »> 
 
 072 
 
 )» 
 
 »» 
 
 11 
 
 12* 
 
 ?) 
 
 >> 
 
 15 
 
 ii 
 
 >! 
 
 171 
 
 ii 
 
 j; 
 
 20 
 
 ■1±\ 
 
 ii 
 
 >) 
 
 ii 
 
 >> 
 
 25 
 
 ii 
 
 »» 
 
 30 
 
 11 
 
 The diameter of the mouth is usually twice that of the neck, but with very 
 large down-casts the mouth is somewhat less. If d is the diameter of the 
 neck and D that at the mouth, then — 
 
 D = 1-75^ + 5 inches. 
 
 The maximum flow of air in feet per minute through the neck (d inches) 
 should be 1,000 Vd. 
 
7 IS MANUAL OF MARINE ENGINEERING. 
 
 CHAPTER XXVII. 
 
 WEIGHT AND OTHER PARTICULARS OF MACHINERY RELATING THERETO. 
 
 The Weight Of the Engines and Boilers of every ship is a matter of extreme 
 importance, although certain classes of shipowners rarely give a second 
 thought to it, and probably would not pay anything extra for a saving of 
 5 per cent., although it would enable them to carry more cargo and gain 
 yearly a respectable sum of money. In the case of small steamers of high 
 speed it is of the very first importance ; in fact such speeds as they are run 
 at to-day are only attainable by the most careful designing of the engines, 
 whereby the weight is reduced to a minimum. The best modern instances 
 of this are to be found in the Torpedo Boat and Torpedo Boat Destroyer ; 
 some good ones are also observable in the cross-channel steamers, both 
 paddle and twin screw. In the case of the modern Atlantic steamers the 
 margin for weight of cargo is exceedingly small, so that a ship may have 
 50,000 tons displacement, and only a margin of 500 tons for cargo pure and 
 simple. The machinery of such a ship will weigh 8,500 tons. It will be seen, 
 therefore, that the saving of a small percentage of its weight would add a 
 very large percentage to the dead-weight cargo capacity. (Of course there 
 is plenty of room for stowing cargo in these ships, so that they can carry 
 plenty of goods, whose specific gravity is low, and thereby obtain high rates 
 of freight if there is surplus displacement sufficient.) In the case of cross- 
 -channel steamers the present speeds were only possible by the reduction in 
 size of boiler per I.H.P., consequent on the introduction of the triple and 
 ■quadruple engine and forced draught. In their case, however, it was not 
 merely a question of weight, but one of capacity as well. In the case of the 
 ■cargo steamer the weight of the machinery is only a comparatively small 
 fraction of the total displacement, and hence is, as already stated, too generally 
 disregarded; but here, too, the question is worthy of attention, as in the 
 times of bad freights the difference of a few tons in weight of cargo carried 
 may make the difference between a profit and a loss. 
 
 Most modern engineers, however, have their attention called to the 
 question of weight by other considerations than obtain with the shipowner. 
 
 In the ease of warships, weight of machinery again is of prime importance, 
 inasmuch as the armament and fuel supply is in no small measure dependent 
 on the margin of displacement left, after providing for the whole equipment 
 and machinery. In other words, the lower the weight of the engines and 
 boilers, the larger may be the armament, the armour, and the fuel supply. 
 In the case of battle ships, whose speed is comparatively moderate-, and the 
 displacement on the dimensions, consequently larger than is the case with 
 cruisers whose speed is very high, the machinery need scarcely be so light : 
 ard although, as a rule, the design of both engines and boilers is similar to 
 that v,t the cruiser the rate of revolution* is not nearlv so hi>'h. 
 
THE MODERN FIRST-CLA.SS CRUISER. 719 
 
 Again, the small cruisers, whose speed is to be as high as that of the 
 big ones, have not the same margin for their machinery ; it is, therefore, 
 imperative that their engines should be lighter per H.P. The saving in 
 weight is generally effected in this case by making the engines to run at a 
 very high number of revolutions, with a high effective steam pressure, and 
 designing the various parts with the greatest care possible, using only material 
 of the greatest strength for the purpose. The factor of safety, therefore, is 
 practically as high as that of other ships, because these ships may have to 
 run in Avar -time continuously at or near full speed for long periods. 
 
 The machinery of the Torpedo Boat and Destroyer is, of course, also 
 designed with the greatest care, and made of the very best material, but 
 here the so-called margin of safety is reduced ; in other words, engines are 
 made so that' at their ordinary rates of speed there will be as good a margin 
 of safety as in any other ship ; spurting they will have to do in time of war, 
 but their fuel supply does not permit of long duration of it, and the exigencies 
 of the service in peace time would not require it to be done often in the course 
 of a twelve-month ; hence, although the factor of safety at high speeds may 
 be low, it is still w r ell within the safe-working limit of the material, and no 
 great risk of accident is run. A reference to Tables lxxxv. to lxxxviii. will 
 show the progress of modern engineering in both the Navy and mercantile 
 marine during the past thirty years, and the lines on which modern engineers 
 have advanced to attain their present success in coal consumption and the 
 weight of machinery. 
 
 In 1870 Battleships and Cruisers were still fitted with simple horizontal 
 engines and box boilers of low pressure. They were on the expansive prin- 
 ciple, steam being cut-off at a little past half stroke, and exhausted into a 
 surface condenser. The boilers were designed for a pressure of 30 lbs., 
 although sometimes 1 lb., or even 2 lbs., more was attained. The maximum 
 piston speed was about 600 feet per minute on trial trips with large engines, 
 and about 50 to 500 feet with the smaller ones. The mean pressure was 
 about 20 lbs. per square inch in battleships and 20 to 25 lbs. in cruisers. 
 The I.H.P. per ton total weight machinery was about 7-0; 8-5 I. H.P. was 
 developed for each square foot of grate, and 33 per 100 square feet of total 
 heating surface ; the consumption of coal was about 3| lbs. at full speed, 
 and about 2| at half power. 
 
 The Battleship of 1905 was fitted with reciprocators, had the four-cylinder 
 triple-compound engine balanced on the Schlick Tweedy system, and running 
 at high revolutions gave a trial speed of 19 knots. To-day all British battle- 
 ships and most foreign are driven by turbines operating on four screws, 
 developing 35,000 to 60,000 S.H.P., and attaining a speed on trial of 25 knots 
 with 330 to 380 revolutions per minute. 
 
 The Modern First-elass Cruiser differs from the battleship chiefly in speed, 
 and consequently in length. The turbine-driven ones of to-day attain a 
 speed on trial of 33 knots, and some are said to do even more. Steam is 
 supplied by water-tube boilers, and the revolutions at full power are about 
 300 per minute. The ships previous to the adoption of the turbine were 
 of very considerable size and power, and attained speeds of 23 to 24 knots 
 with twin screws running from 110 to 135 revolutions per minute. Their 
 engines were cenerallv of the four-crank triple-compound, like those of the 
 battleship. 
 
720 MANUAL OF MARINE ENGINEERING. 
 
 The Second-class Cruiser of to-day is larger and much faster than thos. 
 of the first-class of ten years ago. Their speed, when driven by turbines, i? 
 26 knots at about 500 revolutions. 
 
 The Light Cruiser and scout are modern creations, having a speed of 25 to 
 26 knots when driven by reciprocators at 200 to 210 revolutions ; but when 
 driven by turbines 30 knots is general, and by many later ones 36 knots has 
 been attained. 
 
 The Destroyers are now of two classes ; the one for coastal defence and 
 attack, the other ocean-going as videttes to a fleet. The early destroyer 
 of twenty years ago was 180 to 190 feet long, had twin-screw compound 
 engines of about 4,000 I.H.P. at 360 to 400 revolutions : they were followed 
 by similar but somewhat larger vessels, which attained a speed of 30 knots 
 with triple engines indicating 6,000 to 7,000 H.P. at about 400 revolutions. 
 A few faster ones were built of the same type, whose speed was 32 to 33 knots. 
 To-day the destroyer for shore work has a speed of 25 to 26 knots, with a 
 power of about 7,000, and the sea-going craft 35 knots driven by three screws 
 and turbines of 25,000 S.H.P., running at about 750 revolutions per minute. 
 The high speeds attained by this class of ship are due to the very light 
 construction of both hull and machinery, but the later reciprocators were 
 designed 30 per cent, heavier than the early ones, with which some trouble 
 was experienced on account of lightness. The tendency to-day is to revert 
 back to twin screws for small craft driven by compound turbines, each com- 
 plete in itself, and capable of quite independent treatment, or by turbines 
 connected to the screw shafting by helical wheel gearing or electrical 
 generators and motors ; the impulse turbine being preferred as more con- 
 venient. 
 
 Piston Speeds. — In 1870, prior to the introduction of the compound 
 engine, the piston speed of an Atlantic passenger steamer was 400 feet per 
 minute, of smaller passenger steamers 350 feet, while fast cruisers had a 
 piston speed of 550 feet ; battleships the same, while small cruisers had 
 only 450 feet. 
 
 In 1880, before the advent of triple -compound engines, Atlantic and 
 ocean steamers generally had advanced to 600 feet, the smaller passenger 
 steamer to 450 feet, the cruiser to 600 feet, the battleship to 600 feet, and 
 small cruisers to 550 feet per minute. 
 
 In 1882 forced draught was introduced into the Navy, and by this means 
 the increased supply of steam permitted of higher speeds, so that in 1885 
 large cruisers had a piston speed as high as 660 feet, and small ones as high 
 as 750 feet per minute. At that time the triple engine had got a foothold 
 in the mercantile marine, and permitted of a piston speed of 750 feet in 
 ocean mail steamers and small passenger steamers 550 feet per minute. 
 
 By 1890 higher pressures of steam had been employed, and confidence in 
 the triple engine established ; consequently the piston speed for cruisers and 
 battleships was raised to 800 feet, ocean liners to the same speed, and small 
 passenger steamers in some few instances exceeded this. 
 
 Five years later battleships had a piston speed of 875 feet, cruisers a 
 little more ; while the newcomer, Destroyer, attained its very high speed by 
 moving the pistons at a velocity of 1,200 feet per minute. The Atlantic mail 
 and the small passenger steamer had made only very slight advances on 1890. 
 
 In 1905 battleships and cruisers were fitted with four-crank carefully 
 
BOILER AND MKAN PRESSURES. 721 
 
 balanced engines running at 960 feet, and small cruisers at 1,000 feet per 
 minute ; the Atlantic mail steamer at 975, and the small express steamer 
 at 950 feet. The cargo steamer pure and simple makes her voyages with 
 a piston speed of 500 to 660, depending on the size of the engine* 
 
 The increase in piston speed had been obtained by quickening the revolu- 
 tions from 70 to 120 in battleships, and from 75 to 140 in large cruisers ; 
 the greatest difference is, however, in small cruisers, the change being from 
 90 to 250 revolutions. In the mercantile marine the revolutions of an ocean 
 liner are 80 against 55, and in ordinary small passenger steamers 150 against 
 80 to 90 of thirty-five years ago. It is therefore not surprising to find that 
 the I.H.P. per ton of machinery in ocean mail steamers is only 6-75 as against 
 5 or 35 per cent, increase, while in large cruisers it is 12 against 7, or an increase 
 of 70 per cent. In small cruisers the I.H.P. was 19 per ton against 6 of 
 thirty-five years ago, or an increase of 216 per cent. It may be taken as 
 a rule that a reduction in weight per I.H.P. is obtained by increased piston 
 speed if it is due to high revolutions ; on the other hand, if piston speed is 
 got by lengthening the stroke, no saving is effected, but, on the contrary, 
 an increase may be looked for. 
 
 Boiler and Mean Pressures. — The general tendency of increase in boiler 
 pressure is to increase the weight of the engines as well as the boilers. If, 
 however, an increase of mean pressure is obtained by the increase of boiler 
 pressure, and the rate of increase be greater than that of the boiler pressure, 
 the weight of the engines per I.H.P. may be really decreased. Such, how- 
 ever, was not generally the case in the history of marine engineering progress. 
 In 1870 with a boiler pressure of 45 lbs. absolute, the mean pressure was 
 as much as 24 lbs. in the expansive surface condensing engines ; or 53 per 
 cent, of the initial pressure. In the compound engine of about that time 
 the referred mean pressure was only about 22 lbs. with a boiler pressure 
 75 lbs. absolute, or only 29 per cent. Later, with a boiler pressure of 105 lbs. 
 absolute, the mean in express steamers was only 30 lbs., or 28-5 per cent. 
 Later still, with the triple engine and a boiler pressure of 170 lbs. absolute, 
 the mean in cruisers of high speed was 38 lbs. or 22-4 per cent, and 32 or only 
 18-8 per cent, in merchant ships. Finally, with initial pressure of 250 lbs. 
 in the Navy the referred mean pressure was about 20 per cent., while in the 
 mercantile marine now with 215 lbs. it is 17 per cent. 
 
 The effect of increase of pressures is best seen by comparing the second- 
 class cruiser of 1898 with that of 1896. These engines developed practically 
 the same power at the same number of revolutions ; the design of each was 
 identical, but the cylinders were made smaller in diameter for the increase 
 in mean pressure, the referred mean pressure being 50-9 lbs. in one case and 
 40-3 in the other. As might be supposed, the weight of engines is practically 
 the same in each case. If, however, the boiler pressure had been increased 
 without an increase of referred mean pressure, the I.H.P. would have been 
 the same, but the weight of engines greater. If the boiler pressure had 
 remained unaltered but the referred mean pressure increased, there would 
 be an increase in I.H.P. without a corresponding increase in weight. It 
 may therefore be assumed that increases in boiler pressure only add to the 
 weight of the engines when there is no substantial increase in the referred 
 mean pressure. Also, that when there is a substantial increase in the referred 
 
 * P.y the N.E. Coast Standard Specification the piston speed ranges from 148 with 3 feet stroke to 
 560 with 4-75 feet stroke on service of cargo boats. 
 
 4G 
 
722 MANUAL OF MARINE ENGINEERING. 
 
 mean pressure, the weight of machinery per I.H.P. is not necessarily increased 
 by the increase in the boiler pressure. 
 
 Materials. — The weight of machinery is, of course, very considerably 
 influenced by the materials of which it is made. For example, the ordinary 
 mercantile engine is largely made of cast iron of a quality whose strength 
 does not exceed 7 tons per square inch. The engines of naval high-speed 
 ships have little cast iron in their construction, but what there is can with- 
 stand 11 tons per square inch, and some of it even as much as 13 tons ; 
 much of them is of cast or wrought steel, having a strength of 28 tons in 
 placo of the merchants ships' 7-ton cast iron. Again, in naval ships the 
 bronzes are extensively used in lieu of cast iron, and although Admiraltj 7 
 bronze is but little stronger than best cast iron, it is much more so than the 
 7 -ton castings, and consequently may be, and is from one-half to two-thirds 
 the thickness of the corresponding iron castings of the mercantile marine. 
 The light weight per I.H.P. of the cross-channel mail and passenger steamer 
 ia largely due to their having engines made of the same materials as naval 
 engines. New materials are frequently brought to the notice of engineers 
 by metallurgists and manufacturers which promise great things in the way 
 of saving weight, and some of them fulfil the promise, but others, from one 
 cause or another, disappoint both their discoverers and users. Aluminium 
 is an instance of this, and was very disappointing till Vickers, Ltd., introduced 
 Duralumin, which, while being very little heavier than aluminium, has a 
 strength and toughness far in advance of it, and sold in bars and sheets at 
 quite a moderate price; it cannot, however, be cast. Nickel steel and 
 venadium steel, also chrome nickel, are serving the turn of the naval designer 
 by withstanding the impact of shot ; they also serve the turn of the engineer 
 by withstanding high stresses — especially alternating ones — and any number 
 of repetitions of such shocks as come on many of the moving parts of the 
 marine engine. 
 
 Annealing and oil tempering, too, may be more extensively used to 
 improve steels as now made, or at anyrate to make them more uniform in 
 nature and with more of the valuable character as usually described by their 
 friends than as found always in practice. 
 
 The superior bronzes, too, may be used with advantage to save weight, 
 inasmuch as the ultimate strength of some is equal to that of steel ; more- 
 over, castings may be made of them much thinner than if of steel. Most 
 of these bronzes can be forged and rolled, when their strength is improved, 
 and they are rendered even more suitable for high-class engineering than 
 in the cast state. 
 
 Design has great influence on the weight of marine machinery, as may 
 have been supposed ; and what can be done by the most careful designing is 
 seen in the case of the machinery of the Destroyer, when 45 I.H.P. per ton 
 is developed, or even in that of the third-class cruisers, when 20 I.H.P. was 
 developed under forced draught, and even 145 under natural draught per 
 ton weight of reciprocating engines and boilers. 
 
 In the case of the Destroyer it will, of course, be urged that the saving 
 was effected largely at the expense of strength. This cannot be said of the 
 cruiser's engines any more than it can be of those of the cross-channel steamer. 
 
 It is true that not much saving can be effected in the principal working 
 parts, except by such means as boring the shafts and rods. No doubt uow, 
 
ENGINE PARTS SUBJECT TO INTERMITTENT STRESSES. 723 
 
 when improved tool steel and appliances admit of boring at a cheap rate. 
 hollow shafts will be used as freely in express steamers as in naval ones, and 
 with the same satisfaction. The great saving will be effected by change in 
 design of the fixed portions and in most of the smaller details being dealt 
 with drastically. It astonished many marine designers who turned from 
 large engines to consider the quick ones iu Destroyers to find bow much 
 material was practically wasted in the numerous unconsidered fittings and 
 fixings of an engine and engine-room. 
 
 Not much in any case can be done in reducing the weight of the parts 
 subject to alternating stresses, except by the use of materials better calcu- 
 lated to withstand them, such as special steel, etc., oil tempering, etc. 
 Some saving, however, can be made on those parts liable only to intermittent 
 stresses, especially when of a definite and easily calculable nature ; and 
 still more may be done in this direction on the parts subject to simple, steady 
 loads. 
 
 Table lxxxiii. gives the safe limit of the various metals used by engineers 
 under tUe several conditions obtaimng in practice, and the following observa- 
 tions may be taken as guides in dealing with the question of design : — 
 
 Effect of various Loads and Stresses. — The various parts of marine 
 machinery are subject to loads ot different natures, and each one, as a rule, 
 has to bear a repetition of its own particular load constantly and continu- 
 ously. The nature of the load should determine the magnitude of the greatest 
 stress w-hich the part subject to it should sustain. If it be a steady one 
 all materials will withstand much higher stresses than are possible with a 
 load more or less suddenly applied and removed continuously. Experience 
 has shown that in structures on land metal can be loaded so that the stress 
 is considerably more than half its elastic limit, if it is steady — that is ; if 
 there is no variation or infrequent application; on the other hand, it has 
 been found that if a load of such magnitude is applied and removed quickly 
 and continuously the metal soon gives w r ay. If the stress is lowered either 
 by reducing the load or increasing the section of material to sustain it, the 
 number of applications of load to produce fracture is larger ; if further 
 lowered, it takes a larger number still ; finally, it is found that if the load 
 does not stress the material beyond a certain fraction of the elastic limit, 
 the number of applications may be so large without producing fracture that 
 the load may be considered quite a safe one, and the corresponding stress 
 used for calculations when employing that material under similar conditions 
 Such loads as these may, for convenience, be called intermittent, and the 
 stresses produced intermittent stresses. 
 
 The parts of the engine subject to intermittent stresses are the piston-rod 
 ends in and beyond the pistons and cross-heads, the piston-rod and connecting- 
 rod caps and bolts, the cylinder cover studs, main bearing caps and bolts, the 
 bolts and studs of the valve gear, the guides, etc., the crank-shaft, the high- 
 pressure cylinder and its covers, the medium-pressure cylinder and its covers 
 generally, and all the column and framing bolts. The tunnel or intermediate 
 shafting theoretically has a steady load, but really it is moderately intermittent 
 — that is to say, the variation is not from zero to the maximum, but from a 
 minimum considerably above zero to the maximum. 
 
 If the load varies from a positive to a negative value — that is, from tension 
 to compression and from compression to tension — the load is said to be an 
 
724 
 
 MANUAL OF MARINE ENGINEERING. 
 
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 continuously is nearly the same in its destructive effect as a positive load of 
 the same magnitude as the two jointly. That is, if a piston-rod is subject to 
 
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ELASTIC LIMIT OR YIELD POIN*. 725 
 
 a thrust of 20 tons and a pull of 15 tons or vice vend, the effect on the material 
 is nearly the same as if it were subject to an intermittent load of 35 tons. 
 It is, however, not quite the same, and further experiment with modern 
 metals may show that at moderate loads the difference between alternating 
 and intermittent is not so great. In any case, material subject to alternating 
 loads must not be stressed so highly for any one direction as if those loads 
 were intermittent ; in fact, from Wohler's experiments the stress should not 
 be much more than half. The rapidity of intermission and alternation is 
 also an important factor in determining safe stresses for materials. 
 
 Further, the shape of the part subject to load is of importance ; if ther» 
 are incisions similar to that of a screw thread, which suddenly change and 
 concentrate stress and invite fractures the stresses should be much lower than 
 those of a plain uniform section, such as that of a bar could safely bear. 
 
 The parts of an engine subject to alternating stresses are the piston-rods, 
 connecting-rods, cross-heads, columns, pistons, low-pressure cylinder, its 
 cover and bottom, the medium-pressure cylinder, etc., when at low powers, 
 paddle shafts and screw shafts at their outer ends due to the weight of pro- 
 pellers and the reaction of floats, or blades, especially in the case of a screw 
 ship when racing or running light. The valve motions, the air-pump levers, 
 rods, etc., are subject to partially alternating stresses. 
 
 Table lxxxiii. gives the safe limits to which an engineer may go under the 
 various conditions, if weight is of prime importance in designing engines for 
 continuous work. If weight is not an object of first importance a reduction 
 of 10 per cent, may with advantage be made. If the article is nicked, sud- 
 denly changed in section, or has a screw thread, a reduction of 25 per cent, 
 should always be allowed. In the case of alternating stresses the equivalent 
 force one way only is taken, for calculation — viz., tension. 
 
 Shearing and Torsion. — Certain metals have a very high power of resist 
 ance to tension and compression without the like ability to resist shearing ; 
 for example, naval brass has a coefficient of shear of only 55*8 per cent. — that 
 is, its strength against shearing is little more than half what it possesses 
 against tension ; on the other hand, cast iron has a coefficient of 162 per cent., 
 and phosphor bronze castings 113 per cent. The coefficient of wrought steel 
 and iron is 80, while that of most of the strong zinc bronzes is, as a rule, only 
 little more than naval brass. They are consequently not very fit for such 
 purposes as shafting, where shearing forces are always of importance, especially 
 so in the case of tunnel shafting. 
 
 The Elastic Limit or Yield Point. — It is probable that every material 
 which has never been subject to direct tension will have some degree of 
 permanent set after application of any load ; it will, of course, be very small, 
 and generally microscopic. If, however, the material is stressed nearly to 
 what we call the elastic limit, released, and again subjected to stress, there 
 should be no further permanent set till that point (the elastic limit) is exceeded. 
 
 When this limit is found there is reason to suppose that so long as no 
 •portion of the material is subject to a stress in excess of it, the material should 
 endure for ever. Unfortunately materials do give way from what is called 
 " fatigue '- without apparently having been subject at any time to a stress 
 nearly approaching the supposed elastic limit. It follows, then, that the 
 safe iimit of continuous stress is considerably below the elastic limit, or that 
 the loads on some -portion of such structures as have collapsed, are much 
 
726 MANUAL OF MARINE ENGINEERING. 
 
 greater than generally supposed. It is highly probable that the latter is the 
 case, and to a great extent the cause of accident. In a complex structure 
 it is nearly impossible to exactly proportion each part to its load, and 
 absolutely impossible to spread the sustension of load evenly over every inch 
 of area of section. Moreover, in the process of manufacture initial stresses 
 are often set up in many places, all of which are of uncertain magnitude 
 and direction ; to these are added the stresses produced by working load, 
 which may not be, and probably are not, distributed in accordance with 
 theory. It is, therefore, more than improbable that a structure designed 
 with a theoretical margin or factor of safety of ample magnitude has in one 
 or more places material stressed dangerously near to the elastic limit, and 
 perhaps sometimes even beyond it in some cases. 
 
 Safe Working Stresses. — The limits given in Table Ixxxiii. may be used 
 when the maximum stresses are calculable and distributed uniformly over 
 the sections. For example, a weight suspended by a bar having an eye at 
 each end, so as to hang freely, will cause a uniform stress on the transverse 
 sections ; if instead of eyes forged with the rod the rod was screwed and 
 fitted with nuts whose faces are rough and the surfaces on which the nuts 
 sit are uneven, the stress will not be uniform, and may be much greater 
 on one side than the other. Again, the sections of a round or square bar 
 loaded and subject to bending either as a beam or cantilever are not uni- 
 formly stressed, the outer layer, or that furthest from the neutral axis, being 
 much more so than those nearer it ; but at the same time, the layer being 
 firmly attached to its neighbour, receives support from it in a different way 
 from what is the case of a fagot of strips free to slip one on the other as in 
 a carriage spring. The rules for estimating are based, however, practically 
 on the latter supposition, hence what would seem to be a too high stress can 
 be safely borne by such parts. 
 
 In the case of a complex article like a crank-shaft subject to complicated 
 loads, it is impossible to estimate the maximum intensity of stress at any 
 and every part of a section. It is certain that at very few, if any, places 
 the stress is uniform over the whole section at such places ; and at some 
 the intensity of stress must be great on portions of the sections. More- 
 over, the quality of the material is not uniform throughout such a forging, 
 the strength across the direction of flow of metal being less than given in 
 Table Ixxxiii. Under such circumstances the nominal stress adopted for 
 calculations should be much less than those given in that table. For similar 
 parts of light machinery the nominal stress may be taken at 60 per cent, 
 of those given in this table, and for those of ordinary machinery 35 to 40 per 
 cent. In practice, therefore, the designer must carefully consider each 
 case by itself, with all its conditions and surroundings, to determine the 
 nominal stress on which to calculate the dimensions of the parts which must 
 sustain the load. 
 
 Stretch under Tension. — Eigidity of structure is often of prime import- 
 ance, and generally so in the engine-framing, guides, etc. All materials 
 stretch more or less under stress, and the lower the stress the less is the 
 stretch ; consequently it is often found necessary to make details of the 
 structure larger than necessary for mere strength, in order to avoid the 
 stretching of them, which would be inconsistent with rigidity or stiffness of 
 form. It is generally forgotten that up to 7 or 8 tons' tension per square 
 
HIGH-SPEED CRAFT. 727 
 
 inch cast iron stretches more than steel or wrought iron, and that the material 
 has practically no elastic limit, for it has no permanent set up to fracture. 
 
 Again, there are other conditions which must not be overlooked, as, for 
 instance, in the case of castings as distinguished from forgings or other 
 wrought forms of material. The latter are practically free from internal 
 defects * when machined and made part of an engine ; the castings, however, 
 may be solid, but are often not so, and must not be assumed to be. To allow 
 for such contingencies a reduction of 10 to 20 per cent, should be made from 
 the stress given by sound bars in the testing machine as the elastic limit of 
 castings. The figures in Table lxxxiii. for castings are based on this allow- 
 ance. Again, rough forgings often have incipient flaws arising from the 
 indenting of the hammer in forging, and, in the case of steel, from small gas 
 bubbles and scoria near the surface of the ingot, and sometimes from slag 
 getting forged on the metal. From the bright working parts such imper- 
 fections are machined away or are cut out before finish forging is done. 
 
 Another serious consideration in determining the nominal stress for 
 calculation is the purpose the engine has to serve. Express steamers, such 
 as passenger and mail boats, have to exert 80 to 90 per cent, of the trial or 
 maximum power for 50 to 70 per cent, of the year ; cross-channel steamers 
 (short runs) steam about 25 to 35 per cent, of the year ; while excursion and 
 river steamers do from 30 to 50 per cent. They all have occasionally, and 
 are always liable to have, to run at full power. Every part of the machinery 
 of such ships should, therefore, be much lower stressed at full power than the 
 corresponding parts of machinery which is seldom or never worked to the 
 full power for which it was designed, whose full speed on service is done with 
 only 75 per cent, of such power, and that for only 10 per cent, of each year. 
 
 The speed of the engine also affects the problem. A slow-working paddle- 
 engine, with few reversals of motion and time for " cushioning " and other 
 means of avoiding shock, does not jar its component parts as does a high- 
 speed one, the momentum of whose moving parts materially influences the 
 stresses on them and their connections, which stresses are most suddenly 
 applied and removed or reversed many more times in a minute, and which 
 may be increased by water in the cylinder to an unknown extent, and from 
 want of time and other causes may not get the cushioning at the right periods 
 to do good and ease the concussions. Such engines, however, are generally 
 required to run at full speed for very short spells and those far apart, and 
 may, therefore, have lighter scantlings than would be necessary for a similar 
 engine on shore, such as an electric light or power engine, which has to run 
 every day for five to nine hours, or even more. 
 
 Paddle Engines, when required to be very light, may be designed by 
 referring to the stresses given under 60 revolutions and modifying on the 
 lines laid down above. When weight is not of so much importance as endur- 
 ance, the stresses may be reduced by 10 per cent, all round. 
 
 Naval Engines, Express Steamer Engines, and Passenger Cargo Steamer 
 Engines may be designed by taking the figures under 200 revolutions and 
 allowing 10 per cent, reduction all round, and modifying on the lines laid down. 
 
 For" High-speed Craft, such as Destroyers, Scouts, Third-class Cruisers, 
 etc., the column under 300 revolutions, with the modifications, but without 
 the all-round 10 per cent, reduction, may be followed. 
 
 * This is not true of larse shafts, for the steel near the axis is never as good as that near the surface 
 and sometimes is very poor indeed, hence the boring out Is really no practical loss of strength. 
 
728 MANUAL OF MARINE ENGINEERING. 
 
 For Tramp Steamers, whose power is small, so that the engines are virtually 
 always working at full power, and made of material untested and generally of 
 common quality, a reduction of 20 per cent, from column 100 or 200 revolutions 
 is advisable when calculating the various parts of the engines and machinery. 
 
 Boilers, etc., form the most important part of the machinery from the 
 weight point of view, and the possible saving is greater than from any other 
 part, as has been seen when water-tube boilers have been substituted for 
 cylindrical, and earlier when steel shell plates took the place of iron ones. 
 The Admiralty have reduced the thickness of these plates till little remains 
 possible in that direction with ordinary steel, and it is understood that the 
 Board of Trade, Lloyd's, and other authorities will shortly follow suit* It 
 is also equally possible now that a material as superior to the ordinary mild 
 steel as it was to iron will soon be placed on the market for the use of boiler- 
 makers. It is possible, however, to make a much lighter boiler with the 
 ordinary steel without running any risk beyond what already exists. For 
 example, the steel bars for stays are, when the boilers are in use. subject to 
 a steady continuous load, gently applied and released ; they are tested 
 mechanically under similar conditions to a stress of 30 tons per square inch ; 
 but when at work are stressed only to 4 tons, whereas 6 tons would not be 
 excessive, especially in these days of non-corrosion in the inside of boilers. 
 With a reduction in the size of stays a saving would be made on the nuts, 
 etc. The inner plating is also heavier than it need be under the above con- 
 ditions. The heads and snaps of rivets are made regardless of weight, and 
 of the fact that most of the rivets are subject to shear only. For cleaning 
 purposes boilers were and are still designed with room in various parts suffi- 
 cient to admit a man and permit him to do the work. This means a larger 
 boiler and much more water than is necessary. Now, with salt water cut 
 off and practically pure water used, there cannot be the risk of scaling, 
 especially in short voyages, to warrant such extravagance in space. Railway 
 locomotives use hard water and water containing much solid matter, but no 
 such provision is made as in ordinary marine boilers for excavating scale. 
 Of course where there is a strong probability of salt water being used some 
 such provision must be made, but this would be only in ships where weight 
 of machinery is unimportant and long voyages the rule. 
 
 Cost. — Although less material is used when the weight of an engine is 
 cut down, the value of that saved is far exceeded, as a rule, by the cost of 
 labour employed in effecting the reduction. The substitution of one material 
 for another with the object of saving weight invariably means a more costly 
 one. The removal of surplus materials from all forgings, and some of that 
 even from castings, is a costly process, although the modern machine tool, 
 with the high-class self-hardening steels now in use, permits of it being done 
 at a much cheaper rate than formerly obtained. Also, it is generally found 
 that the reducing to a minimum of the sizes of flanges, facings, etc., entails 
 much more care in moulding, marking off, and machining, with the conse- 
 quent greater cost of labour ; besides which, more costly methods of doing 
 work are necessitated by the want of room in carrying it out. Then, too, 
 the much higher number of revolutions resorted to with the object of reducing 
 weight per I.H.P. requires much greater care in the manufacture and ad- 
 justment of all the working parts, and the use of fittings and appliances 
 which were done without before. 
 
 * The water-tube boiler is exclusively used now in naval ships, and is gradually being supplied to 
 express steamers of all kinds where weight is of serious consequence. 
 
AUXILIARIES AND APPURTENANCES. 729 
 
 In 1880 a compound naval engine cost about £12 per I.H.P., or £75 
 per ton. Five years later triple engines of large size were £6-4 and small 
 £8 per I.H.P., and still about £75 per ton. In 1890, the cost of cruisers' 
 machinery was about £7 per I.H.P. and £90 per ton. Later on large cruiser 
 machinery with Belleville boilers cost about £12 per I.H.P. and £130 per 
 ton. The more modern with less expensive water-tube boilers cost somewhat 
 less. 
 
 The very light naval engines of the past few years illustrate what can 
 be done in the way of saving weight when cost is of secondary considera- 
 tion. At the same time it will be seen that the cost per I.H.P. of trial at 
 full power is really low, and even if the lower power or that developed without 
 forcing be taken, it is seen that these engines by comparison are by no means 
 costly. For example, the third-class cruiser engines of 1891, with cylindrical 
 boilers, cost £6-75 per I.H.P. full power, at natural draught it was only £10, 
 or £93 per ton. The third-class cruiser of 1897 had water-tube boilers and 
 engines costing £6-5 per I.H.P. of full power, or £9 of that at natural draught ; 
 this was, however, £118 per ton. The destroyer engines of 1895 cost only 
 £6 per I.H.P., but £200 per ton, and later ones cost as much as £236 per 
 ton. 
 
 The machinery as now fitted into naval ships is much more elaborate 
 than formerly obtained, and turbines are employed to drive the screws. 
 Battleship machinery, including auxiliaries, in 1913 cost about £9-5 per 
 S.H.P., while that of first-class cruisers cost £8-5 to £7, depending largely 
 on the rate of revolution. The machinery of second-class cruisers and large 
 scouts cost about £6-8 to £7 per S.H.P., While that of destroyers was somewhat 
 less, due to the higher rate of revolution. The cost per ton was about £120 
 for battleships and £145 for cruisers. Destroyers and similar craft had then 
 machinery costing about £220 to £240 per ton. 
 
 Taking similar periods for the mercantile marine, the average cost of a 
 compound engine in 1880 was £7-5 per I.H.P., or £35 per ton. Five years 
 later the triple engine was costing £6-5 per I.H.P., or £36 per ton. In 1890, 
 when prices were high in consequence of the Naval Defence Act having filled 
 most of the shipyards and engine works with orders, engines could be got 
 for under £6 per I.H.P., but still £36 per ton. Ten years later the engines 
 of cross-channel steamers, whose weight would be only about 220 lbs. per 
 I.H.P., could be purchased at £5-5 per I.H.P., or £53 per ton, and heavy 
 engines at about £6-5 per I.H.P., and £35 per ton. 
 
 Comparing the cross-channel boats" machinery with that of the 1891 
 third-class cruiser, it will be seen that per I.H.P. they cost the same money, 
 but the cruiser engines weigh only 156 lbs. against the 220 of the merchant 
 ship. The latter cost, therefore, only £53 per ton against the £93 of the 
 Naval ship. 
 
 Auxiliaries and Appurtenances. — Forty years ago the marine engine 
 was practically self-contained ; it had as a parasite an auxiliary feed donkey 
 pump, and in large ships two such pumps. The extensive use of water ballast 
 later on necessitated another and a larger donkey pump, and feed heaters 
 were found to be good things ; centrifugal circulating pumps also came to 
 be generally used in large ships somewhat earlier. In the Navy, in conse- 
 quence of the high revolutions, it was found better to have independent 
 feed pumps, and, later, to take from the main engine all pumps except the 
 
730 MANUAL OF MARINE ENGINEERING. 
 
 air pump. This practice was soon followed in the mercantile marine with 
 fast running engines. Then fresh water for making up the water from the 
 boilers was deemed necessary, and evaporators or distillers were used. 
 Starting and turning engines were known then, but have been used very 
 generally since ; they are now regular fittings in ships of all sizes. Feed- 
 water filters are common where ships are engaged in long voyages and use 
 oil in the cylinders. All these things add to the weight of the machinery 
 without contributing anything directly to the I.H.P. used as divisor. The 
 I.H.P. developed in the auxiliary machinery of a warship is now enormous. 
 In a large merchant steamer, especially long voyage express ones, as already 
 fully set forth in Chap, xx., the auxiliaries add very much to the weight 
 and steam consumption. 
 
 Spare Gear also adds to the weight, and is more in extent to-day than 
 was formerly the case in the mercantile marine ; but less in the Navy, where 
 in old days nearly half an engine was carried (v. Appendix D). 
 
 Fuel. — The consumption of fuel is another most important factor in all 
 considerations of weight ; especially is this the case where, as in the North 
 Atlantic, the distance between coaling stations is long. It is of little use 
 paring down the weight of the machinery of ships on this service, if by so 
 doing the fuel consumed per I.H.P. is increased. The great cry of the oppo- 
 nents of the water-tube boiler was that the increased consumption of fuel 
 outbalanced the saving in weight, so that, for a given voyage, the total weight 
 of machinery and fuel together was greater with a water-tube boiler instal- 
 lation than with a cylindrical one with Howden's or Ellis' forced draught. 
 Forty years ago 2-5 lbs. per I.H.P. was the general rate of consumption of 
 good coal ; a little later it was as low as 2 lbs. with the very large three- 
 crank compound engines and high pressure of steam. The triple engine 
 in its early days consumed 1-7 lbs. in a general way, and 1-5 under best 
 conditions. Now the best engines with Howden's system, superheaters, 
 feed-heaters, etc., consume if triple-compound 1-4, if quadruple 1-25, if geared 
 turbines 1-15 lbs. of good mercantile coal per hour. 
 
 In the Navy the consumption of coal (including auxiliaries) was with 
 triple-compound engines and the cylindrical boiler 1-8 lbs. of Welsh coal 
 when running full speed. Now with turbines and water-tube boilers, etc., 
 it is from 1-55 to 1-70 on service. If the I.H.P. of the auxiliaries were included 
 the rate would, of course, be somewhat lower in both cases. 
 
 The modern naval engine when reciprocating has a low rate of expansion 
 when working at full power, as already explained, consequently the fuel 
 consumption per I.H.P. compares unfavourably with that of a merchant 
 ship ; but the conditions are dissimilar in more ways than pertain to the 
 main engines ; if the main engines and their own particular auxiliaries only 
 are using steam from the boilers the consumption of good Welsh coal does not 
 exceed 1-85 lbs. per I.H.P. The consumption of fuel in recent ships having 
 turbines only is about 1-87 lbs. per S.H.P., which would be equivalent to 
 1*758 lbs. per I.H.P., assuming an efficiency of 94 per cent. In the cruiser 
 "Dartmouth" it was only 1-48 lbs. at full speed, and 1-72 lbs. at 90 per 
 cent, of full power. The water consumption in scouts running at 500 to 
 600 revolutions per minute was only 14-85 lbs. at full power, and 15-5 lbs. 
 at half power. H.M.S. "Gloucester" consumed 1-48 lbs. of fuel at full 
 power of 19,000 S.H.P., and at 14,000 it was 1-59 lbs. 
 
WEIGHT OF MARINE MACHINERY. 731 
 
 The Weight of Modern Marine Machinery Installations of all kinds varies 
 with the rate of revolution of the motor, whether it be a reciprocator or a 
 turbine, when all other things are pari passu ; it may be arrived at with a 
 very fair approach to accuracy by the following empirical formula? where 
 I.H.P. and S.H.P. are the mean powers developed on trials at full speed for 
 a period exceeding one hour continuous steaming at revolutions R per minute. 
 Weight found in this way will form a sufficient guide for the preliminary 
 design of a ship, but must not be taken as being quite so accurate as to pre- 
 clude the necessity for a proper estimation. In naval ships and express 
 passenger steamers having light or small tube water-tube boilers of the 
 Yarrow or other similar type, and with the arrangement and supply of 
 auxiliaries usual in such ships. 
 
 , • , t vi v S.H.P. X 16 ' 85 * 
 
 The total weight of machinery with turbines = - — p _ im - tons. 
 
 , . , , ,. , . . I.H.P. x 7-2 L 
 
 The total weight of machinery with reciprocatois = _ inn 
 
 Naval ships with reciprocating engines and large or heavier type of 
 water-tube boilers, such as the Babcock, Belleville. Niclausse, etc. 
 
 m * i ■ ^ t u- I.H.P. X 145 
 
 Total weight of machinery = — — tons. 
 
 Mercantile ships having cylindrical boilers and the installation of auxili- 
 aries usually found in them and large naval ships with similar boilers. 
 
 Total weight of machinery = ' ' ' — tons. 
 
 (1) Mail steamers with reciprocatois of heavy design, . Q = 20 -5 
 
 (2) Mail and naval ships with reciprocators of light single-ended 
 
 boilers, . . . . . . . . Q = 20 "0 
 
 (3) Mail and naval ships with reciprocators of light double-ended 
 
 boilers, - Q - 19*5 
 
 (4) Mercantile ships having turbines and cylindrical boilers, . Q = 44 '0 
 
 (5) Mercantile ships, such as tramp steamers with heavy 
 
 machinery, . . . . . . . Q = 27 
 
 Example (1). — To find the weight of machinery of a battleship having 
 turbines and water-tube boilers, the S.H.P. is 25,000 at 300 revolutions per 
 minute. 
 
 _ . ,, 25,000 x 16-85 01rt _, 
 Wei § ht = 300-100 = 2 ' 10 ' t0nS - 
 
 Example (2). — The weight of a scout's machinery developing 16,000 S.H.P. 
 at 750 revolutions per minute. 
 
 m . Ui 16,000 x 16-85 , 1K , 
 " "ght = 750 _ioo = *15tons. 
 
 Example (3). — An Atlantic steamer has turbine machinery and cylindrical 
 boilers, the S.H.P. is 30.000 at 180 revolutions. 
 
 „ T . , ± 30,000 x 44 _ „ Q , 
 Weights 18Q+50 =5,, 39 tons. 
 
732 MANUAL OF MARINE ENGINEERING. 
 
 Example (4). — An express steamer having double-ended boilers and engines 
 •of 7,000 (.HP. at 180 revolutions per minute. 
 
 , r , .'.. 7,000 x- 19-5 KQftx 
 
 The weight is = -~k- ■—,, - = o30 tons. 
 
 n 180 + o0 
 
 Example (5). — A tramp steamer of 1.750 LH.P. at 70 revolutions per 
 minute. 
 
 The weight of machinery = -^ j~Kn~ " ^4. tons. 
 
 Relation of Weight to Tonnage. — Taking the same period as before, it is 
 interesting to note that in Atlantic mail steamers the machinery was 13 per 
 ■cent, of the displacement, and the I.H.P. at the rate of 1*1 per ton of gross 
 register. In small passenger steamers it was about 13 per cent. ; the I.H.P. 
 1 "35 per gross register ton. 
 
 Ten years later the speed of ocean steamers had increased by 20 per 
 cent. The weight of the machinery was still about 13 per cent, of the dis- 
 placement, but the I.H.P. was 1*8 per gross register ton. 
 
 The increase of speed with the smaller steamers was 25 per cent., the 
 weight of machinery 15 to 19 per cent, of the displacement, and the I.H.P. 
 1 "5 to 1 "7 per gross register ton. 
 
 In 1895 the Atlantic express steamer had a speed of 21 knots obtained 
 with machinery whose weight was 20 per cent, of the displacement, and 
 whose I.H.P. was at the rate of 1 *9 per gross register ton. Twin screw cargo 
 and passenger boats with a speed of 15 knots of this period had a weight 
 of machinery of only 7*3 per cent, of displacement and less than 1 I.H.P. 
 per gross register ton, while in similar boats of 13 knots speed the proportions 
 were 5 per cent, and # 5 I.H.P. 
 
 Fast Channel screw steamers of 17| knots were at the same time on 
 service with machinery weighing 27 per cent, of the displacement, and 
 indicating 3 to 5 H.P. per gross register ton. 
 
 The North Atlantic Service is carried on to-day by express steamers driven 
 by turbines at a speed of 25 knots per hour on the New York route, and by 
 others driven by the reciprocating engine at 22 '5 to 23*5 knots, and, lastly, 
 by even larger steamers, like the " Olympic," driven by a combination of 
 reciprocators and turbines at 22*5 knots. The Boston and Canadian service 
 ■employs ships somewhat smaller, having a speed of 19 to 20 knots and driven 
 by turbines, while the older steamers preceding these more modern ones were 
 driven by reciprocators at about the same speeds. 
 
 The machinery of the largest and fastest Atlantic expresses of the turbine 
 type requires about 30 per cent, of the displacement to carry it, and will 
 develop a gross power equal to 2*15 S.H.P. per gross register ton, or 1*90 
 S.H.P. per ton of gross displacement. That of the " Olympic," develops 
 1-1 S.H.P. per displacement ton, and the weight is 17 per cent, ot the 
 displacement. 
 
 The largest Atlantic expresses having reciprocating engines require 23 per 
 cent, of the displacement to carry the weight of the machinery and develop about 
 2-20 I.H.P. per gross register ton, or about 1 -50 I.H.P. per ton of gross displace- 
 ment. The turbine-driven Canadian and other expresses running at somewhat 
 less speeds require only about 16 per cent, of the displacement, and develop 
 •only 1-20 S.H.P. per gross register ton, or 0-75 per ton of gross displacement. 
 
 The " Lusitania " and "Mauritania" of the Cunard Line were a great 
 
THE CRUISERS OF LARGE SIZE. 735 
 
 advance on any steamship then existing, being 760 feet long, 87-5 feet beam, 
 and having a displacement of 36,440 tons on 32-5 feet mean draught of water ; 
 " Olympic " and her sister ship " Britannic," whose dimensions are, length 
 853 feet, beam 92-75 and 94 feet, and the latter's displacement is 53,000 tons. 
 The machinery of these ships is the combination of two reciprocating triplex 
 engines, having cylinders 54 inches, 84 inches, and two low pressures of 
 97 inches diameter, with a piston stroke of 75 inches running at about 80 
 revolutions, with a huge low-pressure turbine driving the middle screw, as shown 
 in figs. 43 and 44, making about 170 revolutions per minute. 
 
 Express steamers of still larger size are the " Aquitania," of 52,000 tone 
 displacement, 870 feet long, and 97 feet beam, and the " Vaterland," 55,000 
 tons, 908 feet long, 100 feet beam, and a speed about 25 knots. 
 
 The Mail and Passenger Steamers on service to India, The Cape, and 
 Australia are also much larger than those engaged in these routes only a few 
 years ago, and although their speed is not so high as in the New York service 
 there is a marked increase. They are generally about 10,000 to 12,500 tons 
 gross register, and steam about 18 knots, being driven by twin-screws and 
 quadruple -compound engines. Their length is usually 500 to 550 feet, with 
 a beam about 60 to 64 feet, and draught of water 27 to 28 feet for the Suez 
 Canal, and rather more on the Cape route. Recently some ships have been 
 placed on the Indian route driven by turbines, and the Union Company, of 
 New Zealand, adopt the mixed method of propulsion in their new steamers. 
 
 In 1880 a Naval First-class Cruiser attained a speed of 18-5 knots with 
 machinery whose weight was 27 per cent, of the displacement. Ten years 
 later the same speed was recorded with triple-compound engines only 15-5 per 
 cent, in weight. In 1910, for a speed of 24 knots, 17-7 per cent, sufficed, 
 while in second-class cruisers it was 15 per cent., and in third-class cruisers 
 of 24 knots with reciprocators and water-tube boilers 16-8 per cent. Now, 
 with turbines and water-tube boilers wholly, the machinery of a Battleship 
 is 15 per cent, of her displacement, develops 2-11 S.H.P. per displacement 
 ton, and drives her at 25 knots. The Battle Cruiser of 30 knots requires 
 20 per cent, of the displacement, and develops 3 S.H.P. The 30-Jcnot Light 
 Cruiser with turbines, etc., has machinery whose weight is 44 per cent, of her 
 displacement and S.H.P. is 11 per displacement ton, while the somewhat 
 larger 26-knot Cruiser has machinery 22 per cent., giving 5-5 S.H.P. per ton. 
 The 30-kvot Destroyer with reciprocating engines required 46 per cent, of the 
 displacement for machinery which developed 17 I.II.P. per displacement ton. 
 The modern 32- to 34-knot destroyers have turbine machinery, whose weight 
 is only 30 per cent, of the displacement, and gives 18 S.H.P. per ton displace- 
 ment and about 60 S.H.P. per ton weight of machinery. 
 
 The Modern Battleship is much beyond what was considered unwieldy 
 a few years ago, being now 650 feet long, 92 feet beam, 28,000 tons displace- 
 ment, about 90,000 S.H.P., giving a speed of 25 knots. 
 
 The Battle Cruiser is larger still, being 660 feet long, 90 feet beam, and 
 30,000 tons displacement ; her speed is 30 knots by 87,000 S.H.P. Even 
 larger Light Cruisers have been recently built over 900 feet long, with a speed 
 of 35 knots. Turbines and water-tube boilers are in all these modern warships. 
 
 The Light Cruiser is now quite a big ship, being 420 feet long, 40 feet beam, 
 3,800 ton6 displacement, having engines developing more than 40,000 S.H.P., 
 and driving them at over 30 knots. 
 
 The Cruisers of large size of 1907 were 520 feet long. 74-5 feet beam, 14.600 
 
734 MANUAL OF MARINE ENGINEERING. 
 
 tons displacement, haviug engines of 28,500 I.H.P., triple-expansion, with 
 water-tube boilers, their speed being 23 knots. The I.H.P. per ton of dis- 
 placement 1-98, and tbeir weight 16-3 per cent, of the displacement. 
 
 Scouts and Destroyers are also of much larger size now than obtained a 
 few years ago, and some of them are of exceedingly high speed. H.M.S. 
 " Swift " is a good example of the development that has taken place of late 
 years to attain high speed with the ability to keep the sea in any weather ; 
 she is 345 feet long, 34-2 feet beam, and has a displacement of 2,170 tons 
 on a mean draught of 10*5 feet; she has four screws worked by turbines 
 developing over 30,000 S.H.P., which drive her at about 36 knots. H.M.S. 
 " Tartar," an ocean-going destroyer of 860 tons displacement and a length of 
 270 feet, has attained a mean speed of 35*4 knots with a little over 14,000 
 S.H.P. driving three screws. Destroyers are now over 300 feet long. 
 
 The Cross-Channel Express Steamer remains at about the same speed 
 as obtained before the advent of the turbine with the fastest of their class ; 
 but here, too, there has been a very considerable increase in size of ship 
 during recent years ; this is especially so in the English Channel routes. 
 In the Irish Channel express routes the new turbine steamers are 351 feet X 
 41*1 feet X 24 - 5 feet deep, of 2,500 tons gross register, and attain a speed of 
 22 "5 knots with three screws ; the older boats on the Holyhead and Kingston 
 route are quite as large — viz.. 360 feet X 41*5 feet X 27*3 feet deep, with 
 a gross tonnage of 2,641, and a speed of 23*5 knots, obtained with twin screws 
 and reciprocating engines. In the Isle of Man service the turbine steamer, 
 " Ben-ma-Chree," is of 3,353 tons displacement, no less than 375 feet long, and 
 46 feet beam, and has a speed of 23 -27 knots on service, whereas the paddle 
 steamer, " Empress Queen" (fig. 8), was 360 feet long, with a displacement of 
 3,015 tons, and a speed of 21'7 knots. In the Dover-Calais and Folkestone- 
 Boulogne routes the turbine steamers are now of 2,400 tons displacement, 
 350 feet long, and 42 feet beam, and attain a speed of 23 # 5 knots against 
 the 1,700 tons, 338 feet and 34 '7 feet beam with 21*5 knots of 1900. 
 
 In the Clyde Estuary and other similar service the three-screw turbine 
 steamer has been taking the place of the paddle steamer, and fig. 6 is a good 
 instance of this kind of ship. The speed on this service is quite high enough 
 at 20 to 21 knots, but it is very likely to be increased to even 24 on the longer 
 routes requiring the passage in a day. 
 
 In Shallow Water of Rivers, Estuaries, and Lakes the paddle-wheel, especi- 
 ally as fitted at the stern, continues to be the favourite for high speed or 
 great power for towing, especially in America, where much of the passenger 
 traffic is still water-borne. On the great lakes there are huge paddle-boats 
 similar to that shown in fig. 5, but the twin-screw ship is entering into com- 
 petition, and sooner or later the three- and four-screw turbine ship will enter 
 the field. On the big rivers the stern wheeler has a strong hold on the respect 
 of the shipowner. On the Mississippi there are such boats 300 feet long. 42"6 
 feet beam, and only 7 feet deep, having large passenger accommodation, and a 
 speed of nearly 18 miles per hour. The engines are simple expansive-condensing, 
 having cylinders 30 inches diameter and piston stroke of 120 inches. 
 
 The weight of paddle-wheel machinery is greater than that of the screw, 
 but with modern designs 8 to 9 I.H.P. is developed per ton, as against the 
 6 to 7 I.H.P. of the light engines and low-pressure steam of thirty years ago. 
 
 When circumstances permit of the adoption of the special stern arrange- 
 ments, as shown in figs. 1 and 2, twin screws of small size and high revolutions 
 can be used with great advantage, so far as weight is concerned. 
 
MARINE MACHINERY. 
 
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36 
 
 MANUAL OF MARINE ENGINEERING. 
 
 TABLE LXXXV. — Particulars of Naval Ships' Machinery 
 
 Designation. 
 
 When 
 
 Built. 
 
 00 
 
 s 
 
 p 
 
 u 
 GO 
 
 
 
 Cylinders. 
 
 Full Spued 
 
 
 Kind of Engine*. 
 
 
 
 
 ~ 4» 
 
 •6 
 I 
 
 "2 s 
 
 
 
 
 3 
 V 
 
 
 
 Diameter. 
 
 
 II 
 
 s. 
 QQ 
 
 1 1 
 
 
 
 
 — 
 
 s 
 
 
 J3 
 
 a 
 
 
 9 
 
 a 
 
 •si 
 
 > 
 
 I 
 
 ~ 3 
 
 
 
 
 3 
 
 
 3 
 7i 
 
 
 s 
 
 0Q 
 
 
 £ 
 
 9 
 
 Lbs. 
 
 
 
 
 
 
 Ins. 
 
 Ins. 
 
 
 Ft. 
 
 
 Battleship 
 
 1870 
 
 1 
 
 Horizontal Trunk Surface-Condensing, 
 
 2 
 
 127 (= 118) 
 
 54 
 
 71-5 
 
 643 
 
 20-1 
 
 
 Do., 
 
 1870 
 
 1 
 
 ( Do. Return Connecting Rod Surface \ 
 ( Condensing, - - - -/ 
 
 2 
 
 120 
 
 54 
 
 627 
 
 564 
 
 19-5 
 
 
 Do., 
 
 1870 
 
 1 
 
 Do. do. do. do., 
 
 2 
 
 98 
 
 48 
 
 68-4 
 
 547 
 
 19-6 
 
 
 Do., 
 
 1S70 
 
 2 
 
 Do. do. do. do.. 
 
 4 
 
 72 
 
 36 
 
 72-2 
 
 433 
 
 230 
 
 
 Do., 
 
 ISsO 
 
 2 
 
 Vertical Direct-Acting Compound, 
 
 
 4 
 
 60-104 
 
 42 
 
 82-1 
 
 575 
 
 22-3 
 
 
 Do., 
 
 1S85 
 
 2 
 
 Do. do. do., 
 
 
 6 
 
 77-55-77 
 
 48 
 
 i-0-0 
 
 640 
 
 25-8 
 
 
 Do., 
 
 1888 
 
 2 
 
 Do. do. do. , 
 
 
 6 
 
 
 45 
 
 1000 
 
 750 
 
 
 
 Do., 
 
 1890 
 
 2 
 
 Do. do. Triple, - 
 
 
 6 
 
 43-62-96 
 
 51 
 
 950 
 
 807 
 
 34 8 
 
 
 Do., 
 
 1895 
 
 2 
 
 Do. do. do., - 
 
 
 6 
 
 40-59-88 
 
 51 
 
 103-0 
 
 875 
 
 33 1 
 
 
 Do., 
 
 1897 
 
 2 
 
 Do. do. do., 
 
 
 6 
 
 40-59-88 
 
 51 
 
 103-0 
 
 875 
 
 38-4 
 
 
 Do., 
 
 1899 
 
 2 
 
 Do. do. do., 
 
 
 6 
 
 30-49-80 
 
 51 
 
 108-5 
 
 922 
 
 49-0 
 
 
 Do., 
 
 1901 
 
 2 
 
 Do. do. do., 
 
 
 6 
 
 31J-51J-84 
 
 51 
 
 1102 
 
 936 
 
 50-5 
 
 
 Do., 
 
 1903 
 
 2 
 
 Do. do. do., 
 
 
 8 
 
 33J-54J-63-63 
 
 48 
 
 120-0 
 
 960 
 
 50-2 
 
 
 Cruisers — 
 
 
 
 
 
 
 
 
 
 
 
 1st Class, 
 
 1870 
 
 1 
 
 Horizontal Trunk Surface-Condensing, 
 
 2 
 
 (= 104J) 
 
 48 
 
 74-1 
 
 593 
 
 24-0 
 
 
 Do., 
 
 1877 
 
 2 
 
 Do. Return Connecting- Rod Compound, 
 
 8 
 
 41/75-41/75 
 
 36 
 
 97-2 
 
 553 
 
 24-8 
 
 
 Do., 
 
 1S85 
 
 2 
 
 Do. do. do. do., 
 
 4 
 
 46-86 
 
 39 
 
 95-1 
 
 618 
 
 26-1 
 
 
 Do., 
 
 1888 
 
 2 
 
 Do. Direct-Acting Triple, - 
 
 6 
 
 36-51-78 
 
 42 
 
 116-0 
 
 812 
 
 37-1 
 
 
 Do., 
 
 1890 
 
 2 
 
 Vertical do. do., 
 
 12 
 
 36-52-80 
 
 48 
 
 105 
 
 840 
 
 41-8 
 
 
 Do., 
 
 1894 
 
 2 
 
 Do. do. do., 
 
 6 
 
 40-59-88 
 
 51 
 
 106-0 
 
 900 
 
 33-4 
 
 
 Do., 
 
 1897 
 
 2 
 
 Do. do. do., 
 
 8 
 
 45-70-76-76 
 
 48 
 
 113-0 
 
 904 
 
 51-7 
 
 
 Do., 
 
 1899 
 
 2 
 
 Do. do. do., 
 
 S 
 
 34-55J-64-64 
 
 48 
 
 116-6 
 
 932 
 
 46-7 
 
 
 Do., 
 
 1901 
 
 2 
 
 Do. do. do., 
 
 8 
 
 34-55i-64-64 
 
 48 
 
 118-9 
 
 951 
 
 50-3 
 
 
 Do., 
 
 1902 
 
 2 
 
 Do. do. do., 
 
 S 
 
 36-59"-68-68 
 
 48 
 
 120-0 
 
 960 
 
 50-6 
 
 
 Do., 
 
 1902 
 
 2 
 
 Do. do. do., 
 
 8 
 
 43£-71-81J-81i 
 
 48 
 
 122-4 
 
 979 
 
 50-8 
 
 
 Do., 
 
 1900 I 2 
 
 Do. do. do., - 
 
 8 
 
 43£-6"9-77-77 
 
 42 
 
 135 
 
 935 
 
 44-1 
 
 
 Do., 
 
 1908 2 
 
 Do. do. do.. 
 
 8 
 
 40j-65£-75-75 
 
 48 
 
 127-0 
 
 1016 
 
 50-7 
 
 
 TABLE LXXXVa. — Particulars of Naval Ships' Machinery 
 
 Cruisers — 
 2nd Class, 
 Do., 
 Do., 
 Do., 
 Do., 
 Do., 
 Do., 
 Do., 
 Do., 
 Do., 
 Do., 
 
 ( bruisers— 
 
 3rd Class, 
 
 Do., 
 Do., 
 Do., 
 Do., 
 Do., 
 Do., 
 Do., 
 Do., 
 Do., 
 Do., 
 Do., 
 Do., 
 Do., 
 
 1870 
 
 1 
 
 1870 
 
 1 
 
 1877 
 
 1 
 
 18S2 
 
 2 
 
 18S6 
 
 2 
 
 1889 
 
 2 
 
 1889 
 
 2 
 
 1891 
 
 2 
 
 1896 
 
 2 
 
 1898 
 
 2 
 
 1900 
 
 2 
 
 1870 
 
 1 
 
 1875 
 
 1 
 
 1878 
 
 1 
 
 1&84 
 
 1 
 
 1S85 
 
 2 
 
 1887 
 
 2 
 
 1888 
 
 2 
 
 1890 
 
 2 
 
 1890 
 
 2 
 
 1891 
 
 2 
 
 1898 
 
 2 
 
 1901 
 
 2 
 
 1903 | 2 
 
 '. 1904 2 
 
 1 
 
 
 Horizontal Trunk Surface-Condensing, 
 Do. Direct- Acting Surface-Condensing, 
 Return Connecting-Rod Compound, 
 Direct-Acting Compound, 
 
 Do 
 Do. 
 Do. 
 Do. 
 Vertical 
 Do. 
 Do. 
 Do. 
 Do. 
 
 do. 
 do. 
 do. 
 do. 
 do. 
 do. 
 do, 
 
 do. 
 Triple, 
 do., 
 do., 
 do., 
 do., 
 do., 
 
 Horizontal Return Connecting-Rod Surface ) 
 Condensing, - - - - | 
 
 Return Connecting-Rod Compound, 
 do. do do., 
 
 do. do. do., 
 
 Direct-Acting Compound, 
 
 Do. 
 Do. 
 Do. 
 Do. 
 Do. 
 Do. 
 Vertical 
 Do. 
 Do. 
 Do. 
 Do. 
 Do. 
 Do. 
 
 do. 
 do. 
 do. 
 do. 
 do. 
 do. 
 do. 
 do. 
 do. 
 
 do., 
 Triple, 
 do., 
 do., 
 ''"■. 
 .do., 
 do., 
 do., 
 do., - 
 
 ( = 86) 
 
 88 
 
 88-72-88 
 
 42-80 
 
 38-64 
 
 34J-51-76J 
 
 334-49-74 
 
 33J-49 74 
 
 33-49-74 
 
 26-42-68 
 
 26-42-48-48 
 
 55 
 
 57-90 
 
 46-64 
 
 42-72 
 
 26-46 
 
 27-50 
 
 26-37-57 
 
 26J-39-57 
 
 30J-45-68 
 
 30H5-68 
 
 20J-33-54 
 
 :0J-33 -38£-38£ 
 
 24i-38£-42i-42j 
 
 31J-51J-57-W 
 
 79-1 
 
 74-2 
 
 70-0 
 
 90-0 
 
 120-0 
 
 140-0 
 
 140-0 
 
 140-0 
 
 144-2 
 
 141-0 
 
 181-3 
 
 90-3 
 
 95-8 
 960 
 80-3 
 150-0 
 150-0 
 150 
 220-0 
 160 
 160-0 
 218 
 250 
 250-0 
 
 30 211 
 
 593 
 520 
 560 
 675 
 780 
 820 
 910 
 910 
 936 
 916 
 906 
 
 451 
 
 527 
 
 570 
 
 4S1 
 
 750 
 
 825 
 
 825 
 
 990 
 
 880 
 
 880 
 
 9S1 
 
 1000 
 
 1000 
 
 1055 
 
 21-5 
 20-3 
 24-0 
 267 
 40-4 
 41-6 
 40-0 
 38-8 
 40-3 
 50-9 
 492 
 
 20-0 
 
 21-1 
 22-1 
 26 3 
 444 
 33-2 
 36-4 
 300 
 24-0 
 37-1 
 52-7 
 52 
 57-3 
 440 
 
NAVAL SHIPS MACHf-NERY. 
 
 737 
 
 (Battleships and 1st Class Cruisers) — Reciprocating Engines. 
 
 
 
 Full Power 
 
 4/5 or Natural 
 
 Weight 
 
 I.H.P. 
 
 «J 
 
 
 
 
 Trial. • 
 
 Draught. 
 
 of Machinery. 
 
 Full Speed per 
 
 s 
 
 
 Number and Kind. 
 
 a 
 
 00 
 
 B 
 
 cd 
 CO 
 
 6t 
 
 B 
 
 '■§«' 
 
 s § 
 
 — 3 
 *7l 
 
 I.K.P. 
 
 Speed. 
 
 I.H.P. 
 
 Speed. 
 
 Total. 
 
 u 
 
 a. 
 
 Eel; 
 
 A 
 
 - — 
 
 — — * 
 
 5a 
 
 
 
 •ia 
 
 O . 
 
 O 4) 
 
 *■« 
 
 CO t- 
 
 cl" 
 
 "3g 
 
 E-i t. 
 "cm 
 
 fe 3 
 
 S 
 
 A 
 
 "5. 
 
 s 
 
 + 
 
 .C 
 ft© 
 
 
 
 u 
 
 
 a 
 
 
 
 
 
 
 
 d 
 
 
 O V 
 
 
 
 
 s 
 
 O 
 
 b* 
 
 
 
 
 
 
 fi 
 
 S3 
 
 g 
 
 00 
 
 
 Is 
 
 
 
 Lbs. 
 
 Sq.ft. 
 
 Sq. ft. 
 
 
 Knots. 
 
 
 Knots. 
 
 Tons. 
 
 Lbs. 
 
 Lbs. 
 
 
 
 
 
 
 Box, - 
 
 30 
 
 840 
 
 22,200 
 
 8,539 
 
 14-69 
 
 7,187 
 
 13-8 
 
 1090 
 
 286 
 
 339 
 
 7-83 
 
 10-2 
 
 385 
 
 0126 
 
 
 Do., - 
 
 316 
 
 
 
 7,842 
 
 14-97 
 
 7,470 
 
 H-7 
 
 1043 
 
 300 
 
 313 
 
 7-48 
 
 
 
 0-129 
 
 
 Do., - 
 
 30-0 
 
 570 
 
 14,388 
 
 4,960 
 
 13-75 
 
 
 
 714 
 
 327 
 
 
 6-86 
 
 8-6 
 
 34-0 
 
 0-103 
 
 
 Do., - 
 
 317 
 
 515 
 
 14,115 
 
 4,914 
 
 14-0 
 
 4,311 
 
 13-2 
 
 750 
 
 342 
 
 389 
 
 6-55 
 
 9-5 
 
 33-9 
 
 0-125 
 
 
 Cylindrical, 
 
 65 
 
 608 
 
 17,500 
 
 6,624 
 
 14-4 
 
 
 
 998 
 
 337 
 
 
 6-64 
 
 10-9 
 
 379 
 
 0131 
 
 
 Oval and cylindrical, 
 
 90 
 
 667 
 
 22,784 
 
 10,184 
 
 16-75 
 
 8,000 
 
 16-2 
 
 1266 
 
 278 
 
 354 
 
 8-04 
 
 15-3 
 
 44-7 
 
 0-151 
 
 
 S.E. cylindrical, 
 
 90 
 
 790 
 
 20,587 
 
 11,271 
 
 16-75 
 
 8,277 
 
 16-0 
 
 1209 
 
 241 
 
 331 
 
 9-30 
 
 14-2 
 
 54-7 
 
 0-114 
 
 
 Cylindrical, 
 
 135 
 
 629 
 
 18,930 
 
 12,465 
 
 16-75 
 
 8,813 
 
 16-2 
 
 1015 
 
 189 
 
 258 
 
 12-20 
 
 19-7 
 
 65-S 
 
 0-015 
 
 
 S.E. cylindrical, 
 
 155 
 
 731 
 
 20,348 
 
 11,500 
 
 17-5 
 
 9,430 
 
 16-75 
 
 1163 
 
 224 
 
 276 
 
 10 00 
 
 16-0 
 
 56-6 
 
 0-082 
 
 
 S do. do., 
 
 155 
 
 817 
 
 25,233 
 
 12,414 
 
 18-0 
 
 10,404 
 
 17-0 
 
 1341 
 
 241 
 
 282 
 
 9-30 
 
 15-2 
 
 49 3 
 
 0-090 
 
 
 Belleville, 
 
 300 
 
 1050 
 
 33,770 
 
 13,763 
 
 18-5 
 
 10,454 
 
 17-2 
 
 1290 
 
 209 
 
 280 
 
 10-70 
 
 13-1 
 
 40-8 
 
 100 
 
 
 Do., 
 
 300 
 
 1170 
 
 37,120 
 
 15,503 
 
 18-5 
 
 11,626 
 
 
 1400 
 
 202 
 
 270 
 
 11-10 
 
 13-2 
 
 41-8 
 
 0-093 
 
 
 Mixed, 
 
 
 1390 
 
 43,260 
 
 18,220 
 
 19-0 
 
 13,670 
 
 
 1580 
 
 196 
 
 260 
 
 11-40 
 
 13-1 
 
 42-1 
 
 0-113 
 
 
 Box, - 
 
 30 
 
 900 
 
 24,868 
 
 7,361 
 
 16-5 
 
 
 
 1043 
 
 317 
 
 
 7-06 
 
 8-18 
 
 29-6 
 
 0-180 
 
 
 S.E. oval, • 
 
 65 
 
 690 
 
 19,500 
 
 7,714 
 
 18-5 
 
 6,658 
 
 17-0 
 
 1012 
 
 294 
 
 340 
 
 7-62 
 
 11-18 
 
 395 
 
 0-271 
 
 
 8 S.E. cylindrical, - 
 
 90 
 
 546 
 
 15,160 
 
 5,661 
 
 16-6 
 
 
 
 776 
 
 307 
 
 
 7-29 
 
 10-37 
 
 37 2 
 
 0-180 
 
 
 4 D.E. cylindrical, - 
 
 130 
 
 492 
 
 15,190 
 
 8,738 
 
 18-75 
 
 6,090 
 
 18-0 
 
 772 
 
 198 
 
 279 
 
 11-30 
 
 17-7 
 
 57-5 
 
 0-138 
 
 
 D.E. cylindrical, 
 
 155 
 
 1135 
 
 31,073 
 
 21,411 
 
 21-6 
 
 14,924 
 
 20-4 
 
 1543 
 
 162 
 
 231 
 
 13-80 
 
 18-8 
 
 68-5 
 
 0-171 
 
 
 J 4 D.E. 1S.E. cylin-) 
 j drical, ) 
 
 155 
 
 812 
 
 24,908 
 
 12,851 
 
 20-5 
 
 10,517 
 
 19-9 
 
 1161 
 
 202 
 
 247 
 
 1110 
 
 15-2 
 
 514 
 
 0-158 
 
 
 Belleville, 
 
 300 
 
 2200 
 
 67,800 
 
 25,774 
 
 22-0 
 
 22,547 
 
 21-5 
 
 2230 
 
 195 
 
 224 
 
 11-50 
 
 11-7 
 
 38-0 
 
 0157 
 
 
 Do., 
 
 300 
 
 1460 
 
 40,990 
 
 16,961 
 
 20-5 
 
 12,785 
 
 19-8 
 
 1540 
 
 204 
 
 270 
 
 11-00 
 
 11-6 
 
 42-6 
 
 0-140 
 
 
 Do., 
 
 300 
 
 1390 
 
 47,300 
 
 19,156 
 
 21-0 
 
 14,049 
 
 19 9 
 
 1577 
 
 185 
 
 250 
 
 12-10 
 
 13-8 
 
 40-5 
 
 0-143 
 
 
 Do., 
 
 300 
 
 1650 
 
 51,600 
 
 21,400 
 
 21-75 
 
 16,440 
 
 
 1800 
 
 188 
 
 245 
 
 11-90 
 
 13-0 
 
 41-3 
 
 0-150 
 
 
 Do., 
 
 300 
 
 2314 
 
 71,964 
 
 31,409 
 
 24-0 
 
 23,103 
 
 
 2500 
 
 178 
 
 242 
 
 12-60 
 
 13-6 
 
 43 6 
 
 0178 
 
 
 Do., 
 
 300 
 
 1610 
 
 50,300 
 
 22,360 
 
 22 p 7 
 
 16,100 
 
 
 1750 
 
 175 
 
 243 
 
 12-80 
 
 13-8 
 
 44-4 
 
 
 
 Mixed, • . 
 
 205 
 
 1760 
 
 62,250 
 
 23.500 | 
 
 23-6 
 
 
 . . 
 
 2250 
 
 215 
 
 
 10-44 
 
 13-3 
 
 37-7 
 
 
 
 Do., 
 
 210 
 
 
 72,000 
 
 27,000 , 
 
 235 
 
 19,000 
 
 21-3 
 
 
 
 
 
 
 
 •• 
 
 (2nd and 3rd Class Cruisers) — Reciprocating Engines. 
 
 
 Box, - - - - 
 
 30 
 
 504 
 
 12,748 
 
 4,500 
 
 15-12 
 
 
 
 611 
 
 304 
 
 
 7-36 
 
 9-00 
 
 35-4 
 
 0-200 
 
 
 Do., .... 
 
 30 
 
 455 
 
 12,420 
 
 3,878 
 
 14-82 
 
 
 
 640 
 
 370 
 
 
 6-06 
 
 8-52 
 
 31-3 
 
 0-209 
 
 
 S.E. cylindrical. 
 
 70 
 
 510 
 
 12,700 
 
 4,950 
 
 14-5 
 
 
 
 873 
 
 395 
 
 
 5-67 
 
 9-70 
 
 38-9 
 
 0-252 
 
 
 8 do. do., 
 
 90 
 
 542 
 
 14,924 
 
 5,500 
 
 16 6 
 
 4,100 
 
 
 824 
 
 336 
 
 448 
 
 6-67 
 
 10 16 
 
 37-0 
 
 0-219 
 
 
 Gunboat, • 
 
 110 
 
 388 
 
 11,565 
 
 6,151 
 
 18-0 
 
 4,217 
 
 
 559 
 
 204 
 
 300 
 
 11-0 
 
 15-8 
 
 53-0 
 
 0-157 
 
 
 4 D.E. cylindrical, - 
 
 155 
 
 570 
 
 13,830 
 
 9,653 
 
 190 
 
 6,215 
 
 18-0 
 
 649 
 
 151 
 
 234 
 
 14-8 
 
 16-9 
 
 70-0 
 
 0-220 
 
 
 Do. do., 
 
 155 
 
 525 
 
 12,682 
 
 9,435 
 
 19-0 
 
 6,144 
 
 18 
 
 624 
 
 148 
 
 228 
 
 15-3 
 
 17 9 
 
 74-4 
 
 0-221 
 
 
 Do. do. , 
 
 150 
 
 575 
 
 15,641 
 
 9,271 
 
 20-5 
 
 7,423 
 
 193 
 
 748 
 
 181 
 
 226 
 
 12-4 
 
 16-1 
 
 59-4 
 
 0-220 
 
 
 8 S.E. do., 
 
 155 
 
 624 
 
 18,579 
 
 9,846 
 
 20-3 
 
 8,307 
 
 19-6 
 
 915 
 
 208 
 
 247 
 
 10-8 
 
 15-8 
 
 53-0 
 
 210 
 
 
 Belle\-ille, 
 
 309 
 
 869 
 
 25,606 
 
 10,272 
 
 20-0 
 
 7,155 
 
 18-7 
 
 825 
 
 180 
 
 258 
 
 12-8 
 
 11-9 
 
 40-1 
 
 0-142 
 
 
 Do., 
 
 300 
 
 792 
 
 24,426 
 
 9,770 
 
 20-5 
 
 7,660 
 
 19-3 
 
 832 
 
 190 
 
 243 
 
 11-7 
 
 12-3 
 
 40-0 
 
 0149 
 
 
 Box, - 
 
 30 
 
 210 
 
 5,980 
 
 2,170 
 
 12-8 
 
 , , 
 
 
 236 
 
 244 
 
 
 9-19 
 
 10-3 
 
 36-3 
 
 0-152 
 
 
 6 S.E. cylindrical, - 
 
 60 
 
 244 
 
 5,850 
 
 2,139 
 
 13-0 
 
 
 
 350 
 
 374 
 
 
 6-01 
 
 8-9 
 
 36-6 
 
 0-165 
 
 
 Gunboat, - 
 
 60 
 
 246 
 
 6,312 
 
 2,400 
 
 13 
 
 
 
 394 
 
 365 
 
 
 
 
 38-1 
 
 0-165 
 
 
 S.E. cylindrical, 
 
 90 
 
 301 
 
 9,254 
 
 4,030 
 
 14-6 
 
 3,131 
 
 
 517 
 
 300 
 
 370 
 
 7-8 
 
 
 
 
 
 Gunboat, - 
 
 120 
 
 213 
 
 6,400 
 
 3,365 
 
 16-5 
 
 2,201 
 
 
 296 
 
 197 
 
 301 
 
 11-3 
 
 15-8 
 
 52-6 
 
 0-207 
 
 
 Do., 
 
 130 
 
 222 
 
 6,836 
 
 3,754 
 
 16-5 
 
 2,462 
 
 
 353 
 
 210 
 
 321 
 
 10-6 
 
 16-9 
 
 55-1 
 
 0-216 
 
 
 Do., 
 
 140 
 
 244 
 
 7,878 
 
 4,613 
 
 17-5 
 
 2,647 
 
 
 395 
 
 192 
 
 334 
 
 11-6 
 
 18-9 
 
 58-4 
 
 0-242 
 
 
 Locomotive, 
 
 155 
 
 244 
 
 7,088 
 
 4,561 
 
 18-5 
 
 3,592 
 
 17-8 
 
 274 
 
 135 
 
 170 
 
 16-7 
 
 18-7 
 
 64-2 
 
 0-149 
 
 
 4 D.E. cylindrical, - 
 
 155 
 
 365 
 
 9,938 
 
 
 
 4,615 
 
 17-2 
 
 515 
 
 
 249 
 
 
 
 46-5 
 
 0-200 
 
 
 Do. do., 
 
 155 
 
 410 
 
 11,025 
 
 7,469 
 
 19 V 5 
 
 5,016 
 
 17-6 
 
 539 
 
 156 
 
 257 
 
 13 V 8 
 
 18 V 2 
 
 69-7 
 
 0-209 
 
 
 Express water-tube, 
 
 300 
 
 352 
 
 18,100 
 
 7,152 
 
 20-7 
 
 5,388 
 
 195 
 
 383 
 
 120 
 
 159 
 
 18-9 
 
 20-3 
 
 39-7 
 
 0-179 
 
 
 Do. do., 
 
 300 
 
 352 
 
 17,760 
 
 7,331 
 
 20-8 
 
 5,218 
 
 195 
 
 370 
 
 113 
 
 159 
 
 19-8 
 
 20-8 
 
 41-7 
 
 0-168 
 
 
 Do. do., 
 
 300 
 
 490 
 
 26,000 
 
 9,800 
 
 
 
 
 . , 
 
 . . 
 
 
 
 . . 
 
 37-7 
 
 
 
 Do. do., 
 
 
 752 
 
 43,000 
 
 14,330 
 
 25-2 
 
 
 
 
 
 
 
 •• 
 
 333 
 
 
 47 
 
T38 
 
 MANUAL OP MARINE ENGINEERING. 
 
 TABLE LXXXV6. — Particulars or Machinery of 
 
 
 
 CD 
 
 
 
 Cylinders. 
 
 Full 
 
 Speed. 
 
 
 Designation. 
 
 Wlien 
 Buiit 
 
 V 
 
 u 
 u 
 Jj 
 
 i— 
 
 u 
 
 S 
 
 Kind of Engines. 
 
 li 
 
 CD 
 
 a 
 
 3 
 55 
 
 Diameter. 
 
 o 
 u 
 
 s 
 
 u 
 
 CO 
 
 D. 
 
 get, 
 
 > 
 
 Q 
 
 05 
 
 V 
 
 a 
 
 ID 
 
 a 
 o 
 
 m 
 
 p 
 
 a 
 
 Lbs. 
 
 
 
 
 
 
 
 Ins. 
 
 Ins 
 
 
 Ft. 
 
 
 Gunboat, 
 
 1876 
 
 1 
 
 J Horizontal Return 
 ( Compound, 
 
 Connecting - Rod ) 
 
 2 
 
 31-48 
 
 18 
 
 137 
 
 411 
 
 23-8 
 
 
 Do., 
 
 1889 
 
 1 
 
 Horizontal Direct-Acting Triple, - 
 
 3 
 
 20-30-45 
 
 24 
 
 184 
 
 736 
 
 366 
 
 
 Torpedo gun-) 
 boat, ) 
 
 1886 
 
 2 
 
 Vertical do. 
 
 do., - 
 
 6 
 
 22-33-49 
 
 IS 
 
 310 
 
 930 
 
 
 
 Do., 
 
 1890 
 
 2 
 
 Do. do. 
 
 do., • 
 
 6 
 
 22-33-49 
 
 21 
 
 250 
 
 876 
 
 38'4 
 
 
 Do., 
 
 1890 
 
 2 
 
 Do. do. 
 
 do., - 
 
 6 
 
 24-34-51 
 
 21 
 
 250 
 
 875 
 
 34 2 
 
 
 Vidette boat, 
 
 
 1 
 
 Do. do. 
 
 Compound, • 
 
 2 
 
 8J-13 
 
 9 
 
 460 
 
 690 
 
 54-8 
 
 
 Torpedo boat, 
 
 1877 
 
 1 
 
 Do. do. 
 
 do., 
 
 2 
 
 9J-15 
 
 95 
 
 435 
 
 689 
 
 47 6 
 
 
 Do., 
 
 1877 
 
 1 
 
 Do. do. 
 
 do., 
 
 2 
 
 12|-21 
 
 12 
 
 354 
 
 708 
 
 465 
 
 
 Do., 
 
 1886 
 
 1 
 
 Do. do. 
 
 do., 
 
 2 
 
 18-30 
 
 16 
 
 402 
 
 1072 
 
 44 3 
 
 
 Do., 
 
 1887 
 
 1 
 
 Do. do. 
 
 Triple, • 
 
 3 
 
 15-22-34 
 
 16 
 
 402 
 
 1072 
 
 36 4 
 
 
 Do., 
 
 1894 
 
 1 
 
 Do. do. 
 
 Quadruple, - 
 
 4 
 
 14-20-27-36 
 
 16 
 
 410 
 
 1093 
 
 47-2 
 
 
 Do., 
 
 1893 
 
 1 
 
 Do. do. 
 
 Triple, - 
 
 3 
 
 18-26-39i 
 
 18 
 
 370 
 
 1110 
 
 65-8 
 
 
 Destroyer, 
 
 1893 
 
 2 
 
 Do. do. 
 
 do., • 
 
 6 
 
 18-26-39* 
 
 18 
 
 360 
 
 1078 
 
 43-8 
 
 
 Do., 
 
 1894 
 
 2 
 
 Do. do. 
 
 do., • 
 
 6 
 
 18-26-39J 
 
 18 
 
 393 
 
 1178 
 
 44'6 
 
 
 Do., 
 
 1894 
 
 2 
 
 Do. do. 
 
 do., • 
 
 8 
 
 19-27-27-27 
 
 16 
 
 392 
 
 1046 
 
 60-0 
 
 
 Do., 
 
 1894 
 
 2 
 
 Do. do. 
 
 do., • 
 
 6 
 
 19-29-43 
 
 18 
 
 366 
 
 1098 
 
 46-4 
 
 
 Do., 
 
 1895 
 
 2 
 
 Do. do. 
 
 do., • • 
 
 6 
 
 18-27i-42 
 
 18 
 
 366 
 
 1097 
 
 42-2 
 
 
 Do., 
 
 1896 
 
 o 
 
 Do. do. 
 
 do., • 
 
 6 
 
 18-27-42 
 
 18 
 
 397 
 
 1190 
 
 437 
 
 
 Do., 
 
 1897 
 
 2 
 
 Do. do. 
 
 do., • 
 
 8 
 
 20-29-30-30 
 
 18 
 
 395 
 
 1186 
 
 56-9 
 
 
 Do., 
 
 do., 
 
 2 
 
 Do. do. 
 
 do., • 
 
 6 
 
 21-32J-48 
 
 18 
 
 365 
 
 1095 
 
 53-8 
 
 
 Do., 
 
 do., 
 
 2 
 
 Do. do. 
 
 do., • 
 
 6 
 
 19-29H6 
 
 18 
 
 395 
 
 1185 
 
 52-1 
 
 
 Do., 
 
 do., 
 
 2 
 
 Do. do. 
 
 do., • 
 
 8 
 
 20J-31-34-34 
 
 18 
 
 392 
 
 1176 
 
 60-0 
 
 
 Do., 
 
 1904 
 
 2 
 
 Do- do. 
 
 do., • 
 
 8 
 
 23-35-88-38 
 
 20 
 
 342 
 
 1140 
 
 " 
 
 
 Gunboat, 
 
 1898 
 
 2 
 
 Do. do. 
 
 do., • 
 
 6 
 
 11-17-27 
 
 16 
 
 2S6 
 
 763 
 
 65-3 
 
 
 Sloop, • 
 
 1901 
 
 2 
 
 Do. do. 
 
 do., - 
 
 6 
 
 11J-18-29J 
 
 24 
 
 200 
 
 800 
 
 42 6 
 
 
MACHINKEY OF SMALL NAVAL SHIPS. 
 
 r39 
 
 Small Naval Ships (Reciprocating Engines). 
 
 1 
 
 Boilers. 
 
 Full Power 
 Trial. 
 
 4/5 or Natural 
 Draught. 
 
 Weight 
 of Machinery. 
 
 I.H.P. 
 
 Full Speed per 
 
 CD 
 
 S 
 
 « 
 
 o 
 c3 
 
 CO 
 
 5 
 + 
 
 -ta 
 
 "53 
 
 
 Number and Kind. 
 
 c 
 
 3 
 
 00 
 DO 
 
 0> 
 
 u 
 Ph 
 
 oj 
 
 u 
 CS 
 
 ta 
 
 B 
 
 — 3 
 
 O 
 
 I.H.P. 
 
 Speed. 
 
 i 
 I.H.P. 
 
 Speed. 
 
 Total. 
 
 Pi 
 
 ~5 
 
 fa 
 
 43 
 JS 
 
 Is 
 
 ° a 
 a -3 
 o-p. 
 
 o 
 
 43 
 
 o 
 
 &H-S 
 
 £o 
 
 3 
 C 
 CO 
 
 w a oi 
 
 O.eS 
 
 «!? 
 
 O 4) 
 
 O — 1 
 
 
 
 Lbs. 
 
 Sq.ft. 
 
 Sq. ft. 
 
 
 Knots. 
 
 
 Knots. 
 
 Tons. 
 
 Lbs. 
 
 Lbs. 
 
 
 
 
 
 
 2 Gunboat, 
 
 60 
 
 45 
 
 1,018 
 
 519 
 
 10 
 
 •• 
 
 .. 
 
 79-7 
 
 344 
 
 
 6-51 
 
 11-53 
 
 50-9 
 
 0-233 
 
 
 Do., 
 
 150 
 
 68 
 
 2,200 
 
 1298 
 
 14 1 
 
 1033 
 
 12-4 
 
 108-3 
 
 185 
 
 235 
 
 120 
 
 19-10 
 
 59-0 
 
 0135 
 
 
 4 Locomotive, - • • 
 
 150 
 
 121 
 
 4,426 
 
 2696 
 
 18-5 
 
 •• 
 
 
 120-0 
 
 97 
 
 
 23 1 
 
 22-4 
 
 613 
 
 0-219 
 
 
 Do., 
 
 150 
 
 190 
 
 5,330 
 
 3840 
 
 200 
 
 2598 
 
 .. 
 
 171-0 
 
 100 
 
 149 
 
 22-4 
 
 20-2 
 
 72-0 
 
 0-232 
 
 
 Do., 
 
 150 
 
 171 
 
 6,204 
 
 3665 
 
 19 
 
 2631 
 
 
 2110 
 
 130 
 
 178 
 
 17-3 
 
 21-4 
 
 591 
 
 287 
 
 
 1 Locomotive, • 
 
 146 
 
 5-5 
 
 235 
 
 151 
 
 15 
 
 
 • • 
 
 4-6 
 
 68 
 
 
 33-0 
 
 27 5 
 
 64 3 
 
 0514 
 
 
 Do., 
 
 130 
 
 7 
 
 278 
 
 175 
 
 15-2 
 
 • • 
 
 
 5-51 
 
 70-4 
 
 
 31-8 
 
 25-0 
 
 63-0 
 
 0-441 
 
 
 Do., 
 
 120 
 
 155 
 
 630 
 
 400 
 
 17-5 
 
 • ■ 
 
 • • 
 
 11-0 
 
 615 
 
 
 36-4 
 
 26 
 
 63-5 
 
 0-410 
 
 
 Do., 
 
 130 
 
 41 
 
 1,775 
 
 1013 
 
 19-0 
 
 .. 
 
 • • 
 
 31-6 
 
 70 
 
 
 32-0 
 
 24-7 
 
 57-0 
 
 0-500 
 
 
 Do., 
 
 180 
 
 38 
 
 1,660 
 
 1057 
 
 19-5 
 
 • • 
 
 • • 
 
 31-5 
 
 66-9 
 
 
 335 
 
 27-8 
 
 63-7 
 
 0-500 
 
 
 2 Water-tube Yarrow, 
 
 250 
 
 50 
 
 2,350 
 
 1600 
 
 23 
 
 • • 
 
 
 
 
 
 
 32-0 
 
 68-1 
 
 
 
 Do. do., 
 
 220 
 
 75 
 
 3,500 
 
 2300 
 
 23 5 
 
 • • 
 
 • • 
 
 48-8 
 
 47-5 
 
 
 47-1 
 
 30-7 
 
 660 
 
 0-377 
 
 
 2 Locomotive, - 
 
 180 
 
 100 
 
 4,656 
 
 3497 
 
 26-1 
 
 • • 
 
 • • 
 
 94-0 
 
 60 
 
 
 37-2 
 
 35-0 
 
 750 
 
 0-392 
 
 
 Water-tube Yarrow, 
 
 180 
 
 154 
 
 8,216 
 
 3884 
 
 27-6 
 
 • • 
 
 • • 
 
 90-6 
 
 52 
 
 
 42-8 
 
 251 
 
 47 3 
 
 377 
 
 
 Do. Thornycroft, 
 
 210 
 
 189 
 
 9,780 
 
 4368 
 
 28-4 
 
 • • 
 
 
 108-2 
 
 56 
 
 
 40-3 
 
 231 
 
 44-7 
 
 0-438 
 
 
 Do. Normand, - 
 
 175 
 
 150 
 
 8,800 
 
 4484 
 
 27-3 
 
 .. 
 
 • • 
 
 1232 
 
 62 
 
 
 36 3 
 
 29-9 
 
 50-9 
 
 440 
 
 
 Do. Reed, - 
 
 210 
 
 209 
 
 10,164 
 
 3885 
 
 27-9 
 
 .. 
 
 .. 
 
 116-2 
 
 67 
 
 
 33 4 
 
 18-5 
 
 38-7 
 
 0-461 
 
 
 Do. Blechynden, 
 
 200 
 
 176 
 
 10,022 
 
 4367 
 
 27-4 
 
 • • 
 
 • • 
 
 106-0 
 
 54 
 
 
 41-1 
 
 24-8 
 
 43-6 
 
 0-400 
 
 
 Do. Thornycroft, 
 
 220 
 
 196 
 
 11,855 
 
 5795 
 
 30-1 
 
 • • 
 
 
 127-0 
 
 49 
 
 
 45-6 
 
 29-5 
 
 48-9 
 
 0-462 
 
 
 Do. Normand, - 
 
 220 
 
 210 
 
 11,945 
 
 6347 
 
 30-1 
 
 -• 
 
 • • 
 
 143-5 
 
 50 
 
 
 44-2 
 
 302 
 
 53-0 
 
 0478 
 
 
 Do. Reed, - 
 
 250 
 
 255 
 
 13,384 
 
 6318 
 
 30-3 
 
 .. 
 
 .. 
 
 1437 
 
 51 
 
 
 43 9 
 
 24-7 
 
 47-2 
 
 0-479 
 
 
 Do. Thornycroft, 
 
 250 
 
 245 
 
 13,561 
 
 6523 
 
 30-2 
 
 • • 
 
 .. 
 
 126-0 
 
 43 
 
 
 51-7 
 
 27-5 
 
 48-0 
 
 0-420 
 
 
 Do. Yarrow, - 
 
 240 
 
 270 
 
 16,000 
 
 7766 
 
 263 
 
 .. 
 
 .. 
 
 1730 
 
 50 
 
 
 44-8 
 
 28-7 
 
 48-5 
 
 346 
 
 
 4 Yarrow, 
 
 250 
 
 805 
 
 4,000 
 
 1300 
 
 135 
 
 960 
 
 13 
 
 950 
 
 164 
 
 
 13-7 
 
 161 
 
 355 
 
 136 
 
 
 4 Babcock, 
 
 210 
 
 144 
 
 4,040 
 
 1404 
 
 13-5 
 
 1021 
 
 •• 
 
 162-5 
 
 260 
 
 356 
 
 86 
 
 10 
 
 35 
 
 0151 
 
740 
 
 MANUAL OF MARINE ENGINEERING. 
 
 TABLE LXXXVL— Examples of Ocean Express 
 
 Name of Ship. 
 
 When 
 Built. 
 
 Tonnage. 
 
 No. 
 
 of 
 
 Screws. 
 
 Cylinders. 
 
 Revolu- 
 tions 
 per 
 Min. 
 
 Speed 
 
 of 
 
 Piston 
 
 per 
 
 Min. 
 
 Re- 
 ferred 
 Mean 
 
 Pres- 
 sure. 
 
 
 Gross 
 Register 
 
 Displace- 
 ment. 
 
 No. 
 
 Diameters. 
 
 Stroke. 
 
 
 
 
 Tons. 
 
 Tons. 
 
 
 
 Inches. 
 
 Ins. 
 
 
 Feet. 
 
 Lbs. 
 
 
 Oregon, 
 
 1883 
 
 7,353 
 
 13,700 
 
 1 
 
 3 
 
 104-70-104 
 
 72 
 
 65-3 
 
 781 
 
 33-0 
 
 
 Umbria, 
 
 3884 
 
 8,128 
 
 15,500 
 
 1 
 
 3 
 
 105-71-105 
 
 72 
 
 67-5 
 
 810 
 
 33-6 
 
 
 Oroya, 
 
 1886 
 
 6,297 
 
 10,500 
 
 1 
 
 3 
 
 40-66-100 
 
 72 
 
 64-5 
 
 774 
 
 36-6 
 
 
 Victoria, . . 
 
 1887 
 
 6,527 
 
 8,124 
 
 1 
 
 3 
 
 40-5-63-5-102 
 
 72 
 
 63-0 
 
 756 
 
 33-9 
 
 
 Teutonic, . . 
 
 1889 
 
 9,686 
 
 18,000 
 
 2 
 
 6 
 
 43-68-110 
 
 60 
 
 78-0 
 
 780 
 
 37-8 
 
 
 Oruba, . . 
 
 1889 
 
 5,857 
 
 9,700 
 
 1 
 
 3 
 
 30-66-97 
 
 66 
 
 67-8 
 
 746 
 
 33-5 
 
 
 Furst Bismark, 
 
 1890 
 
 8,430 
 
 15,300 
 
 2 
 
 6 
 
 43-5-67-106-3 
 
 63 
 
 81-0 
 
 850 
 
 35-1 
 
 
 Empress of India, 
 
 1891 
 
 5,905 
 
 11,500 
 
 2 
 
 6 
 
 32-54-82 
 
 54 
 
 88-0 
 
 792 
 
 39-9 
 
 
 Ophir 
 
 1891 
 
 6,900 
 
 10,600 
 
 2 
 
 6 
 
 34-51-5-85 
 
 54 
 
 87-5 
 
 768 
 
 31-2 
 
 
 Scot 
 
 1891 
 
 7,815 
 
 14,500 
 
 2 
 
 6 
 
 34-5-57-5-92 
 
 60 
 
 80-0 
 
 800 
 
 36-0 
 
 
 Campania, . . 
 
 1893 
 
 12,950 
 
 22,000 
 
 2 
 
 10 
 
 37/98-79-37/98 
 
 69 
 
 80-9 
 
 930 
 
 33-0 
 
 
 K. Willielm, D.G. 
 
 1897 
 
 14.349 
 
 23,500 
 
 2 
 
 8 
 
 52-89-7-96-4-96-4 
 
 69 
 
 77-2 
 
 887 
 
 38-5 
 
 
 Oceanic. ... 
 
 1899 
 
 17,274 
 
 25,900 
 
 2 
 
 8 
 
 47-5-79-93-93 
 
 72 
 
 79-0 
 
 948 
 
 35-8 
 
 
 Savoia. . . . 
 
 1900 
 
 11,638 
 
 15,400 
 
 o 
 
 8 
 
 44-5-68-5-80-3-80-3 
 
 67 
 
 92-0 
 
 1 027 
 
 36-5 
 
 
 Deutschland, . 
 
 1900 
 
 16.500 
 
 24,500 
 
 o 
 
 12 
 
 (104-36-6/106- ) 
 1 36-6/106-74f 
 
 73 
 
 80-0 
 
 973 
 
 37-4 
 
 
 K. Wilhelm II. 
 
 1902 
 
 19,361 
 
 26,500 
 
 2 
 
 16 
 
 37/49-75-112 
 
 71 
 
 80-0 
 
 947 
 
 35-0 
 
 
 Caronia, 
 
 1905 
 
 19,687 
 
 27,500 
 
 2 
 
 8 
 
 39-54-77-110 
 
 66 
 
 75-0 
 
 825 
 
 46-3 
 
 
 Empress of Britain, 
 
 1906 
 
 14,190 
 
 18,196 
 
 2 
 
 8 
 
 36-52-75-108 
 
 69 
 
 72-0 
 
 828 
 
 40-5 
 
 
 K. Princess Cecil ie, 
 
 1908 
 
 19,400 
 
 27,000 
 
 2 
 
 8 
 
 37-49-75-113 
 
 71 
 
 82-0 
 
 970 
 
 40-8 
 
 
 Laurentic, . . . 
 
 1908 
 
 14,892 
 
 19,160 
 
 3 
 
 8 
 
 (30-46-53-53 and > 
 i 1 L.P. Turb. f 
 
 54 
 
 85-0 
 
 765 
 
 42-4 
 
 
 Olympic, . • 
 Lutetia, . • 
 
 1910 
 1913 
 
 46,359 
 14.580 
 
 49,336 
 15.600 
 
 3 
 4 
 
 8 
 8 
 
 (54-84-97-97 and! 
 ( 1 L.P. Turb. ( 
 J41-57-65-65 and 1 
 \ 2 L.P. Turbs. 1 
 
 75 
 
 44 
 
 78-3 
 
 97S 
 
 42-5 
 
 
 Bergensfiord, . . 
 
 1913 
 
 10,666 
 
 16,177 
 
 2 
 
 8 
 
 26-37-5-53-75 
 
 51 
 
 95-0 
 
 808 
 
 39-3 
 
 
 R. V. Eugenie, 
 
 1913 
 
 9,727 
 
 10,180 
 
 4 
 
 8 
 
 (29-43-47-47 and } 
 \ 2 L.P. Turbs. f 
 
 42 
 
 113 
 
 791 
 
 
 
 Kaisar-i-Hind, . . 
 
 1915 
 
 11,430 
 
 15,706 
 
 2 
 
 8 
 
 30-5-44-61-87 
 
 54 
 
 87-5 
 
 7SS 
 
 40-3 
 
 
 Britannic, . . . 
 
 1915 
 
 47,500 
 
 53,000 
 
 3 
 
 8 
 
 (54-84-97-97 and 1 
 1 1 L.P. Turb \ 
 
 7? 
 
 77-0 
 
 962 
 
 40-0 
 
 
 Justicia, 
 
 1917 
 
 32,234 
 
 41,500* 
 
 3 
 
 9 
 
 (".V5-.-.6 64-<".4 \ 
 Unc\ 1 I.. P. Turb.i 
 
 60 
 
 85-0* 
 
 850 
 
 •• 
 
 
 Approximate figures 
 
EXAMPLES OF OCEAN EXPRESS STEAMERS. 
 
 741 
 
 Steamers (Reciprocating or Mixed Engines). 
 
 
 Boilers. 
 
 Trial Results. 
 
 Machinery. 
 Weight. 
 
 Trial LH.P. per 
 
 Machinery 
 on of 
 :ement. 
 
 
 Number and Kind. 
 
 Steam 
 Pres- 
 sure. 
 
 Grate ™»> 
 a rpa 1 Heating 
 
 Area - Surface. 
 
 Total 
 LHP. 
 
 Mean 
 Speed. 
 
 Total. 
 
 Per 
 LH.P. 
 
 Ton of 
 
 Ma- 
 chinery 
 
 G.R.T. 
 
 Sq. Ft. 
 
 of 
 Grate. 
 
 100 Sq 
 Ft. of 
 T.H.S. 
 
 Weight of 
 
 per T 
 
 Displa< 
 
 
 
 
 Lbs. 
 
 Sq. Ft. Sq. Ft. 
 
 
 Knots. 
 
 Tons. 
 
 Lbs. 
 
 
 
 
 
 
 
 Cylindrical, . . 
 
 . 
 
 110 
 
 1,575 
 
 38,062 
 
 13,300 
 
 18-30 
 
 2,205 
 
 371 
 
 6-03 
 
 1-80 
 
 8-44 
 
 35-0 
 
 •161 
 
 
 i) 
 
 . 
 
 110 
 
 1,680 1 38,817 
 
 14,300 
 
 19-00 
 
 2,330 
 
 365 
 
 6-13 
 
 1-76 
 
 8-51 
 
 39-4 
 
 •150 
 
 
 6 D.E. Cylindrical, . 
 
 160 
 
 6261 17,640 
 
 6,751 
 
 16-50 
 
 1.343 
 
 446 
 
 5-03 
 
 1-07 
 
 10-80 
 
 38-4 
 
 •128 
 
 
 >« >> 
 
 . 
 
 150 
 
 632 15,610 
 
 6,347 
 
 17-00 
 
 • • 
 
 
 .. 
 
 0-97 
 
 10-00 
 
 40-7 
 
 
 
 » *3 
 
 F.D. 
 
 180 
 
 1,154 
 
 40,972 
 
 17,000* 
 
 21-00 
 
 2,778 
 
 366 
 
 6-12 
 
 1-96 
 
 14-70 
 
 41-5 
 
 •103 
 
 
 !« »> 
 
 . 
 
 160 
 
 530 
 
 15,107 
 
 5,525 
 
 16-50 
 
 1,257 
 
 510 
 
 4-40 
 
 0-94 
 
 10-40 
 
 36-6 
 
 •130 
 
 
 » >> 
 
 . 
 
 155 
 
 1,450 
 
 47,000 
 
 16,000 
 
 20-50 
 
 2,430 
 
 340 
 
 6-59 
 
 1-66 
 
 11-00 
 
 34-0 
 
 •159 
 
 
 4 D.E. 
 
 F.D. 
 
 160 
 
 520 
 
 20,193 
 
 10,125 
 
 18-00 
 
 1,414 
 
 313 
 
 7-16 
 
 1-71 
 
 19-40 
 
 49-9 
 
 •123 
 
 
 5 D.E. & 2 S.E. „ 
 
 
 160 
 
 756 
 
 26,004 
 
 8,386 
 
 18-50 
 
 1,290 
 
 342 
 
 6-55 
 
 1-21 
 
 11-09 
 
 32-2 
 
 •122 
 
 
 6 D.E. 
 
 F.D. 
 
 160 
 
 837 
 
 22,964 
 
 11,656 
 
 18-50 
 
 1,720 
 
 330 
 
 6-78 
 
 1-50 
 
 13-90 
 
 50-7 
 
 •112 
 
 
 12 D.E. 
 
 . 
 
 165 
 
 2,630 
 
 82,200 
 
 29,600 
 
 22-00 
 
 4,440 
 
 336 
 
 6-69 
 
 2-28 
 
 11-20 
 
 36-0 
 
 •202 
 
 
 12 D.E & 2 S.E. „ 
 
 . 
 
 178 
 
 2,618 
 
 84,285 
 
 30,000 
 
 22-80 
 
 4,460 
 
 333 
 
 6-73 
 
 2-09 
 
 11-50 
 
 35-6 
 
 •190 
 
 
 15 D.E. 
 
 F.D. 
 
 192 
 
 1,962 
 
 74,686 
 
 28,000 
 
 21-00 
 
 4,414 
 
 353 
 
 6-34 
 
 1-54 
 
 13-60 
 
 35-5 
 
 •170 
 
 
 16 S.E. 
 
 F.D. 
 
 175 
 
 1,225 
 
 45,565 
 
 23,000 
 
 22-60 
 
 2,505 
 
 244 
 
 9-18 
 
 2-00 
 
 18-80 
 
 50-4 
 
 •163 
 
 
 12 D.E & 4 S.E. „ 
 
 F.D 
 
 220 
 
 2,188 
 
 85,468 
 
 38,900 
 
 23-50 
 
 5,670 
 
 326 
 
 6-86 
 
 2-36 
 
 17-90 
 
 45-6 
 
 •231 
 
 
 12 D.E & 7 S.E. „ 
 
 
 225 
 
 3,121 
 
 10,643 
 
 40,000 
 
 23-50 
 
 5,780 
 
 324 
 
 6-91 
 
 2-00 
 
 12-80 
 
 37-2 
 
 •218 
 
 
 D.E. 
 
 F.D. 
 
 215 
 
 1,298 
 
 52,140 
 
 22,000 
 
 18-50 
 
 4,145 
 
 422 
 
 5-00 
 
 1-12 
 
 16-90 
 
 42-2 
 
 •160 
 
 
 6 D.E. & 3 S.E. „ 
 
 F.D. 
 
 220 
 
 1,125 
 
 47,141 
 
 18,508 
 
 19-65 
 
 2,383 
 
 288 
 
 7-77 
 
 1-30 
 
 16-50 
 
 39-3 
 
 •131 
 
 
 7 D.E. 
 
 F.D. 
 
 220 
 
 2,970 
 
 101,900 
 
 45,000 
 
 23-50 
 
 6,530 
 
 325 
 
 6-89 
 
 2-32 
 
 15-2 
 
 44-1 
 
 •242 
 
 
 6 D.E. 
 
 F.D. 
 
 210 
 
 777 
 
 29,300 
 
 13,950 
 
 18-00 
 
 2,330 
 
 380 
 
 6-00 
 
 0-93 
 
 18-00 
 
 47-3 
 
 •122 
 
 
 24 D.E. & 5 S.E. „ 
 
 F.D. 
 
 210 
 
 5,400 
 
 151,000 
 
 54,836 
 
 22-50 
 
 8,580 
 
 355 
 
 6-30 
 
 1-19 
 
 10-15 
 
 36-3 
 
 •163 
 
 
 18 S.E. 
 
 F.D. 
 
 200 
 
 1,205 ' 46,950 
 
 19,000 
 
 20-50 
 
 • • 
 
 
 .. 
 
 1-30 
 
 15-70 
 
 40-5 
 
 
 
 8 S.E. 
 
 F.D. 
 
 220 
 
 590 
 
 23,000 
 
 8,500 
 
 17-50 
 
 1,784* 
 
 465 
 
 4-82 
 
 0-797 
 
 14-40 
 
 37-0 
 
 •176 
 
 
 7 S.E. 
 
 F.D. 
 
 180 
 
 480 
 
 20,660 
 
 10,840 
 
 18-10 
 
 1,743* 
 
 360 
 
 6-22 
 
 1-11 
 
 22-60 
 
 52-4 
 
 •142 
 
 
 4 D.E. & 4 S.E „ 
 
 F.D. 
 
 215 
 
 743 
 
 30,570 
 
 11,530 
 
 18-00 
 
 2,296 
 
 446 
 
 5-02 
 
 1-01 
 
 15-50 
 
 37-7 
 
 •146 
 
 
 24 D.E. & 5 S.E. „ 
 
 F.D. 
 
 210 
 
 5,400 
 
 150,958 
 
 54,800 
 
 22-5 
 
 8,600 
 
 352 
 
 6-37 
 
 1-15 
 
 10-15 
 
 36-3 
 
 •162 
 
 
 12 D.E. 
 
 F.D. 
 
 215 
 
 1,416 
 
 61,344 
 
 25,000* 
 
 18-0* 
 
 4,167* 
 
 380 
 
 6-00 
 
 0-775 
 
 17-60 
 
 40-8 
 
 •104 
 
742 
 
 MANUAL OF MARINE ENGINEERING. 
 
 TABLE LXXXVII.— Examples of Passenger- 
 
 
 
 When 
 
 Built. 
 
 Tonnage. 
 
 No. 
 
 of 
 
 Screws. 
 
 Cylinders. 
 
 Revolu- 
 tions 
 per 
 Min. 
 
 Speed 
 
 of 
 Piston. 
 
 Re- 
 ferred 
 Mean 
 Pres- 
 sure. 
 
 
 Name of Ship. 
 
 Grosa 
 Register 
 
 Displace- 
 ment. 
 
 No. 
 
 Diameters. 
 
 Stroke. 
 
 
 
 
 Tons. 
 
 Tons. 
 
 
 
 Inches. 
 
 Ins. 
 
 
 Feet 
 
 Lbs. 
 
 
 Martello, 
 
 1884 
 
 3,721 
 
 7,800 
 
 1 
 
 3 
 
 3J -50-82 
 
 57 
 
 59-0 
 
 561 
 
 26-2 
 
 
 Cufic, . 
 
 
 1888 
 
 4,827 
 
 10,600 
 
 1 
 
 3 
 
 27-14-5-74 
 
 60 
 
 67-0 
 
 670 
 
 33-3 
 
 
 Cazengo: . 
 
 
 1889 
 
 2,889 
 
 5,180 
 
 1 
 
 3 
 
 32-48-80 
 
 48 
 
 73-0 
 
 584 
 
 34-2 
 
 
 Bovic, . a , 
 
 
 1892 
 
 6,583 
 
 14,500 
 
 2 
 
 6 
 
 22-5-36-5-60 
 
 48 
 
 80-0 
 
 640 
 
 34-5 
 
 
 Kensington, . , 
 
 
 1894 
 
 8,669 
 
 17,850 
 
 2 
 
 8 
 
 25-5-37-5-52-5-74 
 
 54 
 
 87-0 
 
 783 
 
 40-0 
 
 
 Georgic, . , 
 
 
 1895 
 
 10,077 
 
 20,000 
 
 2 
 
 6 
 
 24-39-5-65-5 
 
 51 
 
 77-0 
 
 655 
 
 34-5 
 
 
 K. Wilhelmina, 
 
 
 1896 
 
 4,248 
 
 6,575 
 
 1 
 
 4 
 
 27-39-56-80 
 
 54 
 
 68-2 
 
 614 
 
 36-0 
 
 
 Othello, 
 
 
 1896 
 
 5,059 
 
 11,200 
 
 1 
 
 3 
 
 24-75-40-70 
 
 48 
 
 71-0 
 
 568 
 
 36-7 
 
 
 Cymric, . , 
 
 
 1898 
 
 12,647 
 
 25,000 
 
 2 
 
 8 
 
 25-5-36-5-53-75-5 
 
 54 
 
 81-0 
 
 729 
 
 36-9 
 
 
 K. Wilhelm I. , 
 
 
 1898 
 
 4,446 
 
 7,845 
 
 1 
 
 \ 
 
 27-5-39-55-82 
 
 60 
 
 73-0 
 
 730 
 
 32-6 
 
 
 Sylvania, . , 
 
 
 1898 
 
 5,600 
 
 12,160 
 
 2 
 
 6 
 
 22-5-365-60 
 
 48 
 
 94-0 
 
 752 
 
 41-5 
 
 
 Cleopatra . , 
 
 
 1899 
 
 6,849 
 
 13,900 
 
 1 
 
 3 
 
 32-54-90 
 
 66 
 
 65-0 
 
 715 
 
 34-4 
 
 
 Saxonia, . , 
 
 
 1900 
 
 13,963 
 
 22,580 
 
 2 
 
 8 
 
 29-41-5-59-84 
 
 54 
 
 78-0 
 
 702 
 
 39-0 
 
 
 Runic. , , 
 
 
 1900 
 
 12,482 
 
 24,600 
 
 2 
 
 8 
 
 22-31-5-46-67 
 
 51 
 
 81-0 
 
 689 
 
 35-3 
 
 
 Celtic, . . t 
 
 
 1901 
 
 20,880 
 
 30,000 
 
 2 
 
 8 
 
 33-47-5-68-5-98 
 
 63 
 
 80-0 
 
 840 
 
 35-2 
 
 
 Missourian, . , 
 
 
 1902 
 
 7,914 
 
 17,200 
 
 2 
 
 6 
 
 25-42-5-72 
 
 48 
 
 70-0 
 
 560 
 
 35-0 
 
 
 Tahiti, 
 
 
 1904 
 
 7,585 
 
 12,500* 
 
 2 
 
 6 
 
 30-50-80 
 
 54 
 
 85-0 
 
 765 
 
 42-1 
 
 
 Amsterdam, . 
 
 
 1906 
 
 16,967 
 
 25,550* 
 
 2 
 
 8 
 
 29-41-5-61-87 
 
 6<* 
 
 85-0 
 
 850 
 
 40-8 
 
 
 Orcomo, 
 
 
 1905 
 
 11,546 
 
 18,500* 
 
 
 8 
 
 26-37-5-53-5-76 
 
 54 
 
 80-0 
 
 720 
 
 40-9 
 
 
 Rotterdam, . , 
 
 
 1908 
 
 24,149 
 
 30,500* 
 
 2 
 
 8 
 
 33-47-68-97-5 
 
 60 
 
 85-0 
 
 850 
 
 42-0 
 
 
 Edinburgh Castle, 
 
 
 1910 
 
 13,326 
 
 20.000* 
 
 2 
 
 8 
 
 32-46-66-96 
 
 60 
 
 85-0 
 
 850 
 
 40-2 
 
 
 Francon'a, , 
 
 
 1911 
 
 18,154 
 
 24.290 
 
 o 
 
 8 
 
 33-47-67-95 
 
 60 
 
 77-0 
 
 770 
 
 37-4 
 
 
 Port Lincoln, , 
 
 
 1913 
 
 7,243 
 
 12,266 
 
 1 
 
 4 
 
 27-5-39-56-81 
 
 54 
 
 76-8 
 
 691 
 
 38-0 
 
 
 Orion, . 
 
 
 1914 
 
 .. 
 
 19,600 
 
 2 
 
 6 
 
 27-46-76 
 
 48 
 
 95-0 
 
 760 
 
 33-2 
 
 
 Insulinde, 
 
 
 1914 
 
 9,615 
 
 14,080 
 
 2 
 
 6 
 
 29-47-81 
 
 52 
 
 80-0 
 
 693 
 
 32-4 
 
 
 Berrima, 
 
 
 1915 
 
 11,120 
 
 19,789 
 
 2 
 
 8 
 
 23-5-34-5-48-5-70 
 
 54 
 
 78-0 
 
 702 
 
 32-7 
 
 
 J. P. Coen, . 
 
 
 1915 
 
 11,693 
 
 15,600 
 
 2 
 
 6 
 
 27-44-75 
 
 49 
 
 85-0 
 
 693 
 
 32-4 
 
 
 • Approximate figures. 
 
EXAMPLES OF PASSENGER-CARGO STEAMERS. 
 
 743 
 
 Cargo Steamers (Reciprocating Engines). 
 
 
 Boilers. 
 
 Trial Results. 
 
 Weight of 
 Machinery. 
 
 Trial I.H.P. per 
 
 Weight per Ton 
 
 of 
 Displacement. 
 
 
 Number and Kind. 
 
 Steam 
 Pres- 
 sure. 
 
 Grate 
 Area. 
 
 Total 
 Heating 
 Surface. 
 
 Total 
 LH.P. 
 
 Speed. 
 
 Total. 
 
 Per 
 I.H.P. 
 
 Ton'.of 
 
 Ma- 
 :hinery 
 
 G.R.T. 
 
 Sq. Ft. 
 
 of 
 Grate. 
 
 100 
 Sq. Ft. 
 
 of 
 T.H.S. 
 
 
 
 
 
 Lbs. 
 
 Sq. Ft. 
 
 8q. Ft. 
 
 
 Knots. 
 
 Tons. 
 
 Lbs. 
 
 
 
 
 
 
 
 2 D.E. ft 2 S.E. 
 
 sylindrL, 
 
 150 
 
 192 
 
 7,000 
 
 2,320 
 
 12-0 
 
 528 
 
 505 
 
 4-40 
 
 0-624 
 
 12-14 
 
 33-1 
 
 •0680 
 
 
 
 1) 
 
 
 180 
 
 216 
 
 7.392 
 
 2900 
 
 13-0 
 
 533 
 
 412 
 
 5-44 
 
 0-600 
 
 13-43 
 
 39-2 
 
 •0500 
 
 
 2 D.E. 
 
 If 
 
 
 160 
 
 192 
 
 6,750 
 
 3,036 
 
 14-0 
 
 508 
 
 373 
 
 6-00 
 
 1-050 
 
 15-80 
 
 44-9 
 
 •0980 
 
 
 II 
 
 II 
 
 
 180 
 
 346 
 
 11,512 
 
 3,800 
 
 13-0 
 
 795 
 
 481 
 
 4-76 
 
 0-577 
 
 10-98 
 
 33-0 
 
 •0550 
 
 
 S.E. 
 
 M 
 
 F.D. 
 
 200 
 
 382 
 
 18,549 
 
 8,300 
 
 15-8 
 
 1,300 
 
 352 
 
 6-38 
 
 0-968 
 
 21-73 
 
 44-8 
 
 •0730 
 
 
 Jj 
 
 1 
 
 
 180 
 
 394 
 
 13,298 
 
 4,610 
 
 13-0 
 
 840 
 
 408 
 
 5-49 
 
 0-156 
 
 11-70 
 
 34-7 
 
 •042 
 
 
 D.E. 
 
 ft 
 
 
 210 
 
 344 
 
 11,660 
 
 3,350 
 
 14-0 
 
 754 
 
 504 
 
 4-44 
 
 0-788 
 
 9-74 
 
 28-6 
 
 •115 
 
 
 3 S.E. 
 
 II 
 
 
 200 
 
 182 
 
 6,414 
 
 2,430 
 
 11-0 
 
 434 
 
 400 
 
 5-60 
 
 0-456 
 
 11-46 
 
 32-6 
 
 •039 
 
 
 >» 
 
 »• 
 
 
 210 
 
 593 
 
 21,100 
 
 7,310 
 
 15-0 
 
 1,528 
 
 468 
 
 4-78 
 
 0-580 
 
 12-33 
 
 34-7 
 
 •061 
 
 
 D.E. 
 
 II 
 
 F.D. 
 
 210 
 
 253 
 
 9,750 
 
 3,810 
 
 14-5 
 
 769 
 
 452 
 
 4-95 
 
 0-847 
 
 15-10 
 
 38-5 
 
 •098 
 
 
 2 D.E. 
 
 II 
 
 F.D. 
 
 180 
 
 220 
 
 9,610 
 
 5,356 
 
 15-3 
 
 974 
 
 407 
 
 5-50 
 
 0-950 
 
 25-2 
 
 55-7 
 
 •080 
 
 
 2 D.E. & 2 S.E. 
 
 
 F.D. 
 
 200 
 
 806 
 
 13,080 
 
 4,500 
 
 14-5 
 
 897 
 
 446 
 
 5-02 
 
 0-657 
 
 14-70 
 
 34-4 
 
 •065 
 
 
 S.E. 
 
 • 1 
 
 F.D 
 
 210 
 
 570 
 
 25,803 
 
 9,000 
 
 16-0 
 
 1,985 
 
 490 
 
 4-58 
 
 0-651 
 
 15-90 
 
 35-3 
 
 •088 
 
 
 »> 
 
 II 
 
 
 210 
 
 485 
 
 17,100 
 
 5,200 
 
 13-0 
 
 1,185 
 
 510 
 
 4-39 
 
 0-400 
 
 10-72 
 
 30-4 
 
 •047 
 
 
 8 D.E. 
 
 tJ 
 
 
 210 
 
 1,014 
 
 41,680 
 
 13,500 
 
 17-0 
 
 2,975 
 
 493 
 
 4-54 
 
 0-600 
 
 12-33 
 
 30-0 
 
 •099 
 
 
 2 D.E. & 2 S.E. 
 
 II 
 
 F.D. 
 
 200 
 
 314 
 
 12,500 
 
 4,800 
 
 12-0 
 
 900 
 
 420 
 
 5-33 
 
 0-608 
 
 15-30 
 
 38-4 
 
 •041 
 
 
 3 D.E. & 3 S.E. 
 
 II 
 
 F.D. 
 
 180 
 
 557 
 
 22,464 
 
 9,800* 
 
 17-0 
 
 1,790* 
 
 410 
 
 5-46 
 
 1-292 
 
 17-60 
 
 43-6 
 
 •143 
 
 
 7 D.E. 
 
 f 1 
 
 F.D. 
 
 215 
 
 826 
 
 31,800 
 
 12,500* 
 
 16-0 
 
 2,735* 
 
 490 
 
 4-57 
 
 0-735 
 
 15-13 
 
 39-3 
 
 •107 
 
 
 2 D.E & 3 S.E. 
 
 !» 
 
 
 215 
 
 657 
 
 25,047 
 
 8,100* 
 
 15-0 
 
 1,845* 
 
 510 
 
 4-39 
 
 0-701 
 
 12-33 
 
 32-4 
 
 •100 
 
 
 8 D.E. & 4 S.E. 
 
 J) 
 
 F.D. 
 
 215 
 
 1,116 
 
 46,436 
 
 16,500* 
 
 17-0 
 
 3,604* 
 
 490 
 
 4-57 
 
 0-683 
 
 14-79 
 
 35-6 
 
 •115 
 
 
 6 D.E. 
 
 » 
 
 
 220 
 
 1,019 
 
 38,982 
 
 15,000* 
 
 17-5 
 
 8.350* 
 
 500 
 
 4-48 
 
 1-125 
 
 14-72 
 
 38-5 
 
 •167 
 
 
 >i 
 
 II 
 
 
 210 
 
 1,086 
 
 40,100 
 
 12,349 
 
 16-6 
 
 2,850 
 
 512 
 
 4-33 
 
 0-680 
 
 11-39 
 
 30-8 
 
 •118 
 
 
 S.E. 
 
 » 
 
 S.H. 
 
 220 
 
 240 
 
 11,600 
 
 4,100 
 
 12-7 
 
 1,008* 
 
 550* 
 
 4-07 
 
 0-566 
 
 17-1 
 
 35-3 
 
 •082 
 
 
 3 D.E. 
 
 »» 
 
 F.D. 
 
 200 
 
 450 
 
 18,900 
 
 6,943 
 
 14-6 
 
 .. 
 
 .. 
 
 
 • • 
 
 15-4 
 
 36-7 
 
 • • 
 
 
 3 D.E. & 2 S.E. 
 
 II 
 
 F.D 
 
 200 
 
 460 
 
 19,412 
 
 7,000 
 
 15-0 
 
 1,470* 
 
 470 
 
 4-77 
 
 0-728 
 
 15-2 
 
 36-1 
 
 •104 
 
 
 2 D.E. & 2 S.E. 
 
 »» 
 
 F.D. 
 
 215 
 
 440 
 
 18,188 
 
 5,358 
 
 13-5 
 
 1,360 
 
 566 
 
 2-94 
 
 0-482 
 
 12-2 
 
 29-4 
 
 •0687 
 
 
 8 S.E. 
 
 1 
 
 F.D 
 
 210 
 
 520 
 
 21,175 
 
 6,800 
 
 15-0 
 
 1,410* 
 
 465 
 
 4-82 
 
 •• 
 
 131 
 
 32-4 
 
 •0904 
 
744 
 
 MANUAL OF MARINE ENGINEERING. 
 
 TABLE LXXXVIII.— Examples of Small Passenger- 
 
 
 
 
 When 
 Built. 
 
 Tonnage. 
 
 No. 
 of 
 
 Screws 
 
 Cylinders. 
 
 Revolu- 
 tions 
 per 
 Min. 
 
 Speed 
 
 of 
 Piston. 
 
 Re- 
 ferred 
 Mean 
 Pres- 
 sure. 
 
 
 Name of Ship. 
 
 Gross 
 Register 
 
 Displace- 
 ment. 
 
 No. 
 
 Diameters. 
 
 Stroke. 
 
 
 
 
 Tons. 
 
 Tons. 
 
 
 
 Inches. 
 
 Ins. 
 
 
 Feet. 
 
 Lbs. 
 
 
 Lincoln, 
 
 
 
 1882 
 
 1,075 
 
 2,300 
 
 1 
 
 2 
 
 36-69 
 
 39 
 
 71-0 
 
 462 
 
 26-5 
 
 
 Retford, 
 
 
 
 1882 
 
 961 
 
 1,600 
 
 2 
 
 4 
 
 23-44 
 
 27 
 
 108-0 
 
 486 
 
 23-4 
 
 
 Cambridge, . 
 
 
 
 1885 
 
 1,259 
 
 1,970 
 
 2 
 
 4 
 
 30-57 
 
 36 
 
 89-0 
 
 534 
 
 23-8 
 
 
 Eldorado, . , 
 
 
 
 1886 
 
 1,514 
 
 2,200 
 
 1 
 
 3 
 
 28-43-70 
 
 39 
 
 88-0 
 
 572 
 
 31-2 
 
 
 Gazelle, . 
 
 
 
 1889 
 
 672 
 
 790 
 
 2 
 
 6 
 
 16-5-26-41 
 
 30 
 
 121-0 
 
 605 
 
 34-1 
 
 
 Blarney . 
 
 
 
 1889 
 
 1,250 
 
 2,270 
 
 1 
 
 3 
 
 24-38-62 
 
 42 
 
 83-0 
 
 581 
 
 32-5 
 
 
 Lydia, . 
 
 
 
 1890 
 
 1,059 
 
 1,870 
 
 2 
 
 6 
 
 24-37-56 
 
 33 
 
 165-0 
 
 907 
 
 44-0 
 
 
 Duke of Clarence, 
 
 
 
 1892 
 
 1,458 
 
 1,780 
 
 2 
 
 6 
 
 22-34-51 
 
 33 
 
 150-0 
 
 825 
 
 39-8 
 
 
 Bruno, . 
 
 
 
 1892 
 
 841 
 
 1,775 
 
 1 
 
 3 
 
 23-5-35-57 
 
 33 
 
 104-0 
 
 572 
 
 34-6 
 
 
 Chelmsford, . , 
 
 
 
 1893 
 
 1,635 
 
 2,250 
 
 2 
 
 6 
 
 26-39-5-61 
 
 36 
 
 130-6 
 
 780 
 
 31-8 
 
 
 Ibex, . . , 
 
 
 
 1895 
 
 1,062 
 
 1,270 
 
 2 
 
 6 
 
 22-34-51 
 
 33 
 
 152-0 
 
 836 
 
 41-0 
 
 
 Vienna, . 
 
 
 
 1995 
 
 1,753 
 
 2,342 
 
 2 
 
 6 
 
 26-39-5-61 
 
 36 
 
 134-0 
 
 804 
 
 38-2 
 
 
 Roebuck, . , 
 
 
 
 1897 
 
 1,186 
 
 1,740 
 
 2 
 
 6 
 
 23-36-56 
 
 33 
 
 160 
 
 880 
 
 47-2 
 
 
 Prince Edward, 
 
 
 
 1897 
 
 1,414 
 
 1,260 
 
 2 
 
 e 
 
 19-30-48 
 
 24 
 
 192 
 
 768 
 
 36-7 
 
 
 Munster, 
 
 
 
 1897 
 
 2,632 
 
 2.850 
 
 2 
 
 8 
 
 29-45-48-48 
 
 33 
 
 175 
 
 96 
 
 45-2 
 
 
 Duke of Cornwall 
 
 
 1S98 
 
 1,540 
 
 2,368 
 
 2 
 
 6 
 
 22-5-34-38-5-38-5 
 
 33 
 
 160 
 
 880 
 
 44-4 
 
 
 Duke of Devonshire, 
 
 
 1898 
 
 1,205 
 
 1,720 
 
 2 
 
 6 
 
 21-32-50 
 
 30 
 
 175 
 
 875 
 
 48-1 
 
 
 Dresden, 
 
 
 
 1898 
 
 1.839 
 
 2,400 
 
 2 
 
 6 
 
 26-39-5-63 
 
 36 
 
 135 
 
 810 
 
 36-9 
 
 
 Prince George, 
 
 
 
 1899 
 
 2,041 
 
 2,100 
 
 2 
 
 8 
 
 26-40-45-45 
 
 30 
 
 177 
 
 885 
 
 36-2 
 
 
 Princess Mary, 
 
 
 
 1900 
 
 
 2,147 
 
 2 
 
 6 
 
 23-2-37-4-59-5 
 
 36 
 
 122-5 
 
 735 
 
 34-5 
 
 
 Anglia, . , 
 
 
 
 1901 
 
 1,863 
 
 2,300 
 
 2 
 
 8 
 
 26-40-43-43 
 
 33 
 
 175 
 
 962 
 
 42-4 
 
 
 Colchester, . , 
 
 
 
 1901 
 
 1,160 
 
 1,970 
 
 2 
 
 8 
 
 14-5-23-27-27 
 
 33 
 
 145 
 
 788 
 
 44-6 
 
 
 Antrim, 
 
 
 
 1905 
 
 2,100 
 
 2,150 
 
 2 
 
 8 
 
 23-36-42-42 
 
 30 
 
 195* 
 
 955 
 
 44-0 
 
 
 Princess Juliana, 
 
 
 
 1909 
 
 2,885 
 
 2,760* 
 
 2 
 
 8 
 
 28-43-5-49-49 
 
 33 
 
 170* 
 
 935 
 
 40-2 
 
 
 Patriotic, 
 
 
 
 1912 
 
 2,254 
 
 3,500* 
 
 2 
 
 8 
 
 21-5-35-41-41 
 
 36 
 
 160* 
 
 960 
 
 42-0 
 
 
 Bernstorff, . , 
 
 
 
 1913 
 
 2,310 
 
 3,140 
 
 1 
 
 4 
 
 27-44-5-52-52 
 
 42 
 
 110 
 
 770 
 
 38-4 
 
 
 Jupiter, . , 
 
 
 
 1916 
 
 •• 
 
 3,875 
 
 1 
 
 3 
 
 26-42-69 
 
 42 
 
 106 
 
 742 
 
 36-9 
 
 
 * Approximate figures. 
 
EXAMPLES OF SMALL PASSENGER-CARGO STEAMERS. 
 
 744a 
 
 Cargo Steamers (Reciprocating Engines). 
 
 
 Boilers. 
 
 Trial Results. 
 
 Weight of 
 Machinery. 
 
 Triai I.H.P. per 
 
 Weight per Ton 
 
 of 
 Displacement. 
 
 
 Number and Kind. 
 
 Steam 
 Pres- 
 sure. 
 
 Grate 
 Area. 
 
 Total 
 Heating 
 Surface. 
 
 Total 
 I.H.P. 
 
 Speed. 
 
 Total. 
 
 Per 
 I.H.P. 
 
 Ton of 
 
 Ma- 
 chinery 
 
 G.R.T. 
 
 Sq. Ft. 
 
 of 
 Grate. 
 
 100 
 Sq.Ft. 
 
 of 
 T.H.S. 
 
 
 
 
 Lbs. 
 
 Sq. Ft. 
 
 Sq Ft. 
 
 
 Knots. 
 
 Tons. 
 
 Lbs. 
 
 
 
 
 
 
 
 2 S.E. cylind 
 
 rical, 
 
 83 
 
 100 
 
 3.450 
 
 1,390 
 
 12-7 
 
 233 
 
 376 
 
 6-00 
 
 1-29 
 
 13-9 
 
 40-3 
 
 •101 
 
 
 J» 3 
 
 
 83 
 
 sa 
 
 2,700 
 
 1.065 
 
 13-0 
 
 174 
 
 368 
 
 6-12 
 
 1-12 
 
 12-83 
 
 39-4 
 
 •109 
 
 
 2 D.E. 
 
 
 85 
 
 220 
 
 5,820 
 
 2,373 
 
 15-03 
 
 386 
 
 364 
 
 6-15 
 
 1-88 
 
 10-79 
 
 40-8 
 
 •196 
 
 
 2 D.E. , 
 
 , F.D. 
 
 154 
 
 140 
 
 5,620 
 
 2,082 
 
 15-00 
 
 293 
 
 315 
 
 7-10 
 
 1-45 
 
 14-87 
 
 37-1 
 
 •133 
 
 
 ?S.E. , 
 
 , F.D. 
 
 150 
 
 118 
 
 3,800 
 
 1,650 
 
 17-00 
 
 250 
 
 339 
 
 6-60 
 
 2-45 
 
 14-00 
 
 43-4 
 
 •372 
 
 
 '£ D.E. 
 
 
 160 
 
 126 
 
 4,080 
 
 1,550 
 
 13-00 
 
 289 
 
 418 
 
 5-36 
 
 1-24 
 
 12-30 
 
 38-0 
 
 •123 
 
 
 2 D.E. 
 
 , F.D. 
 
 160 
 
 320 
 
 11,956 
 
 5,977 
 
 19-50 
 
 530 
 
 200 
 
 11-09 
 
 5-64 
 
 18-7 
 
 49-0 
 
 •283 
 
 
 2 D.E. 
 
 , F.D. 
 
 160 
 
 258 
 
 8,000 
 
 4,000 
 
 18-50 
 
 420 
 
 235 
 
 9-51 
 
 2-75 
 
 15-5 
 
 50-0 
 
 •236 
 
 
 2 S.E. 
 
 
 150 
 
 114 
 
 3,920 
 
 1530 
 
 14-25 
 
 192 
 
 281 
 
 7-97 
 
 1-82 
 
 10-20 
 
 39-2 
 
 •108 
 
 
 5 S.E. 
 
 
 160 
 
 357 
 
 11,450 
 
 4,400 
 
 18-00 
 
 593 
 
 302 
 
 7-42 
 
 2-69 
 
 12-32 
 
 38-4 
 
 •204 
 
 
 2 D.E. 
 
 , F.D 
 
 160 
 
 250 
 
 8,000 
 
 4,250 
 
 19-00 
 
 420 
 
 221 
 
 10-12 
 
 4-00 
 
 17-00 
 
 53-1 
 
 •331 
 
 
 5 S.E. 
 
 , F.D. 
 
 160 
 
 357 
 
 11,200 
 
 5,450 
 
 18-00 
 
 587 
 
 241 
 
 9-29 
 
 3-11 
 
 15-27 
 
 48-6 
 
 •251 
 
 
 2 D.E. 
 
 , F.D. 
 
 175 
 
 250 
 
 10,850 
 
 6,200 
 
 20-00 
 
 603 
 
 218 
 
 10-27 
 
 5-23 
 
 24-8 
 
 57-4 
 
 •347 
 
 
 2 D.E. , 
 
 , F.D 
 
 190 
 
 170 
 
 6 200 
 
 3,100 
 
 18-80 
 
 304 
 
 220 
 
 10-20 
 
 2-19 
 
 18-23 
 
 50-0 
 
 •217 
 
 
 4 D.E. , 
 
 , F.D. 
 
 160 
 
 520 
 
 18,500 
 
 9,500 
 
 23-8 
 
 850 
 
 200 
 
 11-17 
 
 3-61 
 
 18-5 
 
 51-3 
 
 •309 
 
 
 4 S.E. 
 
 , F.D. 
 
 ISO 
 
 317 
 
 10,772 
 
 5,520 
 
 19-8 
 
 612 
 
 248 
 
 9-02 
 
 3-59 
 
 17-40 
 
 51-1 
 
 •249 
 
 
 2 D.E. , 
 
 , F.D. 
 
 160 
 
 257 
 
 8,900 
 
 5,000 
 
 19-5 
 
 505 
 
 226 
 
 9-90 
 
 2-63 
 
 19-40 
 
 56-2 
 
 •294 
 
 
 5 S.E. 
 
 , F.D. 
 
 165 
 
 357 
 
 11,160 
 
 5 641 
 
 18-5 
 
 590 
 
 234 
 
 9-56 
 
 3-06 
 
 15-8 
 
 50-5 
 
 •246 
 
 
 4 S.E 
 
 , F.D. 
 
 180 
 
 280 
 
 11,440 
 
 5,892 
 
 20-3 
 
 568 
 
 212 
 
 10-56 
 
 2-89 
 
 21-70 
 
 52-6 
 
 •227 
 
 
 2 D.E. 
 
 
 155 
 
 318 
 
 11,320 
 
 4,250 
 
 17-9 
 
 439 
 
 231 
 
 9-70 
 
 1-34 
 
 13-40 
 
 37-5 
 
 •205 
 
 
 8 S.E. 
 
 , F.D. 
 
 160 
 
 430 
 
 15,000 
 
 7,170 
 
 21-0 
 
 675 
 
 210 
 
 10-7 
 
 3-85 
 
 16-67 
 
 47-8 
 
 •294 
 
 
 3 S.E. , 
 
 
 180 
 
 218 
 
 6,520 
 
 2 ; 377 
 
 16-1 
 
 385 
 
 363 
 
 6-2 
 
 2-05 
 
 10-90 
 
 36-5 
 
 •184 
 
 
 2 D.E & 1 S.E. , 
 
 , F.D. 
 
 150 
 
 402 
 
 12,460 
 
 7,200 
 
 21-9 
 
 826 
 
 257 
 
 8-72 
 
 3-43 
 
 17-9 
 
 57-8 
 
 •384 
 
 
 4 D.E 
 
 , F.D. 
 
 190 
 
 534 
 
 19,450 
 
 9,350* 
 
 22-5 
 
 890* 
 
 200* 
 
 10-50 
 
 3-24 
 
 17-5 
 
 48-1 
 
 •322 
 
 
 2 D.E. & 1 S.E. , 
 
 , F.D. 
 
 195 
 
 386 
 
 14,437 
 
 6,400* 
 
 18-0 
 
 600* 
 
 210* 
 
 10-67 
 
 3-10 
 
 16-6 
 
 44-3 
 
 •171 
 
 
 
 
 210 
 
 
 .. 
 
 3,800 
 
 16-0 
 
 .. 
 
 • • 
 
 • • 
 
 1-64 
 
 •• 
 
 •• 
 
 •• 
 
 
 3 S.E. 
 
 , F.D. 
 
 190 
 
 180 
 
 7,500 
 
 3,100 
 
 14-5 
 
 •• 
 
 • • 
 
 • • 
 
 
 17-2 
 
 41-3 
 
 
7446 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Weight of Machinery. 
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FIX EI! PART3. 74-5 
 
 CHAPTER XXVIII. 
 
 EFFECT OF WEIGHT — INERTIA AND MOMENTUM — BALANCING THE SAME. 
 
 Hitherto in this book little or no consideration has been given to the effect 
 on the engine of the weight of any of its parts ; and although that of each one 
 has some effect on itself and on its immediate surroundings, it may generally 
 be very trifling. The effect of the weight of some of the parts, however, 
 is perceptible throughout the whole engine, and sometimes even makes its 
 presence known from stem to stern of the ship. In the case of the propellers, 
 both screw and paddle, their weight has been taken into account in esti- 
 mating the stresses on their shafts and the pressure on the bearings supporting 
 them ; also the weight of the shafts themselves has not been overlooked in 
 considering the dimensions of their journals, bearings, etc. But in calculating 
 the sizes of shafts and propellers no reference has been made to the fact 
 that their weight does to some extent modify the stresses borne by their parts. 
 As a rule, however, the effect of mere weight is so very small compared 
 with that due to the steam pressure on the pistons, cylinder covers, etc., 
 that in ordinary engines it is and may be generally neglected. But the 
 great increase in rate of revolutions in modern engines, and the demand for 
 greater perfection in working, as well as the absolute necessity for it in many 
 cases, have drawn more particular attention to the subject ; some most 
 interesting investigations have been made by scientists, and from time to 
 time methods devised whereby the ill effects due to the inertia of the moving 
 parts of an engine may be modified or removed altogether, so that it may be 
 run at any speed without distress of any kind to itself or surroundings, and 
 with perfect freedom from vibration. In modern design it is now unusual 
 to neglect the effects of inertia. 
 
 Fixed Parts — that is, those parts the centre of gravity of which have 
 no motion but that due to the motion of the ship — may first be taken into 
 consideration. The effect of their weight is simply that due to gravity, 
 and acts always vertically. For example, the cylinder and its appurtenances 
 of a vertical engine rest on the columns, and by their weight increase the load 
 on the columns ; on the other hand, the stress on the bolts connecting them 
 to the columns is decreased by it. When, however, the ship is id motion 
 they have momentum, so that in case of a sudden stoppage by grounding or 
 collision the tendency would be lor them to move forward and shear the bolts, 
 etc., and to bend the columns as well as strain the column ba?e« and con- 
 nections to the ship. The tendency is also of the same uature when the 
 ship rolls heavily, but in this case motion is transverse to the forward move- 
 ment. It is usual and advisable in all vertical engines, especially in tall oues, 
 to provide for these contingencies by fitting ties of a more or less flexible 
 nature from the cylinders to a suitable part of the ship. In a minor degree 
 
746 MANUAL OP MARINE ENGINEERING. 
 
 the same remarks are applicable to the columns, especially if of cast iron 
 and heavy, but being lower down and sharing in the external support given to 
 the cylinders by the ties, there is no need to do more than brace them together, 
 as is generally done when the columns are of wrought iron or steel. 
 
 By the above definition the shafting is a " fixed part," whose weight sets 
 up in itself bending moments as well as shearing forces which may be suffi- 
 ciently large to take into account with the twisting moment if the bearings 
 are far apart. A rotating shaft has also critical speeds of whipping, at which 
 the centrifugal forces become unbalanced ; and although in the case of 
 marine shafts dangerous stresses may not necessarily be set up, the result 
 is a considerable amount of vibration and racking on the bearings. As a 
 rule, greater stability is secured at speeds above the critical speed, but in 
 marine practice it is perhaps best, if possible, to arrange for the critical 
 speed to be above the maximum engine speed.* 
 
 Moving Parts. — The effect of weight in these, however, demands the chief 
 attention, and in quick-running engines the greatest attention of the designer, 
 not so much, as will be seen, for its static as for its kinetic effect. Some of 
 these have simple rectilineal motion, others have circular motion, while a few 
 have a combination of both, and move in an elliptic orbit ; others, again, 
 reciprocate on curved lines. The static effect of the weight of the piston, 
 piston-rod, crosshead, and connecting-rod, is to increase the load on the piston 
 on the downstroke and decrease it on the upstroke, in both cases uniformly. 
 The effect of the weight of the crank-arms and pin (if unbalanced) is also to 
 increase the down and decrease the up load, but not uniformly in this case. 
 The effect of the former is generally measured by dividing the weight in 
 pounds by the area of the piston in square inches, and adding the result to 
 the effective mean pressures of the steam on the top of the piston and 
 deducting it from that on the bottom. This is shown graphically by drawing 
 the diagram of effective pressures of the top above a reference line and 
 that of the bottom below it : a new reference line is now drawn below the 
 original at a distance representing the pounds thus ascertained and the 
 ordinates measured from it used to determine the load on rods, etc., and 
 the twisting moments on the crank-shaft. 
 
 Fig. 280 is the indicator diagram from the high-pressure cylinder of 
 a triple-compound engine 31 inches diameter ; the weight of the piston, 
 piston-rods, connecting-rods, etc., is 5,728 lbs. ; dividing this by 755 the 
 
 * The critical speed (R.P.M.) of whipping of a shaft of diametei d (feet) and length 
 L (feet) is 
 
 N = 393,000 =-| for shaft supported on bearings. 
 
 Example. — A shaft 10 inches (0834 ft.) diameter is supported on bearings 25 feet 
 apart — 
 
 N = 393,000 X -f£t = 524 K.P.M. 
 o25 
 
 If the shaft be regarded as fixed in the bearings, 
 
 N = 2*28 times tho above value. 
 
 Foi more than two equidistant bearings, the above formulae still hold, but fo» 
 bearings unequally spaced the problem is complicated. These formulae will generally 
 indicate what speeds to avoid. The advanced mathematical reader may consult Dr 
 €hrec on " Whirling of Shafts," Proc. Physical Soc, vol. *ix. 
 
MOVING PARTS. 
 
 U] 
 
 area of the piston in square inches, the constant downward pressure due 
 tc the weight is equal to 7 '6 lbs. per square inch. The figure in plain lines 
 is that taken from the top, that in the dotted lines from the bottom of the 
 
 Bot. 
 
 Fig. 280. — High- 
 pressure Cylinder 
 Diagram. 
 
 C n D Fi S- 281.— Diagram of 
 Effective Pressures 
 Neglecting Inertia. 
 
 Fig. 282.— Diagram of 
 /7 " / » > fi§ Effective Pressures 
 
 including those due 
 to Iuertia. 
 
 cylinder. A B is taken as a base line divided into any number of parts at 
 points 6 6', etc., through which ordinates be, b' e', etc., are drawn, and of 
 which a e, a e, etc.. represent the effective pressure on the top of the piston 
 
748 MANUAL OF MARINE ENGINEERING. 
 
 when in those positions on the downstrokes, and c d, c' d', etc., the effective 
 pressures on the underside of the upstroke. 
 
 Now take C D as a base line (fig. 281) of the same length as A B ; divide 
 in the same way and draw a diagram by taking ordinates through the points, 
 cutting off a e, a' e', etc., the same length as a e, a' e', etc., in fig. 280, etc. 
 On the lower side construct in a similar way the figure V d n d' G from the 
 dotted diagram, fig. 281. 
 
 Now take a point E below C, so that C E represents on the same scale 
 as the diagrams the pressure per square inch on the pi3ton equivalent to the 
 weight of the piston, etc., and draw E F parallel to C D. Then any ordinates 
 on E F, intercepted by Gee' P, etc., represent the true effective pressure on 
 the downstroke, and similarly any ordinates from E F downward, intercepted 
 by P d' G represent the true effective pressure on the upstroke at those points. 
 The maximum ordinate on the lower side indicates the tension on the con- 
 necting-rod and piston-rod bolts for purposes of calculation, and the maximum 
 ordinate on the top side of E F indicates the compressive load on the con- 
 necting-rod, etc., for the same purpose ; the actual load being obtained, of 
 course, by multiplying the number of pounds which these ordinates represent 
 by the number of square inches of piston area. The curve of twisting moments 
 of a vertical engine should also be obtained by using these ordinates. 
 
 A further modification may be made in the case of engines having the air 
 and circulating pumps worked by means of levers, as then the weight of the 
 moving parts of the pumps and their loads tend to balance the weight of the 
 piston. Balance weights on the cranks opposite the arms have also been 
 fitted with the object of balancing the weights of pistons, etc., and have also 
 very properly been fitted with the sole object of balancing the crank weight only. 
 
 The weight of the slide valves and valve motions also modify in a similar 
 way the stresses on their parts but in a minor degree. 
 
 In horizontal engines the pistons, rods, etc., do not by gravity affect the 
 load on the pistons, but the weight of the crank-arms and part of the weight 
 of the connecting-rod do affect the turning moment, but in a different way 
 from that of a vertical engine. In their case the forward effort of the piston 
 was resisted by the weight of crank, etc., till a little past the half-stroke 
 and assisted from the piston to the end of the stroke ; on the back-stroke 
 the assistance continued till nearly half-stroke, when it ceased, and resistance 
 set up and increased to the end. Here balance weights could be, and were, 
 used to advantage, so that most of the horizontal engines had them, while a 
 vertical engine was rarely seen with these appendages. 
 
 Momentum. — After all is said, the effect of mere weight or gravity on 
 the parts ot an engine is very limited and never serious — in fact, may generally 
 be neglected as has been shown. But the forces set up by heavy moving parts 
 when in motion — especially in rapid motion — are quite different both in 
 magnitude and effect. To overcome the inertia of a heavy body and start it 
 into motion a considerable force is required ; to accelerate the motion still 
 moie force is required. A heavy body when in rapid motion has much 
 energy stored in it, all of which must be abstracted in order to bring it to 
 rest ; if the process of bringing to rest is accomplished quickly the force 
 developed vvill be large and produce disastrous results, unless special pro- 
 vision is made foT otbetwi.se absorbing it. This force varies with the mass 
 of the body and its acceleration ; hv/.nce in a quick-running engine the inertia 
 
MOMENTUM. 749 
 
 forces set up in the moving parts are much more serious matters than in a 
 slow-moving engine, not only on account of the stresses set up in the parts 
 themselves and in those adjacent to them, but because of the vibratory effects 
 on the structure of the engine generally, as also on the structure by which 
 it is carried. Inertia forces can be modified by reducing the weight of a 
 moving part, but the velocity is practically fixed, being that required for 
 the full working of the engine. Very little scope is permitted, therefore, 
 to the designer to modify the motion of any one part. 
 
 Dealing first with those parts which have simple linear motion, such as 
 the piston, piston-rod, and crosshead, the valves and valve-rods, etc., the 
 general effect can be appreciated by tracing each phase of movement. For 
 example, at the commencement of the stroke when steam is admitted to 
 the cylinder, it exerts a pressure on the piston and cover alike, hence the 
 thrust on the piston (and from it to the bearings) and the resistance of the 
 engine framing are equal but opposite in direction, and, consequently, exactly 
 balance one another since there is no motion. There is, therefore, no load 
 on the parts external to the engine. To produce motion, the inertia of the 
 piston, etc., has to be overcome ; if this is done by the pressure of the steam, 
 the whole force of the steam on the piston is not transmitted through the rods 
 to the crank, for a considerable portion of it is used up in moving the piston, 
 etc.,* and is thus stored in the piston and its parte as energy ; the load on 
 the cover being, therefore, greater than the part transmitted, it tends to make 
 the engine move in the reverse direction to that of the pistons. Further, the 
 movement of the pistons is accelerated in its course by accessions of force till 
 about the half-stroke ; hence during all this time the engine is tending to 
 move ; in fact, if it is a vertical engine and the piston is on the downstroke, 
 it will jump from its bed unless its weight is greater than the accelerating 
 force, and if there are no holding bolts restraining it. During the latter part 
 of the stroke the momentum of the piston, etc., tends to keep it in motion, 
 and consequently a downward thrust on the rods is created in excess of the 
 load on it and the cover, so that the engine now would move the other way, 
 unless restricted by the bed on which it rests. The same phenomena occur 
 in horizontal engines, and are of the utmost consequence in the locomotive 
 engine ; it was, in fact, with this class of engine that the effects of inertia 
 forces were first impressed on engineers by the disastrous results with some 
 outside cylinder horizontal engines when attempts were made to drive them 
 at higher speeds. Fortunately the practical minds of those days soon found 
 a remedy — one that, from its simplicity and efficiency, is still used in loco- 
 motive engines, in spite of what science has done to reveal its faults. 
 
 The general effect of the variation of momentum of the piston, etc., is to 
 reduce the load on the crank pin at the early part of the stroke, and to increase 
 it at the latter, with the reactive tendency to move the whole engine up and 
 down alternately ; in other words, to reduce the ratio of maximum to mean 
 
 * In fast-running engines, such as are used for generating electricity (in which the 
 revolutions are constant while the load varies), the inertia of the pistons and rods is so 
 great that on light loads the mere steam pressure on them would not overcome it ; in 
 fact, the piston of one engine is started and accelerated by the power transmitted from 
 the others or from the flywheel, even when the load is considerable, hence it is most 
 important that such engines should have good " stern-going " piston-rod guides. If 
 an engine is moved by external means, such as a belt from another engine, the effect of 
 inertia of the moving parts is exactly the same as if it were working under steam. 
 
 4S 
 
750 MANUAL OP MARINE ENGINEERING. 
 
 load, and cause vibration on the bed supporting the engine ; as also to reduce 
 the maximum stresses on the rods and their bolts. The vibration set up in 
 the engine bed is, of course, communicated to the ship. 
 
 The general effect on the crank of the variation of momentum of the 
 piston-rods, etc., will be dealt with later on, but in the meanwhile can be 
 shown graphically by calculating the accelerating forces applied to the piston 
 at a series of points at and from the commencement of the stroke, and taking 
 as ordinates the equivalent pressures per square inch on the piston ; H J K 
 (fig. 281) is such a curve obtained by making E H represent the pressure on 
 the piston required to overcome its inertia, and so also with the other ordinates. 
 If, now, a new base line, L M, fig. 282, is taken and a curve, N e' S P, drawn 
 by taking as ordinates the distances intercepted by the curves HJK and 
 G e e" P, a true representation of the variation in real push and pull of the 
 rods is shown. The ordinates measure the combined effect of steam pressure 
 on the piston, gravity, and inertia of moving parts. The cushioning is shown 
 by the shaded portions at each end, S M P and R N L. In this case it may 
 be mentioned that the diagram, fig. 280, is that of a three-crank triple-com- 
 pound engine with the high-pressure crank leading, and to show the difference 
 in effect when the medium-pressure crank leads the high pressure, the dot and 
 dash line (fig. 282) is the curve of equivalent loads for a similar engine under 
 that condition. 
 
 The general result of the movement of the slide valves, link motions, 
 etc., is similar, and although their weight is less than that of the piston, etc., 
 and their motion much more restricted, their varying momentum can set 
 up a manifest vibration in a quick-running engine. 
 
 The connecting-rod has a complex motion, inasmuch as one end moves 
 in a straight line while the other moves in a circle ; its centre of gravity 
 consequently moves in an ellipse whose major axis is equal to the stroke of 
 the piston. Its motion, therefore, may be viewed as partly linear with the 
 piston-rod, etc., and partly circular with the crank. But when in motion it 
 oscillates like a pendulum, so that its effect on the crosshead will be the same 
 as that of a pendulum on its point of support ; hence it will set up a horizontal 
 force there in addition to the downward one due to gravity. The force will 
 vary with the mass and the horizontal acceleration ; it will set up a horizontal 
 vibration in the engine and increase the stresses on the rod due to cross 
 bending ; hence in very fast-running engines the connecting-rod should be 
 made as light as possible consistent with strength. A short rod is, of course, 
 lighter than a longer one, but it gives a greater thrust, due to steam pressure, 
 and, consequently, tends to increase vibration transversely. 
 
 The crank, including arms and pin, moves in a circle, and consequently 
 centrifugal forces are set up which tend to press the shaft against the journals 
 in the radial direction of the crank pin for the time being. The weight of the 
 crank -pin brasses and that part of the connecting-rod surrounding them, 
 moves in a circle, and causes further centrifugal forces to act on the crank- 
 shaft. These parts by their momentum also set up tangential forces, which, 
 with the resisting force at the bearings, form a couple tending to keep the shaft 
 in motion ; but as the direction changes, so the reacting force changes in 
 direction, and this also tends to rack the bearings and foundation, as well as 
 to set up vibration. 
 
 Balancing. — From the above causes a reciprocating engine of any kind 
 
BALANCING. 751 
 
 when in motion sets up vibrating forces which are only extinguished by 
 doing work of some kind, mostly, however, of a mischievous character. Even 
 the slow-running paddle engine developed the vice, which, when the engine 
 is nearly horizontal and has a single cylinder, as in certain Clyde steamers, is 
 very perceptible fore and aft ways. When the engine was directly coupled 
 to the screw shafting, and its revolutions consequently three or four times 
 that of the older engines, vibration and its effects were soon noticed. Although 
 some general idea of the causes was grasped, and attempts were made to 
 grapple with the evil by fitting balance weights opposite the cranks or in the 
 turning wheel at the end of the crank shaft, no true insight seems to have 
 been gained ; consequently, for more than half a century engineers generally 
 were content to put down the vibration of screw steamers as due to the 
 propeller, and to admit that the paddle wheel was in that respect the superior 
 for passenger steamers. 
 
 Now. any force which has to be extinguished by doing useless work is a 
 force wasted, and if the work done is injurious to engine or hull, the loss is 
 worse still. For this reason the engineers of the past are to blame for not 
 having found means to prevent such waste, and still more so, those who now 
 neglect the efficient methods which have been devised for preventing such 
 forces coming into existence, and for rendering them harmless when developed. 
 
 Mr. Arthur Rigg called attention to the subject in his able work so far 
 back as 1878, but it is due to Sir A. Yarrow's genius for investigation that 
 the question was put in practical form, and to Mr Otto Schlick's analytical 
 ability that methods were devised whereby balancing was no longer an 
 interesting workshop experiment, but an exact science and a fine art; it 
 has been rendered even more so by the later contributions of Prof. Dalby, 
 Mr. Inglis, and others to the Transactions of the Inst of Naval Architects, etc. 
 
 To understand clearly what is involved in the balancing of an engine, 
 it will be best to take, first, the common case of a double-acting single-crank 
 engine. Suppose it to have a flywheel with a weight near its rim of such 
 a size and position that at half-stroke it statically balances the crank, rods, 
 and piston. When this engine is in motion the momentum of the principal 
 parts is roughly balanced for turning by that of the balance weight ; but a 
 couple is formed by joining the centre of the gravity of the weight by a line 
 going diagonally across the engine to its centre line. The shaft and its bearings 
 will, therefore, be racked by this cross-section couple, and the engine tend to 
 tilt and twist on its bed. Moreover, since the momentum of the balance 
 weight is constant tangentially, while that of the pistons, rods, etc., ceases 
 at the ends of the stroke, there must be an excess momentum of the balance 
 weight at those periods which, in the vertical engine, tends to make it move 
 horizontally, and in the horizontal engine vertically. This evil is made 
 apparent in a locomotive engine by what is called "rail hammering," from 
 the blows the rail gets from the wheels with the balance weights when the 
 engine passes the dead centres. 
 
 If the balance weight is placed on one crank-arm mly, as was some- 
 times done for cheapness, similar actions were set up, but less in degree. 
 Two balance weights, one on each arm, make the proper balance, but still 
 fail to avoid the horizontal overbalance when the crank passes the dead 
 points and the consequent horizontal vibration. If only balance weights 
 opposite the cranks can be provided, a compromise must be arrived at, so 
 
752 MANUAL OF MARINE ENGINEERING. 
 
 that tlieir weight is such that the horizontal vibration due to overbalance 
 is reduced to that of the vertical vibration caused by underbalance. If. 
 however, the engine has an air puinp, etc., worked by means of levers so that 
 they move in the opposite direction to that of the piston, their weight may 
 be/ so adjusted that their momentum balances that of the piston and rods, 
 while the balance weights on the crank balance the crank-pin brasses and 
 connecting-rod end. There will, however, still remain a couple tending to 
 rock the engine athwart ship. 
 
 Mr. J.H.MacAlpine designed a quadruple engine (fig. 283) with the cylinders 
 in pairs, one behind the other, and each pair operating on the same crank by 
 means of rocking levers like those of the air pump. This gives, no doubt, a 
 very well-balanced arrangement fairly free from vibration, but the ratio of 
 maximum to mean torque is greater than that of a three-crank engine. 
 
 Further, the motion of the valve and gear in this single engine will create 
 inertia forces which all tend to set up vertical vibration, and should be taken 
 account of especially if it is a fast-running one. This can be done by finding the 
 resultant forces from all the parts, and placing the balance weights accordingly. 
 
 Take now the case of a two-crank engine. If the cranks are opposite one 
 another they will balance for rotation and so require no weights, and if the 
 moving parts are of equal weight their relative momenta will also only tend to 
 make the engine rock in a fore and aft direction on its foundation and not 
 jump on it ; it is true, however, that this tendency to rock will be prevented 
 by the holding down bolts, but with the vibration of the ship as a consequence. 
 Such vibration, however, will not be so troublesome as is the case with an 
 unbalanced single engine tending to jump. In this instance the balancing 
 might be effected by a weight in the flywheel at a slight expense, or else by 
 some form of " bob " weight, suggested bv Sir A. Yarrow, placed beyond 
 one of the cranks so as by its mass and velocity to balance the momentum on 
 the other crank-pin. This, of course, leads naturally to a three-crank engine 
 with the two outer ones opposite the middle crank and the weight of the 
 moving parts of the middle engine equal to those of the two outer- engines 
 together, these being of equal weight. 
 
 This form of engine would seem to be perfectly balanced, and it would 
 actually be so if it had slot crossheads instead of connecting-rods ; as it is, 
 the acceleration of the piston of the middle engine at the commencement of 
 its downstroke is greater than that of the wing engine pistons at commence- 
 ment of their upstrokes with a consequent loss of balance and vibration. 
 
 If a two-crank engine has its cranks at right angles it will be, from the 
 ordinary point of view, a better balanced engine than one with the cranks 
 opposite, although there may be more vibration set up by it. The fact is, 
 that the torque is much more even in this case — that is, the ratio of maximum 
 to mean twisting moment is much less ; and the engine runs more uniformly 
 without a flywheel. This little fact enforces the necessity of clearly dis- 
 tinguishing between the balancing for even circular motion and the balancing 
 necessary to prevent vibration either vertical or horizontal, although generally 
 the balancing to prevent vibration tends to improve the even running of the 
 engine. Balancing of these engines may be done by means of a pair of weights, 
 one at each crank, in such a way that the vertical is not extinguished at the 
 expense of excessive horizontal vibration. 
 
 The same process must be gone through with a three-crank engine whose 
 
BALANCING 
 
 753 
 
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754 
 
 MANUAL OF MARINE ENGINEERING. 
 
 cranks are at an angle of 120° with one another — viz., to ascertain the kinetic 
 resultants of the system, and balance them wholly or partially by means 
 of balance weights placed on the first and third cranks of such a size and 
 position that the engine will neither rock fore and aftways nor jump on its 
 seat ; and, further, it must not have excessive horizontal vibration. The 
 methods of arriving at these results are set out fully later on. 
 
 A four-crank engine, however, is the easier to deal with, as in that 
 case contiguous engines can have their cranks opposite one another so 
 as to balance vertically and nearly balance horizontally and have the 
 minimum rocking moment. Also, as the extreme cranks are at right 
 angles they give a smaller rocking couple than if the angle were larger. 
 Mr. John Tweedy devised an arrangement of cylinders and cranks in a 
 four-crank triple engine whereby a good balance is obtained without the 
 use of any balance weights or other addition to the engines. In his 
 design the low-pressure cylinders with their gear, as being the lightest, 
 are placed at the ends with the high-pressure and medium-pressure 
 cylinders between them. The first and 3econd cranks are nearly, but not 
 quite, opposite one another, as are also the third and fourth; the second 
 is not quite at right angles with the third; in fact, the cranks are all 
 so set as to give the best balance of the whole. This is as perfect an 
 arrangement as can be obtained without the use of balance weights and is 
 such as to be practically without vibration. 
 
 Fig. 72 shows such an engine and fig. 284 the arrangement of the cranks to 
 
 -f-m 
 
 HP CAMK 
 
 TU KNIMd UOHOt TS. 
 m*l.eil£ATCA THAN «£»« iZX 
 
 em less than «»« ejx 
 
 «*! 
 
 *j£Sss»i 
 
 
 HORIZONTAL FOPCLS 
 THC r*HR0W SCHUCK i TWUOY SYSTtU 
 
 Fig. 284.— Position of Cranks and Curves of Force, &c., of Triple-expansion Engine 
 Balanced on the Yarrow-Schlick-Tweedy System. 
 
 suit the conditions of that particular engine. This design is now a favourite 
 one with all requiring a steady running engine with the least vibration of 
 ship, and is being fitted generally into passenger steamers, where the com- 
 fort of the passengers and the preservation of the paint work and decoration, 
 to say nothing of the structure of the ship, is of prime importance. 
 
 In warships, where accuracy of fire may mean victory, well balanced 
 engines must be fitted. The Admiralty therefore insist on all engines 
 being as perfectly balanced as possible — a point in favour of the turbine. 
 
 As, however, a four-crank engine is generally more costly and occupies 
 more space than a three-crank one, and the ratio of its manifold torque 
 (or twisting moment) to the mean torque is greater than that of a three- 
 crank engine, the latter will continue to be largely used. The following 
 investigations and methods of balancing such an engine will be both of 
 interest and importance. 
 
 Figs. 285 and 286 show an engine designed and patented by M. E. 
 Wigzell, whereby a nearly perfect balance is obtained by having one 
 cylinder above another with the pistons moving in opposite directions. 
 Here again, the obliquity of the connecting-rods prevents absolutely 
 
PRELIMINARY DEFINITIONS. 
 
 755 
 
 perfect balance ; in practice, however, the engine behaves as if it were 
 perfectly balanced. This engine has the further merit of occupying very 
 little fore and aft space and is readily started from any position. 
 
 1. Preliminary Definitions. — The direct object of what is termed 
 " balancing an engine," ig to prevent the effect of the motion of its 
 
 Kg. 285. — Wigzsll'a^Engino. 1 rans verse Sectidh. 
 
 moving parts being felt by surrounding objects, nnd not to eliminate any 
 particular stresses in parts of the engine itself. Neither has the object any 
 connection with fluctuation of speed, the theory and action of flywheels not 
 being related to the subject. 
 
756 
 
 MANUAL OF MARINE ENGINEERING. 
 
 It is a natural consequence of balancing an engine that certain stresses 
 are considerably reduced ; that others remain as they were, and that some 
 possibly even may be increased. 
 
 Another point, which may not at first sight appear evident, is that an 
 engine may be very much out of balance in itself, that is, if placed on a 
 rigid foundation, it causes very severe stresses on the holding-down bolts, 
 
 Fig. 2S6.— Wigzell's Engine. Longitudinal Elevation. 
 
 yet if placed in certain positions on a flexible foundation may appear to 
 be in a perfectly balanced condition ; that is to say, it will run without 
 producing any vibrations of the foundation, and in fact can run safely 
 without holding-down bolts. 
 
 The problem of balancing an engine need not necessarily be dependent 
 
REMARKS ON BALANCING. 
 
 757 
 
 on this phenomenon, and very fortunately so, because, although a ship is 
 certainly a flexible foundation for an engine to rest on, so little of the 
 true nature of its flexibility is subject to calculation that, for new *hips 
 particularly, any calculations involving this property are really mere guess- 
 work and therefore better left alone. 
 
 If, however, the engine be perfectly balanced in itself it can cause no 
 vibration to the foundation, whatever may be the nature of its flexibility. 
 
 In practice it is impossible to perfectly balance an ordinary engine nnd 
 it will be shown here within what limits and to what extent an engine 
 can be balanced. All problems will relate particularly to the vertical 
 engine, but the principles involved apply equally well to horizontal and 
 inclined engines. Before proceeding with this, however, attention must 
 be drawn to the fact that the problem of balancing an engine is not in 
 the least concerned either with steam loads and 
 twisting moments due to the same or with the 
 speed of the engine, the dynamical equilibrium of 
 the moving parts being quite independent of how 
 fast or how slow or how varying the speed of the 
 engine may be. Comparative values of the speeds 
 of the various parts are certainly required, for the 
 magnitude of any resultant forces and couples are 
 determined by the speed, but unless those absolute 
 values are required, only comparative values of the 
 speeds of the parts are essential. 
 
 As many engineers, some even of high standing, 
 are apt at times to confuse the matter, and imagine 
 that there is some mysterious connection between 
 steam pressures and balance weights, an enlargement 
 of what has been said under " Momentum," Section 
 2, showing that there is no connection between the 
 two, is given in the following simple demonstration. 
 
 Let fig. 287 represent a diagrammatic engine. 
 Let the steam load on the piston be + P. Then 
 the load on the top cover is - P, and, confining Fig. 2S7. 
 
 attention to the vertical forces, the vertical pressure 
 
 at the main bearings is + P. That is, the steam load causes no pres-ure 
 of the engine on the foundation, A. 
 
 The moving parts of the engine, however, are not weightless, also they 
 never have a uniform velocity in any given direction (in this case, the 
 vertical) — that is, they always have a certain acceleration, positive or 
 negative. Now, since 
 
 force = mass x acceleration 
 
 another vertical force is consequently introduced, and is — 
 
 weight of moving parts x vertical acceleration 
 
 Suppose that when the parts are in the position shown, this force 
 equals + I lbs. acting downwards. Since the whole of the moving parts 
 are concerned, this is equivalent to a load + I pressing the piston down 
 as shown. Now it is self-evident that there cannot be any force - I acting 
 on the cylinder cover, therefore this force + I remains alone and its effect 
 
758 
 
 MANUAL OF MARINE ENGINEERING. 
 
 is consequently to make the engine heavier by + I lbs. At other points 
 the inertia force will act upwards instead of downwards, and it will also 
 have a continuously varying value. 
 
 There are, therefore, up and down forces due to the weight and velocity 
 of the moving parts only, acting on the engine as a whole and recurring 
 every revolution. 
 
 As these are the forces that tend to make the foundation vibrate, the 
 balancing of an engine consists of counteracting the forces so that the 
 moving parts are really made as if weightless when considered dynamically 
 in certain given directions. 
 
 It has also been stated above that the problem of balancing has no 
 direct relation to the speed of the engine. This will be obvious on con- 
 sideration, for any counteracting weights that may be applied must 
 necessarily have their velocity imparted to them by the engine itself, and 
 the factor of velocity is consequently eliminated from both sides of the 
 representative equations. 
 
 2. Harmonic Motion. — The theorem of harmonic motion is stated here 
 because of its great importance and the direct bearing it has on the present 
 subject. 
 
 Suppose a point P (fig. 288) to move uniformly round the circle XYX 1 . 
 Draw two rectangular axes X, O Y, and P N perpendicular to O Y. Then 
 
 Fig. 288. 
 
 if P has a uniform angular velocity N will move along Y O Y 1 , with a motion 
 that is called " Simple Harmonic Motion." 
 
 Let OP = r, POY = 0, 
 
 then O N = r cos 6. 
 
 Let v = linear velocity of P, 
 
 then Velocity of N = —v sin 9 
 
 and 
 
 v z 
 
 Acceleration of N = cos 6. 
 
 r 
 
 The proof of these will be found in most treatises on theoretical mechanics. 
 
 P is called the generating point. 
 
 The angle 8 is called the phase. 
 
 r is called the amplitude. 
 
 The time of one revolution of P is called the period. 
 
HARMONIC MOTIONS. 
 
 759 
 
 A curve of the displacement of N from or of the acceleration of N may 
 be drawn if the abscissae be made proportional to 6 and the ordinates pro- 
 
 Fig. 289. 
 
 portional to the displacement or acceleration. Such a curve is shown in 
 
 fig. 289) by a; cos 6. 
 
 3. To Combine Two or more Harmonic Motions of the same Period- 
 Let P and Q (fig. 290) be the generating points of two harmonic motions 
 
 along the axis Y. 
 
 Then, since the angular velocities of O P and Q are the same and are 
 of the same sign, the angle P O Q is constant. 
 
 Complete the parallelogram POQR. The linear velocities of P and Q 
 are proportional to P, O Q respectively, and are perpendicular to them. 
 
 Therefore OR is proportional to their resultant which is perpendicular 
 to it. 
 
760 MANUAL OF MARINE ENGINEERING. 
 
 Because OP, OQ and angle POQ are constant, O R is constant and 
 has the same angular velocity as P and Q. Therefore R is the generating 
 point of the resultant harmonic motion. 
 
 The combination of any number of harmonic motions of the same period 
 and sign is analogous to the combination of any number of coplanar forces 
 by the polygon of forces. 
 
 4. The Inertia of the Various Parts of an Engine will now be dealt 
 with separately. 
 
 Inertia of Piston, Rod, and Crosshead. — These fittings all being connected 
 together rigidly, and having only one line of motion, the velocity and 
 acceleration of the whole is the same at all points. 
 
 (1) If their movement is controlled by a connecting-rod of infinite length 
 they have a simple harmonic motion, and the acceleration at any angular 
 position 6 of the crank from the top centre is therefore 
 
 v 2 
 f= cos 6 (par. 2). .... (1) 
 
 (2) If their movement is controlled by a connecting-rod of definite length 
 the simple harmonic motion is at once destroyed and the expression for the 
 acceleration is a very complicated one, so much so that it is impossible in 
 practice to use it in its complete form for either algebraical or graphical 
 •calculations. 
 
 The expression correctly stated in definite form is — 
 
 v" 1 / rl 2 cos 2 8 + j- 3 sin 4 d \ 
 
 f= -— I cos 6 + — 775 o — • o m , — (2) 
 
 J r\ (I 1 - r- sin z 6) f / x ' 
 
 •Or, written in the form of a series, 
 
 v ' 2 f „ / 1 1 15 1 V 
 
 f= --\cose + cos2 6(-+j£ + m . ¥ + . . .) 
 
 , rt / 1 1 3 1 \ 
 
 + o«4*(j. ?+ig . 7 + . . .) 
 
 + &c. } . - (3) 
 
 Where I = length of connecting-rod 
 
 and s = length of connecting-rod 4- radius of crank. 
 
 As neither of these expressions are convenient to use, it is necessai-y to 
 
 adopt the nearest approximation that will admit of a tangible interpretation. 
 
 v- 
 The first approximation is/= cos 6, which is the same as (1), and is 
 
 therefore the value of /"to use when the connecting-rod is of infinite length, 
 or, in other words, when its "obliquity" is neglected. 
 The second approximation is 
 
 . v°~( . cos 2 A 
 
 /= - -^cosd + — — \ - - (4) 
 
 and this is sufficiently accurate for all ordinary calculations and makes an 
 almost complete allowance for the obliquity of the connecting-rod, for the 
 other terms of the series diminish very rapidly in value. This expression 
 (4) will therefore be used in the problems of finding the actual values of 
 
 the disturbing forces in an engine. 
 
INERTIA OF PISTON, ROD, AND CROSSHEAD. 761 
 
 The error involved by using this expression instead of the complete series or 
 equation (2) is, if a = 4, 1*56 per cent., and if a = 5, 0-8 per cent. This error is 
 therefore very small indeed, and in any graphical calculations would hardly be 
 appreciable. 
 
 Means of Balancing the Piston and Rod. — Expression (4) may be written 
 
 v 2 v 2 
 
 in the form / = cos 6 cos 2 6. Cos 6 and cos 2 are both simple 
 
 r rs r 
 
 harmonic functions but with this difference, cos has one period per 
 
 revolution of the engine and cos 2 d has two periods per revolution. This 
 
 will be seen at a glance in fig. 289, where both these curves and their 
 
 resultant are drawn. 
 
 Let M x = weight of piston, rod, and crosshead. Now, in order to 
 
 M, v 2 M v 2 
 
 balance the inertia force, F = - — l . — cos 6 } . — cos 2 6, two weights 
 
 g r g rs 
 
 must be fitted so that one of them has an acceleration in the same line of 
 
 v 2 
 motion as the piston, equal to -I cos 6, and the other has an acceleration 
 
 v 2 
 also in the same line of motion, equal to+ — cos 2 0, both weights being 
 
 equal to M x . 
 
 Or, the weights Wj and W 2 may have different values and radii of 
 influence (r a and r 2 ), so long as 
 
 M v 2 W v 2 
 
 - =^- cos 6 = — lil cos (180° + e), i.e., W, r, = M, r 
 gr gr x / i i i 
 
 and 
 
 M v 2 W v 2 
 
 _ ^i_L cos 2 6 = ^ cos 2 (180° + 6), i.e., W„r„= M, r. 
 
 gsr ff sr 2 ■ 
 
 Considering the first force, it will be obvious that if a weight Wj be 
 fitted to revolve with the crank-shaft and opposite to the crank, it will 
 have the required vertical acceleration and, therefore, the required counter- 
 acting inertia force. 
 
 This is the method that is in the majority of cases adopted. 
 
 It has one drawback, however ; another inertia force is introduced in 
 
 a horizontal direction due to its harmonic motion in that direction. 
 
 W v 2 
 This force is equal to - — — - sin 0. 
 
 .9 r \. 
 The main result of this method is really to overbalance the engine in 
 
 a horizontal direction, as will be seen in a subsequent paragraph. 
 
 Practically, two weights would have to be fitted each equal to half 
 Wj, and on each crank web. 
 
 A more correct method is that suggested by Mr. Yarrow to fit two 
 (for practical convenience) "bob" weights, driven by eccentrics, keyed on 
 
 W 
 
 the shaft. —J should equal the weight of the bob, eccentric, and rods, 
 
 'Ji 
 
 and the eccentrics of throw r x must be placed opposite to the crank. The 
 
 only error involved in this method is the slight horizontal inertia force 
 
 due to the weight of the eccentrics and obliquity of the eccentric-rods. 
 
 M v 2 
 Considering the second force - — - — cos 2 S, a mechanical difficulty is 
 
 </sr 
 
7G2 
 
 MANUAL OF MARINE ENGINEERING. 
 
 confronted here in producing a fitting that shall move with an acceleration 
 proportional to cos 2 (180° + 6). Two ways in which this could be done 
 are shown in figs. 291, 291a. The figures are self explanatory. Another 
 way would be by means of a cam fixed to the shaft that would cause a 
 weight to move up and down twice per revolution and satisfy the condition 
 that its acceleration would be always proportional to cos 2 (180° + 6). 
 
 There are possibly other mechanical devices for attaining the same 
 object. 
 
 Now, these methods, although very well in theory, would be altogether 
 too inconvenient for actual practice. Thus it is that this part of the 
 disturbing force on an engine is neglected as far as actual balancing ia 
 
 ZxtaUk 
 
 Figs. 291 and 291a. 
 
 Figs. 292 and 202a. 
 
 concerned. Doing so is generally termed " neglecting the obliquity of the 
 connecting-rod." 
 
 5. Inertia of the Connecting-rod In finding the acceleration of the con- 
 necting-rod a problem of extreme complexity is presented, and it is usual 
 and necessary, therefore, in practice to make certain approximations and 
 assumptions in treating the subject. The connecting-rod having its top 
 end moving in a straight line and its bottom end in a circle, has conse- 
 quently a different total acceleration at all points in its length. Adopting 
 the approximation in par. 4, equation 4, the top end has an acceleration the 
 
 same as the piston-rod, which is therefore/ = (cos d+ — -). 
 
 The bottom end moving with a uniform angular velocity about the axis 
 
 of the shaft has a vertical acceleration cos 6 and a horizontal accelera- 
 
 tion — sin A 
 r 
 
INERTIA OF CONNECTING-ROD. 763 
 
 Any point on the connecting-rod between the centres will have an 
 elliptic path, the minor axis of which varies from 1r to O, and the vertical 
 
 acceleration varies from cos 9 to ( cos 6+ ). If the mass of 
 
 r r \ 8 ) 
 
 the whole rod be supposed concentrated at the centre of gravity the accelera- 
 tion of the whole rod vertically will be 
 
 v 2 / . n cos 2 
 
 / . n cos 2 8\ 
 
 \ C ° 3 d + 1 • — — } ( fi S- 292 ) ' ( 5 > 
 
 If it is desired to find the true value of the inertia of the rod in order to 
 combine it with the "steam" twisting moments, or for reasons of other 
 interest, the strict investigation is outlined in the footnote, but as far as 
 balancing is concerned, it is out of the question for practical reasons to use 
 such elaboration. 
 
 Considering expression (5) for the acceleration, it is obvious from the 
 remarks on the inertia of the piston-rod (§ 4) that a balance weight can 
 be fitted to the engine that shall have an acceleration proportional to cos 6, 
 but that it is impossible practically to fit one having an acceleration 
 
 proportional to cos 2 9. The part . j has then perforce to 
 
 be neglected.* 
 
 ■"or 
 
 * Let 9 be angular position of the crank from the top centre and « the angle the 
 connecting-rod makes with the centre line. Then at any instant the angular velocity 
 
 of the connecting-rod is -=r and the angular acceleration of the connecting-rod is -7-^. 
 
 There are three forces on the rod which determine its resultant acceleration (fig. 292a) :— 
 
 Fi = ( 777 ) a °ting along the rod. 
 
 F 2 = -772 acting at right angles to the rod through H. 
 
 g at 
 
 M 
 
 F 3 = — /acting through G parallel to the centre line. 
 
 Where M = weight of the rod. 
 
 / = acceleration of rod vertically. 
 o = acceleration of piston-rod. 
 
 It may easily be shown that — 
 
 da v cos 
 
 dt V P - r* sin 2 d 
 cPa v>(l 2 -r°-)sin6 
 
 dfi~ r(P - r 2 sm 2 d)? 
 
 v 2 / W 2 cos 2 $ + r 3 sin* 0\ 
 
 / = - — ( cos 6 + i — ) 
 
 V (P - r 2 sin 2 0)f / 
 
 r 
 
 By substitution, each of the three forces can be found. In addition to the forces 
 there is the deadweight of the rod = M acting vertically through G. To find the effect 
 on the twisting moment, the moments of the four forces must be taken about for 
 different values of and added algebraically to the remainder of the twisting moment 
 diagram. The proof of the above formulae may be found in advanced treatises on 
 mechanics and steam. 
 
764 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Means of Balancing the Connecting - rod. — Neglecting, therefore, the 
 
 v*- 
 part involving cos 2 8, the vertical acceleration of the rod is — cos 9. To 
 
 balance by a rotating weight, a weight must be placed in the same 
 position as that for balancing the piston, &c. If M 3 equal the weight of 
 connecting-rod, then the weight W 3 is given by W 3 r = M 3 r. This weight 
 will partly balance the rod in a horizontal direction. It will, however, 
 really overbalance horizontally, because, although the whole rod has a 
 
 vertical acceleration - — cos 6, only part of it has an acceleration— sin 6 
 
 T V 
 
 V 2 
 
 (which has the same maximum value or amplitude as — cos 6, its phase 
 
 being 90° behind). Therefore, if it is desired to balance any particular 
 engine horizontally in preference to vertically, the full weight W 3 must 
 not be put on, but only a portion of it, which for a fairly accurate approxi- 
 
 Tit 
 
 mation may be such that — M 3 r = W 3 r. (fig. 292). 
 
 This is one of the assumptions alluded to at the beginning of the 
 
 paragraph, and really amounts to dividing the connecting-rod into two parts, 
 
 n ni 
 
 and assuming the part — M 3 to be concentrated at the crosshead and y- M 9 
 
 to be concentrated at the crank-pin. 
 
 Professor Dalby gives a division of the rod that is said more accurately 
 to represent the case, the division being as follows : — At the crank pin 
 
 (77b X fl\ 
 — j5 — )> t ne remainder at the crosshead, h, being the distance from 
 
 crosshead centre to centre of oscillation. 
 
 This necessitates an experiment with one of the actual rods to find 
 the centre of oscillation. 
 
 The finding of the centre of gravity does not require any special ex- 
 periment, for in ordinary shop transit the slings are usually placed in 
 this position. In case that either observations are not readily available, 
 the following table of typical connecting rods is given : — 
 
 {v. Fig. 296.) 
 
 Counecting-Rod to 
 
 Fig. 52. 
 
 Fig. 5S. 
 
 Fig. 297. 
 
 Position of centre of gravity from top centre, 
 
 ,, ,, oscillation ,, ,, '. 
 
 m x h 
 
 p . ........ 
 
 •675/ 
 •8S/ 
 
 •594 
 
 ■58 1 
 
 •912/ 
 
 •52S 
 
 643/ 
 •99/ 
 
 •637 
 
 If it is desired to balance the engine as far as possible, both vertically 
 and horizontally, then the reciprocating part of the connecting-rod must 
 be balanced by means of reciprocating bob weights as explained and the 
 rotating part must be balanced by a rotating weight. 
 
 6. Inertia of Crank. — The balancing of the crank is the simplest part 
 of the whole problem. The unbalanced part, of course, consists of the 
 
INERTIA OF AIR PUMP. 
 
 765 
 
 crank webs beyond the shaft and the crank-pin. The acceleration of the 
 
 v 2 v 2 
 
 centre of gravity vertically is/ = cos 0. and horizontally / = — sin 6. 
 
 r r 
 
 A 
 
 or, if referred to a horizontal axis, the horizontal acceleration is / = - — 
 
 J r 
 
 cos d. Thus to balance both vertically and horizontally a weight must be 
 placed opposite the crank-pin acting at radius r x such that its weight is 
 given by W 4 r x = M 4 R, where R is the distance of the centre of gravity 
 of the crank webs &c, from the shaft centre. 
 
 7. Inertia of Valve Gear. — The simple valve gear is, of course, a slider- 
 crank chain like the piston, connecting-rod, and crank, and its treatment is 
 therefore precisely the same. 
 
 Practically it is not convenient to fit special balance weights for the 
 valve gear, and they are generally embodied in those for the main parts of 
 engine. As this method introduces a disturbing couple on the engine which 
 must be eliminated, the problem of combination will be treated at length in 
 par. 9. 
 
 Again, most marine engines are fitted with Stephenson's link motion. 
 It is impossible to use the formula for the acceleration of the valve in any 
 method of balancing ; in fact, the solution of the displacement, velocity, and 
 acceleration of the slide valve is an extraordinarily complicated operation, 
 especially if slot links are fitted. 
 
 It is usual, therefore, to adopt the following assumptions in dealing with 
 this part of the engine : — 
 
 1. Reciprocating Masses. 
 
 Ahead, . . Valve. 
 Spindle. 
 Half the link. 
 Link block. 
 
 0*5 eccentric-rod if notch up gear ; or 
 0'4 eccentric-rod if direct gear. 
 Half the bridle-rods = 1 bridle-rod if con- 
 nected to " ahead " end. 
 Half the link. 
 
 - 5 eccentric-rod if notch up gear; or 
 0*4 eccentric-rod if direct gear. 
 1 bridle-rod if connected to "astern" end. 
 
 Remainder of eccentric-rod. 
 Eccentric sheave. 
 Eccentric strap. 
 Remainder of eccentric-rod. 
 Eccentric sheave. 
 Eccentric strap. 
 
 8. Inertia of Air Pump. — The air pump when out of centre with the 
 engines, and having an opposite motion to that of the particular engine that 
 drives it, presents a rather peculiar problem when its effect on the engine 
 support or foundation is considered conjointly with the means of balancing. 
 
 Let W 5 W 6 be the equivalent weights of parts when having an accelera 
 txonf equal to that of the piston-rod (fig. 293). Then A B = B 0. 
 
 49 
 
 Astern, 
 
 2. Rotating Masses. 
 Ahead, 
 
 • • 
 
 Astern, 
 
766 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Let the inertia forces due to W„ W 6 be P, P, respectively (p = W 5 t 
 
 f 
 
 9 
 
 and P x = W 6 — J, and suppose them to act as shown in fig. 293 when the 
 
 lever is in that position. 
 
 The force P is equivalent to a transverse couple 2 PA on the engine and 
 
 a downward force P acting through the rocking shaft bearing (fig. 293a). 
 
 There is thus a total force P + P x acting upward with the piston-rod inertia 
 
 force and a force 2 P acting downward through the bearing. 
 
 Means of Balancing. — There is no practicable method of balancing the 
 
 transverse couple, except perhaps by fitting an equal air pump on the 
 
 opposite side of the engine and 
 driven from the same crosshead. 
 This would, however, be an ex- 
 tremely awkward fitting. 
 
 Now, if h is an appreciable 
 dimension compared with the bread th 
 of the ship, the effect will be to pro- 
 duce torsional vibrations of the ship; 
 but if h be very small compared with 
 the same, there is on practically the 
 same centre line a downward force 
 2 P and an upward force P + P r The 
 difference of these forces is the force 
 causing vibrations of the same kind 
 as those produced by the engine. 
 In practice it will be found that W 5 
 and W 6 are very nearly equal to one 
 another, and in that case the two 
 inertia forces neutralise each other, 
 and the air pump may consequently 
 be neglected. 
 
 In the greater number of fast- 
 running engines h will, however, 
 
 Figs. 293 and 293a. 
 
 be of appreciable magnitude, and it is, therefore, usual to make the 
 addition W 5 + W 6 to the engine reciprocating weights. 
 
 Note. — When in the remainder of this chapter the term " perfectly 
 balanced " is used without comment, it may be taken as meaning that the 
 parts under consideration are perfectly balanced as far as the limits 
 just discussed enable them to be. 
 
 9. Single-crank Engine and Valve Gear balanced by Rotating Weights 
 only. — Eeduce all weights of moving parts to equivalent weights driven by 
 cranks of the same radius, say the crank radius r. 
 
 Let M x = weight of engine reciprocating parts at radius r. 
 M 2 = 
 M 3 = 
 M 4 = 
 
 Imagine a plane O X Y (fig. 294) through any point O on the crank 
 shaft and at right angles to it. This plane will be called the reference 
 plane. 
 
 )> 
 
 ,, rotating 
 
 » 
 
 )) 
 
 valve reciprocating 
 
 )> 
 
 >) 
 
 ,, rotating 
 
 >i 
 
Vi 
 
 Fig. 294. 
 
 1 
 
 cv 
 
 *-d- 
 
 BALANCING. 
 Fig. 294a. 
 
 VERTICAL FREE FORCE 
 
 T67 
 
 Fig. 2946. 
 
 VERTICAL MOMENT 
 HBCnST REFERENCE PLANE 
 
 it 
 
 IF 
 
 Mm 
 
 % 
 
 <# 
 
 
 
 
 J._ 
 
 — - 
 
 J-/SP 
 
 
 % 
 
 
 
 
 
 7 
 
 
 
 
 
 r> 
 
 
 
 
 
 h 
 
 
 
 
 
 (fl 
 
 
 
 3 
 
 
 a 
 
 
 
 Hi 
 
 5 
 
 
 
 ^& 
 
 ;: 
 
 c 
 > 
 
 
 
 
 (A 
 
 
 ./v 
 
 «1 / 
 
 
 
 
 
 T 
 
 5 
 
 ► / 
 
 
 
 
 
 
 
 r 
 
 
 P 
 
 C 
 
 
 jsr 
 
 Fig. 294c. 
 
 Fig. 294d. 
 
 Fig. 294e. 
 
 Case 1. — Suppose it is desired to balance the engine as completely as 
 possible in a vertical direction in preference to the horizontal direction. 
 
 By the theory of couples any number of coplanar forces can be reduced 
 to a single force acting at any point accompanied by a couple. 
 
 In the engine under consideration there is at any time a free vertical 
 
 v- 
 
 force (M l + M 2 ) cos 6 - — (M 8 + M 4 ) cos (6 + a), and a couple 
 
 referred to the reference plane 
 
 2 «2 
 
 = (M\ + M 2 ) a cos 6 (M 3 + M 4 ) b cos (9 + «) 
 
768 MANUAL OF MARINE ENGINEERING. 
 
 As the force and couple cannot be balanced by a single force, because 
 they are not in the same phase, two weights at least must be employed to 
 balance the engine. A couple as well as a force can be represented by a 
 straight line of length proportional to its moment. It is usual in theoretical 
 mechanics for this line, which is called the axis of the couple, to be drawn 
 anywhere at right angles to the plane of the couple, but in order not to 
 confuse the subject it will be found much more convenient to draw this 
 line in the plane of the couple. 
 
 Project the lines of cranks on to the reference plane (shown in front 
 elevation on figs. 294a and 2946). and measure off parts O A and O B equal 
 to (Mj + M 2 ) and (M 3 + M' 4 ) respectively. Measure off O A l and OB 15 equal 
 to (M, + M 2 ) a and (M 3 + M 4 ) b respectively. Each of these quantities 
 
 multiplied by , of course, gives the maximum inertia forces and 
 
 couples due to them, and the angle a is also the phase of those forces 
 
 v 2 
 and couples of the crank behind those of the valve gear. - — being 
 
 common to all, balance weights included, need not therefore enter into 
 any calculations. Then A, B, A T and B 1? are the generating points of 
 simple harmonic motions along OX, O Y. Compound these by the problem 
 in par. 3. Their resultants are C and X respectively. 
 
 Suppose it is convenient to place the rotating balance weights at points 
 G and F on the shaft. Then, stated definitely, the problem is to find the 
 values of the weights W l W 2 , and the angles at which they are to be fixed 
 relatively to the crank such that the resultant of the vertical free force 
 produced by their rotation is equal and opposite to O 0, and that the 
 resultant of their moments about the reference plane is equal and opposite 
 to C r 
 
 For convenience let M x + M 2 = P, and M 3 + M 4 = Q. Let the angles 
 Wj W 2 make with the crank be a Y a 2 respectively. 
 
 Then these values may be found by solving the equations — 
 
 P + Q cos a = - (Wj cos aj + W 2 cos a 2 ) - . - - - - (1) 
 
 ~Pa+ Qbcosa = - (WjC cos a x + W 2 d cos a 2 ) (2) 
 
 W x 2 (c - df = P 2 (a - df + Q2 (6 - df - 2 P Q (a - d) (b - d) cos a (3) 
 
 W 2 2 (c - df = P 2 (c - a) 2 + Q 2 (b - cf - 2 P Q (c - a) (b - c) cos a (4) 
 
 Care must be taken that the distances are positive or negative according 
 to their sense. These are very easily solved if the known values of the 
 constants be substituted immediately. 
 
 The graphical solution, however, is more quickly arrived at. 
 
 Draw a base line H K, and a pair of axes O X, O Y (figs. 294c, 294c?, 
 and 294e). On H K mark off the engine centres, &c, to scale, and erect 
 perpendiculars as shown. 
 
 On either of the balance weight centres (say that of W.) mark off 
 G R = P, and G U - Q. 
 
 Draw horizontals through R and U. Join R F, cutting D L in S. Join 
 U F, cutting ET in N. Then, by the principle of the "lever," the part of 
 D L or P that must be apportioned to Wj is D S, and the part to W 2 is S L. 
 
 Similarly, the part of E T to be apportioned to W 1 is E N and the part 
 
BALANCING. 769 
 
 to "W 2 ig TN. EN is necessarily greater than E T, because the valve 
 centre is outside both the centre of Wj and W 2 . 
 
 To find W l mark off (fig. 211c?) os = D* and «n = EN. Join no 
 and produce. Then on = Wj and o n v its direction relatively to the crank 
 line oy. 
 
 To find W 2 mark offoUSL (fig. 294e), and It = -TN. TN being 
 negative the direction must be in the opposite direction to sn. 
 
 Then W 9 = ot and o t is its position relatively to the crank line O Y. 
 
 If the work has been done correctly, the resultant of the two weights 
 Wj W 2 will be found to be equal and opposite to O C, and the resultant of 
 their moments about the arbitrary reference plane will be found to be equal 
 and opposite O Cj (figs. 294a and 2946). 
 
 The centre of gravity of Wj and W 2 must be at radius r, but, if 
 necessary for it to be at radius r v then W 1 r = wjr v and W 2 r = w 2 r 2 . 
 Thus the engine is balanced perfectly in a vertical direction for both the 
 distributing couple and the free up and down force. 
 
 Case 2. — Suppose it is desired to balance the engine as completely as 
 possible in a horizontal direction in preference to the vertical direction. 
 
 Here the rotating parts of the engine 
 have simply to be dealt with, because the 
 reciprocating parts have no resolved part 
 horizontally. 
 
 Apply the same process as for Case 1, 
 but instead of using Mj + M 2 and M 3 + M 4 
 use M 2 and M 4 respectively. 
 
 It will now be seen, as intimated in the 
 latter part of par. 5, that the engine is 
 over-balanced horizontally when completely 
 balanced vertically by rotating balance 
 weights. 
 
 Case 3. — Suppose it is desired to balance 
 the engine as completely as possible in both 
 directions, the balance weights being rotary -pj 2 95. 
 
 only. 
 
 The only thing that can be done here is to effect a compromise. 
 
 Find, as in Oases 1 and 2, the weights necessary for balancing in 
 both directions separately. Let these weights be W 1 " and W 2 r for the 
 vertical balance and W,* and W 2 * for the horizontal balance. Draw their 
 centre lines as in fig. 295. They will most probably not be on the same 
 
 y W,* 
 
 centre line. Divide the angle between them so that -* = ^— — w h . 
 
 Then Yop is the best angle for the weight to be fitted. Also make 
 op = — i — ~ L . Then op is the value of the weight W l to be fitted. 
 
 a 
 
 Treat W 2 " and W 2 * in a similar manner. 
 
 The engine is thus incompletely, and at the same time as completely, 
 balanced as possible under the condition of having only rotating balance- 
 weights. 
 
 10. Single-crank Engine and Valve Gear Balanced by Bob Weights and 
 Rotating Weights combined. — The engine may be balanced perfectly both 
 vertically and horizontally by the above combination. 
 
770 
 
 MANUAL OF MARINE ENGINEERING. 
 
 In order to do so the bob weights are made to balance the reciprocating 
 parts, and the rotating weights are made to balance the rotary parts. 
 
 To find the bob weights proceed exactly as before, using Mj and M 3 
 
 for P and Q. The 
 eccentrics of throw 
 convenience. If 
 W„r = B„r v B, and 
 
 weight 
 
 found are those which would be driven by 
 This radius will in most cases be too large for 
 
 W x r = Bjrj, and 
 new weights which should include the 
 
 weights 
 
 r x be the convenient throw, then 
 
 . B 2 being the 
 of the eccentric and rod. To find the rotating balance weights 
 
 again proceed as before, 
 
 using 
 
 M 2 and M 4 for P 
 and Q respectively, but 
 also taking into account 
 the weight of the bob 
 weight eccentrics, as they 
 introduce a horizontal dis- 
 turbance which must be 
 eliminated. 
 
 The solution of the 
 problem of balancing by 
 this particular method, of 
 course, depends on the fact 
 that the bob weights have 
 no resolved part horizon- 
 tally. 
 
 11. To Balance an 
 Engine of any number of 
 Cranks. — As the method 
 of solving this problem is 
 the same in principle as 
 has just been described in 
 pars. 9 and 10, the case 
 will be most easily ex- 
 plained by means of an 
 example, having particulai 
 reference to the three- 
 crank engine. 
 
 In the example the 
 engine will be balanced 
 both vertically and hori- 
 zontally, and the combined 
 system of bob and rotating 
 Weights must, therefore, 
 be used. Should, how- 
 ever, the designer of any 
 particular engine decide 
 that he cannot very well adopt the most complete method, he will understand 
 from the previous paragraphs what sacrifices will have to be made if only 
 one kind of balancing device is used. The methods of construction are pre- 
 cisely the same, although more extensive. I 
 The example (fig. 297) is a three-cylinder engine with ordinary link motion 
 for each valve. Two bob weights will be fitted, one at each end of the engine, 
 
 912 I 
 Fig. 296. 
 
BALANCING. 
 
 T71 
 
 and four balance weights, one on each crank web of the outside engines. 
 It will also be decided to balance half the rotating masses by the H.P. and 
 L.P. forward weights, and half by the H.P. and L.P. after weights. 
 
 ©i 
 
 t'a 
 
 £ 
 
772 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Ill «• 
 
 XBBBB BBS 339 
 
BALANCING. 
 
 773 
 
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 MANUAL OF MARINE ENGINEERING. 
 
 §1 
 
 aaajaaaa manna aoa 
 
BALANCING. 
 
 775 
 
776 
 
 MANUAL OF MARINE ENGINEERING. 
 
 stamwsss xnsns bsb 
 
BALANCING. 
 
 777 
 
 8 
 
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 MANUAL OF MARINE ENGINEERING. 
 
BALANCING. 
 
 779 
 
 The weights of the moving parts are tabulated as follows, and the diagrams 
 are drawn in figs. 298 to 302. 
 
 TABLE XCI. — Balancing Engines : Weights of Moving Parts. 
 
 
 Reciprocating Weights. 
 
 Rotating Weights. 
 
 Half 
 
 Sum of 
 Recipro- 
 
 
 Equivalent 
 
 
 Equivalent 
 
 
 Actual Weight of 
 
 Weight at 
 
 Actual Weight of 
 
 Weight at 
 
 Rotating 
 
 C 1 1 L 1 l 1 !_; 
 
 
 Reciprocating 
 
 Radius of 
 
 Rotating 
 
 Radius of 
 
 Weight. 
 
 T?nta t.jTxcr 
 
 
 Parts. 
 
 Crank, 
 
 Parts. 
 
 Crank, 
 
 
 .1 v* ' L 1 1 1. 1 [ J _ 
 
 Weights. 
 
 
 
 13£ Inches. 
 
 
 13J Inches. 
 
 
 
 
 At 
 
 
 
 At 
 
 
 
 
 
 
 Radius, 
 
 
 
 Radius, 
 
 
 
 
 
 Lbs. 
 
 Ins. 
 
 Lbs. 
 
 Lbs. 
 
 Ins. 
 
 Lbs. 
 
 Lbs. 
 
 Lbs. 
 
 H.P. engine, . 
 
 2,000 
 
 ,34 
 
 2,0C0 
 
 1,000 
 
 13J 
 
 1,000 
 
 500 
 
 3,000 
 
 ,, ahead, . 
 
 698 
 
 3 
 
 155 
 
 338 
 
 3 
 
 75 
 
 374 
 
 230 
 
 ,, astern, . 
 
 202 
 
 3 
 
 45 
 
 292 
 
 3 
 
 65 
 
 324 
 
 110 
 
 M. P. engine, . 
 
 2,300 
 
 13* 
 
 2,300 
 
 1,100 
 
 134 
 
 1,100 
 
 550 
 
 3,400 
 
 ,, ahead, . 
 
 855 
 
 3 
 
 190 
 
 338 
 
 3 
 
 75 
 
 374 
 
 265 
 
 ,, astern, . 
 
 202 
 
 3 
 
 45 
 
 292 
 
 3 
 
 65 
 
 324 
 
 110 
 
 L.P. engine, . 
 
 2,800 
 
 m 
 
 2,800 
 
 1,200 
 
 134 
 
 1,200 
 
 600 
 
 4,000 
 
 ,, ahead, . 
 
 1,036 
 
 H 
 
 240 
 
 345 
 
 34 
 
 80 
 
 40 
 
 320 
 
 ,, astern, . 
 
 207 
 
 H 
 
 48 
 
 302 
 
 34 
 
 70 
 
 35 
 
 118 
 
 rorward bob, 
 
 2,538 
 
 5 
 
 940 
 
 Eccentric. 
 
 100 
 
 50 
 
 
 After bob, . . 
 
 3,306 
 
 5 
 
 1,225 
 
 >> 
 
 100 
 
 50 
 
 
 H P. forward, 
 
 J- Balance We 
 
 f 
 
 376 
 
 134 
 
 • • ■ 
 
 
 . ■ . 
 
 , , after, . 
 L.P. forward, 
 
 sights. -| 
 
 378 
 3S1 
 
 134 
 134 
 
 
 
 " 
 
 , , after, . 
 
 I 
 
 384 
 
 13* 
 
 ... 
 
 
 
 Fig. 298, 298a, 2986 are for finding the bob weights ; 
 
 Fig. 299, 299a, 2996 for the H.P. and L.P. forward balance weights. 
 
 Fig. 300, 300a, 3006 for the H.P. and L.P. after balance weights. 
 
 Fig. 301 finds the resultant of the vertical free force by using the sum of 
 the reciprocating and rotating parts. 
 
 Fig. 302 finds the resultant of the vertical free force of all the bob and 
 balance weights. 
 
 If the work is done correctly these two resultants will be found to be 
 equal and opposite to each other. 
 
 The figures and construction will be found self-explanatory after reading 
 pars. 9 and 10. 
 
 12. So far, it has been seen that, in order to balance an engine, it is only 
 necessary to know the relative values of the inertia forces and not the actual 
 value. But, in order to compare one design of an engine with another, it is 
 necessary to know the actual values (or at least relative values referred to a 
 common basis) of the resultant disturbing forces and couples. 
 
 For instance, in comparing submitted designs of engines for any particular 
 ship, it is obvious that, in a general sense, the less the resultant forces and 
 couples are, the less the vibration, and, therefore, the better the engine. 
 
 Referring to par. 4, et seq., dealing separately with the inertia of the 
 various parts, and having balanced the engine as perfectly as practically 
 possible, it will be seen that the disturbing forces left unbalanced are : — 
 
 For the piston, rod, and crosshead a vertical force - 
 For the connecting-rod a vertical force 
 
 M, v 2 cos 2 6 
 
 gr 
 
 M 3 v 2 
 gr 
 
 n 
 
 cos 2 6 
 
780 
 
 MANUAL OF MARINE ENGINEERING. 
 
 For the valve gear there is a very minute quantity, which is a function of 
 2 6 and the higher multiples of 6, and which may very reasonably be 
 neglected. 
 
 The" residual resultants in the cases of the single cylinder engine will now 
 be considered, and the methods and principles of constructing the various 
 resultants will apply equally well to a multi-cylinder engine. 
 
 13. Case 1 (corresponding to Case 1, p. 553). — The residual vertical 
 
 M., + T connecting-rod ) — and is 
 I J grs 
 
 constant. There is no residual couple, seeing that the very small force left 
 
 from the valve gear is neglected. 
 
 Put 6 = 0, then cos 2 6 = 1 and x is the amplitude, par. 2, of the force, 
 and is the quantity that can be used for comparative purposes. 
 
 To draw the curve x cos 2 6, the ordinates for various values of 6 are 
 easily determined graphically. Draw a quadrant of a circle of radius x to 
 any suitable scale (fig. 303). Divide the arc anto a convenient number of 
 parts and drop perpendiculars on to the axis O X as shown. Draw a base 
 line A B and divide it into twice the number of parts into which the complete 
 
 Fig. 303. 
 
 circle is divided, because the curve has two periods per revolution. Mark 
 the divisions and degrees and erect perpendiculars to A B through the 
 division points as shown. Mark off Al = 01, 62 = 6 2, c3 = c3, and so 
 on. Then the curve through E, 3, F is the required curve, the ordinates of 
 which give the residual resultant disturbing or unbalanced inertia force at 
 any position d of the crank from the top centre. The curve of resultant 
 moment about any reference plane may be drawn in a similar manner — for 
 instance, that plane through a node of the ship. As in this case the engine 
 is overbalanced horizontally there is a residual horizontal disturbing force. 
 
 Let the resultant of the balance weights Wj and W 2 be W„ (fig. 304). 
 Also, let the resultant of the rotating parts of the engine be R, both these 
 resultants being drawn in their proper relation to the crank. Find the 
 resultant R^ of W„ arad R. It makes an angle <p with the crank — that is to 
 say, its phase is p° before that of the crank (pars. 2 and 3). 
 
 V*' 
 
 Then the resultant horizontal force is R* — sin (9 + <p). 
 
 gr 
 
 * In the following pages the inclusive minus signs are ignored. 
 
INERTIA FORCE AND COUPLES. 
 
 781 
 
 Since the resultant is <p° in front of the crank, and also since the axis of 
 reference is 90° in front of the vertical axis, the ordinate of maximum 
 amplitude is (<p + 90)° in front of the ordinate B, and is, therefore, drawn 
 through the point C. Divide the base line off again as before, but the 
 number of divisions is the same as that of the complete circle, as there is 
 here only one period per revolution. Draw the circle 4 3 5, 6, and set off 
 the ordinates as before. Then the curve D 4 G (fig. 303) ia the required 
 curve, the ordinates of which give the residual resultant horizontal un- 
 balanced force at any position & of the crank from the top centre. 
 
 Case 2. — Engine balanced perfectly horizontally. Here there is no 
 residual horizontal force, and the curve therefore coincides with the base 
 line, the amplitude always being zero. 
 
 In a vertical direction there is still the residual force x cos 2 6 as in case 
 1, but now there is an additional force due to the difference between the 
 resultant of the weights required for vertical balancing and those that have 
 been fitted for horizontal balancing. 
 
 Let Wj W 2 be the weights as fitted, and W h their resultant (fig. 305). 
 
 Fig. 304. 
 
 Fig. 305. 
 
 O is the resultant of the reciprocating parts of the engine as drawn in 
 fig. 294a (function of 6). Find the resultant of O and W h = R,. 
 
 Then the resultant inertia force vertically is R„ — cos (d + y) + x cos 2 6. 
 
 if 
 
 Although in Oase 1 it was not absolutely necessary to draw the curve of 
 force for comparative purposes, it is necessary here, because the maximum 
 amplitude of the resultant is not the sum of the two amplitudes in the 
 above expression. 
 
 Draw the curve R, — cos (6 + y), noting that its phase is y° in front 
 
 if 
 
 of the curve x cos 2 6. 
 
 Draw the curve x cos 2 6. 
 
 The resultant is the sum of the ordinates. 
 
 These*curves are drawn in fig. 306. 
 
 50 
 
782 
 
 MANUAL OF MARINE ENGINEERING. 
 
 The total amplitude H K is that required for comparative purposes. 
 
 This may seem on first reading a rather laborious process, but if the 
 principles be thoroughly understood it will be found to be many times more 
 rapid than drawing the curves by calculating the ordinates for various 
 positions of the crank. 
 
 Case 3. — Engine balanced completely horizontally and vertically. 
 
 Here the residual horizontal force is zero, and the residual vertical force 
 is the same as that in Case 1. 
 
 The curves just drawn are for a velocity, v, of the crank pin, correspond- 
 ing to a certain number of revolutions, N, per minute. 
 
 The amplitude for any other number of revolutions, Nj, will be that just 
 
 found multiplied by 
 
 SO- 
 
 . &StlZt'f y fr fafttr) 1~ X&20 
 
 Ziegru& from Cctft Centre 
 
 Fig. 306. 
 
 14. Curves of Free Force and Couples for a Multi-Crank Engine. — This 
 will, as in the case of balancing (par. 11), be illustrated by means of the 
 example of a three-cylinder engine — 
 
 Let M 2 = weight of H.P. reciprocating parts. 
 M 9 = „ M.P. 
 
 Mo 
 
 L.P. 
 
 1. Free Force. — As the engine is as perfectly balanced as possible in both 
 vertical and horizontal directions, the vertical free force will only be the 
 
 resultant of the three residual forces M, — cos2 01 9 — cos 2 (6 + 120°), 
 
 2 9 r 8 9 r s 
 
 Mo- cos 2(0+ 240°). 
 
 gr 
 
 v v 
 
 For convenience let the constants M, — , M 9 — , 
 
 1 gr s * gr s 
 
 Mo - = x, y, z 
 3 grs J 
 
 respectively. 
 
 Draw axes O X, O Y (fig. 307) and let the H.P. crank be along O Y. 
 
 In finding the resultant of the above three inertia forces or harmonic 
 functions, note j articularly that the angles between the generating lines are 
 not the same as the angles between the cranks, because the forces are 
 functions of 2 &. From A to B (fig. 308), which is 180°, really represents 
 360°, aR far as the curves are concerned, and the sequence of the generating 
 lines is as if the H.P. were leading instead of the L.P. 
 
 Then in fig. 307 the generating lines are in the order as shown. Find 
 the resultant O R by the polygon as shown (par 3). 
 
FREE FORCE AND COUPLES. 
 
 783 
 
 Then the resultant virtually leads the H.P. generating line by a, or when 
 
 o 
 
 referred back to the real degree scale (fig. 308) by --. 
 
 a 
 
 The ordinate of maximum amplitude of the resultant therefore passes 
 through the point C, and not through the point D, where it would if the 
 generating lines were in the same 
 sequence as the cranks. 
 
 Draw a quadrant of a circle of 
 radius o E, and set off the resultant 
 curve as previously explained. 
 
 The engine being balanced per- 
 fectly horizontally the amplitude of 
 the resultant force is zero, and the 
 curve coincides with the base line. 
 
 If the engine were not com- 
 pletely balanced horizontally, the 
 residual resultant horizontal force 
 would, of course, be a function of 
 4, and the generating lines would 
 have the same sequence as the 
 cranks. 
 
 2. Couples. — As there are three 
 centre lines, and a residual vertical 
 force along each, there will be a 
 residual resultant couple on the 
 engine. The moment of the couple will, of course, vary according to 
 the position of the reference plane, and in order to compare one design 
 of an engine with another some common method of locating the reference 
 plane is necessary. 
 
 The Admiralty, in their recent specifications, require it to be through a 
 point on the shaft midway between the extreme cranks, and, considered 
 
 « & H 
 
 Fig. 307. 
 
 -Degrees A>* &/a At? centre 
 
 £ig. 308. 
 
 generally, this is a very reasonable point to select, because it makes the 
 length of the engine have a minimum influence on the value of the couple, 
 and also causes the plane to be at about the same position in the ship, what- 
 ever the design of the engine may be. This point is thus suited for marine 
 engines, but lor land engines it is advisable to select a point that will enable 
 a stricter comparison to be made. This point should be at what is termed 
 
784 MANUAL OF MARINE ENGINEERING. 
 
 in theoretical mechanics " the central axis " of the couple. To explain this, 
 let F = 00 be the resultant force (figs. 294a and 2946), and O C x the 
 resultant couple referred to any arbitrary plane O Y. Resolve the couple 
 in and at right angles to the direction of the force F. 
 
 Replace the couple Oe — that is, in the same phase plane as the force — by 
 another equal couple but having each of its forces equal to F. Thus F h -Oe 
 (figs. 294a and 2946). Then F and - F on the reference plane cancel each 
 other, leaving a force F on the new reference plane distant h from the 
 original one and a couple eO l having its phase at right angles to the fo'oe 
 and to the original couple. 
 
 Taking the case of the engine in par. 9, let the weights P = 8, Q = 2. 
 ^.lso, let a = 4, 6 = 7, and a as drawn. 
 
 Then the resultant force F= 7"2 x cos (9 + B) 
 and the resultant couple O C x = 27 '8 x cos (d + u). 
 
 V 2 
 
 Where x= — and neglecting the obliquity of the connecting-rod. 
 
 The resolutes Oe and eO x are 27-2 x cos (6 + d) 
 
 and 5 '8 x sin (6 + <3) respectively. 
 
 , Oe 27 -2 a: cos (0+5) „. 7ft 
 
 Then h ~ -=r = ^~~6 /,T7~s\ = 6 *°- 
 
 F r2 x cos (6+o) 
 
 Thus the central axis is 3 - 78 from the reference plane, and the resultants of 
 the system are a force 7*2 x cos (&+&) acting in the plane of the central 
 axis or new reference plane, and a couple 5 - 8 x sin (6 + 8). 
 
 The couple is always a minimum, and the central axis is, of course, 
 always through the same point on the shaft, and therefore a reference plane 
 which contains the central axis of the system is suitable for standard 
 comparisons. 
 
 With this slight digression the consideration of the residual couples on 
 the three-crank engine will now be proceeded with. The couple about an 
 arbitrary plane must be taken first. 
 
 The moments of the residual vertical forces are — 
 
 v 2 
 
 M, — - a . cos 2 6- - = x, siv, 
 1 grs l . 
 
 M 2 — 6 . cos 2 (0+120°) = Vl 
 
 M 3 c . cos 2 (6 + 240 ) = 
 
 *i- 
 
 grs 
 
 Where a, 6, c are the distances of the centres of engines from the reference 
 plane respectively. Their signs must be according to the direction in which 
 they are read. 
 
 Since the curves are all functions of 2 6 the same remarks apply to the 
 sequence of the generating lines in the polygon as for the free forces. The 
 resultant may be found and the curve drawn in a similar manner. 
 
 If it be desired to find the couple about the central axis plane the process 
 explained above may then be proceeded with. The phase of the resultant 
 couple will then be 90° in advance of the original one, and will, of course, 
 have a different amplitude. 
 
FOUR-CRANK ENGINE. 
 
 785 
 
 As there is no horizontal force there is, consequently, no horizontal couple. 
 
 If the engine be uot completely balanced for functions of in the vertical 
 direction, then, in addition to the residual resultants above, the resultant of 
 the " single period" (0) forces and couples must be found separately and 
 their ordinates added / to those of the " double period " (2 0) graphically, as 
 explained in Case 2, par. 13. 
 
 15. The Four-crank Engine — Preliminary.— it will be evident from what 
 has preceded that the three-crank engine (neglecting valve gear differences), 
 where the three reciprocating lines are of equal weight, is in itself prefectly 
 balanced without the addition of balance weights, so far as the free force is 
 concerned, both for the primary and sec ndary periods. Further, that with 
 three unequal lines Of reciprocating weights it is possible to balance the primary 
 vertical free force completely by a modification of the crank angles, so that 
 the projected vectors on a reference plane form a closed triangle. The second- 
 ary period, however, might be c nsiderably out of balance, because it will be 
 found impossible to close the secondary period triangle if the primary closes, 
 and vice versa. Fig. 309a shows a general case. 
 
 When the three lines are unequal, the valve gear can obviously be included 
 without altering the nature of the results. 
 
 Fig. 309a. — Primary Forces. 
 
 Fig. 3096. — Secondary Forces. 
 
 Without balance weights the unbalanced couple will necessarily be large, 
 because the force triangles and couple triangles, either primary or secondary, 
 cannot possibly close for the same crank angles. (Euc. 1-7.) 
 
 The four- (or more) crank engine offers much greater possibilities of 
 balancing the reciprocating parts without balance weights, by virtue of the 
 extra side to the polygons. 
 
 16. Four-crank Engine — Yarrow-Schlick- Tweedy and Allied System*. — 
 Given arbitrary positions of the cranks— for example, cranks forming a 
 rectangular star — a four- or more crank engine may be balanced by means 
 of balance weights and bob weights, as in the foregoing example of the three- 
 crank engine. 
 
 It is, however, possible by arranging the cranks in a certain sequence 
 and at angles differing a little from the rectangular star to obtain an inter- 
 balance between the reciprocating parts themselves without the assistance 
 of balance or bob weights. 
 
 The literature on balancing the four-crank engine i* interesting, but 
 exceedingly voluminous ; most of it is of an academic nature, and capable of 
 much abridgment. 
 
 The point so much overlooked is that (neglecting the complication of the 
 
786 MANUAL OF MARINE ENGINEERING. 
 
 valve gear, which as a matter of fact has in general practice been neglected) 
 it is only necessary to consider two lines as the bob weights to the other two. 
 
 With this convention the problem is simplified, and can be worked out 
 from the primary data without artificial aids. 
 
 It should be clear that the graphic method previously described is applic- 
 able, the example of the reciprocating parts of the three-crank engine being 
 nothing more nor less than that of a five-crank engine if the bob lines be 
 regarded as working lines. 
 
 There are, nevertheless, a few special points to consider when it is desired 
 to balance the secondary periods so far as possible. A little reflection will 
 lead to the following conclusions : — 
 
 1. The natural weights of the reciprocating parts combined with the 
 natural cylinder spacing cannot necessarily be balanced for the couple 
 (or more briefly for " tilting "), or even for the free force, because the 
 magnitudes of the four sides of the polygon may not permit its closure. 
 
 2. The two inside reciprocating lines must be considerably heavier than 
 the two outside lines, if the couple is to be balanced as well as the 
 free force. 
 
 3. Given the cylinder spacing and minimum reciprocating weights (the 
 'two outside lines), it is possible by an arrangement of cranks to balance 
 
 both the primary free forces and primary couple, but in general it is 
 not possible to balance the secondary forces or couple completely. 
 This will be evident from the fact that it is not possible to close two 
 dissimilar polygons — the one the force, and the other the couple — 
 with the angles of the one prescribing the angles of the other. 
 
 4. The main reciprocating weights should only participate in the solution 
 
 of the crank angles. The rotating weights must be balanced by 
 crank web balance weights in the usual way; and, further, as the 
 valve gear design is not usually settled at the early stage when the 
 crank-shaft must be ordered, and as the obliquity effect is very small, 
 the valve gear may be included in the rotating system, the rotating 
 balance weights being determined at a later stage in the design. 
 
 If it is permitted to re-adjust the cylinder spacings, it is possible to balance 
 both primary and secondary forces together with the primary couple, but the 
 four line weights must bear a definite relation to one another. 
 
 There are several methods of solving the problem, according to the data. 
 Neglecting the obliquity of the connecting-rods, the simplest solution is 
 that when the cylinder spacing is given together with one minimum line 
 weight, which must be either of the outside lines. 
 
 The following is apparently due to Prof. Schubert, of Hamburg, and was 
 introduced by Herr Otto Schlick in 1900,* the proof of the problem being 
 given by the direct method instead of the simpler converse proof given below, 
 which, however, is enough for the present purpose. 
 
 Set out the cylinder spacing to scale on a base line A D. 
 
 Take any point above A D. Join to A, B, C, D. 
 
 Draw C P parallel to B and A P parallel to D. Then A P C is a 
 mass quadrilateral giving obviously a balance of the free force if the four 
 line weights be made proportional to the four sides. 
 
 * Paper on " Balancing of Steam Engines," Inst. N. A., 1900. 
 
FOUR-CRANK ENGINE. 
 
 787 
 
 If the four line weights be denoted by M A , M B , M c , M D , acting at A. B, C, D 
 respectively, and if they be allocated as shown in the figure, the couple will 
 be balanced as well as the force. 
 
 The crank angles will be as shown at point 0, and care must be taken 
 that the right and not the wrong weights are applied to the various 
 cranks. 
 
 The diagram admits of simple proof. 
 
 Since the couple should be zero about any point, take moments about 
 
 Fig. 310. 
 
 A ; then, adopting the notation of the figure, we have for any value of 6 
 measured from an arbitrary axis — 
 
 M B a cos 6 + M (a + b) cos {6 + a) + M D (a -f- b + c) cos (0+ y) = 0. 
 
 Put 6 = 90°, then- 
 Mo (a + b) sin a-f M D (a + 6 -+- c) sin y = 0. 
 
 Now, 
 
 sin w M 
 
 which establish the identity. 
 
 sin a a 4- b + c ,sin7 a + 6 
 ^ and -. — '- = m — . 
 
 sin w M D 
 
788 
 
 MANUAL OP MARINE ENGINEERING. 
 
 is any point, and consequently there is an infinite variety of angles and 
 corresponding weights that will give a balance for primary force and couple. 
 
 The selection of an appropriate point is a matter of judgment to suit the 
 conditions, and. in general, it should be selected so as to give cranks as nearly 
 at right angles as is possible consistent with the given minimum line weight. 
 
 The selection of the point so that the secondary forces are balanced 
 (that is, not neglecting the principal obliquity effect) is a problem that cannot 
 be solved by direct mathematical analysis unless the cylinder spacing and 
 reciprocating masses be symmetrical about the middle point. 
 
 The Yarrow-Schliek-Tweedy System is the name by which the symmetrical 
 arrangement is known, and was introduced by Dr. Otto Schlick. of Hamburg, 
 in 1900. 
 
 The following geometrical device will obtain point to give a balance 
 for primary forces and couple and secondary forces : — 
 
 Fig. 311. 
 
 Erect perpendiculars at A and B' (fig. 311). 
 
 Make A' C = A' B. Draw A' E at right angles to A' C. Join E A. and 
 cut off E F = E B'. 
 
 Cut off A G = A F. With the compasses take B' C as radius, and with 
 centre G describe an arc cutting A A' at H. Join GH = B'C. From B 
 draw B K parallel to H G. Make M 1 ■-= B K. 
 
 Then M I bisected gives the point the required apex of the characteristic 
 triangles, as in fig. 310. 
 
THE YARROWSCHLICK -TWEEDY SYSTEM. 
 
 789 
 
 When the spacing is not symmetrical, and the two outside line weights 
 are given, the secondary forces cannot necessarily be balanced, although the 
 crank angles for the minimum resultant can be found by applying the 
 ordinary method to two or three assumed conditions, and interpolating the 
 best position. 
 . The following example will illustrate this : — 
 
 Given cylinder spacing 42, 66, 37 (v. fig. 312a) ; outside line weights 
 M A = 70, M D = 80. 
 
 Fig. 312c. 
 
 Fig. 312d. 
 
 Then, as in the previous examples, fig. 312& gives the magnitude and 
 direction of M B , and M Cl regarded as bob weights to M A , M r , for an assumed 
 crank angle A D, or a v 
 
 The virtual crank angles for the secondary periods are as in fig. '612d, 
 each real crank angle being, of course, doubled. a' b\ c\ d\ (fig. 312e) 
 is the polygon for the secondary forces, the forces being proportional to the 
 line weights acting at the virtual crank angles. 
 
790 
 
 MANUAL OF MARINE ENGINEERING 
 
 This polygon does not close, and the secondary forces are, therefore, not 
 balanced, the resultant being proportional to H v 
 
 Now, assume other values of the angle A D, a 2 , a 3 , as shown in fig. 312c. 
 From these are obtained the secondary resultant forces R, as in fig. 312e. 
 In fig. 312e draw a curve through the extremities of the unclosed polygons. 
 
 Then E obviously has a minimum value at about point d' x , from which it is 
 easy to estimate the corresponding crank angles. 
 
 For this particular combination of weights and cylinder spacings it will 
 be seen that it is not possible to balance the secondary forces completely; 
 
 Fig. 312/. 
 
 but an increase of one of the lesser line weights or a slight modification to the 
 cylinder spacing will effect it, and the balance can be obtained with very 
 little trial and error. 
 
 The method outlined above, and followed also in the earlier part of the 
 
MODIFIED PRINCIPLE OF BALANCING. 
 
 791 
 
 chapter, is one easily to be remembered, and is. therefore, available on any 
 occasion. 
 
 When, however, these investigations are at hand the solution of any 
 problem can be at once effected by means of diagram (fig. 313) prepared 
 by Mr. C. E. Inglis. and given in his interesting paper at the Inst. Naval 
 Architects, 1911.* 
 
 The meshed part of the diagram is a practical zone in which it is only 
 possible to balance primary and secondary forces together with the primary 
 couple. 
 
 The base line A D represents the cylinder spacing and the intersection 
 of correspondingly numbered lines in the meshwork, the apex of the triangles 
 of fig. 310 — that is, point 0. 
 
 For example, let the cylinder spacing be D-5, 5-14, 14-20; then the 
 intersection of curves 5 and 14 give point 0. 
 
 The directions of M A , M B , M c . M D are then as indicated by the dotted 
 lines. 
 
 Further, the relative values of M A ... D , determined by completing the 
 polygon as in fig. 310, are the only possible ones which will give a balance for 
 the given cylinder spacing and the three above static requirements. 
 
 Several simplifications will naturally occur in any problem if the cylinder 
 spacing and line weights are symmetrical, as in the Y.S.T. system. 
 
 A very elegant solution was contributed by the late M'Farlane Gray to 
 the Institution of Naval Architects (vide Transactions, 1900). 
 
 For more than four cranks the foregoing general methods will be found 
 equally applicable, and in this respect it has the advantage of diagrams which 
 are special to the four-crank engine. 
 
 Modified Principle of Balancing.— For an engine on any elastic foundation 
 very large compared with the engine, and capable of vibrating in a manner 
 
 * Paper on " General Propositions and Diagrams relating to the Balancing of the 
 Four cylinder Marine Engine," Inst. N. A., 1911. 
 
792 
 
 MANUAL OF MARINE ENGINEERING. 
 
 analogous to a stretched string or beam spring, it is clear from physical laws 
 that if the engine be placed at a node any unbalanced forces or couples should 
 not be capable of exciting vibration in the foundations. Practically this 
 holds quite true, and is only modified by the fact that the engine is of appreci- 
 able magnitude compared with that of the "loop" or distance between 
 nodes. 
 
 If, however, the various line weights be selected so that the couples about 
 the node are balanced, it is immaterial whether there is any free force left 
 unbalanced, for it will cause no vibration. 
 
 This principle may be followed sometimes with advantage if the positions 
 of the nodes of the ship are fairly accurately known, in which case it is only 
 necessary to take moments about the nearest node, and close the polygon 
 to determine the crank angles. 
 
 Modificatiens from a preliminary solution will probably be required to 
 
 Vibrations of /?.' Order 
 
 ■224Z I. 
 
 SSI6L 
 
 Vibrations of 2 n . d Order. 
 
 •3€>S *-. 
 
 '368/- 
 
 Vibrations of 3 r ? Order. 
 
 Vibrations of 4f. A Order. 
 
 Fig. 314. 
 
 balance the secondary forces or couples. This method was followed some 
 years ago by the author in balancing some new engines fitted in the twin 
 s.s. " Cambridge," which ship certainly appeared to have less vibration than 
 any similar twin-screw ships on the station fitted with three-crank triples. 
 In this case the node was obtained by observations before the old engines 
 were removed. 
 
 It is well known that sympathetic vibrations can be set up in an elastic 
 body by any suitable agent, particularly if connected mechanically to it, 
 vibrating with the natural periodicity of the body itself. 
 
 It therefore follows that the engine rate of revolution should never be 
 allowed to correspond with the natural period of vibration of the ship herself. 
 
 The matter is, however, complicated by the fact that the significant 
 vibrations of a ship may be of several orders, all the periodicities of which 
 should be avoided. Thus, in the vertical plane a ship is capable of vibrating 
 
EFFECT OF INERTIA AND WEIGHT. 793 
 
 in an ascending series as indicated in fig. 314, those of the first two being 
 usually the most important, on account of their generally greater amplitudes, 
 although sometimes the higher orders have given some considerable trouble. 
 For the first order Dr. Schlick finds that the position of the after node from 
 the after perpendicular is -231 to *253 times the length of the ship for ships 
 with very sharp lines, such as cruisers and despatch boats. 
 
 The distance of the fore node from the forward perpendicular varies from 
 about -31 to -365 the length.* 
 
 For the second order there is very little data, and if the ship is not 
 available at the time to supply it, the nearest estimate is that to be obtained 
 by analogy with the known position of the various nodes in a prismatic rod. 
 
 These positions are indicated in fig. 314. The number of vibrations of the 
 second order in the case of the rod is three times that of the first order, but 
 in a ship it is found to be more nearly 2 to 1. The number of vibrations 
 of the first order, according to Dr. Schlick, f are given very approximately 
 by the following formula : — 
 
 N 
 
 = ^Vd173' 
 
 where N = number of vibrations per minute. 
 
 T = moment of inertia in foot-tons of midship section about the 
 horizontal neutral plane. 
 
 D = displacement of ship in tons. 
 
 L = length between perpendiculars in feet. 
 
 <p = a coefficient having values — 
 
 156,850 for vessels with very fine lines, such as torpedo-boat 
 
 destroyers. 
 143,500 for large transatlantic passenger ships with fine lines. 
 127,900 cargo boats with full lines. 
 
 Clearly this method of balancing, in spite of its simplicity in the presence 
 of reliable data, brings many undesirable complications ; and an engine that 
 is balanced within itself may seem a more satisfactory solution of the 
 problem in general. At the same time, it must not be forgotten that 
 large and long marine engines are far from being the rigid bodies they are 
 assumed to be ; and an assumption that the engine is an integral part of 
 the ship, sharing its general elastic properties, is probably in most cases 
 nearer the truth. 
 
 In any case, and whatever be the system of balancing, an examination 
 into the possible coincidence of the proposed speed of revolution with at least 
 the first and second orders of the natural vibration of the ship should be 
 made. 
 
 17. The Effect of Inertia and Weight — General Case, Supplement to the 
 
 Remarks in the Introduction to this Chapter. — It has been shown that the 
 
 . . . . , . ,_. M»V a , cos2 0\ 
 vertical inertia force at any instant is I cos o -\ )• 
 
 * Schlick on " Vibrations of Higher Order, etc.," Inst. N. A., 1895. 
 t Schlick, Paper on " Vibrations of Steamers," Inst. N. A., 1894. 
 
794 
 
 MANUAL OF MARINE ENGINEERING 
 
 If this force is divided by the area of the piston the inertia pressure per 
 square inch will be found. The effective pressure on the piston is, there- 
 fore, the algebraic sum of the steam and inertia 
 pressures, and if the dead weight of the parts be 
 included, the weight divided by the piston area must 
 be added algebraically as well. 
 
 Considering the inertia force alone, M above is the 
 total weight, including the rotating parts ; but the 
 revolving parts have also a horizontal acceleration, 
 and the horizontal inertia force resolved tangentially 
 to the crank path is equal and opposite to the vertical 
 inertia force of the rotating parts resolved tangentially. 
 This, of course, should be so, because the crank has 
 no tangential acceleration. There is, therefore, no 
 alteration produced in the effective twisting moment 
 by the inertia of the rotating parts, an^ the weight 
 of the reciprocating parts need only be used for obtain- 
 ing the inertia pressure on the piston. M, in the above 
 formula, will, therefore, be the weight of the reciprocating 
 parts Mj. 
 
 It is easily shown that the twisting moment on 
 the shaft due to an effective pressure on the piston 
 
 Fig. 315. 
 
 Pis- 
 
 . Prsin(g+a ) (fig. 315), 
 
 cos a 
 
 or, taking the second approximation to the series in terms of 0, as in the 
 case of the acceleration (par. 4), 
 
 T M = P r ( sin + 
 
 sin 2 
 
 In order to combine the inertia curve with the indicator diagram or 
 steam pressure curve, the same spacing of ordinates for each must be adopted. 
 If the inertia curve be drawn with indicator diagram spacing of ordinates 
 (that is, the abscissse are proportional to the distance of the piston from 
 the end of the stroke) instead of abscissae proportional to 6, it will assume 
 the form shown in fig. 3166 by the line C K. If cos 2 or the obliquity of 
 the connecting-rod be neglected, the curve will be a straight line. The 
 ordinates must, of course, be to the same scale as the steam pressure 
 ordinates. 
 
 Considering the dead weight alone of the moving parts, the reciprocating 
 parts having their motion controlled by the connecting-rod must necessarily 
 have their weight added algebraically to the inertia force. But the rotating 
 parts also cause a twisting moment on the shaft which, if M 2 equal their 
 weight at radius r, is equal to M 2 r sin (6-\-tp), where <p is the angle between 
 the resultant M 2 of all rotating masses and the crank. If there are no 
 balance weights and valve gear be neglected, <p = 0. ' M 2 not being con- 
 trolled by the length of the connecting-rod cannot be considered as being 
 
EFFECT OF INERTIA AND WEIGHT. 
 
 795 
 
 Fig. 316« 
 
 Fig. 3167> 
 
 12 stroke 
 
 4000-- 
 
 3000- 
 
 2000- 
 
 iOOO 
 
 120 ISO I8p 2fp 24f) zfp lop 3fp 
 Fig. 8 1 fie. 
 
 Fig. 316c. W# 
 
796 MANUAL OF MAKIA'E ENGINEERING. 
 
 distributed over the piston area (except for a connecting-rod of infinite 
 length), and the effect of the weight of the rotating parts must, therefore, be 
 considered apart from the indicator diagram. 
 
 Then (fig. 3166) the weight of the reciprocating parts being a constant 
 
 quantity, set off H C = *—^ downwards for the downstroke and upwards 
 
 for the upstroke, and draw a curve, H N, through H. parallel to the inertia 
 line. C K. 
 
 Set off the indicator diagram from the straight back pressure line A B. 
 
 Then the effective pressure on the piston at the beginning of the stroke 
 is not A F but H F, and at any other position of the piston the effective 
 pressure. P. is given by the intercept between the curves F G E and H N. 
 And similarly for the upstroke, due note being taken of the algebraic sense 
 of the various pressures. 
 
 The total effective twisting moment at any instant is now — 
 
 TM = Pr ( S[n {d X a) ) + M 2 r sin (6 + p). 
 
 \ COS a J z 
 
 P r - may be calculated in a semi -graphical manner as follows :— 
 
 Let the crank be at H (fig. 315), then 
 
 py sin(0+* ) =spxoa 
 
 COS a 
 
 Also, when <p is 0. then 
 
 M 2 rsin# = M 2 X D 0. 
 
 18. An example of Case 1 will now be considered. 
 
 Although it is very necessary to include the valve gear when balancing, 
 it is rather an unnecessary elaboration to consider its influence on the twisting 
 moment diagram, as it is so very small, unless, of course, the gear be abnor- 
 mally heavy : — 
 
 Let M x = 400 lbs. (reciprocating). 
 M 2 = 200 lbs. (rotating). 
 Revolutions = 200 per minute. 
 12-inch cylinder, 12-inch stroke. 
 Connecting-rod 2 feet centres. 
 Position of fittings on shaft as in fig. 294. 
 
 Then 
 
 v = 10*47 feet per second. 
 v- = 109-6. 
 
 24 inches 
 
 « = -fr-. ^— = 4. 
 
 6 inches 
 
 h* = -24-,. 
 
 Agr 
 
STRESSES DC>: TO INERTIA. 
 
 707 
 
 The values of the inertia pressure, etc.. for every 30 degrees of the crank 
 are tabulated below (fig. 316, a. b, c. d, e) : — 
 
 
 M,i* 
 
 
 
 
 
 
 
 
 - . - X 
 
 Agr 
 
 P 
 
 OC 
 
 P.OC 
 
 M 2 x 
 
 M 2 siu (0 + «) 
 
 TM = 
 
 
 , cos 2 e 
 cos 5 + . 
 
 I 
 
 = GL. 
 
 (Fig. 3166). 
 
 x A = Q,. 
 
 sin (»" + ?). 
 
 x r = Q 2 . 
 
 Q. + Q s . 
 
 
 Lbs. 
 
 Lbs. 
 
 Inches. 
 
 Inch -lbs. 
 
 Lbs. 
 
 Inch -lbs. 
 
 Inch-lbs. 
 
 o° 
 
 - 30-1 
 
 
 
 
 - 130 
 
 - 780 
 
 780 
 
 30° 
 
 - 23 9 
 
 60 
 
 3-64 
 
 24',700 
 
 - 270 
 
 - 1,620 
 
 + 23, 1 80 
 
 60° 
 
 - 9-04 
 
 73 
 
 5-82 
 
 48,000 
 
 - 345 
 
 - 2,070 
 
 + 45,930 
 
 90° 
 
 + 6-02 
 
 67 
 
 6 
 
 45,500 
 
 - 330 
 
 - 1,980 
 
 + 43,520 
 
 120° 
 
 + 15 06 
 
 60 
 
 4-52 
 
 30,660 
 
 - 225 
 
 - 1,350 
 
 + 29,310 
 
 150° 
 
 -f 17 85 
 
 36 
 
 2-35 
 
 9,580 
 
 - 60 
 
 - 360 
 
 + 9,220 
 
 180° 
 
 + 18-08 
 
 
 
 
 + 130 
 
 + 780 
 
 + 780 
 
 210° 
 
 + 17-85 
 
 56 
 
 2-35 
 
 14,800 
 
 + 270 
 
 + 1,620 
 
 + 16,480 
 
 240° 
 
 + 15-06 
 
 61 
 
 4-52 
 
 31,200 
 
 + 345 
 
 + 2,070 
 
 + 33,270 
 
 270° 
 
 + 6 02 
 
 54 
 
 6 
 
 36,700 
 
 + 330 
 
 + 1,980 
 
 + 38,680 
 
 30J° 
 
 - 9 04 
 
 48 
 
 5-82 
 
 31.600 
 
 + 225 
 
 + 1,350 
 
 + 32,950 
 
 330° 
 
 - 23-9 
 
 43 
 
 3-64 
 
 17,700 
 
 + 60 
 
 + 3G0 
 
 + 18,060 
 
 360° 
 
 - 30 1 
 
 
 
 
 - 130 
 
 - 780 
 
 780 
 
 To find <p. let O C x (figs. 294a and 316c) be the resultant of the balance 
 weights and CjK equal weight of rotating parts (200 lbs.) then O R is their 
 resultant. 
 
 Draw the curve R, sin d (fig. SlQd). 
 
 Set off S D = p. then D a = R sin (d + <p), or M 2 sin (e + p). 
 
 Fig. 316e is the effective twisting moment diagram. 
 
 19. The area of the final twisting moment diagram is proportional to the 
 work done, and although its form has been modified by the inclusion of the 
 inertia and weight, the area remains the same as if the steam pressure alone 
 were considered. 
 
 This also follows from the principle of conservation of energy. 
 
 If the engine be balanced completely horizontally, and not vertically as 
 in Case 1. the inertia and weight pressures from the reciprocating parts 
 are the same as above, but the moment of the weight of the rotating parts 
 is zero at all positions of the crank. 
 
 If the engine be balanced by bob weights (two) then the inertia of these 
 must be considered exactly as if there were two extra cylinders in the engine, 
 the only difference being that there is no steam diagram to combine with 
 their inertia diagram. 
 
 The foregoing process applies equally well. to a multi-crank engine. 
 
 The reciprocating weights and inertia pressures must be found separ- 
 ately for each cylinder, and the resultant horizontal weight-twisting moment 
 for the whole engine combined with the resultant twisting moment due to 
 the effective piston pressures. 
 
 20. Stresses due to Inertia. — As the stress in a given piece of machinery 
 is due to the effective load applied to it, it therefore follows that inertia 
 and dead-weight loads should be added algebraically to the steam load? 
 when finding the stresses in those parts that are subject to them. 
 
 In very slow-run ping engine,-: the effect of inertia is inappreciable : and 
 may, therefore, be neglected, but in quick revolution and high-piston-speed 
 engines its effect must not be lost sight of. 
 
 51 
 
798 MANUAL OF MARINE ENGINEERING. 
 
 Starting at the piston, the first section to be dealt with is through the 
 piston-rod thread. The inertia force is that due to the weight and accelera- 
 tion of the piston, and not of the whole of the moving parts. The stress in 
 the thread only comes into play on the upstroke, and the effective pressure on 
 the piston is the intercept between the steam and weight lines as shown above. 
 
 Similarly, the stress at any section between the top of the piston and the 
 centre of gravity of the connecting-rod may be found, the inertia line being 
 constructed from the weight of the parts above the section. 
 
 As a rule, the factor of safety allowed for marine engines of the ordinary 
 type amply covers the additional stresses due to inertia, and an engine that 
 is well designed for static stresses may generally be considered to be .strong 
 enough to provide for dynamical stresses as well. 
 
 Referring to fig. 3166, it will be observed that the inertia line tends to 
 make the effective pressure constant throughout the stroke. 
 
 For engines of normal scantlings and speed the maximum effective pressure 
 usually occurs at about two-thirds the stroke from the beginning. 
 
 As the weights of moving parts are rarely available when an engine is 
 being designed, and as there are also no indicator diagrams to refer to, the 
 following approximate formula for finding the maximum effective pressure is 
 given. This formula includes the total weight of reciprocating parts. 
 
 Let p m = mean pressure of steam for that cylinder 
 from the proposed I.H.P., 
 Pf = maximum effective pressure, 
 L = length of stroke in feet, 
 N = revolutions per minute, 
 then p f = f m + -0000386 N 2 IA 
 
 The formula applies to the low-pressure cylinder for engines where there 
 is only one low-pressure cylinder. If there are two low-pressure cylinders, 
 then the medium pressure should be taken. It is not suitable for high- 
 pressure cylinders of compound expansion engines, but this does not detract 
 from the value of the formula, because the maximum loads generally occur 
 in the above-meantioned cylinders. 
 
 Example. — To find the maximum effective pressure that may be expected 
 on the low-pressure piston of a triple-expansion engine of 1,500 I.H.P., 150 
 revolutions, 54-inch low-pressure cylinder, 2 feet stroke — 
 
 1,500 x 33,000 
 Pm ~ 3 x 2,290 X 2 x 2 X 150 
 = 12*00 lbs. per square inch. 
 
 f f = 12 + -0000386 x 22,500 x 4 
 = 12 + 3-47 
 = 15*47 lbs. per square inch. 
 
 Example. — To find the diameter of the connecting-rod bolts (two) for the 
 above engine, allowing a stress of 5,000 lbs. per square inch — 
 
 Maximum working load on piston = 2,290 X 15-47 = 35,426 lbs. 
 
 35 426 
 ^-— — - = 3-54 square inches at bottom of thread 
 2 X 5,000 ^ 
 
 = 2i-inch bolts. 
 
THE DESIGN OF BALANCE WEIGHTS. 799 
 
 21. The Design of Balance Weights.— As a rule, and more particularly 
 with two- and four-crar.k engines, the great difficulty to contend with is being 
 able to get the required mass in the space available. 
 
 With a uniform thickness and external radius there is a limiting length 
 oi weight (circular arc), beyond which any increase of mass makes the total 
 mass less effective. The limit varies according to the shape of the weight 
 and the angle which it makes with the crank. 
 
 When the weight is homogeneous and fairly opposite the crank its length 
 should not subtend a greater angle than 120°, and in no case more than 140°, 
 at the centre of the shaft. When inclined at a considerable angle to the 
 crank, as shown in figs. 317 and 318, its length should not subtend more 
 than 90°, on account of the long arm for attaching to the crank web — for 
 any extension of length beyond these limits causes the centre of gravity 
 to approach the centre of the shaft at a greater rate than the mass is increased. 
 
 For composite weights, as in figs. 317 and 318, the angles subtended 
 with the centre of shaft may be increased about 5° above those given above. 
 
 The material of which balance weights are made depends, to a great 
 extent, on the manner in which they are attached to the crank-webs. 
 
 In small engines, and when the weights of the moving parts are obtain- 
 able at the time the crank shaft is designed, and when the crank-shaft is a 
 built one, the weights are very often forged as part of the webs. 
 
 Circumstances, however, generally necessitate loose balance weights 
 being fitted. 
 
 A neat and compact method of attachment is shown in figs. 317 and 
 318, and consists of forging small lips on the crank-webs, over which the 
 weights are keyed. 
 
 The material the weights are made of, when this method of attachment 
 is adopted, must necessarily not be of a .short nature, and should either be 
 forged or cast steel of a low tensile strength. 
 
 The Admiralty and a few other owners will only accept forged material, 
 but there is no reason why a suitable cast steel should not be used for them. 
 It has been used successfully by the author for many years. The advantage 
 of a cast weight lies in being able to make it in the form of a shell which may 
 be filled with lead, the materials thus being most advantageously disposed 
 and the weight being of the smallest possible dimensions. It is also, generally, 
 less expensive. This is shown in fig. 317. A forged balance-weight is shown 
 in fig. 318. The most convenient method of manufacture is to shape them 
 with a band-saw out of rolled slabs. Parts of the forging may be drilled 
 out and filled with lead as shown, but at beet this is an expensive addition 
 and hardly worth the extra labour, as the quantity of lead that can be intro- 
 duced is necessarily only a small fraction of that in fig. 317. Cast-iron weights 
 are used to a great extent on high-speed land engines, and to a less extent 
 for marine work. A cast-iron weight cannot be relied on if held on by lips 
 as in figs. 317 and 318, and the usual method of attachment is by strapping 
 it on as shown in fig. 319. The casting may be cored out and filled with lead 
 if the small space available for it makes it advisable. 
 
 Cast-iron weights with its straps and fittings are, on the whole, much 
 cheaper than either cast or forged steel, and there does not seem to be much 
 reason why, beyond being somewhat clumsy in appearance, they should not 
 be used to a greater extent than they are. 
 
800 
 
 MANUAL OF MARINE ENGINEERING. 
 
 u 
 
 I- 
 
 w 
 to 
 
 < 
 
 
THE DESIGN OF BALANCE WETGHTS ttUI 
 
 In all eases the lips or straps must be made strong enough to take the 
 centrifugal force of the balanceweight. together with the deadweight itself. 
 If strong enough for this they will, in general, be found strong enough to 
 resist the impulsive force on them, due to the engine being suddenly started 
 or stopped. 
 
 It is not expedient in a manual of this kind to devote more space than 
 the foregoing to the problem of balancing. Literature on this subject ia 
 exceedingly voluminous and fairly accessible. The aim of this chapter ha* 
 been to place the elementary and practical principles before the student in 
 such a manner that, should he desire it, he would have no difficulty in pursuing 
 the subject more profoundly from the many highly mathematical papers 
 extant. 
 
 The methods, short cuts, etc., oi attacking the problems of actual practice 
 will suggest themselves, as he proceeds, to the designer who has got a thorough 
 grasp of the principles involved. 
 
W02 MANUAL OF MARINE ENGINEERING. 
 
 CHAPTER XXIX. 
 
 MATERIALS USED BY THE MARINE ENGINEER. 
 
 In every ship the weight of the machinery is of some consideration, and in 
 many ships it is of the utmost importance that it be as light as possible. 
 Reduction in weight can only be made by using material of superior strength, 
 and by so designing the engine as to be able to employ such materials. The 
 most marked advances in marine engineering, both in the size and quality 
 of the work turned out, are in great measure due to the superior materials 
 now available, and to the means possessed for converting them to the engineer's 
 L - equirements. Steel, which was, only a few years ago, a comparatively 
 expensive material, has now practically displaced iron, and that is largely 
 on the score of cheapness. Not only is the price of steel now so low that 
 it can compete successfully with iron, but steel can be produced in such 
 large masses at such comparatively small cost, that heavy shafts* of this 
 material made possible the immense engines recently constructed for the 
 large ocean steamers and warships. Now steel alloys are supplied of such 
 excellent quality, and with such high elastic limits, that for high-class 
 engineering they are likely to displace ordinary mild steel, notwithstanding 
 their prices being so much higher. Bronzes, too, have commanded attention, 
 for not only do they hold their own for special purposes, but since certain 
 kinds can be manufactured with as high a tensile strength as steel, and stand 
 working hot or cold, they have displaced steel and iron for such purposes 
 as propeller blades, parts of the engines, etc., in spite of the comparatively 
 high price. 
 
 Cast Iron. — Although steel has superseded iron in many ways, cast iron 
 still remains as the material most largely used by marine engineers. Its 
 appearance is so well known as to need no description. The qualities of cast 
 iron are usually connected with the district from which the ore is obtained. 
 There also still remains a distinction from mode of manufacture, as " cold 
 blast" is distinguished from "hot blast" in other ways than that of cost; 
 but not much cold blast is made now. 
 
 All the pig iron of each district is, as a rule, divided into four to seven 
 sorts, each of which is known by a number ; that containing most free 
 carbon, or graphite, is designated No. 1, that containing least Nos. 4 or 7. 
 
 No. 1 pig, when broken, exhibits a very coarsely granular fracture, having 
 dark grey scales of considerable size ; when melted " it runs very thin," that 
 is, becomes extremely fluid, and on that account is used for ornamental 
 castings and other work which requires a sharp outline, or is of very little 
 thickness. It is not used much by marine engineers, except to mix with 
 other kinds, or where extreme fluidity is necessary. Castings made from it 
 
 •Screw shafts are now made nearly 30 inches diameter and 60 feet long, weighing 64 tons in the rough state. 
 
SHROPSHIRE IRONS. 803 
 
 are very soft unless they are so thin that the metal has changed its character 
 in the mould by chilling. The carbon here exists chiefly in the form of 
 graphite scales, mechanically mixed as it were with the iron crystals. 
 
 No. 2 pig is not so soft nor so fluid when melted as No. 1. but is still not 
 sufficiently close grained for general use. 
 
 No. 3 pig is that usually employed by the marine engineer for general 
 purposes, as by adding some of No. 1, a mixture suitable for a complicated 
 casting is obtained, and by adding to it some of No. 4, a harder and closer- 
 grained casting can be made. 
 
 No. 4 pig is not much used for foundry purposes, except as a means of 
 closing the grain and hardening the metal. It also differs from the other 
 numbers in the appearance of the fracture ; they all present a highly crystal- 
 line fracture of a distinct grey colour, and on that account called grey iron ; 
 it shows as a grey iron at the fracture, but the grain is much finer, and there 
 is an absence of the coarse graphitic scales so strongly marked in the other 
 three numbers. 
 
 Nos. 5 and 6 are not used at all for foundry purposes, but made for manu- 
 facture into wrought iron ; on this account they are called forge irons, and 
 as the fracture is still somewhat grey in places it is called " grey forge " 
 to distinguish it from No. 7, which is called " white forge," as the fracture 
 displays a distinctly crystalline structure, very hard, and silvery white in 
 appearance. The carbon in this case is wholly combined with the iron 
 chemically. Nos. 5 and 6 often present a mottled appearance, as if a grey 
 iron and a white iron had been melted and imperfectly mixed ; it is on this 
 account that it is often known as " mottled pig." 
 
 It is often customary to generalise and recognise only three varieties of 
 pig iron, viz. : — No. 1 as grey iron, No. 2 as mottled iron, and No. 3 as white 
 iron ; Nos. 2 and 3 being, of course, looked on as " forge " iron, and No. 1 
 as " foundry " iron. 
 
 In Great Britain there are many kinds of pig iron used by moulders, each 
 called by the district where the ore is raised and smelted. 
 
 Scotch Iron is one of the best for foundry purposes ; it is very uniform 
 in quality, of good strength, and will mix well. There are various brands, 
 those of Gartsherrie, Glengarnock, Eglinton, Carron, etc., etc., being best 
 known. The No. 3 pig is most generally used by marine engineers, as it will 
 run sufficiently fluid to make any casting, and judiciously mixed with scrap, 
 it can be depended on for both closeness of grain and strength. 
 
 Cleveland Iron is used very much in the Cleveland district for general 
 work ; it is harder than Scotch iron, but does not possess so much strength, 
 nor is it nearly so tough. A mixture of Nos. 1 and 3 is used for large work 
 requiring strength, and this, when melted, possesses fluidity enough for the 
 ordinary marine castings. Cleveland iron alone is not fit for large marine 
 cylinders, columns, etc., but with the addition of haematite and scrap, to a 
 mixture of Nos. 1 and 3, a good casting is obtained for general purposes. 
 
 Lincolnshire Iron is about equal in quality and general description to the 
 Cleveland. 
 
 Staffordshire, Yorkshire, Derbyshire, and Welsh Irons are very good, and 
 used in their own districts generally. 
 
 Shropshire Irons are very good, and from this county, Monmouth and 
 Stafford the famous cold blast irons were obtained. 
 
804 MANUAL OF MARINE ENGINEERING. 
 
 Cumberland Iron is made from hcematite ore, and the pig generally goes 
 by the name of " West Coast hoematite." It is generally used for steel-making, 
 but is also employed to improve other irons for foundry use. it possesses 
 great strength and toughness, but cannot be used by itself for foundry pur- 
 poses, as it does not run well when melted ; 20 per cent, blended with good 
 Scotch and English irons makes a strong mixture. 
 
 Cold Blast Iron. — Iron manufactured with air at the ordinary tempera- 
 ture is called by this name to distinguish it from the generality of irons 
 which are made by heating the blast to about 700° F. The best known 
 and most generally used cold blast irons were " Bkenavon" and " Lilleshall," 
 which owe part of their excellency, no doubt, to the good ore from which they 
 are made. This iron possesses great strength with closeness of grain, and was 
 used to close the grain and strengthen other irons, but is not now purchasable. 
 
 Iron Mixtures. — All cast iron is improved by re-melting, and the improve- 
 ment continues until it has been re-melted as many as twelve times : after 
 this it falls off in strength. No important casting should, therefore, be made 
 from new iron only, and such as cylinders, pistons, cylinder covers, etc., 
 should be made from iron wholly re-melted if the utmost strength is to be 
 obtained. This rule, however, is only carried out when the thickness of 
 ■netal is cut down to the lowest limit to save weight. 
 
 Cylinders, etc., which are subject to shock as well as to changes of tem- 
 perature and severe strains, should be made of strong tough metal, and, 
 since a good surface is required to withstand the wear of pistons and valves, 
 the metal must have a close grain. Such cylinders, etc., used to be made of a 
 mixture of one-third picked scrap, one-third best Scotch No. 3 pig. and one- 
 third Blamavon or Lilleshall. If the casting is not large and complicated, 
 the grain may be closed by adding picked scrap of a hard nature and increasing 
 the amount of Blsenavon. Some moulders add hcematite to this mixture 
 to increase the strength, but this is not often done, as it decreases the fluidity. 
 
 If the cylinder is to have liners and false faces for the valves to work on, 
 it needs only to be strong ; all hard materials, therefore, may be omitted. 
 
 Cylinder liners and false faces require to have a fair amount of strength, 
 and be as hard as consistent with capability of being machined. Scrap iron 
 of close grain and hard nature is selected to add to best Scotch No. 3 pig and 
 Bkenavon ; if hard scrap cannot be obtained, some No. i pig may take its 
 place ; if necessary, even a small portion of " white " iron may be added to 
 give additional hardness. Hardness and strength with closeness of grain 
 may be obtained by mixing scrap steel (shearings and punchings from boiler- 
 plates) to the extent of even 10 per cent. As th • metal becomes much staffer 
 on the addition of the steel, only thick plain castings can be made with it. 
 Propeller blades and bosses may with advantage be made with a mixture 
 containing steel in lieu of haematite ; the quantity of steel which may be added 
 for this purpose depends on the founders being able to melt it. 
 
 Foundations and other large masses which are uecessarily heavier than 
 absolutely needed for strength may be made of a poorer mixture than suffices 
 for cylinders ; but as they are liable to shock, the iron must be of such a 
 description as to resist this. 
 
 Specific Gravity of Cast Iron varies from 6-886 to 7-289 ; the average may 
 be taken at 7 -20. The weight of a cubic foot is. therefore, 450 lbs., and a 
 cubic inch 0-26 lb. A square foot of it 1 inch thick weigh? 37-50 lbs. 
 
Maximum. 
 
 Average. 
 
 14*05 tons. 
 
 7 - 36 tons, 
 
 4-47 ., 
 
 , t 
 
 3-44 „ 
 
 , . 
 
 58-42 ., 
 
 50 00 „ 
 
 BIVET IRON. 805 
 
 Strength of Cast Iron. — The following is the result of some careful experi- 
 ments made at Woolwich Arsenal some years ago : — 
 
 Minimum. 
 
 Tensile strength per square inch, 4-85 tons. 
 
 Transverse „ ., ,, 137 ., 
 
 Torsional „ „ „ 1-74 ,, 
 
 Crushing „ „ „ 22-54 „ 
 
 Shearing „ „ „ .. .. 12-00 ,, 
 
 "Good cast iron made by mixing qualities suitable for marine work should 
 have an ultimate tensile strength of 25,000 lbs. and resistance to crushing 
 of 110,000 lbs. ; a bar 1 inch square and 36 inches long should deflect J inch 
 without breaking with a load of 800 lbs. at the middle. 
 
 Wrought Iron. — This material, which is very nearly pure iron, is made 
 exclusively in this country by the " indirect process " — that is, manufactured 
 from cast iron by the process called puddling. The old methods of obtaining 
 malleable iron direct from the ore, hence called the " direct process," are 
 practised only by untutored tribes, or in regions inaccessible to general trade. 
 The Siemens method of manufacturing steel is to some extent a reversion 
 to the "direct process," and the product was sometimes properly called 
 ingot iron, to distinguish it from iron made from the " bloom " which is obtained 
 by puddling. 
 
 Rolled Bar Iron. — The "blooms" from the puddling furnace, after being 
 squeezed or hammered, are rolled into bars, which are known as "puddled 
 bars " ; this bar iron is not used by engineers, as its tensile strength is some- 
 times as low as 9 tons per square inch ; it is cut into pieces, which are piled 
 crossways into a "faggot" or "pile," reheated, and rolled again into bars, 
 and now called " merchant bar " iron. 
 
 Merchant Bars are not used by engineers for any very important work, as 
 the iron is still of low tenacity and not very uniform in structure or quality. 
 It was used for making gratings, ladders, etc., for fire-bars, bearers, etc. 
 
 Best Bar is made by reheating "faggots" of merchant bar iron, or good 
 wrought-iron scrap " cross piled," and rolling it again into bars. Its strength 
 is now much improved, and its quality more uniform, and it may be used for 
 general smithing purposes. Its tensile strength, if made of good material, is 
 about 24 tons per square inch on the average ; the " Best " bars of some 
 makers will withstand as much as 26 tons per square inch. 
 
 Best Best Bar is made by again rolling bars from faggots of selected Best 
 bars. It has now a very uniform silky fibre, will bend double cold, and has 
 a tensile strength of 26 to 27 tons ; good specimens should elongate 25 per 
 cent, at fracture with a reduction of area of about 50 per cent. Some kinds 
 of iron have even a higher tensile strength than this, especially when rolled 
 into round bars. Bar iron by cold rolling is increased in strength, but the 
 elongation is reduced. Such iron is, however, seldom or never used now 
 by engineers, except for the screwed stays of boilers. 
 
 Rivet Iron, when used, should be soft and of very good quality to with- 
 stand the work put on it in the process of riveting ; its tensile strength is, 
 therefore, somewhat low, about 24 tons per square inch, and its resistance 
 to shearing is even lower than this, being from 20 to 22 tons per square inch. 
 
806 MANUAL Or MARINE ENGINEERING. 
 
 This iron is generally made by rolling " piles " of selected scrap iron into bars, 
 and for best quality rivet iron, the bar is rolled from " piles " of ordinary 
 rivet iron. 
 
 Rolled Scrap Iron. — A superior quality of bar iron is made by " piling " 
 shearings from boiler plates and rolling in the usual way ; the quality of 
 this iron, however, depends very much on that of the plates from which 
 the shearings came. 
 
 Weight of Bars. — The specific gravity of bar iron of good quality is on the 
 average 7*62 ; the weight of a cubic foot is, therefore, 476 lbs., and that of 
 a cubic inch 0*276 of a pound. Yorkshire bar iron is somewhat denser, its 
 specific gravity being 7 - 76 ; so that a cubic foot of it weighs 485 lbs. 
 
 Yorkshire Iron. — The best kinds of boiler iron were made in South York- 
 shire, in the neighbourhood of Bradford and Leeds, and known as " best 
 Yorkshire iron." Krupp, and some other German manufacturers, made iron 
 plates of a quality equal to this ; Swedish and Russian plates are very similar 
 to it, and in some respects of superior quality. 
 
 It was of very uniform quality, and, although not possessing a very high 
 tensile strength in direction of the grain, it was superior to other irons in 
 strength across the grain ; it was very tough, and had great elasticity, so that 
 it was easily flanged and bent, and stretched very considerably before break- 
 ing. For these reasons, it was most valuable for boiler-making, and not- 
 withstanding its high price (on the average three times that of ordinary 
 boiler plates) it was always used for those parts of a marine boiler exposed 
 to flame. Bar iron of this quality is still made in considerable quantities 
 for smithing purposes in railway work, as well as for the screwed stays of 
 boilers. It is practically as strong as boiler steel, and stands alternating 
 stresses better than ordinary steel ; it does not, however, take so good a screw 
 thread. It is also used for crane chains and other high-class work, and was 
 much esteemed for colliery work. * 
 
 Some best Yorkshire plates had a tensile strength of as high as 24 to 25 
 tons with the grain, and 22 to 23 tons across it ; the elongation being 13*5 
 and 8 per cent, respectively. From some carefully made experiments, Mr. 
 Kirkcaldy found the average strength of Yorkshire iron to be 21*3 tons with 
 and 20 - l across the grain, the elongation being 16*7 and 11*2 per cent.; by 
 annealing the plates the strength was slightly reduced, but the elongation 
 was raised, to 18*4 and 12 - 8 per cent. The elastic strength was also found 
 to be 12 '2 tons in tension, and in compression 11 '5 to 13*3 tons. 
 
 Its specific gravity is 7 '76 ; the weight of a cubic foot is 485 lbs., and that 
 of a cubic inch is 0-281 lb. 
 
 Staffordshire Iron. — This iron, although slightly inferior to best Yorkshire, 
 is still of high quality, and was used for boiler shells, domes, etc., and for 
 such parts of the furnaces and chambers as were not exposed to the direct 
 action of flame. It has a high tensile strength with the grain, but is not so 
 strong across the grain as is the Yorkshire iron. Sir William Fairbairn, in 
 1861, found that some Best Best Staffordshire plates had an ultimate strength 
 of 26-7 tons with and 24 - 47 tons across the grain, the elongation being 6*7 
 and 4 per cent. ; that common Staffordshire plates had an ultimate strength of 
 22'7 tons with and 23"5 across the grain, the elongation being 5 and 4*35 per cent. 
 
 For purposes of cumulation, the ultimate strength of Staffordshire quality 
 
 ♦ The Board of Trade allows a forking stress of 9,000 lbs. per square inch on such bar iron, provided 
 It has been tested to 21-5 tons with an elongation of 27 per cent, in 8 inches. 
 
BESSEMER STEEL. 807 
 
 boiler plates may be taken at 51,500 lbs. with, and 43,000 lbs. across, the 
 grain for plates under § inch thick ; and at 50,000 lbs. with, and 40,100 lbs. 
 across, the grain for plates over that thickness. 
 
 The specific gravity is 7*68, the weight of a cubic foot being 480 lbs., and 
 that of a cubic inch 0-277 lb. 
 
 Staffordshire Bar Iron is very largely used for chains and similar gear, as 
 well as for smithing purposes, and is very highly esteemed for them. 
 
 Iron Forgings.— Iron forgings, when used, are made from scrap iron, 
 and their strength depends very much on that of the iron from which the 
 scrap was cut. Sometimes forgings were made from new iron, but, since 
 the general use of steel, this is seldom or never done now. The method of 
 manufacture is similar to that described for making rolled bars from scrap ; 
 the scrap is sorted, piled, brought to a welding heat, and hammered into 
 slabs ; the slabs are piled one on the other, and reheated to form the forging 
 required. The best description of forging is made by rolling the slabs into 
 bars, so as to give the metal grain ; the bars are then cut into short lengths, 
 piled, and hammered again into slabs, which are piled, etc., as before, to form 
 the forging. This rolling into bars, in addition to giving the iron fibre, tended 
 to give a more uniform structure to the forging, and a homogeneity which 
 cannot be obtained by the simple piling process. 
 
 The specific gravity of large forgings was about 7*63, so that the weight of 
 a cubic foot was 477 lbs., and that of a cubic inch is 0*276 lb. 
 
 Steel. — Steel is used in the form of bars, plates, and forgings, and is 
 also very generally employed for castings where great strength is 
 required. 
 
 All steel was originally made from the best qualities of wrought iron by 
 the process of " cementation " ; this consists of exposing pieces of nearly pure 
 iron to a high temperature in the presence of carbon only in a closed vessel 
 for a considerable time, during which some of the carbon is absorbed by 
 the iron, and is thus converted into a rough kind of steel, called blister steel. 
 These pieces are broken, and sorted according to the appearance of the fracture 
 after which portions are placed in a closed crucible, melted, and cast into 
 ingots ; it is now called cast steel. If the blister steel is piled, reheated, 
 and hammered, or rolled into bars, it is called shear steel. The cast-steel 
 ingots are worked into bars, which still retain the name " cast steel," but 
 this is better known as tool steel, as it is now used almost exclusively for 
 cutting tools. Tool steel containing about 1 per cent, of carbon is, of course, 
 very hard, and has a very high tensile strength, ranging from 50 to 65 tons 
 per square inch, with an average elongation of a little over 5 per cent, only ; 
 some of the milder kinds, such as are used for drifts, etc., have a tensile 
 strength varying from 44 to 60 tons per square inch, with an average elonga- 
 tion of 13 per cent. Spring steel is still milder, having a tensile strength of 
 about 45 tons per square inch, with an elongation of 18 per cent. Tempering 
 in oil increases the strength considerably ; Mr. Kirkcaldy found that a certain 
 steel bar, whose strength when " soft " was 54J tons, when heated and cooled 
 in oil had a strength of 96 tons. 
 
 Bessemer Steel. — The modern methods of making steel known as the 
 direct processes obviate the necessity of using the comparatively expensive 
 wrought iron, and are, therefore, capable of producing a very much cheaper 
 material. In the Bessemer process there are essentially two operations, the 
 
803 MANUAL OF MARINE ENGINEERING. 
 
 conversion of molten cast iron into practically pure iron, and, by the addition 
 of a small and definite quantity of manganese and carbon, the converting of 
 pure iron into steel. Cast iron practically free from phosphorus and sulphur 
 is melted and poured into the converter ; a strong blast of air is forced through 
 the molten metal, so that the carbon it contains burns and is consumed, and 
 the temperature of the mass thereby raised ; when the whole of the carbon 
 is consumed, a small quantity of ferro-manganese (spiegeleisen or natural 
 ferro-manganese was used originally), an iron alloy containing a known 
 proportion of carbon and manganese, is added ; the metal now has that 
 small amount of carbon which causes it to differ from wrought iron, and 
 that amount of manganese which seems to be so essential in making good 
 steel. The metal is now run into ingot moulds, and allowed to cool for 
 further use, or is kept in a " soaking pit," reheated and hammered, or rolled 
 into the forms required. By the Thomas-Gilchrist process quite good steel 
 can be made in a Bessemer converter from iron containing phosphorus ; 
 the phosphorus is absorbed by the converter lining of a gannister, which 
 is prepared from magnesian limestone ; the product is known then as 
 basic steel, while that made by the original Bessemer process is called 
 acid steel. 
 
 Siemens-Martin Steel. — The steel in this process is made in the hearth of 
 a reverberatory furnace by exposing for a considerable time molten haematite 
 or pig iron, or mixtures of cast iron and high-class pure (oxide) ores of iron, 
 to the intense violet heat obtained by producer gas until practically the whole 
 of the carbon disappears. If pig containing phosphorus is used, the furnace 
 must be lined with gannister made from magnesian limestone on the Thomas- 
 Gilchrist plan, so that the phosphorus may be absorbed almost wholly by 
 the lining during the roasting process ; ferro-manganese is added as in the 
 Bessemer process, and the liquid mixture run into ingot moulds. Steel 
 made in the ordinary way from good pig free from phosphorus is said to be 
 by the acid process, while that made from pig iron containing considerable 
 quantities of phosphorus in a furnace or converter lined with magnesian 
 limestone gannister is called basic steel, and said to be made by the Thomas- 
 Gilchrist or basic process. Basic steel is generally softer than ordinary steel, 
 but excellent metal is made by this process, having an ultimate tensile strength 
 of not more than 25 tons, with a stretch of over 30 per cent. ; it can be bent 
 cold two double, and is soft to work and easily welded. Large quantities 
 of this material are now made at Middlesbrough, North Lincolnshire, etc., 
 where there is an unlimited supply of suitable iron, and used by shipbuilders, 
 tank and bridge builders, etc. Basic steel of much higher tensile strength 
 can be now obtained and used with satisfaction. 
 
 At the present time steel plates, bars, and forgings are made almost 
 exclusively from ingots run from either Bessemer converters or Siemens 
 furnaces, and the basic is the kind most in use by engineers. 
 
 The Steel for Boiler Construction,* however, is made by the Siemens- 
 Martin process exclusively. The ingots are reheated and hammered into 
 slabs, which are again reheated and rolled into bars or plates. It is necessary 
 that the material of which a boiler is constructed shall have very considerable 
 elasticity as well as strength, and since best Yorkshire iron stretched to as 
 much as 18 per cent, before fracture no steel is used which is not equal to this. 
 
 * Boiler plates may be basic or acid, but in every case the steel must come from au open hearth 
 furnace. 
 
BOARD OF TRADE TESTS OF MATERIAL. 
 
 W\> 
 
 It is found that the lower the ultimate strength of good steel is, the 
 higher is its elasticity, so that while plates of this material, having an ultimate 
 elongation of 20 per cent., possess a tensile strength of 35 tons per 
 square inch, those having an ultimate strength of only 26 tons will 
 stretch 30 per cent.* The former is very suitable for the shell- 
 plating of boilers, as its strength is nearly 50 per cent, higher, and its 
 elasticity nearly double that of the iron formerly used by boilermakers ; 
 the latter does eminently for the internal parts of a boiler which require to 
 be of a somewhat softer material, that it may be flanged, etc., with ease, 
 and stand the rough usage of the boilersmiths with safety. For this purpose 
 plates are used which have a tensile strength of 26 to 28 tons per square 
 inch, with an ultimate elongation of about 25 per cent. For corrugated 
 furnaces a milder steel still is used, for its limit of strength is 25 tons, and 
 it stretches to 30 per cent, in 8 inches. Shell-plates are now made with a 
 tensile strength of 35 to 40 tons, and a stretch of 20 to 25 per cent, with the 
 grain, and 20 per cent, across the grain. 
 
 Subjoined are the tests required by the various authorities for boiler 
 steel. It will be seen that no two of them agree, which is much to be 
 regretted ; it is in the interest of both science and economy that they should 
 all agree :— 
 
 Admiralty Tests of Material. — All steel to be made by the acid open hearth process. 
 Every plate, etc., used is to be tested, and must comply with the requirements stited 
 below :— 
 
 TABLE XCII. — Admiralty Tensile Tests. 
 
 I 
 
 
 Minimum 
 
 Maximum 
 
 Minimum 
 
 
 Ultimate Tensile 
 
 Ultimate Tensile 
 
 Elongation 
 
 Description of Material. 
 
 Strength, 
 
 Strength, 
 
 in 
 
 
 Tons 
 
 Tons 
 
 8 inches 
 
 
 per square inch. 
 
 per square inch. 
 
 per cent. 
 
 
 
 
 per cent. 
 
 Not exposed to flame and not flanged, 
 
 27 
 
 30 
 
 20 
 
 Exposed to flame and flanged, - 
 
 24 
 
 27 
 
 25 
 
 Rivet bars, - - - 
 
 24 
 
 27 
 
 25 
 
 Steam-pipe plates, .... 
 
 24 
 
 27 
 
 33 
 
 Corrugated or ribbed furnace, - 
 
 23 
 
 25 
 
 27 
 
 Tube forgings (annealed), - 
 
 21 
 
 24 
 
 27 
 
 Pieces cut from tubes (annealed), 
 
 ... 
 
 26 
 
 27 
 
 For bending tests the specimens are heated to a low cherry red, and then cooled in 
 water at 82° F. Strips of plate 1| inches wide must bend double in press, inner radius 
 being 1£ times thickness of plate. For pieces of rivet bar inner radius to equal radius of 
 bar ; and for strips from tubes, & inch. 
 
 Board of Trade Tests of Material. — Strips 2 inches wide should be cut from at least 
 one of every four ordinary plates, from each end of each plate over 15 feet in length, and 
 from each corner of each plate over 20 feet x 6 feet, or over 1\ tons in weight ; where 
 more than one test piece is taken from a plate the mean result is to be adopted. 
 
 * The ultimate tenacity in tons plus the percentage of stretch in 8 inches shoulcL 
 oot be less than 52 for boiler steel. For forgings, Lloyd's Register requires this to be 57. 
 
810 
 
 MANUAL OF MARINE ENGINEERING. 
 TABLE XCIII — Boakd of Trade Texsile Tests. 
 
 Description of Material. 
 
 Minimum 
 Ultimate Tensile 
 
 Strength, 
 
 Tons 
 
 per square inch. 
 
 Maximum 
 
 Ultimate Tensile 
 
 Strength, 
 
 Tons 
 
 per square inch. 
 
 Elongation in 10 inches 
 per cent. 
 
 Plates not exposed to \ 
 flame, - - - / 
 
 Plates that are exposed \ 
 to flame, - - - J 
 
 Rivet bars, - 
 
 Stay bars, - 
 
 Tube strips, 
 Rivets, 
 
 27 
 26 
 26 
 
 27 
 
 26 
 27 
 
 32 
 30 
 30 
 
 32 
 
 30 
 32 
 
 20 per cent, in 8 inches. 
 
 f Not less than 23 per cent. 
 \ for annealed plates. 
 
 Not less than 25 per cent. 
 
 ( Not less than 20 per cent. 
 -J and in combustion 
 chambers 18. 
 About 25 per cent., not 
 
 less than 20 per cent. 
 Contraction of area about 
 60 per cent. 
 
 N.B. — Shell plates up to 36 tons tensile may be used by arrangement. 
 
 For bending tests the specimens should be 2 inches wide and 10 inches long, and 
 should be bent double, the inner radius being \\ times the thickness of the plate ; for 
 plates not exposed to flame, the tests should be made on the plate in its ordinary condi- 
 tion, and for plates that are exposed to flame the specimens should be heated to a cherry 
 red, and cooled out in water at 80° F. before bending. 
 
 Lloyd's Tensile Tests of Plates etc. — Steel plates for shells and girders 28 to 32 tons 
 per square inch ; plates for flanging, welding, or exposure to flame 26 to 30 tons. Material 
 for purposes where tensile strength is unimportant bend tests only are required. The 
 elongation of test pieces must not be less than 20 per cent, in 8 inches for material 0-375 inch 
 and over in thickness whose tensile strength is 28 to 32 tons : and not less than 23 per 
 cent, when it is 26 to 30 tons and 0-375 inch and upwards in thickness. 
 
 Stays, angle and tee bars, 28 to 32 tons tensile, with 20 per cent, elongation, while 
 screwed stays are 26 to 30 tons and 23 per cent. Rivet bars, 26 to 30, with 25 
 per cent. 
 
 Material under 0-375 inch thick, the extension may be 3 per cent, below the 
 above. 
 
 Test pieces, not less than \\ ins. wide, from plates or bars to stand bending to a curve, 
 the inner radius not greater than H times thickness, when in normal condition and after 
 having been heated to a low cherry red and quenched in water at 80° F. Steel rivets 
 to be bent cold and hammered until the shank is bent round 'on to itself with- 
 out fracture on outside of bend, and heads flattened hot to 2£ times diameter of 
 shank. 
 
 One tensile, and either one cold or one temper bend test to be made from every 
 plate ; but for plates exceeding 2\ tons in weight, one tensile test from each end, 
 and a cold bend test from one end and a temper bend test from the other are 
 required. 
 
 British Corporation Boiler Steel. — The quality of steel to be used in the construction 
 of boilers must be as follows, viz. : — 
 
 Plates intended for the cylindrical shells and butt straps, and bars intended for 
 stays, are to have an ultimate tensile strength of not less than 28 tons, or more than 
 32 tons per square inch, with at least 20 per cent, extension in 8 inches of prepared section. 
 Plates which have to be flanged or welded, plates and stays for combustion-chambers, 
 plates for furnaces, and bars intended for rivets are to be from 26 to 30 tons per square 
 inch in tensile strength, with not less than 23 per cent, extension in 8 inches, but with 
 rivet bars the extension is to be not less than 25 per cent, in 8 inches, and 31 per cent, in 
 3£ diameters. 
 
 The sample pieces for temper bend tests are to be heated uniformly to a dull red, 
 then cooled in water the temperature of which does not exceed 80° F., and the pieces in 
 
BASIS TBICB FOR BOILER PLATE?. 8] 1 
 
 both cold aud temper bend tests must withstand, without fracture, being Dent double 
 to a curve, the inner radius of which does not exceed i$ times the thickness of the piece. 
 All plates which have been welded, locally heated, or in which rivet holes have been 
 punched, and steel stay bars which have been worked in the fire must be subsequently 
 annealed, and in no case are steel stays to be welded. Slumber of testa same as for 
 Lloyd's, see above. 
 
 John Spencer & Sons, Newburn, and David Colville & Sons, Motherwell, 
 manufacture plates and forgings from patent silicon steels which have an 
 ultimate tensile strength of 40 tons, with an elongation of 30 per cent, in 
 2 inches and a limit of elasticity of 25 tons. Such plates, however, cost 
 a little more to manufacture than the ordinary Siemens steel, but they repay 
 the boilermaker for it. 
 
 Ordinary mild bar steel, such as used for the stays of boilers, for bolts, 
 studs, etc., has a tensile strength of 26 to 32 tons, with an elongation of 25 to 
 20 per cent., and a harder steel for pins, etc., has 35 to 40 tons, with 15 per 
 cent. 
 
 The specific gravity of mild steel plates and bars is about 7 "86 ; the 
 weight of a cubic foot is 491 lbs. ; that of a cubic inch is - 284 lb., and that 
 of a square foot, 1 inch thick, 40*94 lbs. 
 
 Steel Boiler Plates can be obtained now of large size, as follows : — 
 
 Plates \ inch thick, 10 feet wide, or 40 feet long, but area not to exceed 240 sq. ft. 
 
 >> '1 IT >l 
 
 »» % »» 
 
 3 
 
 » t » 
 
 »> B »» 
 
 M ' »» 
 
 »> * 8 »> 
 
 »> -m e i> 
 
 Plates 1£ may be up to 415 square feet, those If to 380, those 1| to 340, and 
 
 those 1| to 300 (J. Spencer & Sons). 
 
 Plates 1| thick may be 12 feet wide or 30 feet long, and the area to 288 sq. ft. 
 „ 2 „ 12 „ 25 „ „ 250 „ 
 
 „ 2\ „ 12 „ 25 „ „ 200 „ 
 
 Steel Flat Bars are rolled from h X | thick to 8 inches X 4 thick. 
 
 Square Bars from | inch to 6 inches, and with round corners to 8 inches. 
 
 Round Bars from £ inch to 8 inches, and from 10 inches to 16-5 inches 
 by Spencer's up to 30 feet long. 
 
 Thin Plates as low as \ inch can be obtained 6 feet wide or 30 feet long, 
 but the area must not exceed 120 square feet. 
 
 Round Plates \ thick may be obtained 6-5 feet diameter ; |-inch plates 
 11 feet diameter ; f-inch plates 12 feet : f-ineh and upwards 13 feet diameter. 
 
 The basis price for boiler plates is about £7 per ton* tested to pass Board 
 of Trade, Lloyd's, or other Registry, but that price does not include very 
 large or very thick ones, and is for the ordinary commercial boiler steel of 
 28 to 32 tons tensile strength. If a minimum of 29 tons is required the extra 
 price is 10s. per ton, if 30 tons it is 20s. Should a high tensile steel be necessary, 
 such as 35 tons per square inch as a minimum, such plates will cost £6 extra 
 to the basis. 
 
 If more than 15 per cent, of the plates are not truly rectangular there 
 is an extra charge for all beyond this quantity of 25s. per ton. 
 
 • In normal times, in 1918 it is much higher. 
 
 10 
 
 9 
 
 50 
 
 11 1 
 
 9 
 
 50 
 
 12-5 
 
 9 
 
 50 
 
 12-5 
 
 9 
 
 50 
 
 12-5 
 
 9 
 
 50 
 
 12-5 
 
 9 
 
 50 
 
 12-5 
 
 9 
 
 50 
 
 99 
 
 >•* 
 
 260 
 
 »> 
 
 » 
 
 300 
 
 M 
 
 M 
 
 325 
 
 *»> 
 
 ft 
 
 35C 
 
 99 
 
 »•» 
 
 430 
 
 •• 
 
 » 
 
 450 
 430 
 
tfl2 MANUAL OF MARINE ENGINEERING. 
 
 Plates will be supplied up to H inch thick at basis price, provided they 
 are not over 8 feet wide ; over that width, there is an extra charge of 2s. 6d. 
 per 3 inches additional width, so that the 12*5 feet wide plates will cost 45s. 
 per ton beyond basis. 
 
 Plates which are over 1| inches to If inches thick will cost 10s. per ton extra. 
 
 1| » 2 „ 30s. 
 
 So that a plate 2 inches thick will cost 60s. extra per ton. If very mild steel 
 is required, so that the maximum limit of tensile strength is 27 tons, a charge 
 of 10s. extra is made, and if as low as 25 tons maximum it is as much as 
 40s. It will be seen, then, that if very large thick plates are used in boiler 
 construction, their cost will be somewhat high ; but, on the other hand, 
 the saving in labour and materials will more than compensate for it, besides 
 permitting of a superior boiler from the users' point of view. 
 
 Steel Forgings. — Shafts, piston- and connecting-rods, valve-rods, gudgeons, 
 etc., are now made of steel forged from ingots manufactured by the Siemens 
 process. The steel is, of course,, generally of a mild kind, and while possessing 
 properties very similar to those of the rolled bars and plates, is not always 
 quite so uniform in structure and strength. 
 
 There is little doubt that the ultimate strength of marine shafts, when 
 made of steel, does not exceed 35 tons on the average, and those over 12 
 inches diameter cannot be depended on for a higher average than 30 tons, 
 however well forged. The material near the centre of a steel forging of 
 large size remains but partially affected by hammering, so that the larger the 
 diameter of the shaft, the less will be the average strength of the material 
 composing it. Forgings as made with a hydraulic press a<re freer from sus- 
 picion, but even they are not good near the axis when of large size. 
 
 Small steel forgings may have as high a tensile strength as bars rolled 
 from similar ingots, but as a rule they have not so great elasticity, and it is 
 safer, therefore, to suppose them to be 10 per cent, weaker, for purposes of 
 calculations, although, as a matter of fact, their tensile is sometimes higher. 
 Lloyd's require the tensile and percentage of elongation to be together 57 . 
 
 Steel Castings. — Many parts of a marine engine which were formerly 
 of forged iron, are now made with advantage of cast steel ; other parts 
 which, for convenience of manufacture, were made of brass, are often of 
 this material. It has superseded cast iron in many parts which must of 
 necessity be cast, and as the cost of production of cast steel is reduced, and 
 the soundness of it improved, so will the demand for it increase. 
 
 The chief obstacle to further employment is usually this uncertainty 
 as to the soundness of the castings. Continued use, however, is dissipating 
 the prejudices which once existed against its application, and the demand for 
 these castings has caused some manufacturers of them to give the closest 
 attention to their production, the day should not be far distant when a steel 
 casting will command the same confidence as to its soundness that now 
 obtains for iron castings. Indeed, an iron casting may be really more treacher- 
 ous than one of steel, because blowholes and spongy places are always near 
 the surface of the latter, and can often be detected, while those in the iron 
 castings are quite hidden. If a steel casting is machined, so that the faults- 
 places are cut away, the part remaining may be depended on as quite soupd. 
 
NICKEL STEEL. 
 
 813 
 
 Propeller bosses, foundations, columns, levers, crossheads for pistor.- 
 rods, pistons of all sizes and shapes, link-motion blocks, eccentric straps, 
 worms, wheels, etc., are now very generally made of cast steel ; also large 
 crank-shafts have been made of cast steel, and did their work very satis- 
 factorily, even when so large as 15 inches diameter ; connecting-rods, also, 
 have been made of this material, and the economy of production of such 
 parts when thus made in quantity is beyond all doubt. 
 
 The probable average strength ■'•of best steel castings is about 28 tons per 
 square inch, with an elasticity of about 25 per cent, in 2 inches ; greater 
 strength is easily obtained with steel castings, but the elasticity will then 
 be much lower, in fact, it is only by using very carefully selected materials, 
 and by annealing after casting, that so low a strength can be obtained. The 
 ordinary steel castings have an average ultimate strength of about 32 tons, 
 with an elongation of about 18 per cent, when sound. To allow for un- 
 soundness the ultimate strength may be assumed to be only 28 tons for the 
 harder varieties, and 25 tons for the softer ones ; for propeller blades it 
 should not be assumed to be higher than 24 tons, although the sound parts of 
 such castings are often found to have a strength of over 30 tons per square inch. 
 
 Nickel Steel. — This material, which is an alloy of steel and nickel, or, to 
 be more precise, of iron, carbon, manganese, and nickel, is often used where 
 lightness is of first importance in engineering structures, or when the strongest 
 possible material is required. Great care, however, has to be taken when 
 forging and otherwise treating this material in the hot state to avoid seriously 
 injuring its physical properties. It is, therefore, better to have the forging, 
 etc., delivered from the maker's works ready for machining, and in all cases 
 it should be most carefully annealed. It is also advisable to have it oil- 
 tempered as well as annealed to get the full advantages from this splendid 
 alloy. Sir Wm. Beardmore read a very interesting paper to the Institution 
 of Naval Architects, in 1897, on the nickel-steel alloys his firm was making 
 for boiler plates ; with an ultimate strength of 74,000 lbs., and an extension 
 of 23 per cent., the elastic limit was 43,000 lbs. ; and a stronger quality 
 whose ultimate strength was 114,000 lbs., with 25 per cent, extension, the 
 elastic limit was 64,000 lbs. Forgings on a large scale have been made in 
 this country by Messrs. Vickers, Sons & Maxim, Beardmore, and others for 
 marine work as well as armour plates ; while in America, at the Bethlehem 
 Iron Works, this material is very largely used for marine work. Steel having 
 0-25 to 0-45 of carbon seems best for this alloy, and for marine purposes 
 3 to 5 per cent, of nickel is used with it. Less than 3 per cent, nickel is of 
 no advantage. * 
 
 Nickel steel as made in Sheffield and in Germany is generally as 
 follows : — 
 
 Place of Manufacture. 
 
 Carbon. 
 
 Silicon. 
 
 Manganese. 
 
 Nickel. 
 
 Sheffield, .... 
 Germany, .... 
 
 280 
 0310 
 
 0123 
 0112 
 
 0516 
 625 
 
 2 95 
 4-18 
 
 Forgings having 0'29 per cent, of carbon and 3-25 of nickel have an ulti- 
 mate strength of 85,000 lbs., with 30 per cent, elongation, and an elastic 
 
 • Chrome nickel steel is now largely used for light forgings where great strength is necessary. 
 
 52 
 
814 
 
 MANUAL OF MARINE ENGINEERING. 
 
 limit of 54,000 lbs. ; with 3*75 per cent, of nickel and oil tempered, 90,000 
 lbs., with 22 per cent, extension, and 60,000 lbs. elastic limit ; with 4'"5 to 
 5 per cent, the elastic limit is as high as 80,000 lbs., ar<d the extensior 20 per 
 cent, after careful oil tempering and annealing. The coefficient of expansion 
 of an alloy containing 25 to 30 per cent, of nickel is exceedingly low, and its 
 electrical resistance is 40 times that of copper ; it is also nearly incorrodible.* 
 Adding 3 to 3| per cent, of nickel increases the cost of a steel by about 7s. 
 per cwt. This material stands percussion shocks exceedingly well, and also 
 resists vibratory stresses equally well. Iron bars stressed to 40,000 lbs. 
 stood only 59,000 alternations, mild steel 170,000, high carbon (0 - 35 per cent.) 
 317,000, while nickel steel (0-25 C, 3-25 Ni) stood 1,850,000, and 5| per cent, 
 nickel as many as 4,370,000 alternations. 
 
 Chrome Vanadium Steel is now extensively used for engineering purposes 
 on shore and to some extent on board ship, where its high qualities make it 
 a most useful addition to the means for effecting special ends where price is 
 of secondary importance. The addition of quite small quantities of these 
 metals, either singly or together, to the ordinary carbon manganese steels of 
 commerce improves them immensely, as may be seen by examining the 
 following table : — 
 
 
 Elastic 
 Limit. 
 
 Ultimate 
 Tensile 
 
 Stress. 
 
 Elongation 
 
 in 
 2 inches. 
 
 Reduction 
 of Area. 
 
 
 
 
 Tons per 
 
 Tons per 
 
 Per cent. 
 
 Per cent. 
 
 
 
 
 square inch. 
 
 square inch. 
 
 
 
 Carbon manganese steel 
 
 crucible, 
 
 160 
 
 27-0 
 
 35 
 
 60-0 
 
 »> 
 
 ii 
 
 + 0'5% chromium, 
 
 22 9 
 
 34 
 
 33 
 
 60 6 
 
 >> 
 
 •t 
 
 + 1-0% 
 
 25 
 
 38-2 
 
 30-0 
 
 " 57-3 
 
 »» 
 
 IS 
 
 + 0*10% vanadium, 
 
 2S-5 
 
 34-8 
 
 31-0 
 
 60-0 
 
 It 
 
 »» 
 
 + 0-15% „ 
 
 30 4 
 
 36o 
 
 26-0 
 
 59-0 
 
 It 
 
 It 
 
 + 0-25% „ 
 
 341 
 
 39-3 
 
 24-0 
 
 59-0 
 
 J» 
 
 t« 
 
 +- 1*0% chromium 
 4- 0"15% vanadium, 
 
 | 36 2 
 
 48-6 
 
 24 
 
 56 6 
 
 »» 
 
 •1 
 
 + 1 "0% chromium 
 + 0*25% vanadium, 
 
 ) 49 4 
 
 i 
 
 60-4 
 
 18-5 
 
 46 3 
 
 Open-hearth 
 
 »1 
 
 plain, 
 
 17-7 
 
 322 
 
 34 
 
 52 6 
 
 tt 
 
 ») 
 
 f 1 0% chromium 
 + , 15% vanadium, 
 
 ) 34-42 
 
 52 6 
 
 25-0 
 
 55 5 
 
 Under torsion some bars of vanadium chrome steel f inch diameter and 6 inches long 
 had an ultimate shearing resistance of 40*6 tons, with an elastic limit of 17 '4 tons, and 
 the final angle through which they were twisted was 1,623 degrees or 4 - 5I turns. 
 
 Some samples of this material containing the following, viz. : — 
 
 Carbcn, 
 
 Silicon, 
 
 Manganese, 
 
 Chromium, 
 
 Vanadium, 
 
 0*297 per cent. 
 
 0-059 
 
 0-394 
 
 1-066 
 
 0-169 
 
 were subject to a series of most exhaustive tests by Captain Sankey and 
 
 * An incorrodible chrome steel is now made having an ultimate tensile strength of about 60 tons, with 
 a yield point of about 50 tons and elongation 18 to 20 per cent. 
 
COPPEE. 
 
 815 
 
 Mr. Kent Smith, of which the following is an interesting extract, as showing 
 how it is affected by heat treatment : — 
 
 Condition. 
 
 Yield 
 Point. 
 
 Ultimate 
 Stress. 
 
 1 
 
 Ratio Yield 
 to Net. 
 
 Elongation 
 
 in 
 
 2 inches. 
 
 Reduction 
 of Area. 
 
 Number of 
 Alterna- 
 tions. 
 
 The steel as delivered 
 
 
 
 
 
 
 
 from the rolls, . 
 
 36-9 
 
 54 1 
 
 0-68 
 
 24 I 
 
 44-9 
 
 1,906 
 
 After annealing for half- 
 
 
 
 
 j 
 
 
 
 ' hour at 932° F.,. 
 
 36 9 
 
 538 
 
 06S 
 
 24 7 
 
 48-9 
 
 ... 
 
 After being water-quen- 
 
 
 
 
 
 
 
 ched from 932° F. , . 
 
 356 
 
 53-5 
 
 0-67 
 
 25 6 
 
 507 
 
 ... 
 
 After being oil-quenched 
 
 
 
 
 
 
 
 from 932° F., . 
 
 35-8 
 
 54-1 
 
 0-67 
 
 26 
 
 49 '6 
 
 
 After being water quen- 
 
 
 
 
 
 
 
 ched from 1472° F., . 
 
 60 3 
 
 74 6 
 
 0-81 
 
 7-5 
 
 166 
 
 174 
 
 After being oil-quenched 
 
 
 
 
 
 
 
 from 1472° F., . 
 
 37 
 
 54-5 
 
 0'6S 
 
 22 
 
 35-2 
 
 296 
 
 After annealing for half- 
 
 
 
 
 
 
 
 hour at 1472° F., 
 
 21 1 
 
 39 '0 
 
 0-54 
 
 34 5 
 
 53-1 
 
 2,237 
 
 After being oil-quenched 
 
 
 
 
 
 
 
 at 1598° F., and re- 
 
 
 
 
 
 
 
 heated to 662° F., 
 
 49 -8 
 
 59 3 
 
 0-85 
 
 23 
 
 50 8 
 
 1,314 
 
 After being water-quen- 
 
 
 
 
 
 
 
 ched at 1652° F., 
 
 56-6 
 
 78-0 
 
 072 
 
 10 
 
 21o 
 
 ... 
 
 After being oil-quenched 
 
 
 
 
 
 
 
 at 1652° F., 
 
 47 9 
 
 60-9 
 
 0-79 
 
 18-5 
 
 39-1 
 
 
 After being oil-quenched 
 
 
 
 
 
 
 
 at 1652° F., and re- 
 
 
 
 
 
 
 
 heated to 1112° F., . 
 
 43 4 
 
 53 6 
 
 0-81 
 
 22 
 
 56-6 
 
 1,938 
 
 Soft Steel for Solid-drawn Tubes is usually of Swedish make or made 
 elsewhere of Swedish ores, which are remarkably pure and free from phos- 
 phorus and sulphur. Such steel usually contains about 0-127 per cent, of 
 carbon, O016 of silicon and 0"332 of manganese. It is always ductile 
 and soft, so that it can be drawn down cold to almost any degree of 
 thinness required, while being free from any blemish, or even mark, that 
 would raise a reasonable doubt as to its soundness. 
 
 Manganese Steel. — The invention of Sir Robert Hadfield is of interest, 
 although it is not so generally useful to the engineer as is to be desired. That 
 best known is an alloy containing 11 to 14 per cent, of manganese, and is 
 largely used for the wearing parts of dredgers, crushing machinery, and ash 
 ejectors. It is very tough and very hard, so that, while it can be forged 
 easily and bent readily when cold, it cannot be machined. In the forged 
 condition it has a tenacity of 60 tons per square inch, with an elongation 
 of 35 to 40 per cent, in 8 inches. At present articles are forged to shape or 
 ground, but it is hoped that soon some method of machining this valuable 
 material will be found. 
 
 Copper. — This metal is used to form alloys more frequently than 
 employed by itself ; it is too soft for general purposes ; but as it is so 
 much improved by the addition of even small quantities of other metals, 
 it is perhaps, next to iron, the most important of the metals used by the 
 marine engineer. 
 
816 MANUAL OF MARINE ENGINEERING. 
 
 In commerce there are three distinct qualities of ingot copper dealt with, 
 viz. : — (1) G.M.B. (good merchantable brand), a general term used as a basis 
 for price ; (2) Tough Ingot, which is generally used by founders for general 
 purposes, and branded with some well-known smelter's name or mark ; (3) 
 Best Selected, is purer and of more uniform quality, and used for high class 
 work. Electrolitic is the purest form, however, for highly important work. 
 
 In its simple state it is employed chiefly for pipes on account of its ductility 
 and strength, and in some measure because it can be joined by brazing, so as 
 to be as strong there as the original sheet, and for plates or sheets with which 
 to make the fire-boxes of the locomotive boiler. It does not generally corrode 
 under the action of sea-water or air, but does sometimes waste by the 
 mechanical action of water and steam moving at high velocities over its 
 surface. In some few localities the water seems to have a destructive effect 
 on this metal, owing no doubt to the presence of free gases mechanically 
 mixed with it ; of these gases, sulphuretted hydrogen is the worst in the 
 water of rivers and ports, and chlorine in sea-water ; the presence of bromine, 
 too, is often detected by the smell so often noticeable in the hot-wells of 
 some engines. 
 
 Copper pipes of large size (6 inches and upwards) are always made from 
 sheets, curved into the required form by rolling or hammering, and brazed 
 at the seams. Copper pipes can be made seamless, practically of any size, by 
 electro-deposits, etc., on Cowper-Coles method of treatment, but steel is now 
 superseding this metal. 
 
 Smaller pipes are sometimes made in the same way, but generally by 
 " drawing." The feed, blow-off, and scum pipes in the Navy are always 
 of solid-drawn copper, because there should be no seam in such important 
 pipes ; and the main steam pipes were also solid drawn when made of 
 copper, and not exceeding 6 inches diameter, but now they are of solid- 
 drawn steel. 
 
 Solid-drawn pipes are very seldom used in the mercantile marine, partly 
 because they are somewhat more expensive, but chiefly because they were 
 not so uniform in thickness as the brazed ones, and thought to be 
 more liable to split unless carefully manufactured from very soft tough 
 copper. 
 
 The strength of copper depends somewhat on its purity, but principally 
 on the amount of work it has undergone, especially in the cold state. Copper 
 castings have an ultimate strength of only about 10 tons per square inch ; 
 when forged its strength is increased to about 15 tons, and when rolled into 
 bars is 16 tons ; if a small proportion of phosphorus is added (about 2 per 
 cent.) the strength is increased to 20 tons ; pure copper when drawn out 
 into wire has a strength of about 28 tons before, and 18 tons after annealing. 
 Sheet copper has an average strength of about 13| tons, and for purposes of 
 calculation may be assumed to be 30,000 lbs. per square inch. Small quanti- 
 ties of arsenic and some other metals improve copper very much. With 
 only h per cent, of arsenic plates have an ultimate strength of 33,420 lbs., 
 with a stretch of 37 per cent. (Roberts- Austen). Wire made of copper with 
 1 per cent, of aluminium stood 78,000 lbs. per square inch, and French wire 
 made of copper with a small addition of silicon 128,000 lbs. Wire with 
 0-529 per cent, of antimony stood 78,000 lbs. as against 49.000 lbs. with 
 pure copper wire. 
 
ZINC OR SPELTER. 817 
 
 The specific gravity of sheet copper is 8 - 805 ; the weight of a cubic foot 
 is 550 lbs., that of a cubic inch 0*318 of a pound, and that of a square foot 
 1 inch thick 45-83 lbs. 
 
 Tin. — This metal, although seldom used alone, is very important, as 
 forming one of the chief constituents of bronze or gun-metal, but now more so 
 as being the chief ingredient of some white metals. The best qualities are 
 obtained from Cornwall, and the chief supply of this metal was from that 
 country ; of late years, however, considerable quantities have been imported 
 from the Dutch East Indies, Malacca, and Australia, and although not so 
 pure as the Cornish tin, the price of the latter has been very considerably 
 affected by the supply. 
 
 Tin is used as a protective covering to other metals on account of its 
 immunity from the corrosive action of salts and acids. The Admiralty 
 formerly required all condenser tubes to be coated with tin when fitted in 
 iron condensers ; and this practice was also followed by some Mercantile 
 Shipping Companies. Since sea-water ceased to be used as feed supply the 
 tinning of condenser tubes has gone out of practice. Sheet tin, which is 
 thin sheet iron or steel coated with tin, is used for " liners " between 
 " brasses," as well as for making oil feeders, lamps, cups, etc. 
 
 Tin mixed with small quantities of copper, antimony, etc., is used under 
 the name of white metal to line and face bearings. 
 
 The tensile strength of tin is too low to admit of its being used alone in 
 construction ; its ultimate strength when cast is only 2-11 tons per square 
 inch. Its specific gravity is about 7*3, consequently the weight of a cubic 
 foot is 456 lbs., and that of a cubic inch 0-264 lb. 
 
 Zinc or Spelter. — This metal also is seldom used alone, but is very largely 
 employed to alloy with copper to form brass, and with it and some other 
 metals in small quantities the well-known strong zinc bronzes. The best 
 kinds come from Australia and the Continent ; the Silesian spelter is the 
 purest, and generally used in making brass for rolling into sheets or drawing 
 into tubes and rods. Ordinary zinc contains lead in appreciable quantities. 
 Electrolitic is quite pure, and is best for high-class alloys. 
 
 1 In its simple state zinc is used by marine engineers to prevent corrosion 
 in boilers, condensers, and hot-wells, and to protect the ship's plating near 
 the bronze propellers. Cast blocks, or, better still, pieces of rolled bar or 
 sheet of this metal are placed in metallic contact with the iron of the boiler 
 in such places as have been found by experience to require protection. The 
 purer the zinc is, the more perfect is the protection afforded ; but unless 
 there are exceptional circumstances affecting the feed -water, common zinc 
 or even " hard spelter " (residuum from the galvanising bath) will form 
 a sufficiently strong galvanic couple to prevent deterioration of the iron 
 surfaces. 
 
 Zinc is also employed as a covering for iron or steel, to protect it from the 
 action of sea-water, etc., and being much cheaper than tin, and easily applied 
 to the surface of the iron, is used on a far more extended scale than tin. Zinc 
 is also used as the principal constituent of certain kinds of white metal 
 made for bushes working in water. 
 
 The tensile strength of zinc is even lower than that of tin, being only 
 1-336 tons per square inch when cast. Its specific gravity is 7 0, conse- 
 quently the weight of a cubic foot is 437 lbs., and that of a cubic inch 0'253 lb. 
 
818 MANUAL OF MARINE ENGINEERING. 
 
 Lead is nearly always used alone, and the purer it id, the more valuable 
 for engineering purposes. In the mercantile marine the bilge piping is 
 often made of lead, and the pipes for emptying and tilling the ballast tanks 
 are also sometimes of lead. It is used for these purposes because of its 
 resisting the corrosive action of sea and bilge water, and being very 
 ductile it can be easily bent to follow the curves and corners of the ship'3 
 bottom. 
 
 Sheet lead is sometimes used to protect the covering of boilers from wet, 
 and also to cover engine-room floors when made of wood, as it gives a better 
 foot-hold than iron plates. It is employed, too, for jointing pipes, etc., 
 when the flanges are rough and uneven. Of late years it has been largely 
 used as the chief constituent of certain white bearing metals, hardened with 
 antimony. 
 
 It is sometimes used by moulders to give a good colour to common brass 
 for ornamental purposes, and to cause it to be readily turned in the lathe ; 
 but even the smallest addition of this metal tends to reduce the strength of 
 brass, and it should be, therefore, generally avoided. 
 
 The tensile strength of sheet lead is only 0-81 ton per square inch, and 
 that of lead pipe is 1 ton. The specific gravity is 11*418, consequently the 
 weight of a cubic foot is 712 lbs., of a cubic inch 0*412 lb., and of a square 
 foot 1 inch thick, 59 *3 lbs. 
 
 Aluminium.* — This metal, formerly so rare and very expensive, is doubt- 
 less destined to become a very important one to marine engineers from its 
 extreme lightness as it has to the automobile engineer. Its price when 
 introduced (about £1,000 per ton) quite precluded its use, even to form 
 an alloy ; in normal times, when it can be purchased at £80, engineers can 
 use it freely to make bronze, which is as strong and elastic as mild steel. 
 Aluminium, with the addition of 6 per cent, of copper, rolled into sheets 
 and bars having a fair strength is still light, and cost only Is. per lb. 
 Such sheets have an ultimate strength of 11 to 12 tons annealed, and 14 to 
 16 tons unannealed, as against 7 to 8 tons in the case of sheets of pure 
 aluminium. Aluminium bronzes in the cast state are very porous, and the 
 crystalline structure very coarse. A little aluminium improves all zinc 
 bronzes, but damages tin bronzes. 
 
 In the pure state when drawn into wire, it has a tensile strength of about 
 8 tons per square inch ; by hammering cold, the strength is raised from 
 about 7 tons (as cast) to about 12 tons. An alloy of 90 per cent, of copper 
 and 10 per cent, of aluminium, when rolled, has an ultimate tensile strength 
 of 32 to 40 tons per square inch, and Professor Arnold found that an alloy 
 of 92-5 of copper and 7*35 of aluminium registered the highest resistance to 
 alternating stresses he had ever observed — viz., 1,359, as against 657 of the 
 9-9 per cent, mixture. 
 
 The specific gravity 'of aluminium is only 2-56, consequently a cubic foot 
 weighs 160 lbs., a cubic inch 0-092 lb., and a square foot 1 inch thick, 13-33 lbs. 
 
 Duralumin is a good alloy made and sold by Vickers, Ltd., at such a 
 moderate cost as ensured its general use in engineering practice, considering 
 its great strength and lightness. It consists of about 90 per cent, of 
 aluminium, and has a specific gravity of 2-8, and, therefore, very little 
 heavier than pure aluminium. Its melting point is about 1,200 J F. 
 ♦The standard price in 1918 is £230 per ton. 
 
BRASS TUBE METAL. 819 
 
 Its tensile strength may be* as high as 40 tons per square inch, with an 
 extension, however, somewhat small ; but at 28 to 30 tons it is quite satis- 
 factory, being as much as 15 per cent, in 2 inches, and with a tensile strength 
 of 25 tons 20 per cent, elongation is obtained. It is not used in the cast state, 
 but was supplied in sheets from 24 to 28 inches wide, and up to 10 S.W.G. 
 thick at 2s. 4d. per lb., which corresponds to 9d. for an equal quantity (volume) 
 of copper or brass, or to lOd. for steel. It was also rolled exceedingly thin 
 in narrower widths for 2s. per pound. Hot-rolled bars from 1^ to 3 inches 
 diameter cost only 2s. Id., and was supplied down to § inch diameter at 
 2s. 4d. per pound. During the war the price is more than doubled. 
 
 Cold-drawn bars were only lid. per pound dearer. This metal can be 
 worked hot or cold, and, therefore, made into tubes, sectional bars, etc., 
 stampings, and forgings. It is non-magnetic, is little affected by sea or fresh 
 water or damp air, and takes a good polish. 
 
 Antimony is used in very small quantities, to harden other metals and 
 alloys, especially the white metals for bearings. 
 
 Alloys. — Strictly speaking, alloys of copper and zinc only can be called 
 brass, but ordinary bronze, as made for bearings, liners, and bushes, is often 
 called brass, and from these circumstances the liners of journals and pin- 
 bearings are called " brasses." Alloys of copper and tin, or those of copper 
 and tin together with zinc or other metal, are called bronze. Alloys of copper 
 and zinc with manganese or other modifier are now generally called zinc 
 bronzes. 
 
 Brass. — The yellow brass used for ornamental castings is usually composed 
 of two parts of copper and one part of zinc ; when carefully made, castings 
 of yellow brass have a tensile strength of 12 to 13 tons, but the ordinary 
 yellow brass, as supplied to founders, has a strength of only 10 to 11 tons; 
 it is fairly tough, but too soft for general purposes. 
 
 Muntz's Metal, composed of three parts of copper and two of zinc, can be 
 rolled out into bars and sheets, so as to have an average tensile strength of 
 22 tons per square inch, and in some cases bars of this metal have a strength 
 as high as 27 tons. It is very ductile, and can be forged when hot ; it will 
 stretch very considerably before fracture, and may be used for springs when 
 hammered or cold rolled, and not annealed. 
 
 Its specific gravity is 8*2, consequently a cubic foot weighs 512 lbs., 
 that of a cubic inch 0'296 lb., and that of a square foot 1 inch thick is 
 42-7 lbs. 
 
 Naval Brass. — By the addition of 1 per cent, of tin to Muntz's metah it 
 has a better power of resisting the action oi sea-water, while retaining all its 
 other properties. An alloy of 62 copper, 37 zinc, and 1 tin is known as naval 
 brass, because of its use originally in the construction of naval composite 
 ships, and for the bolts of the engine-fittings which are exposed to sea-water 
 in warships. This metal can be forged hot, and bent cold two double ; its 
 strength is superior to the ordinary Muntz's metal, and some specimens 
 rolled cold and unannealed have been proved to have an ultimate strength of 
 nearly 40 tons per square inch. As usually supplied, it has an ultimate 
 strength of 27 tons, with an elongation of 19 to 20 per cent, in inches ; the 
 elastic limit is about 19 tons. 
 
 Brass Tube Metal. — The ordinary brass condenser-tubes are made of a 
 composition containing 70 per cent, of copper and 30 per cent, of zinc, but 
 
820 MANUAL OF MARINE ENGINEERING. 
 
 the Admiralty require thern to be composed- of 70 per cent, of best selected 
 copper, 29 per cent, of Silesian zinc, and 1 per cent, of tin,* and all tubes 
 supplied to the Navy have to undergo the test "described on p. 363. The 
 strength of the metal of tubes made with 70 per cent, of best selected copper 
 and 30 per cent. Silesian zinc is as high as 36 tons per square inch. 
 
 Gun-metal or Bronze. — There is no particular mixture to which this name 
 belongs, as it is applied promiscuously to any composition of copper and 
 tin, or copper, tin, and zinc. 
 
 The best known composition, and one which has high strength, is fairly 
 hard, and very tough, is that containing 90 per cent, of copper and 10 per 
 cent, of tin. Its tensile strength, when carefully made, is 17 tons per square 
 inch ; its specific gravity is 8*66, consequently a cubic foot weighs 561 lbs., 
 a cubic inch 0*325 lb., and. a square foot 1 inch thick 46 - 8 lbs. To insure sound 
 castings, however, it is necessary to add a small quantity of zinc. 
 
 A much harder metal is made by mixing 84 per cent, of copper with 
 
 16 per cent, of tin ; its tensile strength is 16 tons, specific gravity 8 "56, and 
 the weight of a cubic foot is 534 lbs. 
 
 For heavy bearings, where hardness is of more importance than strength, 
 although the metal must not by any means lack strength, a good metal is 
 made by mixing 79 per cent, of copper with 21 per cent, of tin : its tensile 
 strength is nearly 14 tons when carefully made, and the average is 13 J tons ; 
 the specific gravity is 8 '73, and the weight of a cubic foot is 544 lbs. 
 
 Admiralty Bronze of 88 copper, 10 tin, and 2 zinc best quality has an 
 ultimate tensile strength of 20 tons, with an elongation in 2 inches of 25 per 
 cent. ; when of good commercial metals test bars should give 18*5 tons with 
 20 per cent., and castings stand at least 16 tons with 15 per cent. A little 
 phosphorus in the tin or copper improves the strength generally. With 05 of 
 lead added, the tensile strength at temperatures up to 550° Fah. is maintained. 
 
 Phosphor-Bronze. — This metal is composed of copper and tin, with a small 
 proportion of phosphorus. It is harder than the ordinary bronze, very close 
 grained, and of superior strength. The average ultimate strength is about 
 
 17 tons per square inch, while that of some specimens is as high as 22 tons. 
 The singular thing about this metal is its high elastic limit, being very nearly 
 that of the ultimate strength. Its resistance to shearing in the cast state is 
 also exceptionally high. Great care is required in melting, and repeated 
 meltings very much reduce its virtue. It may be rolled out into extremely thin 
 sheets, or drawn into wire, when the average tensile strength is 56 tons per square 
 inch. Phosphor bronze sheet is used for the valves of air-pumps (vidzp. 393). 
 
 This metal is also used for bearings, brasses, propeller-blades and bosses, 
 pump-rods, etc. 
 
 The Original Manganese Bronze is now seldom or never used. It contained 
 considerable quantities of manganese, had a dull brown fracture, but a high 
 tensile strength. The present alloy is good zinc bronze improved by the 
 addition of ferro-manganese. The manganese is said to deoxidise any copper 
 oxides which may be mechanically mixed with the copper, so " rendering the 
 metal more dense and homogeneous." The No. 1 quality, which was used 
 for forgings and rolling into rods, plates, sheets, angles, etc., when cast in 
 metal moulds had an ultimate strength of 24 tons, and an elastic limit of 
 14 tons per square inch. 
 
 * Some makers use 2 per cent, of lead instead of the tin, and claim as good or better results. Bengal 
 metal, which is a specially made 70/30 mixture, resists corrosion better still. 
 
TURBADIUM. 321 
 
 Rolled rods, plates, etc., as now made by the Company, have, when mild, 
 an ultimate strength of 28 tons, and an elongation of 40 per cent. ; but when 
 so required it can be made with an ultimate strength of 30 to 32 tons, an 
 elastic limit of 15 to 17 tons, and an ultimate elongation of 15 to 20 
 per cent. 
 
 By cold rolling the strength can be raised to even 40 tons, but the elonga- 
 tion is then reduced to 10 per cent. 
 
 The Manganese Bronze Company manufacture various other qualities 
 of bronze for special purposes, such as follows : — 
 
 Crotarite is designed for such things as boiler stays and other fittings as 
 are exposed to heat and required to sustain heavy stresses. This alloy has 
 a melting point almost as high as copper and a coefficient of expansion no 
 greater. It contains no zinc, is highly malleable, and can be riveted cold. 
 Its ultimate tensile strength is 25 to 26 tons per square inch, with an elonga- 
 tion in 2 inches of 30 to 40 per cent. Its elastic limit is high, being no less 
 than 16 tons. 
 
 Immadium is another special alloy suitable for shafts, spindles, and 
 rods of pumps of all Kinds, inasmuch as it is practically incorrodible in sea- 
 water, and is even unaffected by water containing small quantities of acids, 
 while its tensile strength is exceedingly high. The strongest kind has an 
 ultimate strength of 40 tons per square inch, with an elongation of 20 to 
 25 per cent, in 2 inches, and an elastic limit of 20 tons per square inch. The 
 mild variety has an elastic limit of 18 tons, and an ultimate tensile strength 
 of 36 tons per square inch, with an elongation of 25 to 30 per cent. 
 
 This alloy is valuable for purposes such as boiler mountings and fittings, 
 cylinder fittings, and connections, inasmuch as it maintains a high tensile 
 strength when heated to quite high temperatures, for at 500° the elastic limit 
 is 9 to 10 tons per square inch, while the ultimate tensile strength is 23 - 5 tons, 
 with an elongation of 22*5 per cent., while at 420° F. the temperature of steam 
 of 315 lbs. pressure absolute the elastic limit is as high as 8 J tons, and the 
 ultimate tensile strength 24 "5 tons per square inch, with an elongation of 
 25 per cent, in 2 inches. Castings of this material are usually sound and 
 easily machined, so that they may be substituted for steel ones with 
 advantage, seeing that they are even stronger than steel and no more costly, 
 if risk of wasters and cost of machinery are taken into account. 
 
 Turbadium is a good alloy patented by the Parsons Manganese Bronze 
 Company, and supplied for the casting of propeller blades. It is a zinc 
 bronze consisting of about 55 per cent, of copper, 43 of zinc, 1'44 of iron, 
 and small quantities of manganese and tin. The important point, however, 
 is that the structure of this alloy as made is essentially /9 or polyhedric in 
 character, so as to withstand erosion from which the surface of all bronze 
 propellers suffer when run at very high rates of revolution as in turbine 
 steamers. Propellers of this material have been supplied by the Company 
 to the steamships " Lusitania," " Mauritania," etc., with most satisfactory 
 results. Fig. 178 is a photo of one of those fitted to the " Mauritania," 
 taken after many months of service in the Atlantic ; it shows no signs of 
 erosion, as did those of other bronze almost immediately after going on 
 service. 
 
 The elastic limit of this material is 18 tons, and the ultimate tensile 
 strength 40 tons with an elongation of 18-58 per cent, in 8 inches. It can 
 
822 MANUAL OP MARINE ENGINEERING. 
 
 be treated by press or hammer at a cherry-rea heat, so that all buckles and 
 bends due to striking quays or wreckage can be removed and the blade 
 restored to proper shape and pitch quite easily. 
 
 Stone's Bronze is a zinc and copper composition, possessing all the charac- 
 teristics of modern manganese bronze, and is used very extensively for pro- 
 peller blades and engine castings, where high tensile strength and great 
 toughness is necessary. 
 
 Bull's Metal is another of the zinc bronzes with either nickel or ferro- 
 manganese added, having a high tensile strength with great toughness ; 
 in the rolled state it has resistance to compression as high as 100 +ons per 
 square inch. 
 
 Melloid is a tin bronze with a small percentage of phosphorus. It is 
 made by Bull's Company, and can be rolled and forged hot, and rolled 
 cold. 
 
 Delta Metal. — The invention of Mr. Dick is a high-class bronze, somewhat 
 similar to manganese bronze, but rather harder and with a higher tensile 
 strength. Castings made from it have an ultimate strength of 34 tons, and 
 resist torsion better than the other zinc bronzes, whicTi are, as a rule, disap- 
 pointing in this respect. 
 
 Parsons' White Brass. — This is a most valuable material for facing and 
 lining x bearings ; the composition of No. 2 quality is tin, 68; zinc, 30 - 5; 
 copper, 1 ; and lead, - 5 ; its success as a bearing metal is most unqualified, 
 and it does well in crank-pin brasses at moderate speeds. 
 
 Babbit's White Metal. — This was, for very many years, almost the only 
 white metal used for bearings, and until modern times was without a rival. 
 It is composed of 10 parts of tin, 1 of copper, and 1 of antimony. 
 
 Admiralty White Metal. — A very good white metal is made by mixing 
 6 parts of tin with 1 of copper, 6 parts of tin with 1 of antimony, and adding 
 the two mixtures together, and is used in all Admiralty work. 
 
 Plumtine is a comparatively cheap and very efficient white metal, 
 composed of 48 - 5 per cent, of lead, 40*5 of tin, 11 of antimony, and 0*5 
 of copper. 
 
 Magnolia is also a white metal consisting largely of lead, and is well known 
 as an excellent bearing metal. 
 
 Fenton's White Metal, which is used for stern bushes and the bushes of 
 paddle-wheels, etc., is composed of 8 parts of zinc, 1"66 of tin, and 0'44 of 
 copper ; it is fairly tough and hard, and in sandy water resists wear exceed- 
 ingly well. 
 
 Stone's White Bronze is also an excellent metal for bearings and crank- 
 pin brasses, especially of heavy engines running at high speea. 
 
 German Silver. — An alloy of nickel and copper and zinc, known by this 
 name, is used largely for instruments, and for domestic implements after 
 being electro-plated. It is of yellowish- white appearance, fairly hard and 
 tough, and capable of being worked cold. 
 
 Richard's Plastic Metal. — This is a white metal, which melts at a moder- 
 ately low temperature, and can be worked with a soldering-iron. It is very 
 useful, therefore, for mending bearings and for coating damaged brasses ; 
 it also does for a filling metal, working well in bearings and guides. 
 
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COMPOSITION OF WHITE METALS. 
 
 825 
 
 TABLE XCV. — Composition of White (Bearing) Metals. 
 
 Name of Metal. 
 
 Cu. 
 
 Sn. 
 
 Sb. 
 
 Zn. 
 
 Pb. 
 
 Fe. 
 
 Miscel- 
 laneous. 
 
 Dewrance's Locomotive, . 
 
 22 2 
 
 33 3 
 
 44 4 
 
 
 ... 
 
 
 ... 
 
 For gland packings, . 
 
 7-8 
 
 88-1 
 
 3 5 
 
 ... 
 
 2 6 
 
 
 ... 
 
 Used in the German navy, 
 
 7 5 
 
 85-0 
 
 7-5 
 
 
 
 
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 7 
 
 7-5 
 
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 78-5 
 
 7-0 
 
 
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 55 
 
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 8-5 
 
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 Fenton's, 
 
 4-4 
 
 166 
 
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 790 
 
 ... 
 
 
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 Magnolia, 
 
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 21-0 
 
 
 78-0 
 
 10 
 
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 46 
 
 130 
 
 
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 Kingston's, .... 
 
 6 
 
 88-0 
 
 
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 Parsons' white brass, 
 
 5-6 
 
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 0-8 
 
 761 
 
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 58-5 
 
 2 
 
 39-5 
 
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 68 
 
 
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 .100 
 
 
 100 
 
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 80-0 
 
 
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 , , heavily loaded bearings, 
 
 64-0 
 
 5-0 
 
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 ... 
 
 30 
 
 
 lNi 
 
 Plumtine, ..... 
 
 05 
 
 40-5 
 
 11-0 
 
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 48-5 
 
 
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826 
 
 MANUAL OF MARINE ENGINEERING. 
 
 
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PROPERTIES OF METALS. 
 
 627 
 
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828 
 
 MANUAL OF MARINE ENGINEERING. 
 
 TABLE XCVII— Prices of Materials.* 
 
 Materials at Maker's Works. 
 
 
 Prices. 
 
 1912. 
 
 
 Prices, 
 
 1906. 
 
 Pig iron, Blaenavon, cold blast, No. 3, . .• . 
 
 None 
 
 
 
 £9 
 
 2 
 
 6 per ton.. 
 
 ,, Shropshire. ,, ,, , 
 
 
 . £6 
 
 
 
 
 5 
 
 12 
 
 6 
 
 ,, Hiematite. West Coast, „ . 
 
 
 3 13 
 
 6 
 
 
 3 
 
 5 
 
 „ 
 
 „ ,, East Coast, „ . 
 
 
 3 11 
 
 6 
 
 
 3 
 
 8 
 
 6 , 
 
 „ Scotch, good brand, ,, . 
 
 
 3 6 
 
 0(1917 
 
 , n i5s 
 
 ) 3 
 
 2 
 
 6 ,. 
 
 No. 1, . 
 
 
 3 11 
 
 
 
 
 3 
 
 8 
 
 „ 
 
 ,, Cleveland, ,, No. 3, . 
 
 
 2 14 
 
 
 
 
 2 
 
 9 
 
 „ 
 
 „ ' „ No. 1. . 
 
 
 2 19 
 
 6 
 
 
 2 
 
 10 
 
 » ,, 
 
 , Lincoln, ,, Foundry 
 
 
 3 
 
 6 
 
 
 2 
 
 18 
 
 „ 
 
 Derby, 
 
 
 . 3 2 
 
 6 
 
 
 3 
 
 1 
 
 6 ,v 
 
 Wrought Tron, Cleveland bars, ordinary. 
 
 
 7 15 
 
 
 
 
 7 
 
 5 
 
 „ 
 
 ,, „ best. 
 
 
 8 5 
 
 
 
 
 7 
 
 15 
 
 o „ 
 
 „ ,, ship plates. 
 
 
 8 10 
 
 
 
 
 7 
 
 5 
 
 o ir 
 
 „ angles. . 
 
 
 7 15 
 
 
 
 
 7 
 
 5 
 
 o , 
 
 ,, Staffordshire, best bar, . , 
 
 
 8 10 
 
 
 
 
 8 
 
 
 
 o „ 
 
 ,, sheets, ..... 
 
 
 8 5 
 
 
 
 
 8 
 
 5 
 
 o „ 
 
 Steel, angles and tees, ship quality. 
 
 
 . 7 7 
 
 6 
 
 
 6 
 
 12 
 
 3 ,► 
 
 ,. plates, ship quality, .... 
 
 
 7 15 
 
 
 
 
 7 
 
 
 
 „ 
 
 ,, boiler plates, Board of Trade, etc., , 
 
 
 8 15 
 
 
 
 
 8 
 
 
 
 „ 
 
 ,. forgings, plain shafts, with couplings, 
 
 . £15 
 
 to 20 
 
 
 
 £14 t 
 
 o 18 
 
 
 
 o ., 
 
 Copper ingots G.M.B., .... 
 
 
 70 
 
 
 
 
 79 
 
 5 
 
 ., 
 
 „ ,, tough, 
 
 
 74 10 
 
 0(1917 
 
 , £136) 
 
 84 
 
 5 
 
 o ., 
 
 „ ,, best selected, . . . 
 
 
 74 10 
 
 
 
 
 84 
 
 10 
 
 „ 
 
 ., locomotive plates, .... 
 
 
 84 10 
 
 
 
 (basis) 94 
 
 10 
 
 . 
 
 ,, sheets, . .... 
 
 
 82 
 
 
 
 •J 
 
 92 
 
 
 
 o „ 
 
 ,, pipes, solid drawn. \\ in. to 3 in., . 
 
 
 
 
 10 J 
 
 ) J 
 
 
 
 
 
 10} per lh 
 
 ,, ,, brazed, ..... 
 
 
 
 
 10 
 
 >» 
 
 
 
 
 
 10J 
 
 Brass condenser plates, Mercantile, 
 
 
 
 
 7A 
 
 
 
 
 
 
 73 
 
 Naval, 
 
 
 
 
 71 
 
 
 
 
 
 
 8J .. 
 
 ,, ,. tubes. 66 per cent copper, . 
 
 
 
 8? 
 
 
 
 
 
 
 9 
 
 ,, ,, 70 ,, „ Mercantile, 
 
 
 
 8| 
 
 
 
 
 
 
 91 ., 
 
 , ,, ,, 70 ,, „ Admiralty, 
 
 
 
 9r 
 
 
 
 
 
 
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 Tin. English ingots, ...... 
 
 205 
 
 (1917 
 
 , £300) 
 
 166 
 
 
 
 per ton„ 
 
 Zinc slabs, Silesian, ...... 
 
 25 15 
 
 
 
 
 25 
 
 2 
 
 6 ,. 
 
 ,, ,, best, ...... 
 
 26 15 
 
 0(1917 
 
 £52) 
 
 25 
 
 12 
 
 6 
 
 Aluminium, ingots, Nos. 1, 2, and 4. . . 
 
 84 
 
 0(1917 
 
 £230) 
 
 160 
 
 
 
 „ 
 
 No. 6, 
 
 76 
 
 
 
 
 150 
 
 
 
 „ 
 
 ,, sheets, from Nos. 1 to 30, wire gauge. . 
 
 105 14 
 
 
 
 £150 to 
 
 180 
 
 
 
 „ 
 
 Lead, pig, 
 
 19 15 
 
 (1917 
 
 £32) 
 
 17 
 
 15 
 
 „ 
 
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 21 15 
 
 
 
 
 19 
 
 10 
 
 « ,. 
 
 Nickel ingots, of good quality, . 
 
 1 
 
 
 
 
 
 
 ] 
 
 per lb. 
 
 ,, refined (98 to 99 per cent, purity) 
 
 1 
 
 6 
 
 
 
 
 
 German silver sheet, ..... Is. 3d. 
 
 to 1 
 
 4 
 
 Is. 2d. 
 
 to 
 
 1 
 
 5 „ 
 
 Muntz metal sheet, ....... 
 
 
 
 81 
 
 
 
 
 
 
 7| „ 
 
 Naval brass rods, ....... 
 
 
 
 82 
 
 
 
 
 
 
 «i ,, 
 
 Phosphor bronze, ingots, for general castings. 
 
 106 
 
 
 
 
 102 
 
 
 
 per ton- 
 
 ,, „ suitable for bearing?. 
 
 116 
 
 
 
 
 113 
 
 
 
 „ 
 
 ,, rods, ...... 
 
 1 
 
 o? 
 
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 •) 
 
 1 
 
 2| per lb. 
 
 ,, sheet. ...... 
 
 1 
 
 21 
 
 >» 
 
 
 
 1 
 
 4 
 
 Manganese bronze, ingots, ...... 
 
 
 
 £9ii to 
 
 105 
 
 
 
 per ton. 
 
 Antimony, cakes, ....... 
 
 27 10 
 
 (1917 
 
 £90) 
 
 68 
 
 
 
 „ 
 
 Bismuth, ......... 
 
 10 
 
 
 
 
 
 
 7 
 
 per lb. 
 
 Platinum win-, . . i 
 
 9 5 
 
 
 
 
 4 
 
 LI 
 
 6peroz.Troy 
 
 „ sheets, 
 
 9 5 
 
 
 
 
 4 
 
 11 
 
 6 
 
 ♦ Since August, 1914, the price3 of all metals and alloys have been abnormal, and generally more- 
 sr /ess artificial. 
 
STANDARD TESTS OF STEEL. 
 
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830 
 
 MANUAL OF MARINE ENGINEERING. 
 
 TABLE XCVIIIa. — Steel Forgings (American Standard). 
 
 Process of Manufacture. 
 1. Steel for forgings may be made by the open hearth, crucible, or Bessemer 
 
 process. 
 
 Chemical Properties. 
 
 2. There will be four classes of steel forgings, which shall conform to the following 
 limits in chemical composition : — 
 
 
 Forgings of 
 Soft or Low 
 Carbon Steel. 
 
 Forgings of 
 
 Carbon Steel 
 
 not 
 
 Annealed. 
 
 Forgings of 
 Carbon Steel, 
 Oil Tempered 
 or Annealed. 
 
 Forgings of 
 Nickel Steel, 
 Oil Tempered 
 or Annealed. 
 
 Phosphorus shall not exceed 
 Sulphur „ ,, 
 Nickel ,, ,, 
 
 Per c nt. 
 
 o-io 
 
 010 
 
 Per cent. 
 0-06 
 06 
 
 Per cent. 
 04 
 004 
 
 Per cent. 
 04 
 04 
 3 75 
 
 Physical Properties. 
 3. The minimum physical qualities required of the different sized forgings of each 
 class shall be as follows : — 
 
 Tensile Tests. 
 
 
 Tensile 
 Strength. 
 
 Yield 
 Point. 
 
 Elonga- 
 tion in 
 2 Inches. 
 
 Contrac- 
 tion of 
 Area. 
 
 Lbs. per 
 Sq. Inch. 
 
 Lbs. per 
 Sq. Inch. 
 
 Per cent. 
 
 Per cent. 
 
 Soft Steel or Low Carbon Steel. 
 
 58,000 
 
 29,000 
 
 28 
 
 35 
 
 \ For solid or hollow forgings, no diameter or thick- 
 | ness of section to exceed 10 inches. 
 
 Carbon Steel, not Annealed. 
 
 75,000 
 
 37,500 
 
 18 
 
 30 
 
 \ For solid or hollow forgings, no diameter or thick - 
 \ ness of section to exceed 10 inches. 
 
 80,000 
 
 Elastic 
 Limit. 
 
 40,000 
 
 22 
 
 35 
 
 Carbon Steel, Annealed. 
 
 1 For solid or hollow forgings, no diameter or thick - 
 | ness of section to exceed 10 inches. 
 
 75,000 
 
 37,500 
 
 '23 
 
 35 
 
 { For solid forgings, no diameter to exceed 20 inches, 
 / or thickness of section 15 inches. 
 
 70,000 
 
 35,000 
 
 24 
 
 30 
 
 For solid forgings, over 20 inches diameter. 
 
 90,000 
 
 Yield 
 Point. 
 
 55,000 
 
 20 
 
 45 
 
 Carbon Steel, Oil Tempered. 
 
 1 For solid or hollow forgings, no diameter or thick - 
 / ness of section to exceed 3 inches. 
 
 85,000 
 
 50,000 
 
 22 
 
 45 
 
 1 For solid forgings of rectangular sections not ex- 
 1 ceeding 6 inches in thickness, or hollow forgings, 
 J the walls of which do not exceed 6 inches in 
 ( thickness. 
 
STEEL FORGINGS. 
 
 831 
 
 TABLE XC Villa.— Steel Forcings (Continued). 
 
 Tensile Tests. 
 
 
 Tensile 
 
 Yield 
 
 Elonga- 
 tion in 
 
 Contrac- 
 tion of 
 
 Strength 
 
 Point. 
 
 2 Inches. 
 
 Area. 
 
 
 Lbs. per 
 
 Lbs. per 
 
 
 
 
 Sq. Inch. 
 
 Sq. Inch. 
 
 Per cent. 
 
 Per cent. 
 
 Carbon Steel, Oil Tempered — Continued. 
 1 For solid forgings of rectangular sections not ex- 
 
 80,000 
 
 45,000 
 
 23 
 
 40 
 
 1 ceeding 10 inches in thickness, or hollow forgings, 
 J the walls of which do not exceed 10 inches in 
 ( thickness. 
 
 Nickel Steel, Annealed. 
 
 80,000 
 
 50,000 
 
 25 
 
 45 
 
 J For solid or hollow forgings, no diameter or thick- 
 / ness of section to exceed 10 inches. 
 
 80,000 
 
 45,000 
 
 25 
 
 45 
 
 \ For solid forgings, no diameter to exceed 20 inches, 
 ( or thickness of section 15 inches. 
 
 80,000 
 
 45,000 
 
 24 
 
 40 
 
 For solid forgings, over 20 inches diameter. 
 
 
 
 
 
 Nickel Steel, Oil Tempered. 
 
 35,000 
 
 65,000 
 
 21 
 
 50 
 
 \ For solid or hollow forgings, no diameter or thick- 
 
 ] ness of section to exceed 3 inches. 
 
 ( For solid forgings of rectangular sections not ex- 
 
 90,000 
 
 60,000 
 
 22 
 
 50 
 
 ) ceeding 6 inches in thickness, or hollow forgings, 
 ) the walls of which do not exceed 6 inches in 
 ( thickness. 
 
 
 
 
 
 ( For solid forgings of rectangular sections not ex- 
 
 85,000 
 
 55,000 
 
 24 
 
 45 
 
 ) ceeding 10 inches in thickness, or hollow forgings, 
 
 
 
 ) the walls of which do not exceed 10 inches in 
 
 
 
 
 
 ' thickness. 
 
 4. A specimen. 1 inch by \ inch (1" x \"), shall bend cold at 180° without fracture 
 on outside of bent portion, as follows : — 
 
 Bending Test. 
 
 
 Around a diameter of 
 \ inch, 
 h „ 
 1J „ 
 
 1 „ 
 
 1 „ 
 
 I ., 
 
 1 M 
 
 For forgings of soft steel. 
 
 ,, carbon steel, not annealed. 
 
 ,, ,, annealed, if 20 inches in diameter 
 
 or over. 
 „ ,, ,, if under 20 inches diameter. 
 ,, ,, oil tempered. 
 „ nickel steel, annealed. 
 ,, ,, oil tempered. 
 
832 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Solid Matter in Solution in Water. — Professor Vivian Lewes gave the 
 following analyses as fairly representing the salts, etc., deposited trom (1) 
 ordinary river water, (2) in tidal rivers near their estuaries and well water 
 near the sea, and (3) in ordinary sea water : — 
 
 TABLE XCIX. 
 
 Constituents. 
 
 River. 
 
 Brackish. 
 
 Sea. 
 
 Calcic carbonate, 
 
 ,, sulphate, 
 Magnesic hydrate, . 
 Sodic chloride, . . , 
 Silica, .... 
 Oxides of iron and alumina, 
 Organic matter, 
 Moisture, 
 
 
 
 
 
 75-85 
 3 68 
 2-56 
 0-45 
 7-66 
 2 96 
 3-64 
 3-20 
 
 43-65 
 34-78 
 4-34 
 0-56 
 7-52 
 3 44 
 1-55 
 4-16 
 
 097 
 85-53 
 3-39 
 2-79 
 1-10 
 32 
 trace 
 5-90 
 
 
 
 
 
 
 100-00 
 
 100 00 
 
 100 00 
 
 The following were given by him as analyses of the deposits found in a 
 marine boiler using salt water as its extra supply and the engine getting 
 its internal lubrication in the usual way with valvoline, which is a mineral 
 oil having a specific gravity of 0-889 and the boiling point 700° Fah. 
 It is quite free from free acids, and should be quite free from any animal or 
 vegetable oil or fat, and therefore highly suitable for such a purpose. 
 
 TABLE U Boiler Deposits. 
 
 Scale from Furnace. 
 
 Calcic sulphate, . 
 ,, carbonate, . 
 Magnesic hydrate, 
 Iron, alumina, and silica, 
 Organic matter and oil, 
 Moisture, 
 Alkalies, 
 
 From 
 Top. 
 
 From 
 Below. 
 
 84-87 
 5-90 
 2-83 
 2-37 
 3 23 
 0-80 
 nil 
 
 100-00 
 
 59 
 6 
 
 11 
 2 
 
 19 
 1 
 
 11 
 07 
 29 
 85 
 54 
 14 
 
 nil 
 
 100-00 
 
 Deposit from Tubes. 
 
 Calcic sulphate, . 
 ,, carbonate, . 
 Magnesic hydrate, 
 Iron, alumina, silica, &c, 
 Organic matter and oil, 
 Moisture, 
 Alkalies, 
 
 Scale on 
 Tubes. 
 
 50 
 
 4 
 
 14 
 
 7 
 
 21 
 
 1 
 
 1 
 
 •92 
 •18 
 •12 
 ■47 
 •06 
 •17 
 •08 
 
 100-00 
 
 Deposit 
 above 
 Scale. 
 
 11-60 
 0-82 
 
 22-21 
 9-14 
 
 50-20 
 4-23 
 1-80 
 
 100 00 
 
 Deposit from Bottom of Boiler. 
 
 Calcic sulphate, . 
 , , carbonate, 
 Magnesic hydrate, 
 Silica, alumina, and iron, 
 Organic matter and oil, 
 
 22-52 
 
 nil 
 
 7 09 
 34 S5 
 27 95 
 
 Moisture, 
 Alkalies, 
 
 5-79 
 1-80 
 
 100 00 
 
 On careful examination of the organic matter and oil present in these 
 deposits, it was found that quite one-half of it was " valvoline," in an 
 unchanged condition, which had collected round small particles of calcic 
 sulphate. 
 
EFFECTS OF TEMPERATURE. 833 
 
 The Effect of Temperature on the Strength of Metals. — To-day, with higb 
 -pressure and high superheating likely to become general, it is of very great 
 importance to know how far the metals used by engineers are affected by 
 rise of temperature both in strength and structure. Steam (saturated) at 
 a pressure of 150 lbs. above the atmosphere has a temperature of 365° Fah. ; 
 at a pressure of 200 lbs. the temperature is 387° Fah. ; at 250 lbs., 406° Fah., 
 and at 300 lbs., 421° Fah. Superheating may raise the temperature almost 
 to any degree, and on shore as much as 600° Fah. is considered to be desirable. 
 In any case, however, ivithin the marine H.P. cylinder the temperature will 
 not greatly exceed that of the saturated steam of the same pressure as in 
 the valve-box. It may, therefore, be taken as a rule for such, that the H.P. 
 cylinder and its internal fittings must be suitable for a temperature of 500° 
 Fah. ; the medium-pressure cylinder for 350° Fah., and the low-pressure 
 cylinder for 250° Fah. and the condenser top, 200° Fah. 
 
 Cast Iron is little affected by such temperatures as the above. Tredgold 
 stated that cast iron was by no means reduced in strength by heat up to 
 600° Fah. 
 
 Wrought Iron and Steel (Mild Ordinary) gain in ultimate strength with 
 rise of temperature, so that at 400° Fah. it is 10 to 15 per cent, higher than 
 when at 60° Fah., but with a reduction in the extension. 
 
 Copper. — Pure copper is seldom or never used in modern engineering. 
 Commercially, pure copper, as in pipes and sheets, is not much affected by 
 heat up to 500° Fah. Sir W. Roberts-Austen found that very pure copper 
 having a tensile strength of 29,600 lbs., with 27 per cent, extension at 59° Fah., 
 had 30,900 lbs. with 30 per cent, at 212° ; 30,090 lbs. with 23 per cent, at 
 304° ; 29,020 lbs. with 25 per cent, at 424° ; 28,240 lbs. with 37 per cent, 
 at 480° ; and 25,390 lbs. with 21 per cent, at 617°. Also that copper con- 
 taining only 0-2 per cent, of arsenic had an ultimate strength of 30,150 lbs. 
 with 30 per cent, extension at 64° Fah., and 30,880 lbs. with 30 per cent, 
 at 214° ; 30,960 lbs. with 21 per cent, at 290°, and 28,730 lbs. with 30 per 
 cent, at 480°. Such copper as this is very suitable for steam pipes. Copper 
 containing 0-5 per cent, of arsenic is also good for locomotive fire-box plates 
 as at 835° Fah., as its tensile strength is as much as 20,000 lbs., with 13 per 
 cent, extension. Professor Unwin states that rolled copper, which, at 60° Fah., 
 had a strength of 40.000 lbs., had 38,500 at 212°, 36,000 at 350°, and 33,000 
 at 536°. 
 
 Gunmetal or Admiralty Bronze,* which at 60° may be taken as having a 
 tensile strength of 15-3 tons with an extension of 12-5 per cent., has 31,000 lbs. 
 with 10 per cent, at 200° ; 29,800 lbs. with 8-25 at 350° ; but only 15,700 lbs. 
 with 0-75 per cent, at 400° ; at 500° Fah. the strength is about the same, 
 but with no extension whatever. Bronze with more zinc in its composition, 
 say, 15 per cent., has a tensile strength of 29,000 lbs. with an extension of 
 26 per cent. ; at 400° Fah. it is still 27,750 lbs. with 25 per cent., but at 450° 
 it drops to 9,700 lbs. with 1-2 per cent. only. 
 
 Naval Brass, which at 60° has a strength of 60,000 lbs. with an extension 
 of 19 per cent., still has 52,000 lbs. with 16 per cent, at 400° ; at 500° the 
 strength is about the same, with an extension of over 13 per cent. 
 
 * Mr. J. Dewrance found that an alloy of 875 copper, 10 tin, 2 zinc, and 05 lead had 
 a tensile strength of 158 tons with an elongation of 18 per cent, at 550° F., and at 700° 
 825 tons with 2 per cent. When cold it had 165 tons with 8 per cent. 
 
834 
 
 MANUAL OF MARINE ENGINEERING. 
 
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BRONZKS. 835 
 
 Phosphor Bronze (cast) with a tensile strength of 39,000 lbs. and an 
 extension of 17-5 per cent, at 60° Fah. has about the same strength at 200° ; 
 37,000 lbs. with 15 per cent, extension at 250° ; 36,700 with 12 per cent, 
 at 300° ; 29,400 with 7 per cent, at 350° ; and 27,000 with 5 per cent, at 
 400° Fah., and the same at 500° Fah. Rolled phosphor bronze, with a 
 strength of 68,500 lbs., and an extension of 6-3 per cent, and an elastic limit 
 of 62,000 lbs. has 67,500 lbs. with 8-2 per cent,, and an elastic limit of 
 64,500 lbs. at 400° Fah. ; and at 500° Fah. the strength is as high as 60,000 lbs. 
 with 9-0 per cent, extension, the elastic limit falling to 54,000 lbs. only. 
 
 Bull's Metal, of soft quality, rolled, has a strength of 73,000 lbs. with an 
 extension of 16-4 per cent., the elastic limit being 64,000 lbs. ; at 300° Fah. 
 it stood 60,700 lbs. with 12-7 per cent, extension, and at 400°, 56,750 lbs. 
 with 14-6 per cent., the elastic limit being then as high as 47,000 lbs. 
 
 Aich's Metal, which at 60° Fah. has a tensile strength of 57,300 lbs. and 
 49,600 lbs. at 212° Fah. ; it has still 35,500 lbs. at 482° Fah. 
 
 Aluminium, which at 60° Fah. has a tensile strength of 26,000 lbs., has 
 at 212° Fah. 21,280 lbs., at 300° Fah. 18,100 lbs., at 390° Fah. 14,100 lbs., 
 and at 480° 10,750 lbs. 
 
 Manganese Bronze. — Parsons' No. 6 metal, which has a tensile strength 
 of over 76,000 lbs., with an ultimate extension of 25 per cent, at 100° Fah., 
 still has 56,000 lbs. at 450° Fah., with an extension of 27 per cent., and is 
 specially designed for boiler and cylinder mountings and other castings 
 exposed to high temperatures. 
 
 Melloid, a special alloy made by Bull's Metal Company to resist heat, has 
 an ultimate tensile strength at 60° Fah. of 98,000, with an extension of 
 12 per cent, in 2 inches ; at 600° Fah. it is 79,000 lbs., with 18 per cent. ; at 
 700° Fah. it is 51,900 lbs., with 22 per cent. ; and even at 800° Fah. it could 
 resist 31,800 lbs. and extend 50 per cent. 
 
 Dewrance's Bronze, an alloy of 87-5 copper, 10 tin, 2 zinc, and 0-5 lead, 
 as cast, has a tensile strength of 16-5 tons, or 36,900 lbs. with an extension 
 of 8 per cent. At 450° Fah. the tensile is 36,060 lbs. with an extension of 12-2 
 per cent., while at 550° it is as high as 35,400 with 18 per cent. ; and even 
 at 600 the tensile was 25,200 lbs. with 8 per cent. 
 
 Phosphor Bronze, 96 - l copper, 36 tin, and 03 phosphorus, is good for 
 temperatures up to 450° Fah., having an ultimate tensile of 25 tons, with 
 elongation of 15 per cent, in 6 inches, and a yield point of 21 tons. After 
 exposure for 300 hours to this temperature the ultimate was still 23 to 
 24 tons with 15 per cent. • 
 
 Manganese Copper, 96 - 5 copper, 35 manganese, has a yield point of about 
 18 tons, and is good for temperatures up to 600° Fah. ; the elongation at 
 ultimate is 17 per cent, in 2 inches. 
 
 Nickel Copper is also a useful alloy for things exposed to high temperatures, 
 and may be used for castings or drawn bars. 
 
836 MANUAL OF MARINE ENGINEERING. 
 
 CHAPTER XXX. 
 
 OIL AND LUBRICANTS — ENGINE FRICTION. 
 
 $ 
 
 In these days of large engines and high-speed reciprocators and still higher 
 turbines, the quality of the lubricant used is of first importance. Notwith- 
 standing this, it is by no means an uncommon thing to find that a shipowner, 
 after paying a considerable price to have a first-class engine, will supply the 
 engineer with an inferior oil for the sake of saving 2d. a gallon. The falseness 
 of this economy is apparent when it is known that there may be a difference 
 of as much as 5 per cent, in the efficiency of an engine by using an inferior 
 oil as against a good one, and that with inferior oil the wear and tear of brasses, 
 etc., may be also serious, whereas, with a proper supply of a good lubricant 
 there should be really no wear or tear. At the same time, it must be admitted 
 that the method of lubrication hitherto generally observed on board ship 
 is a primitive and likewise very wasteful one. Surely some system of con- 
 tinuous circulation of the same oil might be followed with the same good 
 results and certainty as obtain with high-speed engines on shore. But it 
 is quite true, however, that to attain this end the use of water on the bearings 
 must first be stopped entirely as a general practice, and only resorted to on 
 emergency ; no doubt this course can be safely followed if the oil is passed 
 through the bearings, pins, etc., as lavishly as is the case with the electrical 
 engine. In the meanwhile it is quite possible and comparatively easy to 
 devise a system whereby the oil from the bearings can be collected, separated 
 from the dirt and water, and so dealt with as to be used over again ; it 
 could in that case be supplied much more freely than is usual now with the 
 lubricating gear at present fitted to modern machinery. 
 
 The Consumption Of Oil for the lubricating of the main engines of a steam- 
 ship should with reasonable care not exceed 1 gallon per 100 I.H.P. per 
 24 hours' steaming. But if the engines are fitted with a good and efficient 
 system of forced lubrication, including coolers and filters, and charged with 
 a supply of oil equal to two or three days' consumption at the above rate, 
 they may be run for months with little or no additional lubricant with the 
 smallest amount of attention from the engine-room staff. The saving of 
 oil alone amply repays interest on the additional cost. 
 
 Oils are of three kinds — animal, vegetable, and mineral — and have pro- 
 bably been discovered and used by mankind in this order. The two former 
 are called fixed oils, because they will not volatilise without decomposition, 
 as do the mineral. 
 
 Of Animal Oils, the ones most used by engineers are — neat's foot. lard, 
 whale, and sperm. 
 
 Sperm Oil is obtained from the head of the sperm whale, and is a very 
 
VEGETABLE OILS. 837 
 
 high-class lubricant, in fact, there is none better for guides and bearings, 
 etc., of engines of all sizes. It is, however, of high price, and there are other 
 mixed oils almost as good for marine purposes to be obtained at a much 
 lower price. 
 
 Whale Oil, better known in other days as train oil, is not often used by 
 itself as a lubricant, although, of course, it could be on an emergency if 
 properly refined so as to be free from any substance that would easily solidify 
 and become hard like a gum. It is now generally used for tempering large 
 steel forgings, etc. 
 
 Lard Oil has not been used on board ship in this country as a lubricant, 
 but in America, as also nowadays in this country, it is known as a good 
 lubricant, having a viscosity equal to that of olive oil, and found to be the 
 best oil for using on the quick cutting modern machine tools when being 
 forced to their utmost power. 
 
 Neat's Foot Oil, obtained from the fat of the feet and hocks of bullocks, 
 was a favourite one with older engineers, and is still used by oil manufac- 
 turers to mix with other oils, usually mineral, so as to form a lubricant suitable 
 for general purposes. 
 
 Vegetable Oils. — The ones best known to engineers are — Olive, castor, 
 colza, rape, palm, cotton-seed, linseed. 
 
 Linseed Oil, obtained from the seed of flax, is not suitable for a lubricant, 
 but is very useful to the engineer in other ways. 
 
 Cotton Oil is expressed from the seed of the cotton plant ; when refined it 
 is colourless and tasteless. It may be used instead of lard oil for cutting 
 tools, as also for mixing with other oils. It is rather too volatile for ordinary 
 marine purposes, and, in fact, from the tendency to dry up, it is hardly 
 looked upon as a safe lubricating oil. 
 
 Palm Oil, in the form of grease, was formerly sometimes used for 
 the outer bearings of paddle ships, and for the filling of the centre 
 recess of tunnel shafting bearing blocks. It may still be used for 
 these purposes instead of tallow on stations where it is cheap. It is, 
 however, now hardly in the purview of the marine engineer as an every-day 
 lubricant. 
 
 Rape and Colza Oils, expressed from the seed of rape or wild turnip, are 
 very much liked, and used generally by locomotive engineers ; apparently 
 generally for the reason that, while being good lubricants, they also are good 
 lamp oils, and not liable to the dangers of fire as is petroleum when an 
 accident occurs, a matter of great consideration before the days of train 
 lighting by gas and electricity. 
 
 Castor Oil, extracted from the castor oil bean, is largely used on board 
 ship in the East Indies, where it is cheap as well as plentiful. It is a very 
 good lubricant, and may be used with advantage, especially in hot climates, 
 where it, as other oils, is apt to run thin. It may here be mentioned that 
 the lubricating value of an oil depends very much on the temperature at 
 which it works. 
 
 Olive Oil, expressed from the fruit of the olive tree, has been used as a 
 iubricant on board ship very extensively, and is a most satisfactory one, 
 being suitable to bearings and guides of all sizes. At one time it was the 
 only oil used in the British Navy, and when replaced by Eangoon oil by an 
 economy-seeking Board, it was shown to be not so extravagant as the price 
 
838 
 
 MANUAL OF MARIA JS KAUiNJiERIN'G. 
 
 seemed to indicate. Olive oil solidifies at 36° Fah., and, like castor, has 
 good viscosity at fairly high temperatures. 
 
 Mineral Oils. — These oils, quite unknown to the old engineers, are now 
 in very general use both for internal and external lubrication, and seem 
 destined to oust other oils from the marine engine-room. The chief supplies 
 are from Southern Russia (Baku) and from the United States of America, 
 although petroleum is known to exist in almost every part of the world. 
 In this country we have a somewhat similar article in our Scotch shale oils, 
 obtained by the distillation of shale and separation from the paraffin con- 
 tained therein. 
 
 The petroleum, as obtained from the wells, is a thick, treacly substance, 
 Russian having a specific gravity as high as - 938, and the American 0*886. 
 After refining, the residue known in Russia as Astatki, is used for fuel on 
 board ship and on locomotives. The American residuum has a specific 
 gravity of 0*928, and is also used for fuel. 
 
 Between that and the highly inflammable light oils, such as petrol, 
 benzaline,.etc, are a large number of grades having varying densities as 
 well as viscosities, and distilled at different temperatures. Among them 
 are oils suitable for the heavy bearings of marine engines as well as the light 
 ones of the smaller kinds ; also oils whose evaporating point is high, and, 
 therefore, suitable for the internal lubrication of marine engines. All these 
 refined oils are pure hydrocarbons — that is to say, they are composed only 
 of hydrogen and carbon. They will not saponify in the ordinary way nor 
 form fatty acids at all, and, therefore, may be used safely for internal lubri- 
 cation without fear of damage to the castings, or to condenser and boilers. 
 At the same time, it is, of course, advisable to prevent the passage of any 
 oily matter whatsoever to the boiler, whether it be some of the pure oil carried 
 along without destruction, or the refuse after use in the engine. 
 
 The following tables give the leading characteristics of the above oils : — 
 
 TABLE CIL— Specific Gravity of Oils when at 60° F. 
 Archbutt and Deeleij. 
 
 Sperm oil, 
 
 Neat's foot oil, 
 
 Lard oil, 
 
 Whale oil, ...... 
 
 Seal oil . 
 
 Porpoise oil, 
 
 Olive oil, 
 
 Rape-seed or Colza oil, . . . 
 
 Cotton-seed oil, . . . . . 
 
 Castor oil, 
 
 Palm oil 
 
 Rosin oil, ...... 
 
 Scotch lubricating oils, . . , 
 
 American dark machinery oils, . . 
 „ light „ 
 
 ,, cylinder „ . 
 
 Russian pale for light machinery oils, 
 heavy 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 0-875 
 0*877 
 0-897 
 0-S85 
 0-895 
 903 
 
 ■880 
 
 915 
 
 916 
 
 922 
 
 925 
 
 926 
 
 915 
 
 915 
 
 923 
 
 965 
 
 923 
 
 •980 
 
 to 895 
 
 „ 0-887 
 
 „ O-9'.'O 
 
 „ 905 
 
 ,, 903 
 
 ,, 0910 
 
MINERAL OILS- 
 
 839 
 
 TABLE GUI. — Density of Oils at Various Temperatures. 
 
 Archbutt and Deeley. 
 
 Name of Oil 
 
 Sperm oil 
 
 Olive oil, ....... 
 
 Rape or Colza oil, ..... 
 
 Castor oil, . . ..... 
 
 Scotch mineral oil, " 865," . . . 
 
 ,, syo, ... 
 
 Russian mineral light machinery oil, . . 
 „ heavy ,, 
 
 American mineral spindle oil, , . . 
 red engine oil, . . 
 Bayonne engine oil, 
 dark medium machinery oil, 
 
 "Valvoline," 
 
 Dark filtered cylinder oil, "900," 
 
 »> 
 »> 
 
 60' F. 
 
 0-87S3 
 09159 
 09151 
 
 0-8683 
 8905 
 0-8978 
 0-9096 
 8677 
 0-9162 
 09113 
 0-8839 
 
 100' F. 
 
 0-8637 
 0-9011 
 9005 
 0-9473 
 0-8533 
 0-8761 
 8837 
 8957 
 0-8535 
 0-9020 
 8973 
 0-8695 
 8757 
 0-8883 
 
 150' F. 
 
 0-8455 
 
 0-8826 
 0-8822 
 0-9284 
 0-8347 
 0-8581 
 08661 
 0-8784 
 0-8358 
 0-S843 
 0-8797 
 0-8514 
 0-8587 
 
 212' F. 
 
 8229 
 0-8596 
 8595 
 9050 
 
 8442 
 
 0-8568 
 08138 
 0-8624 
 0-8579 
 0-8291 
 8377 
 0-8490 
 
 The Flash Point of American cylinder oil (sp. gr., 0-902) is 585° F. 
 
 ( „ 0-893) is 424° F. 
 
 „ „ machinery oil ( „ 0897) is 402° F. 
 
 Russian „ ( „ 0909) is 380° F. 
 
 The Boiling Point of mineral oils is unusually high, and in the case of refined distilled 
 oils seldom less than 600° F., that of Unseed oil being 602° F. The American oils 
 are seldom fluid at a temperature under 25° F., while the Russian generally remain 
 so as low as 0° F. 
 
 Rape oil solidifies at freezing point (32° F.), lard oil at 40° F., sperm oil at 
 39° F., and olive oil at 36° F. 
 
 1'ABLE CIV. — Viscosity of Oils at Various Temperatures (Water 
 at 68° F. being 0-01028)— Archbutt and Deeley. 
 
 Name of Oil. 
 
 60' F. 
 
 100° F. 
 
 150° F. 
 
 212° F. 
 
 
 0-420 
 
 0185 
 
 085 
 
 0046 
 
 Olive ,, 
 
 1-008 
 
 0-377 
 
 154 
 
 0-070 
 
 Rape or Colza oil, 
 
 1118 
 
 422 
 
 0177 
 
 0-080 
 
 Castor oil, 
 
 • ■ < 
 
 2-729 
 
 0-605 
 
 0169 
 
 Scotch mineral oil " 865," .... 
 
 0146 
 
 0066 
 
 036 
 
 ••• 
 
 ,, ,, 890, .... 
 
 509 
 
 0183 
 
 069 
 
 ... 
 
 Russian light machinery oil, ... 
 
 1-156 
 
 307 
 
 099 
 
 043 
 
 „ heavy „ ... 
 
 3-592 
 
 0-762 
 
 0196 
 
 0-066 
 
 American spindle oil, 
 
 0-727 
 
 236 
 
 0-086 
 
 039 
 
 „ pale „ "903/7,". 
 
 1-138 
 
 0342 
 
 0115 
 
 049 
 
 „ "907/12," . . 
 
 1-479 
 
 0419 
 
 ... 
 
 ... 
 
 red engine oil "910/20," . 
 
 1-915 
 
 0-496 
 
 0150 
 
 058 
 
 "Bayonne" engine oil " 912/15," 
 
 2-172 
 
 0572 
 
 0173 
 
 063 
 
 ,, dark medium machinery oil, 
 
 3 046 
 
 0-705 
 
 0210 
 
 0076 
 
 "Valvoline," 
 
 ... 
 
 2-406 
 
 605 
 
 0187 
 
 American filtered cylinder oil " 888," 
 
 ■«. 
 
 3-702 
 
 ... 
 
 238 
 
 „ dark .. ., "900," 
 
 ... 
 
 6-264 
 
 ■•■ 
 
 0-314 
 
840 
 
 MANUAL OF MARINE ENGINEERING. 
 
 TABLE CV. — Boiling, Setting, and Flash Points. 
 
 Name 
 
 or Brand of Oil. 
 
 Sets at 
 
 Flashes at 
 
 Boils at 
 
 
 
 Degrees F. 
 
 Degrees F. 
 
 Degrees F. 
 
 American cylinder, 
 
 sp. gr. -902, 
 
 25 
 
 585 
 
 650 
 
 
 •893, 
 
 
 
 
 
 25 
 
 424 
 
 600 
 
 Valvoline ,, 
 
 ... 
 
 
 
 
 
 ... 
 
 • •• 
 
 700 
 
 American machine, 
 
 sp. gr - S97 
 
 
 
 
 
 25 
 
 402 
 
 600 
 
 Russian ,, 
 
 „ 909 
 
 
 
 
 
 to 10 
 
 3S0 
 
 ... 
 
 Scotch shale oil, 
 
 • • * 
 
 
 
 
 
 32 
 
 • •• 
 
 ... 
 
 Cotton-seed oil, 
 
 • 
 
 
 
 
 
 
 34 to 50 
 
 ■ •• 
 
 5S0 
 
 Castor oil, . 
 
 • 
 
 
 
 
 
 
 14 
 
 ... 
 
 556 
 
 Lard oil, . . 
 
 • 
 
 
 
 
 
 
 25 to 42 
 
 • •• 
 
 
 Linseed oil, 
 
 • i 
 
 
 
 
 
 
 5 „ 17 
 
 • •• 
 
 574 
 
 Neat's foot oil, . 
 
 • 
 
 
 
 
 
 
 32 „ 50 
 
 • •• 
 
 ... 
 
 Olive oil, . 
 
 • 
 
 
 
 
 
 
 21 „ 39 
 
 , . 
 
 ... 
 
 Rape or Colza oil, 
 
 • 
 
 
 
 
 
 
 10 „ 28 
 
 . .. 
 
 625 
 
 Sperm oil, 
 
 • l 
 
 
 
 
 
 
 32 
 
 450 
 
 600 
 
 Palm oil, . 
 
 • 
 
 
 
 
 
 
 SI to 10G 
 
 ... 
 
 ... 
 
 Whale oil, . 
 
 • 
 
 
 
 
 
 
 32 
 
 440 
 
 620 
 
 Vaseline, . 
 
 • 
 
 
 
 
 
 
 104 to 120 
 
 640 
 
 . . . 
 
 Tallow (Russian), 
 
 • 
 
 
 
 
 
 
 107 ,, 119 
 
 540 
 
 660 
 
 TABLE CVI. — Price per Gallon of Various Oils.* 
 
 Name of Oil. 
 
 Price 
 per Gallon. 
 
 Name of Oil. 
 
 Price 
 
 per Gallon. 
 
 Castor oil, .... 
 Colza oil, .... 
 Rape oil, . • 
 Cotton-seed oil, . 
 Linseed oil, . . 
 Olive oil, , . . . 
 
 2/3 to 2/6 
 2/1 
 
 2/- 
 21- 
 
 2/- to 2/3 
 3/3 „ 3/9 
 
 Palm oil, .... 
 Sperm oil (common), . 
 Whale oil, .... 
 Neat's foot oil, . . 
 Lard oil, .... 
 American lubricating (good), 
 
 2/9 to 3/- 
 2/3 „ 2/6 
 
 2/- 
 3/- to 3/3 
 4/- „ 4/3 
 1/3 ,, 1/6 
 
 Compound Oils. — The oils most generally used on shipboard for external 
 lubrication are compounded of a mineral oil and one or more of the fixed 
 oils. Mineral oil of itself, although, on the whole, a good lubricant, is not 
 suitable for bearings or guides which have, or'rnay have, a greater pressure 
 per square inch than 70 lbs. Moreover, increasing the viscosity of a mineral 
 oil by other denser oils does not improve it for the purpose, and, in fact, is 
 often the means of causing greater friction. If, however, a fixed or fatty 
 oil be compounded with the mineral oil in a proper way so that they do not 
 separate out when standing still, the compound will form a suitable lubricant 
 for higher pressures than the above. The higher the pressure per square 
 inch, the greater viscosity must be given to the lubricant by adding fatty 
 oils of higher viscosity. It will also be observed by reference to Table civ. 
 that mineral oils whose viscosity is very high at 60° Fah. are the very ones 
 whose viscosity rapidly decreases with an increase of temperature, and 
 although the fatty oils decrease in viscosity as the temperature rises, their 
 
 • These are all pre-war prices, now they ure abnormal and variable. 
 
ADULTERATION OP OILS. 841 
 
 rate of fall is not so great. In other cases, however, pure mineral seems 
 to be lacking in what Messrs. Archbutt and Deeley term " oiliness." 
 
 Mineral oils are sometimes thickened in viscosity by dissolving certain 
 soaps in them, the favourite for this purpose being aluminium soap, which 
 is made by saponifying whale, cotton, or even lard, oils, and caustic soda, 
 and pouring into it when hot a solution of common alum. This is, of course, 
 an adulteration, but not altogether a bad one, inasmuch as the oil then sticks 
 to the journal much better than it did in the pure state, and gives a better 
 lubrication to the bearing. The soap, however, is liable to start out and 
 obstruct the oilways. When mineral oil has been treated in this way it 
 can generally be detected by the tendency to threadiness when the cork is 
 drawn from the bottle of the mixture. 
 
 The oils most generally used for compounding with mineral oils are — 
 Rape, olive, lard, neat's foot, whale, and sperm. Castor oil will not dissolve 
 in and compound with a mineral oil unless it is first treated with an equal 
 volume of tallow oil. It is, however, not often used for the purpose. Rosin 
 oil is sometimes used to increase the specific gravity of a compound oil as 
 well as to improve its viscosity. It does not, however, improve its lubri- 
 cating powers, and as it increases the tendency to guminess it should be 
 avoided. Marine engineers usually prefer the compound oil to a pure mineral 
 one on account of the tendency to lather — that is, to form grease by mixing 
 with the water applied to the bearing. This formation is hastened, of course,, 
 by the use of a little soda in the water, and when formed it generally prevents 
 the oil running freely out of the bearing or guide. Besides which, should 
 the bearing or guide become warm, this grease melts and adds to the 
 quantity of lubricant so as often to check any further heating. It is 
 evident, therefore, that for a system of forced lubrication the oil must 
 not be able to saponify or even to form an emulsion easily. Certain 
 kinds only of mineral oil can be depended on for the purpose, as being 
 free from these vices. 
 
 Soaps. — Pure mineral oils are unaffected by alkalies and have little or 
 no effect on metals. They, therefore, do not saponify, and for this reason 
 are used for internal lubrication. Formerly, when tallow and fatty oils were- 
 used indiscriminately for internal and external lubrication, the condenser 
 very soon got covered with a soapy deposit more or less hard. All vege- 
 table and animal oils contain fatty matters, which combine with alkalies and 
 form what are called soaps. Some oils will also combine with certain salts 
 of metals and form a soapy mixture; for example, white lead and linseed 
 oil. Oils combined with caustic soda make hard soaps, and, when 
 combined with caustic potash, form what is known as soft soap. Solu- 
 tions of these soaps will act as lubricants, but are, of course, not so- 
 efficient as the oils. They are sometimes temporarily used to cleanse oil 
 ways of the gummy greasy matter which may have been deposited from 
 the oil. 
 
 Adulteration Of Oils. — Compound oils are generally valued by their 
 density and viscosity. It is, therefore, the aim of the dishonest merchant to 
 add some cheap ingredient which shall effect this object. That most gener- 
 ally used for the purpose is resin oil, being cheap and easily mixed with 
 other oils. It is, however, itself inferior as a lubricant, and has the fatal 
 vice of drying hard like varnish. Its presence in oils is easily detected 
 
842 MANUAL OF MARINE EM4INEEHJNG. 
 
 by the pungent taste, which is unmistakable,* and its failure to saponify- 
 when an alkali is added. 
 
 Solid Lubricants. — There are certain substances which act as lubricants, 
 although not at all on the same principle as does a liquid. The chief of 
 these are powdered soapstone, flour of sulphur, and plumbago. 
 
 Soapstone and French chalk, finely powdered to a flour, has a greasy 
 feel, and acts as a very efficient lubricant to wood fittings, and also, to some 
 extent, to metallic ones. It is, however, of little practical value as a modern 
 engine lubricant. 
 
 Flour of Sulphur was frequently used by the older engineers, whose 
 acquaintance with hot bearings was much more intimate and frequent than 
 is that of their succ ssors to-day ; they always carried it among their stores, 
 and, when a bearing was troublesome, poured a small quantity down the 
 oil holes. Its efficiency probably varied with the heat of the bearing. Its 
 function seems to have been to combine with the brass dust and brass ridges, 
 and thereby to form a sulphide or sulphides of the metals which was much 
 softer than the metals themselves, and, in a plastic state, permitted of the 
 production of a web which lined the bearing and provided a new surface for 
 the journals to run on. If this theory is a correct one, the material, of course, 
 cannot be considered as a true lubricant. 
 
 Graphite or Plumbago. — This material, when pure and so highly sub- 
 divided that it forms an impalpable powder, has an extremely greasy feel to 
 the fingers, and is an exceedingly good lubricant for both metals and wood- 
 work. It may be used together with oil, grease, or vaseline for large and 
 heavy bearings revolving at moderate velocity. It is also used for the 
 internal lubrication of some engines, and doubtless may, in the near future, 
 play an important part in that direction when superheated steam is used on 
 shipboard. At present the means of applying it to the internal parts are 
 crude and unsatisfactory. Ingenuity, however, will soon find a method to 
 overcome the present difficulties, and, whether this material is carried over 
 in a dry state or mixed with vaseline or some other semi-liquid carrier, it 
 will be properly distributed and do good service. A small quantity of graphite 
 may be used with advantage on bearings which are lubricated with oil, as 
 it tends to keep the surface of both journal and bearing smooth and polished. 
 When a brass has become rough or scored, this material will fill the recesses 
 and present a new and good surface to the journal. 
 
 Tallow and certain other fats, both vegetable and animal, are, at ordinary 
 temperatures in temperate zones, in the solid state, but a slight rise of tem- 
 perature converts them into a liquid, and, therefore, it is idle to deal with 
 them as if they were solid lubricants. Moreover, none of them are of great 
 importance to the marine engineer of to-day. 
 
 Grease. — If oils whose specific gravity is nearly that of water be mixed 
 with pure warm water and agitated, an emulsion is formed, and the oil will 
 separate from the water very slowly indeed, even when the mixture is allowed 
 to stand. If the water is made slightly alkaline and carefully stirred into 
 some animal and vegetable oils, especially if at a moderately high tempera- 
 ture, they combine and form grease — the value of which as a lubricant 
 depends in great measure on the quality and quantity of the oil used in 
 
 * For further particulars of oils, Bee Lubrication xni Lubricants, by Archbutt and 
 Deeley, published by C. Griffin & Co., Ltd. 
 
COEFFICIENTS OF FRICTION. 
 
 843 
 
 making it. Tallow, palm oil, aud other such fatty substances will take 
 up very considerable quantities of water. 
 
 Rule. — Let S be the velocity of sliding in feet per minute, P the 
 maximum steady working pressure in lbs. per square inch on a slide 
 lubricated in the common way (that is, without force). 
 
 Then P = 50,000 - S + 100 
 
 if intermittent as in piston-rod guide, maximum value is P, then 
 
 P = 65,000 -=- S + 100. 
 
 For journals with steady load 
 
 P = 65,000 -«- d x R + 30. 
 
 d being diameter in feet, and R the revolutions per minute. 
 For intermittent circular as in crank-pins 
 
 P = 100,000 -=- d x R + 30. 
 
 For intermittent oscillating as in cross-head brasses 
 
 P = 250,000 + R + 130. 
 
 TABLE CVII. — Coefficients of Friction. 
 
 Nature of the Surfaces. 
 
 Wood on wood (dry), , 
 
 Metal on hard wood, 
 
 t • « 
 
 Metals on oak, . • » 
 
 ,, metals only, . > 
 
 ,, ,, SmOOtb, r r 
 
 Bronze on lignum vitae, . . 
 
 Iron ,, 
 
 Bronze on bronze sliding, 
 
 . • 
 
 l ) n 
 
 Steel 
 
 Wrought iron on bronze sliding, 
 Last ,, ,, ,, 
 
 Steel on cast iron 
 
 . (Rk) 
 
 . (S) 
 . (Rn 
 
 Quite 
 Dry. 
 
 0-25 
 to 
 0-5 
 
 0-5 
 
 to 
 
 6 
 
 Greasy 
 Dry. 
 
 015 
 
 24 
 
 to 
 
 0-26 
 
 [015 
 
 to 
 
 (0-2 
 
 0-07 
 
 to 
 
 0-08 
 
 0-175 
 0139 
 135 
 0141 
 0*151 
 
 Well Lubricated with 
 
 Oil 
 
 007 
 
 to 
 
 0-08 
 
 05 
 
 Tallow 
 or Soap. 
 
 2 
 to 
 04 
 
 2 
 
 0-14 
 
 Water. 
 
 3 
 
 05 
 05 
 0-38 
 
 Con- 
 stant 
 Flow 
 of 
 Oil. 
 
 05 
 
 003 
 
844 
 
 MANUAL OK MARINE ENGINEERING. 
 
 TABLE CVII. — Coefficients of Friction — Continued. 
 
 
 
 
 
 Well Lubricated with 
 
 Con- 
 
 Nature of the Surfaces. 
 
 
 Quite 
 Dry. 
 
 Greasv 
 Dry. 
 
 
 
 
 stant 
 
 Flow 
 
 
 
 
 
 
 
 
 Oil. 
 
 Tallow 
 
 Water. 
 
 of 
 
 
 
 
 
 
 or Soap. 
 
 
 Oil. 
 
 Steel on wrought iron, .... 
 
 (R 
 
 1 
 n) .. 
 
 0-189 
 
 
 
 
 
 Wrought iron on cast iron, , . 
 
 i 
 
 
 0170 
 
 ... 
 
 ... 
 
 * • . 
 
 ... 
 
 n t> tin, . . 
 
 i 
 
 j • •• 
 
 0181 
 
 «. ■ 
 
 
 • •• 
 
 
 „ magnolia, . . (RH 
 
 S) ... 
 
 060 
 
 ... 
 
 • • . 
 
 ■ ■• 
 
 ... 
 
 
 
 
 
 
 
 
 1 01 
 
 ., „ white metal,. , . 
 
 
 • *M 
 
 *•• 
 
 0-025 
 
 ... 
 
 
 I to 
 ( 0-015 
 
 Loco axles, white metal, . • • • 
 
 (? 
 
 V) ... 
 
 • •• 
 
 0-017 
 
 • • * 
 
 .. . 
 
 
 Shafts on bronze, 
 
 (R 
 
 n) ... 
 
 
 ... 
 
 0O28 
 
 
 
 Leather on metals, . 
 
 (R 
 
 k) 0-56 
 
 0-23 
 
 0-15 
 
 a • ■ 
 
 0-36 
 
 
 Wrought-iron shaft in cast-iron bearing, 
 
 
 
 
 
 
 
 
 speed 10 feet per minute, . 
 
 
 . M 
 
 ■ > • 
 
 0-079 
 
 ... 
 
 ... 
 
 • •• 
 
 Wrought-iron shaft in cast-iron bearing, 
 
 
 
 
 
 
 
 
 speed 60 feet per minute, . 
 
 
 
 
 0-052 
 
 • ■• 
 
 >'»■ 
 
 ••• 
 
 Wrought-iron shaft in cast-iron bearing, 
 
 
 
 
 
 
 
 
 speed 100 feet per minute, . 
 
 
 • •■ 
 
 ... 
 
 0-050 
 
 • ■ • 
 
 • •> 
 
 ... 
 
 Load 50 lbs. per square inch, speed 50 feet 
 
 (< 
 
 >) ... 
 
 ... 
 
 ... 
 
 .. . 
 
 ... 
 
 •0032 
 
 „ 50 „ „ „ 110 „ 
 
 
 * . . . 
 
 . . . 
 
 ... 
 
 ■ •• 
 
 • •• 
 
 •0064 
 
 ,, 50 190 „ 
 
 
 I •• • 
 
 .. . 
 
 ... 
 
 , , 
 
 
 •0106 
 
 „ 150 „ „ „ 50 „ 
 
 
 1 • • • 
 
 ... 
 
 ... 
 
 ... 
 
 ■ •• 
 
 •0017 
 
 ,, 150 „ „ „ 110 „ 
 
 
 * ... 
 
 .. . 
 
 • ♦# 
 
 .. t 
 
 ... 
 
 ■0024 
 
 » 150 „ „ ,, 190 „ 
 
 
 
 .. • 
 
 • t* 
 
 ■ ■■ 
 
 • •• 
 
 •0047 
 
 Speed of 10 feet per min. , load 50 lbs. , 
 
 
 > 
 
 • • ■ 
 
 ... 
 
 • ■ • 
 
 • •• 
 
 •0009 
 
 ,, 10 ,, ,, 7o ,, . 
 
 
 J •• . 
 
 .. . 
 
 ... 
 
 • •• 
 
 • •• 
 
 •0007 
 
 10 „ „ 150 „ 
 
 
 ) •• • 
 
 
 ... 
 
 • •• 
 
 ... 
 
 0250 
 
 ,, 15 ,, ,, 50 ,, . 
 
 
 » •■• 
 
 
 .. . 
 
 ... 
 
 ... 
 
 •0012 
 
 ,, 15 ,, ,, 150 ,, 
 
 
 J ••• 
 
 
 
 ... 
 
 • •• 
 
 •0051 
 
 ,, 7 8 feet per minute, moderate load, 
 
 
 1 ••• 
 
 0-094 
 
 0-072 
 
 ... 
 
 • •< 
 
 •064 
 
 .. ,, heavy ,, 
 
 
 « ■•■ 
 
 0-051 
 
 046 
 
 ... 
 
 • •• 
 
 •047 
 
 Railway axles and bearings (static), . 
 
 
 ( 0-36 
 
 rl) \ tO 
 
 •■■ 
 
 o-ioo 
 
 067 
 
 ... 
 
 • •• 
 
 Cast-iron steel tyres, . . « 
 
 (G 
 
 > 5 to 
 
 50 mile 
 
 3 per h< 
 
 uir. 
 
 
 
 
 f -048 
 
 ) 
 
 
 
 
 
 
 
 ( 011 
 
 1 
 
 
 
 
 
 Steel tyres on steel rails, .... 
 
 » 
 
 , I to 
 /004 
 
 | 10 to 
 
 50 „ 
 
 it 
 
 
 
 
 
 ' 
 
 
 
 
 
 Note. — Rk stands for Rankine ; Rn, for Renme ; S, for Summers; W, for Wood; 
 G, for Goodman ; RHS, for Prof. Smith ; Gl, for Gallon and Westinghouse. 
 
TESTS AND TRIALS 84f 
 
 CHAPTER XXXI. 
 
 TESTS AND TRIALS : THEIR OBJECTS AND METHODS. 
 
 That there must be tests and trials of ships and machinery goes without 
 saying, but their nature and extent is a matter of considerable controversy, 
 and one which has given rise to much discussion. In the case of tests of 
 material only, samples are, as a rule, dealt with, which can be and generally 
 are subject to trials ending in destruction, during which process certain 
 facts and figures are observed and noted, so that from them conclusions 
 may be drawn as to the fitness of the material of which they formed portions 
 for the purposes designed. Seldom or never is any actual portion of an 
 engine or boiler subject to any mechanical test until fitted in place and under 
 trial conditions, hence the stress on it is never greater than the working 
 stress, except when the trial is by water pressure greater than that when 
 working with steam. 
 
 It is a moot question as to which is the best criterion of the value of a 
 material, the ultimate tensile strength, the elastic limit, or the yield point. The 
 ultimate tensile may be taken as a very crucial test in demonstrating the 
 elasticity, as well as the power of resistance ; but it is after all the prelude 
 to a destruction which is irresistible that is being observed, and, therefore, 
 while being interesting, it bears no anology to what should take place in 
 practice ; in fact, it is not the barometer, but the storm and its effects. The 
 elastic limit is really only a phase or milestone in the course of a test, 
 and that the change in the relations between stress and strain is anything 
 more than an interesting phenomenon remains to be proved, except that the 
 stretching in becoming greater points to a giving way of tenacity for the 
 time, for if the load is relieved and for a time withdrawn, the material can 
 better resist the load when it is again applied. It is pretty certain that the 
 yield point or that stress at which the material ceases to regain its original 
 length when relieved of it is open to some objection as a criterion of its 
 ability ; and that it is more than probable that all metals after manufacture 
 will have permanent set after the application of load far below those which 
 afterwards fail to produce that phenomenon, just as a rope will stretch under 
 quite light loads when quite new, and will continue to stretch until its com- 
 ponent parts have been forced into the position which enables them to be 
 efficient, and whereby the fibres can all take their proper share of resistance. 
 In metallic structures a somewhat similar process goes on, but to a less 
 notable extent, and some parts of each component is probably stressed 
 beyond the yield point of the material ; but gradually such stressing decreases 
 until, like a built-up gun made scientifically, each lamen takes its fair share 
 of the load which comes on the component when under working conditions. 
 
 On the whole, it would seem, from the practical engineer's point of view, 
 
 54 
 
346 MANUAL OF MARINE ENGINEERING. 
 
 thnt the yield point is the safest criterion and guide, especially if it is ascer- 
 tained in ways analogous to service conditions. If an engineer desires to 
 employ a particular metal, which will, when in use, be subject to certain 
 maximum stress, then it should be tested with a load which produces double 
 that stress, and remain applied for a considerable time, say 10 hours, after 
 which there should be no permanent set whatever on its removal. Before 
 this crucial trial, however, there should be the application of the load and 
 its removal after, say, five minutes to take out the "initial set." This would 
 show its ability to take continuous load ; but for intermittent and alternating 
 stresses such a load should be applied and released suddenly, say a dozen 
 times, and the effect of each carefully noted. These are only rough and ready 
 means of proving the general ability of the material to endure loads which, 
 while being quite moderate, are much beyond those coming on in practice. 
 
 It has been also a subject of controversy whether engineers shall work 
 with factors of a safety or margins of safety ; that is, shall the test stress 
 which proves the strength of a structure be a multiple of the working stress, 
 or shall it be a fixed excess of it ? The Board of Trade, Lloyd's Register, and 
 some other public institutions stick to the factor of safety — in fact, so long 
 as the Board of Trade follow this method, these others cannot help but follow. 
 The British and other Admiralties, and most of the foreign institutions 
 governing shipping, adopt the margin or a modification of it. That is, past 
 a certain working pressure the addition to it for testing boilers, high-pressure 
 cylinders, etc., is a constant quantity. If the object of a test is to prove 
 the ability of a structure to withstand the effect of the greatest load which 
 can come on it, there seems no good reason why the test should be exactly 
 double the normal maximum pressure, although under some circumstances 
 even that test might be too small. If the test is one appreciably beyond 
 the maximum normal with the generality of engines and boilers, it should 
 prove all that is necessary when there is nothing unusual in the design and 
 construction. If their design and construction is abnormal, is complicated, 
 so as to be impossible of being carefully calculated, and, moreover, when at 
 work, is liable, as oil and gas engines are, to heavy shock and unexpected 
 conditions, the test should be far beyond, and may be justifiably more than 
 double the supposed working pressure. It is obvious, therefore, that in 
 all tests and trials there must be elasticity in the making and application 
 of rules, and each case must stand by itself ; the great thing being to bear 
 in mind what is the object of the trial, and to make the means subservient 
 to the true end. 
 
 Tests and Trials are made of the materials and the parts of an engine, 
 and further of the engine itself when complete in a more or less haphazard 
 way, and generally without clear ideas as to the object or nature of each of 
 them, and consequently they are often not so satisfactory or conclusive 
 as they should be. The object of a trial of any kind should be kept clearly 
 in view all the time when making it, and the end rather than the means 
 should have the first attention ; the nature of the trial or test must be always 
 subservient to the particular object desired, and the mixing up of policies 
 strictly avoided. In every-day practice there are rough and ready ways 
 by which the engineer can satisfy himself as to the adequacy of the means 
 he adopts and the safety of the materials he employs on them, but moie 
 than that is necessary for success in modern marine engineering where high 
 
TESTS AND TRIALS. 847 
 
 speed and huge output of power involve forces and stresses complicated 
 and heavy, continuously applied throughout the whole 24 hours without 
 intermission or stoppage of machinery, which must not fail or even be feared. 
 
 The tests applicable to such mechanism are not necessarily severe, but 
 they must be very circumspect, and used with discretion. The endurance 
 of a material is a quality highly to be desired in most parts, while in others 
 it is secondary to high elastic limit. 
 
 In some parts of an engine necessity compels the minimum of material 
 with the maximum of stress as a consequence ; in such a case it is imperative 
 that the metal employed should have a high elastic limit, and ample proof 
 of it must be insisted on from the testing machine. If the stresses are to 
 be suddenly applied the metal must be shown to be capable of withstanding 
 them by " drop " tests and other similar means. If they are intermittent, 
 or if alternating, then simple tensile or drop tests are not sufficient ; there 
 must be a series of bending backwards and forwards tests, or other similar 
 tests to prove the endurance of the metal. On the other hand, for such 
 parts as are of necessity large to give bearing surface, and on which the 
 stresses are always low, it seems folly to waste time and money on any tests 
 beyond those which are involved in machining, etc., in process of manu- 
 facture ; the nature of steel can be gauged fairly correctly by observing 
 how it cuts in the lathe and planing machine. Such things as are materially 
 affected by heat or other treatment, while undergoing manufacture into the 
 finished article, should have test pieces removed after rather than before 
 such processes. But where such a process as oil tempering is carried out 
 with the object and certainty of improving them, no such test is necessary. 
 
 In the case of cast iron some tests are necessary to prove the quality 
 rather than the actual tensile strength; that is, the ability to bear bending 
 and some degree of shock. Here, again, it is necessary to bear in mind the 
 object of the test, and so to devise the best means for attaining the end in 
 view, and not something more or different. A steam cylinder is subject to 
 shock from the sudden admission and emission of steam ; the H.P. cylinder 
 walls are always in tension, varying from a maximum at cut off to the mini- 
 mum at commencement of exhaust, the metal will be at a temperature some 
 what but not much below that of the steam at admission ; the stress on the 
 material must not be high — not high enough at anyrate to produce perceptible 
 stretch, or there will be an increase in diameter of bore and a consequent 
 leakage past the piston. Moreover, the metal must be such that, while 
 capable of being bored and machined, it must have the quality which admits 
 of taking a polish when rubbed, and that without much lubricant, and must 
 in many cases continue to be rubbed without a lubricant at high temperatures 
 without abrasion of the surface. 
 
 Such a metal must have, therefore, qualities that no tensile testing machine 
 can prove to exist ; in fact, high tensile material is desirable rather for the 
 water tests put on the cylinder than those applied in every -day work. Cross- 
 bending of a bar freely supported is still the most satisfactory method of 
 proving its strength and capacity for taking a sudden load — that is, with- 
 standing shock. Chemical analysis is perhaps the safest test for making 
 sure of its other virtues — in fact, the successful iron moulder to-day 
 establishes his reputation by the care with which he selects his irons for 
 mixing, so that the chemical composition shall be definite, and one known 
 
848 MANUAL OF MARINE ENGINEERING. 
 
 to be co-existent with high mechanical properties. But since the most 
 important parts of an engine which are made of cast iron always work when 
 at a high temperature, mere cold tests are not sufficient ; such tests as are 
 made should be with the metal at a temperature 5 to 10 per cent, higher 
 than that at which they are exposed when at work. These remarks apply 
 with even greater force to the bronzes, some of which, while being very 
 strong at 160° F., are weak and unreliable at 420°, while at 600° are of no 
 use whatever. 
 
 It is usual to test by water pressure the more important parts of a marine 
 engine before placing them in the ship. Here, again, it is not clear, judging 
 by results, what is aimed at by the authorities in making such tests. Prim- 
 arily safety is the first consideration, and was probably the only one originally, 
 when the ability to calculate strains and stresses was limited, and the experi- 
 ence to guide the designer equally limited. To-day the water test prescribed 
 for cylinders appears to be with the object of proving the strength of the 
 casting, and incidentally its soundness. Unfortunately, these two things 
 are often confounded by inspectors and surveyors, so that a casting which 
 is quite strong enough and even sound enough for its work is condemned 
 because at test pressure there has been a leakage through local sponginess 
 although at working pressure there was no leakage. The same test for boilers 
 is open to the same misunderstanding, for it is not unknown to find parts 
 of a boiler which at the pressure at which the safety-valves are to be set 
 are unaffected, do actually bulge and deform, so as sometimes to produce 
 leakage at the test pressure, especially when that test is at double the working 
 pressure. Tightness is one thing, and a very desirable thing, but it should 
 be guaranteed under working conditions. Therefore, that a casting leaks 
 on testing with cold water at a pressure far beyond that at which it 
 works under steam is no proof of unfitness for its work. The high pressure 
 proves its strength ; its ability to sustain steam without loss can only be 
 proved with steam. 
 
 With a complicated casting like a cylinder it is necessary to prove the 
 following, viz., that — 
 
 (a) It is strong enough to sustain the loads, shocks, and thermal changes 
 
 coming on it internally and externally. 
 
 (b) Under working pressure there is no deformation of its body or adjuncts 
 
 that will interfere with the good working of the piston and valves. 
 
 (c) Under working pressure there is no leakage which will cause any 
 
 loss of steam by escape to the air or exhaust. 
 
 (d) And the metal is such as can be machined satisfactorily, and leave 
 
 a good working surface for the rub of pistons, valves, etc. 
 
 The low-pressure cylinder differs from the others, inasmuch as at exhaust 
 the internal pressure is much less than that external, and consequently 
 that the metal is during that period in a state of compression, and the body 
 exposed to collapse. There is no test ever applied to prove its ability to 
 withstand this state, so that the water test is limited to a question of tightness 
 rather than strength. 
 
 The casings and stators of turbines must be strong enough to safely 
 sustain the internal pressure of the steam, which may be that of the boilers 
 
ADMIRALTY RULES FOR TESTING MACHINERY. 840 
 
 in a complete turbine, as it is in the high-pressure portion of a compound 
 one. In actual work it is only at the entry of the H.P. end that is exposed 
 to full steam pressure, but in case of a hitch it might extend throughout. 
 But in the Parsons turbine there must be no increase in diameter when 
 working under pressure which would materially increase the clearance and 
 the loss due to it. nor must the body change from its true circular section ; 
 therefore, when it is tested it should be carefully gauged when at working 
 pressure, as the cylinders of reciprocating engines require doing. 
 
 The L.P. portion of a turbine case is liable to an absence of pressure 
 throughout its length when standing, and, therefore, like the L.P. cylinder, 
 is subject to collapsing forces, its ability to resist which is never tested, and 
 only guessed at by testing it under internal pressure high enough to prove 
 the soundness of the casting. 
 
 The testing of a condenser body is equally unsatisfactory for the same 
 reason, for in designing it to resist such a test material is added, and the 
 form adopted which is quite useless to help it against working strains. If 
 the stiffening ribs of a casting are on the right side for test purposes, they 
 are on the wrong one for working loads, and vice versd. The waterways and 
 ends of a condenser are subject to slight internal pressures, against which 
 provision can be made without such objection, as anything added for testing 
 really helps these parts to sustain the latter rather than working loads. The 
 British Admiralty require the folio whig water tests carrie out satisfactorily : — 
 
 Admiralty Rules for Testing Machinery. 
 
 615. Water Pressure Tests Generally.— The water pressure tests required 
 are to be carried out after the parts are machined, and before they are painted 
 or lagged, or covering of any kind is placed on them. 
 
 616. Tests before Fitting on Board. — The boilers, turbines, and their 
 connections, condensers, air pumps, and all filter, main, receiving, lime, 
 drip, and oil tanks, and the feed tanks, when not built into the ship, are 
 to be tested by water to the required pressures before being put on board 
 the ship ; the steam and exhaust pipes, all boiler mountings and connections, 
 and other important eastings are to be tested before being fixed in their work- 
 ing position. The tests for the turbine cases other than that after completion 
 are to be made after grooving, but before the final cut is taken over them. 
 
 The turbine cases are to be tested in the shop after blading for several 
 hours under steam pressure, and they are to stand this test and all water- 
 pressure tests without any distortion which may be objectionable. 
 
 617. Tests after Fitting on Board. — After the machinery is fitted, complete, 
 on board the ship, a water test is to be made of all steam and exhaust pipes, 
 feed pipes, boilers, condensers, and air pumps, with the pipes and connections 
 of the preceding parts, to the pressures mentioned, and the turbine casings 
 to such pressures as may be approved ; this test is to be applied to the main 
 and auxiliary machinery, air reservoirs, air service and oil fuel pipes and 
 fittings generally. For this rest the lagging may be in place on all boilers 
 and castings, but all pipes and all joints are to be uncovered. The steam 
 pipes are to be tested separately to twice the full boiler pressure. They 
 should be tested up to the full boiler pressure with all important valves 
 shut, valves which may be subject to pressure on either side having the 
 
obu 
 
 MANUAf. OF MARIVE ENGINEERING. 
 
 pressure applied independently on either side. In continuing the test up to 
 the maximum test pressure specified, the valves should be left open. 
 
 Steel steam pipes may be lagged for water test on board ; but the flanges 
 are to be left uncovered. 
 
 Note. — During the test of the oil fuel suction pipes the strainer box covers 
 are to be removed to prevent the possibility of the double bottom compart- 
 ments being subjected to pressure. 
 
 618. Particulars of Tests for Various Parts. — The various parts, as well 
 as the corresponding spare gear, are to be tested by water to the following 
 pressures per square inch : — 
 
 Air reservoir bottles, separator and charging columns and air 
 service piping, with their spare gear, and the parts of the 
 pumps, etc., subjected to the maximum air pressure, 
 
 Each solid-drawn boiler tube and oil fuel heater tube of and 
 below 2 inches external diameter, .... 
 
 Each solid-drawn boiler tube above 2 inches external diameter, 
 
 Each condenser, distiller, and oil cooler tube, 
 
 The feed pump chambers, feed discharge pipes, valves, and any \ 
 apparatus subject to feed pump pressure, and main steam J- 
 pipes if of welded wrought iron, . . . . . ) 
 
 The water chambers of ash ejector pumps, discharge pipes, j 
 valves, etc., up to and including the shut off cock on .- 
 hopper, if so fitted, . . . . . . . J 
 
 Blown-out pipes and valves, the main steam pipes, stop valves," 
 strainers, and all fitting up to the high pressure ahead, 
 cruising, and high-pressure astern turbines ; the auxiliary 
 steam pipes, stop valves, separators, steam side of oil 
 fuel heaters, all boiler mountings and castings, and all 
 fittings which are subject to the full boiler pressure, . J 
 
 Boiler-room oil fuel pump barrels, with suction and delivery 
 chambers, suction and delivery pipes, and valves (ex- 
 cepting sluice valve and strainer on inner bottom ), filters, 
 and oil side of heaters, . ..... 
 
 The boilers when erected in the shop, after fitting on uoartl, 
 and in any case before raising steam, .... 
 
 Cylinders and all parts of auxiliary engines and evaporators, 
 subject to boiler pressure, ...... 
 
 The suction passage of the feed pumps, , , 
 
 The low-pressure cylinders of compound auxiliary engines, 
 
 The pump ends, discharge valve-boxes, and air vessels of the 
 fire and bilge engines, excepting those used in connection 
 with See's ash ejectors (if fitted), ..... 
 
 High -pressure ahead turbine 
 
 4,000 lbs. 
 
 2,500 „ 
 1,500 ,. 
 1,000 „ 
 Three times the 
 full boiler 
 pressure. 
 600 lbs., unless 
 otherwise 
 ordered. 
 
 Twice the full 
 boiler pressure. 
 
 400 lbs. 
 One and one- 
 half times the 
 full boiler 
 pressure 
 300 lbs. 
 235 M 
 
 250 
 
 Steam end and inlet cover, . . , 
 
 * • < 
 
 255 „ 
 
 The remainder, .... 
 
 1 • 
 
 200 „ 
 
 After completion, . 
 
 • 
 
 170 „ 
 
 Cruising turbine : — 
 
 
 
 Steam end and inlet cover, . . , 
 
 . t 
 
 255 ,. 
 
 The remainder, , , 
 
 1 • 4 
 
 200 „ 
 
 After completion, , 
 
 . t 
 
 170 „ 
 
 High-pressure astern turbine : — 
 
 
 
 Cylinder and inlet cover, . . , 
 
 . , 
 
 255 „ 
 
 Exhaust cover, if separate, . .. , 
 
 ( 
 
 200 „ 
 
 After completion, . . . , 
 
 • i 
 
 170 „ 
 
 Low-pressure turbine : — 
 
 
 
 Astern cylinders and inlet cover.", , 
 
 , , 
 
 50 , 
 
 Ahead cylinders and inlet covers, , 
 
 , , 
 
 30 ,. 
 
 Exhaust chamber and passages, . 
 
 , . 
 
 SO . 
 
 After completiou, 
 
 » 
 
 30 , 
 
STEAM TESTS. 851 
 
 Receiver pipe and valve between cruising and high-pressure 
 
 ahead turbine, 255 lbs. 
 
 Feed suction pipes, ........ 250 „ 
 
 Air compressing chambers of compressor for cleaning boiler 
 
 tubes, 150 „ 
 
 Air service piping for cleaning boiler tubes, .... 100 ,, 
 
 The pump chambers of distilling, forced lubrication, and 
 
 engine-room oil fuel pumps, and lime tanks, . . . 100 „ 
 
 All underwater valves up to and including 12 inches diameter 
 attached to the skin of the ship or shipbuilders' tubes, 
 excepting the boiler blow-out valves, on the working 
 side of the valve with the valve closed, and all pipes 
 except where otherwise specified, ..... 100 „ 
 
 Receiver pipes between the high-pressure ahead and low- 
 pressure turbines, ....... 60 „ 
 
 The air pumps, ........ 60 „ 
 
 The auxiliary exhaust pipes, evaporator cases, and the bilge 
 
 suction pipes, ........ 60 „ 
 
 Sluice valves and strainers in oil fuel suction pipes immediately 
 
 above inner bottom, ....... 60 ,. 
 
 Receiver pipes between high-pressure astern and the low- 
 pressure astern turbine, ...... 50 ,, 
 
 The working side of all underwater valves above 12 inches 
 diameter, with the valve closed, ..... 
 
 The air pump pipes and connections, ..... 
 
 Lubricating oil suction and drain pipes, .... 
 
 The main eduction pipes, ....... 
 
 Grease extractors — As may be ordered, but not less than 
 
 Hoppers of See's ash ejectors (if fitted) on maker's premises, . 
 
 The steam chambers of the condensers, the auxiliary engine 
 drain tanks if fitted, the steam chambers of distillers, 
 and distiller cases, ....... 
 
 The shaft casings before being fitted on the shafts, . 
 
 The shaft casings after being fitted on the shafts . . 
 
 (This test to be made with boiled linseed oil.) 
 
 The water chambers of the condensers, the circulating pump 
 cases, all valves and pipes between the corresponding sea 
 inlet and outlet valves, ...... 
 
 The feed, distiller test, drip, and oil tanks, .... 
 
 See's ash ejector (if fitted) and piping complete when erected) 
 
 in place on board, 1 above the hopper. 
 
 620. Valves and Valve Boxes. — All steam stop valve boxes are to be 
 tested to the full specified water pressure with the valves open and also 
 with the valves closed, when they must be tight on their seats. A subse- 
 quent test of these valve boxes is to be made under steam of the specified 
 working pressure, with the valves closed, and the pressure on the working 
 side of the valves. These tests are to be carried out in the shops before the 
 valves are fitted on board the ship, to show that the valves are steam tight. 
 The efficiency of the opening and closing gear is to be tried during these 
 tests. Underwater valve boxes are to be tested as a whole with tbe valves 
 open, and are also to be tested above the valve with the valve closed to the 
 full specified water pressure. 
 
 621. Steam Tests. — All high -pressure pistons and slide valves of the 
 auxiliary machinery, all reducing valves, are to be tested under steam on 
 the Engineer's works to the satisfaction of the Overseer before they are fitted 
 on board ship. 
 
 50 
 
 99 
 
 50 
 
 99 
 
 50 
 
 99 
 
 30 
 
 )> 
 
 30 
 
 99 
 
 30 
 
 99 
 
 30 
 
 99 
 
 30 
 
 . 
 
 30 
 
 9 » 
 
 25 
 
 9$ 
 
 10 
 
 99 
 
 To a head of 
 
 feet of 
 
 water 
 
852 MANUAL OF MARINE ENGINEERING. 
 
 623. Tests of Springs of Safety Valves.— Before the steam trials, all the 
 safety valves and their springs are to be taken out an replaced for exami- 
 nation, measurement, and test, to ascertain the exact load on the valves, 
 and that they are quite free in their seats ; in the case of the boiler safety 
 valves, this test is to be supplemented by raising steam in the boiler to full 
 pressure, the springs being ajudsted while the full pressure is maintained 
 in the boilers. The stops are then to be fitted ; after which, the accumu- 
 lation test of the valves is to be made. 
 
 624. Tests of Gauges. — Before proceeding on the steam trials the pressure, 
 vacuum, and compound gauges are to be removed, tested, and replaced by 
 the Engineers. 
 
 The Italian Government require the following tests to be made in the case 
 of machinery fitted in vessels registered in that country. W.P. is the working 
 pressure in the boilers : — 
 
 (a) High-pressure or 1st cylinders and valve chests, where W.P. is less 
 than 142 lbs. per square inch are to be tested to 1*5 W.P., and when W.P. 
 is over 142 lbs. to (W.P. + 71 lbs.). 
 
 (b) Second cylinders and valve chests are to be tested to W.P. 
 
 (c) Third „ „ „ 0-66 W.P. 
 
 (d) Fourth „ „ „ 0-33 W.P. 
 
 If the receivers are fitted with safety valves sufficiently large, and loaded 
 to the maximum W.P. of each cylinder, the test may be limited to twice 
 the loa 1 pressure. 
 
 (e) The steam jackets of all cylinders to twice the pressure they are 
 subjected to in working condition. If reducing valves are fitted, safety 
 valves must be fitted on the reduced pressure en . 
 
 (/) H.P. casing or cylinder of a turbine to 133 W.P. 
 
 {g) Astern-going turbine casing to W.P. 
 
 (h) L.P. turbine, admission end, 0'33 W.P. 
 
 (i) ,, exhaust end, 28 - 5 lbs. per square inch. 
 
 (?) Built condenser and all condensers for turbines are to be tested to 
 28 J lbs., while for cast-iron condensers for reciprocators 21 lbs. is sufficient. 
 
 (k) Safety valves and feed pumps to 2 W.P. 
 
 (I) Air and circulating pumps to 28| lbs. 
 
 (m) When all the connections are made on board and fitted complete, a 
 trial under steam in the usual way. 
 
 The British Corporation require the following water tests : — 
 
 (a) All evaporators and high-pressure feed heaters tested satisfactorily 
 at maker's works. 
 
 (b) Feed-water filters between pumps and boilers to 2*2 W.P. 
 
 (c) All auxiliary steam pipes excee ling 4 inches in diameter, and all main 
 and refrigerator machine steam pipes, not less than 2 W.P. when of copper, 
 and 3 W.P. when of iron or steel. 
 
 (d) All boiler feed pipes 20 per cent, higher than required for steam pipes 
 —that is, 2-2 W.P. 
 
 Lloyd's Register of Shipping require no water testing, except on the boilers, 
 etc., which are, as already stated — viz., the test pressure is 1*5 the W.P. 
 plus 50 lbs., and on pipes twice W.P. for copper, and three times for steel. 
 
TRIALS UNDER STEAM. 853 
 
 The Board of Trade also requires very little water testing beyond that 
 of the boilers and their pipes. H.P. turbine eases to 133 W.P., L.P. con- 
 denser end 30 lbs and the astern-going case to W.P. 
 
 Tests required by the Rules of the Germanischer Lloyds— The boilers 
 must be tested by hydraulic pressure 1*5 W.P., if the working pressure is 
 not more than 142 lbs. ; above that pressure it need be only (W.P. -f 71 lbs.), 
 where W.P. is the working pressure above atmospheric. Main steam pipes, 
 to twice W.P. 
 
 H.P. cylinders of all engines to (W.P. -j- 71 lbs.). 
 
 L.P. „ a compound engine to . . . .42*7 lbs. 
 
 L.P. „ a triple- and quadruple-compound engine to 28 5 
 
 M.P. ., a triple-compound engine to . . O66W.P. 
 
 1st M.P. „ a quadruple-compound engine to . .0*75 
 
 2ndM.P. „ „ „ . o-40 „ 
 
 Trials under Steam are made for various purposes ; some are of the 
 simplest kind, others are long, variable, and costly. In any case each trial 
 should have a definite object or objects, and when there are more than one 
 they should not be confused, so that one or more may be obscured. As a 
 rule, the primary object common to all concerned is to demonstrate the 
 actual and satisfactory fulfilment of a contract, whereby the engine builder 
 has undertaken to deliver machinery capable of performing certain functions 
 satisfactorily. The following are the principal things involved in such a 
 contract as is usual : — 
 
 (1) That the machinery is actually as described in the contract and 
 specification attached thereto, and accords with such plans as have been 
 submitted and approved. 
 
 (2) That it is approved by and passed by the Surveyors of the Board of 
 Trade, Lloyd's, or other body, where sanction is necessary for the ship to have 
 the status intended by her purchaser. 
 
 (3) That the engines will work satisfactorily at the full working speed 
 with only such attention and lubrication as is usual in similar ships on service, 
 and that they can be depended on to do so for a reasonable period. 
 
 (4) That the boilers will generate and supply a sufficient quantity of 
 dry steam at a pressure 1 per cent, less than the load on safety valves for the 
 above purpose, and for domestic and other needs, as designed. 
 
 (5) That the consumption of steam by the engines is not in excess of 
 that usual with similar machinery, and that the expenditure of fuel in the 
 boilers is also as low as usual. In other words, that the efficiency of both 
 boilers and engines is what it should be. 
 
 (6) That the power developed is in accordance with the contract, and 
 is obtained on conditions conducive to good and economic working. 
 
 (7) That the engines neither vibrate more than such engine should when 
 properly balanced, nor set up vibration in the hull of the ship. 
 
 (8) That they are easily under control, and can be started, stopped, and 
 reversed without fail in quite a quick way. 
 
 (9) That the condenser and its connections are all absolutely air-tight 
 and capable of maintaining a vacuum within 2 inches of the barometer when 
 the sea-water is not above 75° F. without an excessive quantity of cooling 
 water, and with the not-well temperature not less than 100° F. 
 
%"A MANUAL OP MARINE ENGINEERING. 
 
 (10) That the hot surfaces of boilers, pipes, cylinders, and all fittings 
 are properly and securely covered with a non-conducting material, so that 
 the temperature at the surface of the lagging or covering is not more than 
 120° F., nor 25° F. above that of the atmosphere, and generally that there is 
 no loss of heat by conduction or radiation that can be avoided by reasonable 
 and practicable means. 
 
 (11) That the ventilation of the engine-room, stokeholds, and all other 
 places where attendants have to keep watch or remain in for more than 10 
 minutes at a time shall be such that there is a plentiful supply of fresh air, 
 and the temperature does not exceed that of the external air by 20° F. nor 
 above 115° F. 
 
 (12) That the means for access to and escape from all such parts is prac- 
 tically free from risk to life or limb, and can be effected in a reasonably short 
 time in case of emergency. 
 
 (13) That all such places are or can be sufficiently lighted without notice 
 or loss of time to permit of the attendants performing their duties efficiently 
 and quickly. 
 
 (14) That the fire appliances are conveniently placed and accessible at 
 all times, so as not to be beyond use in case of fire in their neighbourhood, 
 and that they are efficient for the service desired from them. 
 
 (15) That in case of storms or heavy seas when there may be necessity 
 for " battening down," there is sufficient ventilation and means of access 
 and egress in machinery spaces for the well-being of the staff on watch. 
 
 (16) That the pumping arrangements generally, and especially those 
 affecting the safety of the ship, such as the bilge pumps, pipes, and fittings. 
 are satisfactory in size and position, accessible and easy to examine and 
 clean, and maintain generally in an efficient state. 
 
 (17) After the trials the boilers should be carefully examined to find 
 if there are any leaks or tendency to leak at any place, and if there have 
 been any general or local straining due to heat and pressure. 
 
 The machinery should be inspected with equal care to ascertain if there 
 be any signs of weakness or inefficiency in any part, or such indications of 
 wear or tendency to wear at the guides, slides, journals, pins, valve motion, 
 etc., or any need of modifying the adjustments of any of the working parts, 
 or necessity to improve the lubricating arrangements. 
 
 (18) The auxiliary machinery should be tested to prove that the capacity 
 and efficiency of each and every part is such as specified, and necessary for 
 the good working of the ship, and that they may be depended on to fulfil 
 their several duties on a prolonged voyage at least 25 per cent, longer than 
 usual on the service intended. 
 
 (19) That the arrangements for taking in fuel and supplying it to the 
 boilers are satisfactory, and that the bunkers are ventilated in such a way 
 as to preclude the possibility of gas accumulation with explosions. 
 
 So much for the trials of the machinery and the inspection of it after- 
 wards, so far as the manufacturer is concerned ; but it is only after ail a means 
 to an end, which is the propulsion of the ship at a certain speed in the most 
 efficient way possible — that is, to enable the highest speed to be attained 
 with the least expenditure of fuel, stores, wear and tear, and attention. The 
 saving of a fraction of a pound of fuel per T.H.P. per hour is a desirable thing. 
 
PROGRESSIVE TRIALS. 855 
 
 but it may be purchased dearly if there is a material increase in consumption 
 of oil, in the repair bill or in the staff charges. 
 
 The Admiralty Ship Trials up to thirty years ago were quite simple and 
 inexpensive, and it may be added by modern standards equally unsatis- 
 factory, chiefly on account of their incompleteness. The engines were tried 
 at moorings and moved at moderate speeds for a few hours to give every- 
 thing a " rub down." After the adjustments found to be necessary from 
 such a trial, the ship was taken to sea for a preliminary run, and if full power 
 was developed for even a short time it was deemed sufficiently satisfactory 
 to warrant the attempt of the official trial two days after. This official 
 trial was with the object of demonstrating the ability to attain rather than 
 to maintain the I.H.P. contracted for. and incidentally to find the speed of 
 the ship by making six runs on the measured mile, during which the revo- 
 lutions were determined by mechanical counters, sets of diagrams taken 
 from the cylinders by tested indicators, and the pressure of steam at boilers 
 and engines noted, as also the vacuum in condenser. If the six runs, half 
 with and half against the tide were performed satisfactorily, steam was 
 allowed to drop until half the boilers could be shut off without blowing of 
 the safety valves. The engines were then reduced in speed by notching 
 up the valve links or putting expansive valves in gear, so that full pressure 
 could be just maintained on the half set of boilers. Four runs on the measured 
 mile were taken under these circumstances on the same conditions as before 
 as to observations. The trial was then practically over, especially if there 
 were no novelties requiring special tests. Sometimes there would be a run 
 afterwards to let the pressure down, as there would be a run preliminary 
 to the measured mile to get the pressure up, and generally to "tune up/' 
 as it is now called, the machinery. The mean or true speed of the ship was 
 ascertained as it is to-day by taking a mean of means and the I.H.P., steam 
 pressure, revolutions, vacuum, etc., by taking the mean of the records made 
 on the measured mile of 6,080 feet. If the bearings had run fairly cool without 
 the application of the hose, the vacuum was within 4 inches of the barometer, 
 and the I.H.P. not materially less than contracted for the trial was satis- 
 factory, so far as the engineers were concerned. 
 
 The Trials of the Ordinary Merchant Ship were then pretty much the same, 
 and to-day those of a cargo boat are no more elaborate and often even less 
 so. In fact, the chief object of such trials now is to satisfy the owners 
 and the staff that the machinery is to be relied on to work satisfactorily in 
 its primary duty of causing the transport of the ship from port to port without 
 exceeding the daily consumption of fuel promised by the builders. If the 
 speed talked of was 10 knots, there is not any question raised if the speed 
 is 9-8 or 10 - 2 actually attained. 
 
 The Modern Steamship Trial, as introduced by the late William Denny, 
 of Dumbarton, is quite a different affair, and is probably due to the unique 
 facilities for such provided by nature in the Clyde estuary, where there 
 is ample room for the trials of the largest and fastest ships in deep water 
 with very little current, and sheltered from wind and waves, all within 
 easy reach of the shipbuilding yards and docks. 
 
 Progressive Trials, as made by Mr. Denny, and now by all builders of ships 
 whose speed and efficiency are of importance, are comparatively simple, and 
 involve little, if any. more cost than the old Admiralty ones. Incidentally 
 
856 MANUAL OF MARINE ENGINEERING. 
 
 these latter gave the two speeds and sets of facts appertaining to them : 
 an ! as at times the Admiralty wanted to know what a ship could do with 
 a quarter or a thir I of the boilers in action the trials thus made had the 
 desiderata of Mr. Denny's proposals — viz., three speeds with their facts and 
 figures were noted, and with them curves of I.H.P., revolutions, slip, and 
 efficiency could be plotted with the speeds as base lines. This was what 
 Mr. Denny showed was necessary to demonstrate the efficiency of the shij 
 under trial, and guide the designer in dealing with the future problems. It 
 also showed most clearly and .unmistakably the efficiency of the 
 machinery, as well as that of the ship, and perhaps what was of equal 
 and possibly greater importance, the efficiency or otherwise of the screw 
 propeller. 
 
 It is, therefore, desirable that all ships should be tried at three differing 
 speeds always, and when time and circumstances permit at four, in order 
 that there may be no question as to the trend of the different curves. If 
 such problems as arose on the effect of the depth of water on the speed and 
 power of a ship, even a greater number of trials are necessary. If three 
 only can be made, they should be at full power, at 78 per cent, of full power 
 revolutions, and at 40 per cent. If four trials can be done, they should be 
 at 25 per cent., 50 per cent., and 80 per cent. If a set of runs can be made 
 satisfactorily at a lower speed than 25 per cent., so much the better, but 
 to do so there should be practically no current or wind, and great care taken 
 to maintain both steam and vacuum with the slightest possible variation. 
 The revolutions should be counted mechanically during the whole time 
 the ship is on the mile, and for the exact time if possible, so that if the slip is 
 constant there should be always the same number of revolutions per mile 
 at every speed. If the engine is fitted with a speed indicator, it should be 
 watched for variation, especially . during the full speed runs for thereby 
 cavitation so called, or tendency to race from any cause, will be noted ; in 
 fact, it is always desirable to have an automatic recording revolution 
 indicator, which would show these variations and their magnitude. 
 
 Progressive Trials at Sea can be taken quite easily in such a way as will 
 enable those interested in checking the efficiency of ship and machinery from 
 time to time. Such trials should be made, because they are really of commer- 
 cial advantage, on all important oceangoing steamships, and may be carried 
 out at trifling expense by means of a good self-recording patent log kept 
 for the purpose, so that it may be looked on as a standard. Its error can 
 be easily found and noted, and so long as it is always used it will not matter 
 as to the amount if it is constant. The ship should run for twenty minutes, 
 or longer if possible, at three differing rates of revolution, which do not vary 
 during each trial. Indicator diagrams are carefully taken, or the torsion meter 
 is put into requisition. Revolutions are counted, steam pressure and vacuum 
 noted, as on the official trial. 
 
 It need hardly be said that a series of trials such as these should be made 
 in comparatively smooth water with the wind steady and as low as possible, 
 and that either in deep water or water whose depth does not vary much. The 
 time occupied need not be more than H hours, and the loss of distance run 
 will be so inconsiderable as to be no bar to the performance of such a useful 
 trial. From the data obtained a set of curves can be plotted in the usual 
 way, and compared with thoso furnished by the builders as the results of 
 
ENDURANCE TRIALS. 857 
 
 the trials at the measured mile, or with those made by the ship's staff when 
 in known thorough good order, with the bottom clean and fresh painted. 
 
 Endurance Trials are now deemed to be as necessary as the speed trials 
 for ail important ships. The British and most foreign naval authorities 
 require all ships to enduie satisfactorily prolonged trials at voyage and 
 cruising speedy as well as at full power, during which careful observation 
 is made of the consumption of fuel as well as the consumption and waste of 
 water ; the performance of all auxiliary engines and appliances are as 
 carefully noted as thai of the main machinery, and altogether the trials are 
 of an exhaustive as well as exhausting nature for all concerned, and very 
 different in every w r ay from the old speed trial trips of last century. 
 Destroyers and torpedo boats had formerly a prolonged run at 10 knots, 
 during which the coal was carefully measured and the ship's radius of action 
 gauged by the results. Latterly 1 4 knots is the usual speed for craft of this 
 kind to be tried at for this purpose. 
 
 Larger craft, including all cruisers, are subjected to a series of trials after 
 having had a basin or dock trial and a preliminary trial at varying speeds 
 to enable adjustments to be made by the contractors. 
 
 (a) The first trial is of 30 tours' duration, with the engines running so 
 as to develop about \ the horse-power ; during this trial, as at all others, 
 the amount of waste of water is noted by measuring the make up supplied 
 by the evaporators, and it is limited to 5 tons per 24 hours for each 1,000 
 H.P. developed. The coal is carefully weighed, and the total consumption 
 of water estimated from data obtainable. During this trial there are tests 
 made of the ability of the cruising turbines when so fitted, each lasting about 
 three hours. 
 
 (b) The second trial is also of 30 hours' duration, of which the first eight 
 is with the engines developing about four-fifths the full power, and the 22 
 hours remaining at about three-fifths. During this trial the water waste 
 must not exceed 4 tons per 24 hours for each 1,000 H.P. 
 
 (c) The third trial is at full power for eight consecutive hours, when the 
 waste must not exceed 3 h tons per 24 hours per 1,000 H.P. 
 
 (d) There are then various short trials to demonstrate the ability of the 
 machinery to respond quickly and efficiently to the orders from on deck. 
 To find what is the speed of the ship going astern when using half the boilers 
 as well as all of them. To note the time taken after the order is given to 
 stop the ship to bring her to a standstill. Also, the time taken in starting 
 at rest to attaining the maximum stern-going speed. 
 
 The pumping capacity and other duties of auxilaries. 
 
 (e) The engines are then opened up throughout and carefully examined, 
 the condensers tested by water, and after closing up again and adjusting. 
 
 (/) A final trial of 24 hours' duration, of which 12 must be a continuous 
 run at half power and 12 hours at such powers and revolutions as may be 
 directed by the Admiralty, with all the boilers alight ready for use, if not 
 actually in use during the whole time. 
 
 The allowance of extra feed above stated as about 2 33 per cent, of the 
 total consumption of the water of the main engines at the time. 
 
 Such trials as these are verv crucial and very costly, the fuel alone will 
 cost £1,300 to £1,500* with a cruiser of 25,000 I.H.P., and the wages and 
 keep of the staff together with the stores a further large sum ; but they 
 
 • At pre-war rates for S. Wales best coal. 
 
858 MANUAL OF MARINE ENGINEERING. 
 
 have been found necessary, and are continued to avoid the possibility of a 
 ship being taken over from contractors with latent defects which only by a 
 prolonged trial, such as would take place on service, would be made manifest. 
 
 In the Mercantile Marine no such crucial trials are deemed necessary, 
 except in the case of large and important ships having machinery novel 
 in design, size, or type, and intended for a long or important voyage service, 
 where break-down or even temporary stoppage would be fatal to her character 
 and damaging to the service and owner. There is, however, beyond all 
 this the interest of the builders of both ship and machinery to consider ; 
 their character and reputation are equally involved in the success or failure 
 of such ships, and they, therefore, prefer, or should do so, these drastic trials 
 which are nowadays insisted on. In the case of cross-channel steamers 
 a trial trip of 100 miles made at full speed without a stoppage on service 
 conditions is considered sufficient when the route on which she is' to run 
 does not exceed that distance. Before commencing this endurance test, 
 it is requisite and usual to have progressive measured mile trials, so that 
 by carefully recording the revolutions the actual speed can be checked against 
 that given by the patent logs, or that estimated by the navigating staff 
 from observation. 
 
 Since turbines have been employed to drive marine propellers, a new 
 set of trials has been devised to prove their ability to manoeuvre the ship. 
 For this purpose it is usual to note how long after the order is given to stop 
 and reverse the engines the ship comes to rest, and the distance run. This 
 is done at two or three rates of speed, as is also the turning of the ship by 
 moving one wing engine ahead and the other astern. The introduction 
 by Mr. Blenkinsop of his special patented reversing gear, whereby the middle 
 engine under these circumstances is always running ahead has shown that 
 the ship turns much quicker than when the centre screw is at rest. Perhaps 
 the reversion to twin screws from the triple and quadruple ones is partly 
 due to the inability of such small screws to quickly control the motion of the 
 ship, as well as to the advantages derived by geared turbines. 
 
 The machinery of express passenger steamers is usually opened up after 
 an endurance trial, but not to the same extent as in the naval ships ; indeed 
 it is only where trouble has been experienced or suspected that a careful 
 examination is deemed necessary, except that the cylinders and main valves 
 are in any case examined, as also the internal parts of the boilers, otherwise 
 it depends very much on the custom of the superintending engineer how 
 much more will be distributed for his inspection. It was the custom formerly 
 to permit of the water service being in action throughout a trial, rather as 
 a precautionary measure than a remedial one. To-day trials should be, 
 and generally are, run without the application of any water, and no engine 
 can be deemed satisfactory that will not run without heating of bearings, 
 pins, or guides when lubricated only ; it may be that the oil is bad or that 
 the lubricating arrangements are faulty, these defects can be remedied or 
 at least excused and put right as soon as possible : lack of bearing surface, 
 poor quality of white metal, or faulty finish to journal and brasses cannot 
 be treated lightly. Forced lubrication, however, should be followed as the 
 most successful method for avoiding trouble and ensuring efficiency. 
 
THE DIESEL OIL ENGINE. 859 
 
 APPENDIX A. 
 
 The Diesel Oil Engine. 
 
 The Diesel Type of Oil Engine is, no doubt, the best of the oil engines for use 
 on board ship, inasmuch as it works on a small consumption of a safe and 
 comparatively cheap fuel, so that it is economic in working, and premits of 
 a large radius of action on a given weight of fuel ; the engine-room staff is 
 no larger than required with steam engines, and there is no boiler-room crew 
 at all. Moreover, so far, it is the only internal combustion engine that has 
 been subjected to exhaustive trials at sea in large sizes * The space occupied 
 by an oil engine installation is considerably less than that of a steam engine 
 one, but there is no corresponding saving in weight. 
 
 Single-acting 3-cylinder 4-cycle heavy type slow-running Diesel engines, 
 250 to 700 B.H.P., weigh 460 lbs. per B.H.P. 
 
 Single-acting 4-cylinder 4-cycle heavy type slow-running Diesel engines, 
 300 to 900 B.H.P., were 400 to 435 lbs., while the 2,000 H.P. modern are 265. 
 
 Single-acting 4- to 6-cylinder 2-cycle marine type at 200 revolutions per 
 minute, 150 to 224 lbs. per I.H.P. The larger at 150 revolutions, 140 lbs. 
 
 Single-acting 4- to 6-cylinder 2-cycle marine high-speed engines, 90 to 
 100 lbs. per I.H.P. 
 
 The very lightest marine Diesel engines, as made in Germany for sub- 
 marines and torpedo boats, are 35 to 40 lbs. per I.H.P., which is about the 
 same as that of the steam turbine installations in destroyers, including boilers, 
 etc. British machinery in " oilers " is 50 to 56 lbs. per S.H.P. 
 
 The Efficiency of the Diesel Engine is not so high as that of steam engines, " 
 qua engine, for the thermal is 42 to 48 per cent, at most, and the mechanical 
 about 84 per cent, at full power, and at dead slow only 33-4, so that in com- 
 paring it with a turbine or steam reciprocator it is not sufficient to take the 
 consumption per I.H.P. as the criterion. Moreover, the ratio of maximum 
 to mean pressure (5 to 6) is so great, and as there is only one impulse per 
 revolution with the two-cycle engine, the torque is very irregular ; conse- 
 quently the efficiency of the screw driven by it must be less than when driven 
 by a steam reciprocator, and much less than with a turbine. 
 
 But to compare the efficiency of the whole installation of an oil plant 
 with that of a steam it is necessary, of course, to take into account the effi- 
 ciency of the boilers. For this purpose consumption of fuel is the criterion, 
 and efficiency is arrived at by comparing the output with the theoretical 
 possibilities of the fuel. 
 
 The heat equivalent of a horse-power per hour is — 
 
 ^^ x 60 = 2,547 B.T.U. 
 
 778 
 
 The marine Diesel engine consumes 0-4 lb. of fuel per S.H.P., the thermal 
 
 value being 18,500 B.T.U. , the S.H.P. requires 7,400 per hour. The possible 
 
 H.P. with this quantity is 7,400 -f- 2,547, or 2-905 ; the efficiency, therefore, 
 
 * The so-called semi-Diesel, but really a low-pressure, heavy oil engine with ignition 
 attachments is now very popular, and made in larger sizes than formerly. 
 
860 MANUAL OF MARINE ENGINEERING. 
 
 of such an engine is 344. By the same test a modern turbine consuming, 
 with its auxiliaries 10'5 lbs. of steam requiring 0-7 lb. of oil fuel per S.H.P. 
 hour will have an efficiency of 19-7 ; and a quadruple reciprocator worked 
 with high superheated steam, about 17 per cent., or only half that of the 
 oil engine. 
 
 The Rate of Revolution of this Oil Engine is, as a rule, higher than that 
 of a steam engine ; in fact, it must be high to be efficient and trustworthy 
 with a limited number of cylinders ; but it is, of course, considerably less 
 than that of the turbine. For standard light engines up to 500 H.P. about 
 500 revolutions is a common speed; up to 1,200 H.P. they are about 400. 
 The ordinary heavy engine of the mercantile marine is run at 275 to 300 
 revolutions per minute up to 500 H.P., 215 to 260 up to 1,200 H.P., the large 
 engines in ocean-going craft 150 revolutions on trial conditions is usual, and 
 in the very largest, as in fig. 325, 100 to 125 revolutions. Naval oil engines 
 as in small craft are usually run at about 400 revolutions, the piston speed 
 being 960 feet per minute, while the piston speed of large merchant ships is 
 750 to 850 feet. 
 
 Every Diesel Engine must have an air-compressing pump of some kind 
 to provide the air for pulverising and injecting the fuel into the cylinder, 
 whose contents have been compressed to 500 or 600 lbs., and even higher 
 in some cases, where the fuel has a high flash point, such as coal-tar oil, 
 which has 400° F., whereas that of the ordinary residues is about 200° F. 
 Sometimes these pumps are driven by an independent steam engine or electric 
 motor, when steam or electricity is available for the purpose ; latterly, 
 especially for marine purposes, the pump is driven from a crank at the fore 
 end of the main engine. The quadruple Reavell is a favourite one for this 
 purpose, but often there is only an ordinary two-stage pump, as shown in 
 the illustration (fig. 320). 
 
 The Two-cycle Engine is bound to have a low-pressure pump to supply 
 the air necessary to scavenge the oil cylinder and fill it with fresh air prior 
 to compression. These pumps are quite simple, and not unlike the old 
 air pump of the surface condenser, and may be, and are, worked in the 
 different ways, as found convenient formerly for them. They may be worked 
 by levers, by a crank or cranks, or direct from the piston or crosshead of the 
 engine itself (v. fig. 323). The English engineers as a rule prefer to work 
 these pumps with levers, etc. Sulzers have adopted the plan of having the 
 pump in line with the oil cylinders, and driving it from a crank below it. The 
 Nuremberg, Augsburg Company drive direct from the oil trunk piston, as- 
 shown in fig. 321, where the pump chamber is immediately below the oil 
 cylinder, and the ratio of their diameters 1-4 to 1-6, so that the capacity 
 permits of a 25 per cent, excess of air to be delivered. By this latter 
 arrangement the oil cylinder trunk is exposed to the cooling action of the 
 air, and thereby prevented from getting overheated, and, on the other hand, 
 the air is somewhat heated before entering the oil cylinder. The air-pump 
 trunk in this case contains the connecting-rod end and its gudgeon pin, so 
 that it is really the guide, and takes the thrust of the connecting-rod ; it 
 is, however, cool, and the space in it large enough to permit of an ample 
 surface of gudgeon and means for efficient lubrication, which is not the case 
 when the ordinary piston trunk of the oil cylinder, as in fig. 49. The fuel 
 consumption of the two cycle is 6 per cent, higher than that of the four. 
 
THE DIESEL OIL ENGINE. 
 
 861 
 
 In the more Modern Designs of most makers of these engines there is a 
 trunk, with a piston-rod, and its crosshead, slipper guide, etc., as found in the 
 
 f^TfpE^,C^33Sli__jC 
 
 Fig. 320. — Two-cycle Single-acting Marine Engine for Heavy Oil — Diesel System. 
 
 (German Design.) 
 
 steam engine (v. figs. 324-5). The engine is two-cycle, single-acting, and the 
 oil cylinder quite open at the bottom, so as to be subject to the cooling action 
 
 55 
 
862 
 
 MANUAL OF MARINE ENGINEERING. 
 
 of the ventilating air. All large marine engines, whether single- or double- 
 acting, should be of this design, so that the connecting-rod end and its gudgeon 
 are free from the surrounding trunk, and in view, where they can be examined 
 from time to time, as well as be efficiently lubricated. 
 
 The Double-acting Oil Engine has not come yet, and although on the 
 Continent it has been tried on quite a large scale, there is reason to believe 
 it has not been altogether satisfactory. If it is on the two-cycle system 
 
 Fig. 321. — Cylinder and Ait Pump of Two-cycle Single-acting Diesel Engine 
 
 (German Design.) 
 
 the cylinder is necessarily of great length, and the engine itself so high (or 
 long if horizontal) as to be prohibitive on most ships. It is also almost an 
 impossible thing for an engine working at such high temperatures as the 
 Diesel to keep reasonably cool with an explosion or conflagration at each 
 stroke, even with water-jackets and water-cooled pistons ; then, too, there 
 are quite serious difficulties with the piston-rod glands even on the double- 
 
THti JUNKER ENGINE. 
 
 863 
 
 acting four-cycle system, which must be magnified when the two-cycle is 
 adopted. The mean pressure in a Diesel two-cycle engine is usually 100 tc 
 110 lbs. per square inch, and may be in some cases as high as 125 lbs. ; the 
 referred mean pressure of a quadruple reciprocating steam engine is 40 to 
 50 lbs., or less than half that of the oil engine ; but the pressure in their L.P. 
 cylinder is only 10 to 12 lbs., so that the size of the oil cylinders may be 
 quite small compared with those of the steam ones, in spite of being single- 
 
 Fig. 322. — Two-cycle Single-acting Marine Engine for Heavy Oil — Diesel System. 
 
 (Italian Design.) 
 
 acting. It will be seen, therefore, that with a piston speed of 800 feet a 
 cylinder 30 inches in diameter can develop 933 I.H.P., so that a six-cylinder 
 engine would produce 5,600 I.H.P. when running on the two-cycle system. 
 The maximum temperature is over 1,000° F. 
 
 The Junker Engine (fig. 323) is an ingenious adaptation of the Wigzell 
 engine (v. fig. 285) to obtain a virtual double-acting marine oil engine. It 
 
864 
 
THE DIESEL OIL ENGINE. 865 
 
 will be seen that to each engine there are two cylinders and three cranks ; 
 each cylinder has two pistons, which move in opposite directions, the one 
 operating on the centre crank, the other on the outer pair of cranks, which 
 are on each side of and opposite to it. Air is compressed between each pair 
 of pistons as they approach each other ; oil is sprayed in between them, and 
 ignites in the way usual with the Diesel. At about seven-eiehths the stroke 
 one piston opens a port to exhaust, and very shortly after another port at 
 the opposite end of the cylinder is opened by the other piston, to admit the 
 blow through or scavenging charge of air, which is supplied by air pumps 
 worked direct from the crosshead as shown ; this blast clears the cylinder 
 space between the pair of pistons of foul air, and leaves it filled with pure 
 air to be compressed as before. In this case each cylinder is single-acting 
 and on the two-cycle system, but there are two cylinders in line tandem 
 operating alternately, so that the combustion is equivalent to one double- 
 acting cylinder. There are three such engines to one shaft, and the air- 
 compression cylinder is worked from the fore end of the crank-shaft in the 
 common way. This engine, besides being ingenious, has some good points, 
 such as no cylinder covers to give trouble at the joints ; no air admission or 
 exhaust valves to cause noise and wear and tear ; the cylinders are ventilated, 
 and so kept reasonably cool. On the other hand, the crank-shaft is an 
 expensive one, and the number of side rods and connecting-rods large. This 
 engine, like the Wigzell, will be perfectly balanced against vertical inertia 
 forces, and nearly so for the horizontal ones, and so ought to run quietly, 
 without vibration, and have an even torque ; but also, like the Wigzell, it 
 is exceedingly high, so that with a long stroke of piston it would be very 
 inconvenient in most ships. 
 
 The Reversing of these Oil Engines is generally accomplished by using 
 the compressed air in one or more of the oil cylinders to press on the pistons 
 with the oil supply shut off. This requires storage of air in bottles, or the 
 continuous charging of air vessels by the pumps ; the latter is the obviou? 
 plan for large engines. The admission of the air can be regulated by hand, 
 or the cam shaft can be shifted through an angle sufficiently great (about 
 30°) to time the delivery of both compressed air and fuel for the reverse 
 motion of the engine. The starting and reversing in each case being effected 
 by compressed air acting as the moving force on the oil cylinder pistons. 
 
 In Manoeuvring with Oil Engines there is not the same elasticity observ- 
 able that is possible with steam engines. With a steam engine the starting 
 and reversing are quickly and more certainly done ; and whereas it wi'l run 
 dead slow for hours with the cranks just passing the dead points, the ordinary 
 oil engine cannot be trusted in the same way, for it is getting an impulse 
 very occasionally, and should a sudden load be put on the propeller, as is 
 the case often in a seaway, or by the action of the rudder in a twin-screw 
 ship, there is great danger of the engine stopping at a critical moment and 
 failing to start again promptly. Then there may be trouble with deposit 
 in the cylinders and pistons when working under these conditions, *o that 
 in foggy weather great care will have to be exercised, for a temporary resort 
 to a light and highly inflammable oil is hardly permissible on shipboard. 
 No doubt, however, the methods now employed are sufficiently good for all 
 practical purposes, even if not so convenient as those of the steam engine. 
 The Diesel Oil Engine requires better workmanship and material i.han 
 
866 MANUAL OF MARINE ENGINEERING. 
 
 is sufficiently good for the best of steam engines, inasmuch as the cylinder, 
 3tc, is subject to both high pressure, high temperature, and to a certain 
 or rather uncertain amount of shock ; there must be perfect accuracy in 
 the fit of the valves and gear by which the fuel supply is controlled, and the 
 amount of compression must be constant for steady and economic working ; 
 hence, while a steam engine may and often does perform its work with a 
 fair degree of efficiency when in a state of disrepair, an oil engine cannot 
 do so. A leaky piston and distributing valve is a source of loss of a sort 
 in a steam engine, but it is one of danger to the oil engine. The consumption 
 of fuel in a steamship does not materially increase in rate during a long 
 voyage, say to Japan and back, but it is more than doubtful if the same will 
 be said of the internal combustion engine driven ship. The difference in pres- 
 sure per square inch on the sides of the pistons of a triple and quadruple 
 engine will not exceed 120 lbs., while in the Diesel engine it is 500 to 600 lbs., 
 and sometimes even higher ; if, therefore, the pistons are not perfectly 
 tight, the loss in the latter case will be much more severe than in the former, 
 and flame escaping into the engine-room, a more serious thing than steam 
 passing from one cylinder to another, or even to the condenser. For this 
 reason the two-cycle oil engine is safer than the single-acting four-cycle engine. 
 and any oil engine having cylinders open to the engine-room should be 
 avoided, or when used should be kept in perfect repair. 
 
 The Double-acting Engine on the two-stroke cycle is much to be desired 
 for large power, inasmuch as the cylinders may be fewer in number or less 
 in diameter ; this type of engine is, however, necessarily less efficient than 
 the four-cycle single-acting one. Dr. Diesel stated that such an engine with 
 three cylinders, developing 850 I.H.P., had been made at Nuremburg, and 
 undergone its trials satisfactorily ; and that there was in course of manu- 
 facture a similar engine with cylinders so large that over 2,000 J.H.P. is 
 developed in each. This engine has cylinders 31| inches in diameter, with 
 a piston-stroke of 41§ inches, and it is intended to run it at 160 revolutions 
 per minute. 
 
 Dr. Diesel recommended that engines of this kind should have six cylinders, 
 " to ensure a regular turning moment and balancing," and added that " this 
 number cannot be considered abnormally high ; on the contrary, it must 
 be accepted as the most suitable and proper number," and concluded that 
 " the day of the large marine engine is already very near at hand." The 
 s.s. " Selandia," of 2,500 I.H.P., had had her trials, and completed a maiden 
 voyage fitted with the Diesel engine ; this and subsequent voyages with her 
 and other similar ships have proved the reliability of these engines in capable 
 hands, and justified the optimism of the inventor and his friends. 
 
 It is interesting to note that these engines are capable of working satis- 
 factorily with other than mineral oils and extracts ; at a trial an engine 
 was fed with nut oil fuel, and worked quite well in every way. The calorific 
 value of this oil is 15,510 B.T.U., and the consumption of it was only 0-53 lb. 
 per hour. The efficiency of this engine was, therefore, 31 per cent., and 
 its thermal efficiency probably about 40 per cent. 
 
 It is possible to work a Diesel engine with any inflammable liquid which 
 has a flash point under 450° F., and will consume completely, so that gaseous 
 product only remains, which will wholly escape through the exhaust without 
 leaving any solid or semi-solid deposit. Mineral oils, as pumped or dis- 
 
OIL ENGINE DRIVEN SCREW SHIPS. 
 
 867 
 
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868 MANUAL OF MARINE ENGINEERING. 
 
 charged from the wells, are unsuitable, as they contain highly volatile con- 
 stituents, which not only render them dangerous to store, but dangerous 
 to use in the engine. Their flash point should not be under 200° F., and 
 they should be free from water, sulphur, and asphaltum, otherwise they 
 will give trouble. Asphaltum is thrown down in the cylinder as hard 
 refractory coke, which cuts and scores the pistons, cylinders, and valves ; 
 the sulphur deposits tend to glue up the piston rings, valves, etc., so that 
 when the engine has cooled down after working it is found set fast, and 
 cannot be started without cleaning* 
 
 The mineral oil residues, after the distillation of petrol, paraffin, and 
 lubricating oils, are good for engine fuel ; other crude oils are treated by 
 exposure to the air, to get rid of the volatile parts, and reduce the sulphur, 
 etc., to small proportions. It is, therefore, correct to speak of the Diesel 
 as a heavy oil engine, rather than a crude oil one, as is often done. 
 
 The reversing is accomplished by means of a sliding cam shaft, on which 
 are four cams to each cylinder for " ahead " going, and four for " astern " 
 going. These cams operate the admission of air, fuel, compressed air, and 
 exhaust. There is a lever by means of which the cam levers are lifted out 
 of gear while the shaft is moved longitudinally, and another lever which 
 shuts off the fuel and the highly compressed air for feeding it during the 
 time the engine is being worked by compressed air at 300 lbs. pressure. The 
 cam shaft is driven by special gearing from the crank-shaft, and can slide 
 about 2 inches from " ahead " gear to " astern." It is said the reversing 
 can be accomplished in 10 seconds by a skilled attendant. The air for this 
 purpose is compressed to 300 lbs pressure by a two-stage pump driven by 
 an independent 4-cylinder oil engine, which also is arranged to drive a dynamo 
 for the purpose of generating electricity, whereby the other auxiliaries 
 throughout the ship may be worked. There are two such auxiliary oil 
 engines, each capable of developing 250 B.H.P. at 230 revs, per min. The 
 air for delivering the fuel is taken from the reservoirs into which these auxili- 
 aries deliver it, and further compressed to 750 lbs., or even more, by a small 
 pump worked by a crank- pin at the fore end of each crank-shaft- 
 
 These engines are lubricated throughout by means of force pumps, and 
 the pistons and rods are kept cool by the oil circulating through them. 
 
 The Westgarth-Carel engine, as was fitted in s.s. " Eveston," is shown 
 in fig. 324. It was a four-cylinder four-crank arrangement on the single- 
 acting two-stroke system, having the air pumps for supplying the air to 
 scavenge and charge the cylinders at the back of the engine, and worked 
 by means of levers in the usual marine fashion : the high-pressure air for 
 manoeuvring the engine and spraying the oil fuel is obtained from a quad- 
 ruple Eeavell pump worked from the fore end of the main crank-shaft. 
 Fig. 326 is a section of one of its cylinders illustrating the methods adopted 
 in this case for fitting the liners, the water jacketing and piston cooling, and 
 the arrangement of the valves, all of which will be seen to be simple and 
 efficient. The starting and reversing of these engines are effected by means of 
 compressed air ; the valves for this purpose are operated by means of levers 
 and cams on an independent shaft driven by the main crank-shaft by the 
 interposition of a vertical shaft and skew-wheel gearing, which permits of 
 
 * It is said that with a higher compression (70 atmospheres) these fuels can be com- 
 pletely consumed 
 
WESTGATITH-CAREL OIL ENGINE. 
 
 869 
 
 change of Tear] by raising or lowering this vertical shaft through a few inches. 
 Mr. Westgarth elected to work the deck and auxiliary machinery by steam 
 generated in a small boiler bv means of the waste heat from the exhaust 
 
 50 
 
 > 
 
 -3 = 
 
 X 
 
 a 3 
 
 
 of the main engines, thereby getting over some practical difficulties with 
 economy in working as well as prime cost. 
 
 Fig. 325 is the vertical transverse section of the four-stroke single-acting 
 engines of the twin-screw ship, " Fiona," 5,219 tons and 4,000 I.H.P. They 
 
870 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Fig. 325.— Sectional View of Main Engine of Twin s.s. " Fiona," 4,000 I.H.P., 835 N.H.P. 
 
 Each eneine has 6 cylinders 4-cvcle, 29 - 9 diameter x 43 - 5 stroke. 
 
WESTGARTH-UAREL OIL ENUIN K. 
 
 871 
 
 are the largest yet made for the mercantile marine, being collectively of 
 835 N.H.P., each engine having six cylinders, 29-9 inches diameter and 
 43-5 inches stroke of piston, and running at 140 revolutions per minute when 
 at full speed. Their consumption of fuel oil is 0-32 lb. per I.H.P. hour of 
 good quality. 
 
 Doxford's Double-piston Oil Engine is shown on fig. 328 as a single unit. 
 It is on the two-cycle single-acting system in general principle like that of 
 
 EL VALVE 
 
 SCAVENCINCAIR VALVE 
 
 SCAVENCINC AIR VALVE 
 
 WATER JACKET 
 
 EXHAUST BELT 
 
 Fig. 326. — Section of Cylinder of Westgarth-Carel Oil Engine. 
 Diameter, 20 inches ; stroke, 36 inches. 
 
 fig. 323, but with only one cylinder per unit. It is, therefore, without the 
 double-acting effort of that engine, but nevertheless it possesses the other 
 good features, such as working balance, without heavy stress on the columns 
 and framing, and the absence of the " heads," which become so hot and 
 
872 
 
 MANUAL OF MARINE ENGINEERING. 
 
 dangerous. There is, however, a multiplicity of working parts per unit 
 which are costly, heavy, and require attention, but in spite of these the engine 
 is an attractive one, and should prove satisfactory on service 
 
 Scavenging Air, 
 Valve 
 
 Exhaust 
 
 Fuel Injection 
 Valve 
 
 Exhaust 
 
 Scavenging, 
 Air Valve 
 Starting 
 Air Valve 
 
 Scavenging 
 Air Valve 
 Starting 
 Air Valve 
 
 Fuel Injection 
 Vj/ve 
 
 Fig. 327. — Cylinder and Heads of Double-acting Engine. 
 
 Starting^ . Relief Valve Fuel Injection 
 
 Valve 
 
 Scavenging 
 Valve 
 
 Transverse Sections. 
 
 Fig. 327a. — Cylinder and Head of a Two-stroke Sbgle-acting Engine. 
 
 The Semi-Diesel Engine has been in use now for some years ; it has beer 
 well tried and proved to be specially adapted for using the safe high flasi 
 point oils and residuals in small power units where an ordinary Diesel engine 
 
Fig. 328. — Doxford's Double-piston Three-crank Oil Engine Unit. 
 
874 
 
 MANUAL OF MARINE ENGINEERING. 
 
THE SEMI-DIESEL ENGINE. 875 
 
 would be unsuitable as well as costly. It is now made by several well-known 
 engine builders substantially, in all essentials, as shown by fig. 328. They 
 may be on the four-stroke cycle, but are nearly always on the two-stroke 
 system. The air for scavenging and recharging the cylinder every time is 
 drawn in through valves to the air-tight casing of the engine itself, the lower 
 side of the piston acting as the pump ; the air thus stored is allowed to flow 
 into the cylinder through a port which is opened on the final passage of the 
 piston, after it has unclosed the exhaust ports ; the air thus admitted is 
 diverted upwards by the form of the piston so as to thoroughly drive out the 
 products of combustion. It is, of course, obvious that only the bare amount 
 of air is driven into the chamber, so that the charge is not very pure nor very 
 dense, and. therefore, the combustion is not so efficient as that in the Diesel 
 engine itself. It has, however, the merit of simplicity and consequent cheap- 
 ness, inasmuch as the two-cycle operation is obtained without the use of 
 separate air pumps, the compression is moderate, and the ratio of maximum 
 to mean pressure comparatively small. On the other hand, the thermal 
 efficiency is lower than that of the Diesel, and an igniter is necessary, but there 
 is no need for such small piston clearance, and any variation in it due to wear 
 of brasses, etc., is of no moment ; the valves and gear are simple, and the 
 fuel is injected with less pressure. Altogether, this engine possesses qualities 
 which make it a cheap, useful, and economic one for the propulsion of com- 
 paratively small craft in the hands of an unskilled and partially trained staff. 
 Moreover, it can be put in operation on shortest notice, being easily started 
 with a little petrol or ether. 
 
 In Beardmore's engines the compression is limited to 150 lbs., so that 
 after explosion it is only 300 lbs. The air for feed fuel is, therefore, 
 only 400 to 500 lbs. Bolinders have made these engines extensively with 
 units up to 16-5 inches diameter and 20 inches stroke. Beardmore's units 
 are up to 14 inches diameter and 15 inches stroke. Plenty & Son have made 
 units up to 13-2 inches diameter and 14 inches stroke. 
 
 In these engines the mean pressure is about 52 lbs., and the ratio of 
 maximum to mean pressure about 5, practically the same as in the Diesel 
 engine. 
 
876 MANUAL OF MARINE ENGINEERING. 
 
 APPENDIX B. 
 
 Lloyd's Kegister Rules for Internal Combustion Engines 
 for Marine Purposes (other than Diesel Type). 
 
 general. 
 
 Section 1. — In vessels propelled by internal combustion engines, the rules 
 as regards machinery will be the same as those relating to steam engines so 
 far as regards the testing of material used in their construction and the fitting 
 of sea connections, discharge pipes, shafting, stern tubes, and propellers. 
 
 CONSTRUCTION. 
 
 Section 2. — 1. The following points should be observed in connection 
 with the design of the engines. 
 
 2. The shaft bearings, connecting-rod brasses, the valve gear, the inlet 
 and exhaust valves must be easily accessible. 
 
 3. The reversing gear and clutch must be strongly constructed and easily 
 accessible for examination and adjustment. 
 
 4. In engines of above 60 B.H.P. which are not reversible and which are 
 manoeuvred by clutch, a governor, or other arrangement must be fitted to 
 prevent racing of the engine when declutched. 
 
 5. Efficient positive means of lubrication (preferably sight feed) must 
 be fitted to each part requiring continuous lubrication. 
 
 6. If the engines are of the closed-in type, they must be so fitted that the 
 contained lubricating oil can be drained when necessary, and in wood vessels 
 an easily drained metal-lined tray must be fitted to prevent leakage of either 
 fuel oil or lubricating oil from saturating the woodwork. 
 
 7. Carburettors, where petrol is used, and vaporisers, where paraffin is 
 used, should be so designed that when the engine is stopped the fuel supply 
 is automatically shut off. If an overflow is provided in the carburettor or 
 vaporiser, a gauze covered tray with means of draining it must be fitted 
 to prevent the fuel from flowing into the bilges. 
 
 Strong metallic gauze diaphragms should be fitted either between the 
 carburettor (or vaporiser) and cylinders, or at the air inlets. 
 
 8. If the ignition is electric, either by magneto or by coil and accumulator, 
 all electric leads must be well insulated and suitably protected from mechanical 
 injury. The leads should be kept remote from petrol pipes, and should not 
 be placed where they may be brought into contact with oil. 
 
 The commutator must be enclosed ; and the sparking coils must not 
 be placed where they can be exposed to explosive vapours. 
 9. No exposed spark gap should be fitted. 
 
 10. In paraffin and heavy oil engines where lamps are used for ignition 
 or for vaporising, these lamps should be fixed by some suitable bracket, and 
 the flames enclosed when in use. 
 
 11. The circulating pump sea suction is to have a cock or valve on the 
 vessel's skin placed on the turn of the bilge in an easily accessible position, 
 
LLOYD'S REGISTER RULES FOR INTERNAL COMBUSTION ENGINES. 877 
 
 and the circulating pipe is to be provided with an efficient strainer inside 
 the vessel. The discharge overboard is to be fitted with a cock or valve od 
 the vessel's skin if it is situated under or near the load line of the vessel. 
 
 12. A bilge pump worked by the engines or an independent power-driven 
 bilge pump is to be fitted, to draw from each part of the vessel. In open launches 
 this bilge pump may be omitted provided suitable hand pumps are fitted. 
 
 13. The cylinders are to be tested by hydraulic pressure to twice the 
 working pressure to which they will be subjected. The water jackets of the 
 cylinders to 50 lbs. per square inch, and the exhaust pipes and silencer to 
 10 lbs. per square inch. 
 
 14. The exhaust pipes and silencer should be efficiently water cooled 
 or lagged to prevent damage by heat, and if the exhaust is led overboard 
 near the water-line, means must be arranged to prevent water from being 
 syphoned back to the engine. 
 
 15. The machinery must be tried under full working conditions, the 
 report stating the approximate speed of vessel, the number of revolutions 
 of the engines at full power, both ahead and astern, and the lowest number of 
 revolutions of the engines which can be maintained for manoeuvring purposes. 
 
 Rules for determining sizes of shafts (v. Chap. xii.). 
 
 Fuel Tanks and Connections. 
 
 Section 4. — 1. Separate fuel tanks are to be tested with all fittings, to a 
 head of at least 15 feet of water. If pressure feed tanks are employed, they 
 are to be tested to twice the working pressure which will come on them, 
 but at least to a head of 15 feet of water. If the tanks are made of iron or 
 steel they should be galvanised. 
 
 2. Strong and readily removable metallic gauze diaphragms should be 
 fitted at all openings on petrol tanks. 
 
 3. Paraffin or heavy oil tanks, not used under pressure, are to be fitted 
 with air pipes leading above deck. Pressure-feed tanks and tanks containing 
 petrol, should be provided with escape-valves discharging into pipes leading 
 to the atmosphere above deck. The upper ends of all air pipes are to be 
 turned down and pipes above 1 inch diameter are to be provided with gauze 
 diaphragms at the end. 
 
 4. No glass gauges are to be fitted to fuel tanks containing either petrol, 
 paraffin, or heavy oil. 
 
 5. Filling pipes are to be carried through the deck so that the gas dis- 
 placed from the tanks has free escape to the atmosphere. 
 
 6. Separate fuel tanks should be provided with metal-lined trays to 
 prevent any possible leakage from them flowing into the bilges, or saturating 
 woodwork. Arrangements are to be provided for emptying the tanks and 
 draining the trays beneath them: For petrol tanks the trays must have 
 drains leading overboard where possible or they should be gauze-covered 
 travs with means for draining them. 
 
 7. All fuel pipes are to be of annealed seamless copper with flexible bends. 
 Their joints are to be conical, metal to metal. A cock or valve is to be fitted 
 at each end of the pipe conveying the fuel from the tank to the carburettor 
 or vaporiser. The fuel pipes should be led in positions where they aie pio- 
 
 56 
 

 80 
 
 N 
 
 x D 2 VS 
 
 
 40 
 
 N 
 
 x D 2 VS 
 
 878 MANUAL OF MARINE ENGINEERING. 
 
 tected from mechanical injury, and can be exposed to view throughout their 
 whole length. 
 
 8. The engine-room, and the compartment in which the fuel tanks are 
 situated, are to be efficiently ventilated. 
 
 9. An approved fire extinguishing apparatus must be supplied. 
 
 Lloyd's Eegister Rules for Diesel Engines. 
 
 Nominal Horse-Power. — The following rule is to be used for determining 
 the power of Diesel engines in regulating the fees for their survey, viz. : — 
 
 N x D 2 VS 
 N.H.P. = - Qft - in the case of single-acting engines of the 4-cycle type. 
 
 „ » ,. 2-cycle „ 
 
 „ double-acting „ 2-cycle „ 
 
 Where D = diameter of cylinder in inches, 
 
 S = stroke of piston in inches in ordinary reciprocating engines, 
 = twice the stroke of piston in the case of engines of the " Junker '• 
 type, 
 N = number of cylinders. 
 
 Rules for the Construction and Survey of Diesel Engines and 
 
 their Auxiliaries. 
 
 Section 1. — In vessels propelled by Diesel oil engines, the rules as regards 
 machinery will be the same as those relating to steam engines, so far as regards 
 the testing of material used in their construction and the fitting of sea con- 
 nections, discharge pipes, shafting, stern tubes, and propellers. 
 
 construction. 
 
 Section 2. — 1. In vessels built under Special Survey and fitted with Diesel 
 engines, the engines must also be constructed under Special Survey. 
 
 5. Any novelty in the construction of the machinery is to be reported 
 to the Committee and submitted for approval. 
 
 6. The auxiliary engines used for air compressing, working dynamos 
 and ballast, or other, pumps, are also to be surveyed during construction. 
 
 7. In cases where the designed maximum pressure in the cylinders does 
 not exceed 500 lbs. per square inch, the diameters of the crank shaft of the 
 main engines are not to be less than those given by the following formula : — 
 
 Diameter of crank shaft = ^D 2 X (AS-f-BL), 
 
 where D = diameter of cylinder, 
 S = length of stroke, 
 
 L = span of bearings adjacent to a crank, measured from inner edge 
 to inner edge. 
 
Lloyd's register rules for dtesel engines. 
 The values of (A S -f B L) are as given in the following table : — 
 
 TABLE I. 
 
 879 
 
 4-C'ycle Single-acting Engine. 
 
 2-Cycle Single-acting Engine. 
 
 Values of the Coefficient. 
 
 4 or 6 cylinders. 
 
 8 cylinders. 
 10 or 12 cylinders. 
 16 cylinders. 
 
 2 or 3 cylinders. 
 
 4 cylinders. 
 
 5 or 6 cylinders. 
 8 cylinders. 
 
 •089 S + -056 L 
 •099 S + -054 L 
 •111S+ 052 L 
 •131 S + -050 L 
 
 For auxiliary engines of the Diesel type the diameters may be 5 per cent. 
 less than given by the foregoing formula. 
 
 8. In solid forged shafts the breadth of the webs should not be less than 
 1-33 times and the thickness not less than 0-56 times the diameter of the shaft 
 as found above, or, if these proportions are departed from, the webs must be 
 of equivalent strength. 
 
 9. Where no flywheel is fitted, the diameter of the intermediate shaft 
 must not be less than given by the formula : — 
 
 Diameter of intermediate shaft = coefficient VT) 2 x S, 
 
 where D = diameter of cylinder, 
 S = stroke of piston, 
 
 and the value of the coefficient is given by the following table : — 
 
 TABLE II. 
 
 4-Cycle Single-acting Engine. 
 
 2-Gycle Single-acting Engine. 
 
 Value of the Coefficient. 
 
 4 cylinders. 
 6, 8, 10, or 12 cylinders. 
 16 cylinder*. 
 
 2 cylinders. 
 3, 4, 5, or 6 cylinders. 
 8 cylinders. 
 
 •456 
 •436 
 •466 
 
 Where the stroke is not less than 1*2 times nor more than 1-6 times the 
 diameter of the cylinder, (-735D + -273 S) maybe taken instead of ^D 2 x S. 
 
 10. In cases where flywheels are fitted, the following value of the co- 
 efficient may be taken for determining the size of the intermediate shaft 
 abaft the flywheel shaft. 
 
 TABLE III. 
 
 4-Cycle Single-acting Engine. 
 
 2-C'ycle Single-acting Engine. 
 
 Value of the Coefficient. 
 
 4 cylinders. 
 
 6 ' „ 
 
 8 „ 
 10 „ 
 12 
 16 „ 
 
 2 cylinders. 
 
 3 " „ 
 
 4 „ 
 
 5 „ 
 
 6 „ 
 8 „ 
 
 •405 
 •400 
 •409 
 •420 
 •427 
 •461 
 
880 MANUAL OF MARINE ENGINEERING. 
 
 11. The diameter of the flywheel shaft must be at least equal to that 
 of the crank shaft 
 
 12. The diameter of the thrust shaft measured under the collars must be 
 at least f ^ths that of the intermediate shaft. The diameter may be tapered 
 off at each end to the same size as that of the intermediate shaft. 
 
 13. The diameter of the screw shaft must not be less than the diameter 
 
 / "03 P'n 
 of the intermediate shaft (found as above) multiplied by ( -63 -\ = 
 
 but in no case must it be less than 1-07 T, ^ -"- 
 
 where P = the diameter of the propeller in inches, 
 
 T = the diameter of the intermediate shaft in inches. 
 
 The size of the screw shaft is intended to apply to shafts fitted with con- 
 tinuous liners the whole length of the stern tube. If no liners are used, or 
 if two separate liners are used, the diameter of the screw shaft should be 
 fiths that given above. 
 
 The diameter of the screw shaft is to be tapered off at the forward end 
 to the size of the thrust shaft. 
 
 14. If the designed maximum pressure in the cylinders exceeds 500 lbs. 
 per square inch, the diameters of the shafting throughout must be increased 
 
 ,, j. 3 /maximum pressure in lbs. per square inch 
 in the proportion of ./ — ^— . 
 
 15. Where the cylinder liners are made of hard, close-grained cast iron 
 of plain cylindrical form, accurately turned on the outside as well as bored 
 on the inside so that their soundness can be ascertained by inspection, and 
 their thickness at the upper part is not less than ^§- of the diameter of the 
 cylinder, they need not be hydraulically tested by internal pressure. If, 
 however, they are made of complicated form, the question of testing must 
 be submitted. 
 
 16. The water jackets of the cylinders, and the water passages of the 
 cylinder covers and pistons, must be tested by hydraulic pressure to 30 lbs. 
 per square inch, and must be perfectly tight at that pressure. 
 
 17. The exhaust pipes and silencers must be water-cooled or lagged by 
 non-conducting material, where risk of damage by heat is likely to occur. 
 
 18. The cylinders are to be fitted with safety valves loaded to not more 
 than 40 per cent, above the designed maximum pressure in the cylinders 
 and discharging where no damage can occur. 
 
 19. The air-compressors and their coolers are to be made so as to be easy 
 of access for overhaul and adjustment. 
 
 20. In single-screw vessels, an auxiliary air-compressor is to be provided 
 of sufficient power to enable the main engines to be kept continuously at 
 work when the main compressor is out of action. 
 
 If the manoeuvring gear is arranged so that the engines can be kept con- 
 tinuously at work with some of the cylinders out of action, the auxiliarv 
 compressor need only be of sufficient power to enable the engines to be 
 kept at work under these conditions. 
 
 In twin-screw engines in which two sets of compressors are fitted, the 
 auxiliarv compressor must be of such size as to enable it to take the place 
 of either of the main compressors. If in such engines each main compressor 
 
LLOYDS REGISTER RULES ON DIESEL ENGINES. 881 
 
 is sufficiently large to supply both engines, a smaller auxiliary compressor 
 will be sufficient. 
 
 A small auxiliary compressor, worked by a steam engine, or by an oil 
 engine not requiring compressed air, is to be fitted for first charging the air 
 receivers. 
 
 21. At least one high-pressure air receiver is to be arranged with con- 
 nections to enable it to be used for fuel injection, in case the working receiver 
 of either main engine is out of use from any cause. 
 
 22. The circulating pump sea suction is to be provided with an efficient 
 strainer which can be cleared inside the vessel. 
 
 AIR RECEIVERS. 
 
 Section 3. — 1. Compressed-air receivers for starting air are to be supplied 
 of sufficient capacity to permit of twelve consecutive startings of the engines 
 without replenishment. 
 
 2. Cylindrical receivers for containing air under high pressure, used either 
 for starting or for the injection of fuel in oil engines, may be made either of 
 seamless steel or of welded, or riveted, steel plates. 
 
 3. Quality oj Material. — If made of welded, or riveted, steel plates, the 
 ordinary rules regarding steel material for boilers apply, which provide that 
 where welding is employed, either in the longitudinal seams or at the ends, 
 the material must have a tensile strength not exceeding 30 tons per square 
 inch (Section 4, par. 7, Eules for Engines and Boilers). In these cases the 
 welding must be lap welding ; neither oxy-acetylene nor electric welding 
 will be permitted. 
 
 4. In the case of seamless receivers, the rules for material will be the 
 same as for boiler shells, but the permissible extension may be 2 per cent, 
 less than that required with boiler plates. 
 
 5. Tensile and Bend Tests are to be made from the material of each receiver. 
 When they are welded or riveted, the tests may be made, and the thickness 
 verified, before the plates are bent into cylindrical form. In the case of seam- 
 less receivers, the thicknesses must be verified by the Surveyor before the 
 ends are closed in, and at this time the Surveyor shall select and mark the 
 test pieces required from either of the open ends of the tube. The test pieces 
 are to be annealed before test, so as to properly represent the finished material. 
 
 6. The permissible working pressure for welded or seamless receivers is 
 to be determined by the following formulae : — 
 
 Maximum working pressure in lbs. per square inch 
 
 C x S x (T 2) 
 
 = — ^ - for thicknesses of f inch and above. 
 
 = =^ — — — ' for thicknesses below f inch, 
 
 where S = minimum tensile strength of the steel material used, in tons per 
 square inch, 
 T = thickness of the material, in sixteenths of an inch, 
 D = internal diameter of cylinder, in inches, 
 C = coefficient as per table : — 
 
882 MANUAL OF MARINE ENGINEERING. 
 
 Coefficient 77 for seamless receivers of thickness of f inch and above, 
 „ 69 „ „ „ „ below f inch. 
 
 ,, 54 „ welded „ „ of § inch and above. 
 
 „ 48 „ „ „ „ below f inch. 
 
 7. For flat ends welded into the cylindrical shells, the thickness must 
 not be less than 
 
 T = r7 VF, 
 
 where T = thickness, in sixteenths of an inch, 
 D = internal diameter in inches, 
 P = working pressure, in lbs. per square inch. 
 
 8. The permissible working pressure for receivers made of riveted steel 
 plates is to be determined by the rules regulating the working pressure of 
 boilers. 
 
 9. Each welded or seamless receiver shall be carefully annealed after 
 manufacture, and before the hydraulic test. 
 
 10. Each welded or seamless receiver shall be subjected to a hydraulic 
 test of twice the working pressure, which it shall withstand without permanent 
 set. 
 
 11. Each receiver made of riveted steel plates is to be tested by hydraulic 
 pressure to twice the working pressure for pressures up to 200 lbs. per square 
 inch. Where higher working pressures are used, the test pressure need not 
 be more than 200 lbs. per square inch above the working pressure. 
 
 12. All receivers above 6 inches internal diameter must be so made that 
 the internal surfaces may be examined, and, wherever practicable, the openings 
 for this purpose should be sufficiently large for access. Means must be pro- 
 vided for cleaning the inner surfaces by steam, or otherwise. 
 
 13. Each receiver which can be isolated must have a safety valve fitted, 
 adjusted to the maximum working pressure. If, however, the air compressor 
 is fitted with a safety valve so arranged and adjusted that no greater pressure 
 than that permitted can be admitted to the receivers, they need not be fitted 
 with safety valves. 
 
 14. Each receiver must be fitted with a drain arrangement at its lowest 
 part, permitting oil and condensed water to be blown out. 
 
 PUMPING ARRANGEMENTS. 
 
 Section 4. — 1. The requirements of the pumping arrangements for the 
 various holds, double bottoms or other ballast tanks, etc., are to be the same 
 as required in steam vessels of the same size. 
 
 2. The engines are to be fitted with two bilge pumps, which are to be so 
 arranged that either can be overhauled while the other is at work. In twin- 
 screw vessels one bilge pump upon each engine will be approved. These 
 pumps are to be arranged to draw from all compartments. Independent 
 power-driven pumps may be fitted in lieu of these, if desired. 
 
 3. A steam pump, or equivalent power-driven pump, is also to be pro- 
 
Lloyd's register rules for diesel engines. 883 
 
 vided with connections to enable it to draw from all compartments and 
 from the sea. It must be arranged to discharge overboard and also on deck 
 to the fire service pipes. It must have at least one suction to the engine- 
 room bilge distinct from those connected with the bilge pumps, so that it 
 may be used for pumping from the engine-room when the bilge pumps are 
 being used upon other parts of the vessel. 
 
 4. In addition to the above, where water ballast is used, the water-ballast 
 pump must have one direct suction from the engine-room bilges. (This is 
 in lieu of the bilge injection required with steam engines.) 
 
 GENERAL. 
 
 Section 5. — 1. For the ordinary fuel tanks the requirements of Section 49 
 will apply. The daily service and other separate tanks must be tested, with 
 all their fittings, with a head of water 12 feet above their highest points. 
 They must be fitted with air pipes discharging above the upper deck. If they 
 are fitted with glass gauges for indicating the quantity of oil contained in them, 
 arrangements must be made for readily shutting off the gauges in the event 
 of the breakage of the glass, and from preventing any damage from leakage 
 of oil. 
 
 2. Special attention must be given to the ventilation of the engine-room. 
 
 3. If the auxiliaries are worked by electricity, the cables in connection 
 with them must be in accordance with the rules for cables for electric light. 
 
 4. It is recommended that all pipes conveying fuel oil should, as far as 
 possible, be made of steel or iron, rather than copper, owing to the rapid 
 corrosion of copper pipes when using oil containing sulphur. 
 
 spare gear. 
 
 Section 6. — The articles mentioned in the following list will be required 
 to be carried, viz. : — 
 
 1 cylinder cover complete for the main engines, with all valves, valve 
 
 seats, springs, etc., fitted to it. 
 In addition, one complete set of valves, valve seats, springs, etc., for one 
 
 cylinder of the main and of the auxiliary Diesel engines, and fuel 
 
 needle valves for half the number of the cylinders of each engine. 
 1 piston complete, with all piston rings, studs, and nuts for the main 
 
 engines. 
 In addition, one set of piston rings for one piston of the main and of the 
 
 auxiliary Diesel engines. 
 
 1 complete set of main skew wheels for one main engine. 
 
 2 connecting-rod, or piston-rod, top-end bolts and nuts, both for the 
 
 main and the auxiliary Diesel engines. 
 2 connecting-rod bottom-end bolts and nuts, both for the main and for 
 
 the auxiliary Diesel engines. 
 2 main bearing bolts and nuts, both for the main and for the auxiliary 
 
 Diesel engines. 
 1 set of coupling bolts for the crank shaft. 
 
884 MANUAL OF MARINE ENGINEERING. 
 
 I set of coupling bolts for the intermediate shaft. 
 
 1 complete set of piston rings for each piston of the main and of the 
 
 auxiliary compressors. 
 1 half set of valves for the main and for the auxiliary compressors. 
 1 fuel pump complete for the main engine, or a complete set of all the 
 
 working parts. 
 1 fuel pump for the auxiliary Diesel engine, or a complete set of all working 
 
 parts. 
 1 set of valves for the daily fuel supply pump. 
 1 set of valves for the water-circulating pumps. 
 1 set of valves for one bilge pump. 
 
 1 set of valves for the scavenge pump, where lift valves are used. 
 A quantity of assorted bolts and nuts, including one set of cylinder cover 
 
 studs and nuts. 
 Lengths of pipes suitable for the fuel delivery and the blast pipes to the 
 
 cylinders, and the air delivery from the compressors to the receivers, 
 
 with unions and flanges suitable for each'. 
 
RULES FOR INTERNAL C0MEUS170N ENGINES. 885 
 
 APPENDIX C. 
 
 Bureau Veritas Rules for Internal Combustion Engines on 
 
 Shipboard. 
 
 Propelling engines above 300 H.P. to be of reversible type. All engines 
 to be provided with, a friction brake and turning gear worked by hand. Above 
 1,000 H.P. the turning gear must be worked by a motor. 
 
 Cylinder covers must be fitted with safety-valves loaded to twice the 
 maximum working pressure. 
 
 The pistons are to be cooled by efficient means. 
 
 In double-acting engines the piston-rod metallic packing is to be auto- 
 matically adjustable. 
 
 Cylinder water-jackets are to be provided with a pressure gauge and 
 drain cocks. 
 
 Each main engine to be provided with an air compressor capable of starting 
 the engine 16 consecutive times without renewing supply of compressed air. 
 
 In twin-screw engines each air compressor should be of sufficient capacity 
 to maintain the necessary pressure in the reservoir for working the two engines ; 
 an auxiliary compressor must be provided sufficient to keep up the air supply 
 in reservoirs when the engines are stopped, or when other compressors are 
 not available ; it must, therefore, be driven by an independent motor. 
 
 Fuel oil pumps are to be strongly constructed, and easily overhauled 
 when the engine is working. 
 
 A hand pump is to be provided for pumping fuel oil from the bunkers 
 through filters to the gravitation tanks. 
 
 A governor is to be fitted to prevent racing of main engines. 
 
 Spare circulating pump, capable of passing 2-25 gallons per H.P. per 
 hour, to be fitted independently of the others. 
 
 The tanks must have a steam-heating pipe near suction pipe, and be 
 fitted with a safety-valve when worked under pressure. The suction pipes 
 are to have duplicate independent filters, having triple perforated plates with 
 gauze wire between them, and with covers easily removed for examination 
 and cleansing. 
 
 All pipes for fuel oil and compressed air are to be of steel or solid-drawn 
 copper with approved joints. 
 
 The cylinders are to be fitted with forced lubrication, and when of high 
 revolutions the main bearings are to have forced lubrication. 
 
 Diameter of crank-shaft = C v^D 2 x S inches. 
 D is the diameter of cylinder and S the piston stroke, both in inches. 
 
886 
 
 MANUAL OF 11ARINE ENGINEERING. 
 
 s 
 
 C. 
 
 S 
 5* 
 
 ! 
 
 2-00 
 
 0-4930 
 
 1-40 
 
 0-."250 
 
 1-95 
 
 0-4975 
 
 1-35 
 
 0-5275 
 
 1-90 
 
 0-5000 
 
 1-30 
 
 0-5300 
 
 1-85 
 
 0-5025 
 
 1-25 
 
 0-5325 
 
 1-80 
 
 0-5050 
 
 1-20 
 
 0-5350 
 
 1-75 
 
 0-5075 
 
 115 
 
 0-5375 
 
 1-70 
 
 0-5100 
 
 MO 
 
 0-5400 
 
 1-65 
 
 0-5125 
 
 1-05 
 
 0-5425 
 
 1-60 
 
 0-5150 
 
 1-00 
 
 0-5450 
 
 1-55 
 
 0-5175 
 
 0-95 
 
 0-5500 
 
 1-60 
 
 0-5200 
 
 0-90 
 
 0-5550 
 
 1-45 
 
 0-5225 
 
 0-85 
 
 0-5000 
 
 The above rules and values of C apply to four-cycle single-acting engines 
 having 3, 4, 6, and 8 cylinders, or single-acting two-cycle ones with 4 or 6 
 cylinders, or double-acting two-cycle ones with 1, 2, or 3 cylinders. 
 
 The diameter of the intermediate shafts must be not less than — 
 
 0*85 x diameter of crank-shaft — 0*2 = d. 
 
 The diameter of thrust shaft at collars = 106 X d. 
 
 The diameter of propeller-shaft = d x -f- 
 
 diameter of screw in inches 
 
 loo ' 
 
TURBINES ON SHIPBOARD. 887 
 
 APPENDIX D. 
 
 TURBINES ON SHIPBOARD. 
 
 Turbines are now employed very extensively for skip propulsion and, where 
 high speed is necessary, almost to the exclusion of the reciprocator. So 
 long as it drove the propeller shaft direct, its rate of revolution was very 
 high for propeller efficiency, and too low for its own efficiency ; consequently, 
 its consumption per unit of power was much higher than obtained in land 
 service. Further, while at full speed it compared quite favourably with the 
 triple and quadruple reciprocator, at slower speed it failed to do so, and at 
 cruising speed it was so much worse that special engines, either turbines or 
 reciprocators, had to be fitted to Destroyers and Cruisers for use at those 
 times when low power only was required for lengthy periods. Moreover, 
 direct driving turbines were useless for slow ships of any kind, but especially 
 were they out of the running in cargo ships whose speed is generally about 
 10 knots. It was, the r efore, imperative, if the turbine was to be of general 
 service on shipboard, that its speed must not be restricted to that of the 
 propeller, and thereby be prevented from working with the minimum amount 
 of steam, nor must the propeller be compelled to run at a rate of revolution 
 inconsistent with efficiency on service and good handling of the ship in 
 harbour, etc. The driver and the driven, therefore, had to be separated, 
 so that each could move at its appropriate and most efficient rate of revolu- 
 tion. This has been effected, and the connection between them made in three 
 different ways. 
 
 Sir Charles Parsons, as has been shown on p. 100, etc., introduced the 
 spur wheel and pinion, formerly so common with the early screw ships, 
 and not altogether unknown on a paddle ship. With helical teeth machine- 
 cut and accurately designed and fitted, the objectiens formerly obtaining 
 with wheel gearing have practically ceased, so that now even the most powerful 
 turbines running at high rate of revolution (4,000 per minute) can be connected 
 by double gear as shown in Figs. 329, 330, and 331, whereby the rate of the 
 revolution of propeller is little more than would be the case with the reci- 
 procating engines. 
 
 Experience has proved that the wear and tear on these helical steel teeth 
 is very small indeed, and although there have been some broken teeth, it was 
 attributable to the kind of steel first used being unsuitable, and to their 
 faulty form at their roots ; both these defects were removed, and the 
 difficulties of the lubrication have also been surmounted. 
 
 The Ratio of Turbine to Propeller Revolution is necessarily great in cargo 
 ships, inasmuch as the screw seldom exceeds 75 revolutions per minute, 
 while the turbine may be run at 3,500 and even higher when of short lengths, 
 as the Curtis generally is. The screw of the " Cairncross " on voyage makes 
 only 62 revolutions ; the ratio of the gearing is 26, so that the turbines make 
 only 1,612 revolutions, which is quite too slow for modern practice. Fig. 329- 
 
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 MANUAL OF MARINE ENGINEERING. 
 
 is an example of what obtained on fast twin-screw express steamers with 
 geared turbines. In this case the screws run at 310 revolutions per minute ; 
 the gearing ratio is quite small (6-4), inasmuch as the H.P. turbines ran only 
 at 1,984, while the L.P. ran at 1,380. It will be seen that the pair of pinions 
 are side by side in this case, and are not long like the " Vespasian's " (vide 
 p. 100), nor had they the bearing interposed to support them, as is now 
 generally necessary for good working. The frontispiece shows this arrange- 
 ment in the E.M.S. " Tuscania," whose gearing is single with a ratio of 12-4 — 
 the screw running 137 revolutions at full speed, the efficiency being 97 per 
 cent., -with steam pressure 200 lbs., and vacuum 28-3 inches, the total S.H.P. 
 is 10,900, and the consumption of steam only 11-25 lbs. per S.H.P. hour, and 
 speed 17-65 knots. 
 
 Fig. 331.— Double Reduction Gear of the Battleship " Nevada," U.S.A. Navy. 
 
 The spur wheels are 10 feet in diameter, and 60 inches long on teeth 
 faces ; the pinions being 10 inches diameter, their revolutions are 1,700. 
 
 To Avoid such Large Spur Wheels, to economise space, and generally to 
 ensure good working, double gearing has been employed in the U.S. Navy with 
 considerable success. It has now come into favour on this side of the Atlantic, 
 and admits of employing turbines with a revolution as high as 4,000 per minute 
 with screws at 81, the ratios being each 7, so that the combine ratio is 49. 
 The main spur wheels and pinions for very large power are each as much as 
 36 inches long, with a good bearing between them. 
 
 Fig. 331 shows the ingenious and very compact arrangement applied 
 
TURBINES ON SHIPBOARD. 
 
 891 
 
 on the U.S.A. battleship " Nevada " for coupling the cruising turbines to 
 the screw shafts, the ratio in this case being 23-85. 
 
 Fig. 332 illustrates the MacAlpine arrangement of oscillating frames in 
 which the pinions run, and permit of an automatic adjustment of each pair 
 of pinions, so that they divide the work evenly between them. Considerable 
 skill is also necessary in adjusting the teeth so that they bear evenly throughout 
 their length when running. 
 
 The Helical Teeth of the wheel and pinions as originally made by Sir 
 Charles Parsons were cut to an angle of 22-5°, with face line parallel to the 
 axis ; experience with these dictated a larger angle as being generally more 
 satisfactory in working and less noisy. In fact, they are now practically 
 noiseless when cut to 30°. The pinions are ' generally of nickel or chrome 
 nickel steel, while the spur wheels have a band of ordinary forged mild steel 
 as a tyre secured to the cast-steel or cast-iron centres. There should be no 
 sharp angles at the roots of the teeth, their bases being well-rounded ; in fact, 
 the space between them may be semi-cylindrical. 
 
 Fig. 332. — Double Gearing with Floating Frames as in U.S.A. Naval Ships. 
 
 The Turbine with Geared Connection can be run at any speed desired, 
 and, consequently, there is no need for the very numerous stages the designer 
 is compelled to put in the slow-working reaction turbine of the Parsons' 
 type hitherto generally found on shipboard ; it may be of the Curtis impulse 
 variety with quite a few rows of blades, etc., and as economic of steam as 
 those on central stations on shore. In the case of the M Normannia," the 
 steam consumption per S.H.P. is only 12-0 lbs., as against 15-1 lbs. of her 
 competing ship, the " Sarnia," with ordinary direct-coupled turbines on the 
 same service. The large cargo steamer, wk Cairncross," with geared turbines 
 on a single screw shaft steamed side by side of the sister ship with triple 
 compound engines, and used only 12-57 lbs. of steam per hour, as against 
 the 15-18 lbs. with the triples. There is, of course, no question now as to the 
 economic superiority of the geared turbines over simple ones or over quadruple 
 engines so far as steam per H.P. is concerned. The mechanical efficiency is 
 also quite satisfactory, as the gearing accounts for only 1-5 to 2 per cent. 
 
892 MANUAL OF MARINE ENGINEERING. 
 
 of the gross power, and as that is S.H.P. taken beyond the gearing, the com- 
 parisons with the others are quite fair ones. The reversing has still to be 
 done with a special ' k stern-going " turbine, as no one has yet ventured on 
 trying a " gate " arrangement as in motor cars, or even the change gear 
 as in small craft at sea with oil engines. 
 
 Electrical Driving of a Propeller is the second system of connection, and 
 was suggested by many thoughtful minds in the early days of turbine pro- 
 pulsion, but the objections on the score of prime cost, space occupied, risk 
 of danger to motors and generators, etc., sufficed to prevent experiment on 
 a large scale. In the U.S.A., however, the advocates of this system were 
 more fortunate, inasmuch as the naval authorities were persuaded to fit 
 their new collier, " Jupiter," with such an installation. 
 
 The " Jupiter " is of 20,000 tons displacement, and has attained a speed 
 of 15 knots by means of turbines whose S.H.P. is 6,940. The propeller runs 
 at 116 revolutions per minute, and the steam consumption is only 12 lbs. 
 per hour, against 14 lbs. of a sister ship with reciprocators. It is, however, 
 claimed for the " Ljungstrom " turbine that with it and a similar electrical drive 
 the consumption will be only 7-4 lbs. of steam per hour on the main engines 
 only, as against 10-1 for double-geared and 9-0 for hydraulic drives, all of them 
 using superheated steam ; the Ljungstrom by 260° added and the hydraulic 
 by 200° F. These are compared with a quadruple engine using 12-71 lbs. with 
 no superheat, and an ordinary turbine drive with no superheat of 11-86 lbs. 
 
 With the Ljungstrom Turbine there are, as shown on p. 140, Fig. 58, 
 two rotors having their blades interlaced and moving in opposite directions ; 
 each drives an alternating dynamo, which produces a three-phase current 
 of about 50 alternations per second at 800 volts. The alternators of each 
 are electrically locked so as to ensure an exactly equal speed and power 
 with each half of the turbine. It is intended that there shall be two such 
 sets to avoid dangers from breakdowns. The two alternators in each set 
 will work in parallel. There is only one pair of poles to each of them ; they 
 supply current to two motors having five pairs of poles, which consequently 
 revolve at a fifth the speed of the turbine. The motors have pinions with 
 helical teeth, which drive a spur wheel on the screw shaft. In this way the 
 ratio of revolution of turbine to propeller is nearly 50, so that a high "ate 
 obtains with the turbine (3,500 revs.). The screw can be reversed by means 
 of the motors, so that there is no need of an astern-going turbine or reci- 
 procator for this purpose. But again, the installation is not so elastic in 
 operation as is the steam one with reciprocators, as slowing down requires 
 additions in the way of resistances, etc. The same remarks and arrange- 
 ments more or less prevail with all the electrically driven schemes ; more- 
 over, with the short Curtis turbine, 3,500 or even a higher rate of revolution 
 is possible. A ship of considerable size is being fitted with such an instal- 
 lation as the above of 5,400 S.H.P. with twin screws. 
 
 In the case of the Ljungstrom turbine the electrical is certainly the most 
 convenient for coupling it to a screw shaft ; but it can, and perhaps will, 
 be effected generally by wheel gearing, inasmuch as by the interposition 
 of a hunting wheel between one pinion and the spur wheel of the first pair 
 of gears, their original reverse motion is converted into combined action 
 on the spur wheel shaft, which carries the second pinion geared into the 
 spur wheel on the screw shaft. 
 
TURBINES ON SHIPBOARD. 893 
 
 The Ljun gstro m turbine is a very interesting instrument, and has been 
 most carefully designed and developed, so that the objections to it from a 
 practical point of view have been fairly met and most of them ingeniously 
 got over ; time, however, alone will permit of coming to definite conclusions 
 as to its place in marine engineering. The efficiency of the system is fairly 
 high, and from experiments made in Sweden with sister ships the consump- 
 tion was 35 per cent, in its favour, as against the direct turbine drive, being 
 ■ only 0-89 lb. of coal per S.H.P., while the consumption in another ship with 
 triple expansion engines was 1-09 lbs., when added superheat was 280° F. and 
 Howden's forced draught. 
 
 Hydraulic Transmission is the third system. It was introduced by Dr. 
 Fottinger, and extensively used in Germany. Fig. 332 shows the impeller 
 and motor of the centrifugal pumps, and Fig. 333a the arrangement of the 
 system in a ship. Dr. Fottinger claims that while mechanically it is not 
 so efficient as either the wheel gearing or the electrical transmission, it is 
 practically nearly as economic, inasmuch as he is able to supply warm feed 
 water to the boilers ; and for prolonged astern movement, it will be much 
 more economic and safer than the reversed turbine. The efficiency, however, 
 when admitting the heat gain is at best 92 per cent, in the large installations 
 and less in smaller ones. 
 
 The Modern Turbine on Shipboard is either altogether of the impulse 
 order or is a reaction with an impulse leading portion. In this country the 
 Parsons and the Curtis are those most generally used ; there are a considerable 
 number of engine builders who construct turbines, and each has a design 
 with some distinctive features of its own. The Ljungstrom is gaining favour 
 with those who set economy of steam consumption as the criterion, but its 
 employment with copper ingot (g.m.b.) at £130 per ton, the electrical method 
 for the present is somewhat handicapped. 
 
 In the U.S.A. the Curtis is now a favourite turbine, but the improved 
 Parsons has still a good following. The Zoelly. Rateau, and some others 
 have also their supporters. 
 
 France has stuck to the Rateau as a rule, and as it is a good turbine, 
 they do well to support their very able countryman and distinguished 
 engineer. 
 
 Germany before the war used the Parsons turbine freely — the Zoelly, 
 A.E.G., and some other forms of the Parsons or Curtis in Teuton guise were, 
 however, coming into general use. But Germany, like the U.S.A., stuck 
 to the reciprocators for warships much longer than the British did ; in fact, 
 at one time the U.S.A. naval authorities reverted to the reciprocator, but 
 that was largely due to their adopting methods and designs which were less 
 suitable to naval conditions than our designs were. 
 
 Italy has employed the Parsons very generally, as indeed other countries 
 have naturally followed the lead of Great Britain. 
 
 Japan has generally adopted the Curtis for naval purposes. 
 
 The Weight of a Turbine Installation complete, direct-driven and with 
 cylindrical boilers, is about 300 lbs. per S.H.P., as against about 400 lbs. 
 of a triple compound engine one, both in ocean-going ships of fairly high 
 speed. 
 
 In cross-channel express steamers a direct-driven turbine installation 
 will weigh about 200 lbs. per S.H.P. with cylindrical boilers, and about 
 
 57 
 
894 
 
 MANUAL OF MARINE ENGINEERING. 
 
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TURBINES ON SHIPBOARD. 
 
 895 
 
 160 lbs. with water-tube boilers. A geared-turbine set with water-tube 
 boilers will be as low as 140 lbs. per S.H.P., with propeller revolutions 270 per 
 minute. 
 
 Mr. A. Cleghorn* states that if the fuel consumption of a direct-coupled 
 turbine installation using saturated steam is taken as 100, that of single- 
 geared installations will be — in express ocean liners 89, in ocean mixed 
 passenger and cargo steamers 93, in cross-channel expresses 91 '5, and in 
 ordinary cargo ships 835. With single gearing and steam superheated 100° 
 F., the relative consumption is 84 - 5, 88, 86'5, and 79. With double gearing 
 and 200° F. superheat the figures are 75, 74, 73'5 and 66 respectively. 
 
 The weight of such installations as the above, taking the direct-coupled 
 turbines using saturated steam as 100, he gives the figures for the express 
 ocean steamers as 93, 935, and 88 - 5 ; for mixed passenger and cargo ships, 
 80, 81, and 72 ; for the cross-channel expresses, 75, 76, and 68 ; and for the 
 tramp steamers, 79, 79 - 5, and 72'5. Mr. Cleghorn gives as actual examples 
 the following : — 
 
 Actual Steam Consumption on Two Large Liners. 
 
 Description of Machinery. 
 
 S. H. P. 
 
 Steam 
 at Tmbines. 
 
 Consumption in Lbs. per Hour. 
 
 Main 
 Turbines. 
 
 Auxiliaries 
 and other 
 Purposes. 
 
 Total 
 
 for all 
 
 Purposes. 
 
 4 Shafts Direct-coupled Turbines, 
 
 2 Shafts Geared Compound "| 
 
 Turbines, 1 H.P. and 1 L.P. V 
 
 to each, . . . .J 
 
 23,000 
 14,500 
 
 Saturated, . 
 
 / 85° F. Super- \ 
 t heat, J 
 
 112 
 
 97 
 
 22 
 23 
 
 134 
 120 
 
 TABLE CIX. — Methods op Transmission from Turbine to Screw 
 
 Compared. (Cleghorn.) 
 
 Points of Comparison in a Ship of 
 20,000 S. H.P. 
 
 Method of Transmission. 
 
 Double 
 Wheel Gearing. 
 
 Hydraulic 
 Transformer. 
 
 Geared Electrical 
 
 Generator. 
 and Motor Driven. 
 
 Revolution of turbines per min., 
 
 „ screws „ 
 Transmission efficiency, 
 Relative fuel consumption, 
 „ machinery weights, 
 
 2,400 and 1,500 
 90 
 0965 
 100 
 100 
 
 1,000 (about) 
 140 
 
 092 
 123 
 109 
 
 2,400 and 1,500 
 90 
 0-885 
 109 
 108 
 
 The Lubrication of the Gearing is very important, seeing how much of 
 the energy transmitted through it is retained and transformed into heat, 
 which, if not removed in some way, might intensify and become dangerous. 
 * Presidential Address to the Institution of Engineers and Shipbuilders in Scotland 
 
896 MANUAL OF MARINE ENGINEERING. 
 
 Primarily the oiling of the surfaces of the teeth is to minimise the friction 
 due to contact rubbing, but since in spite of it some 1£ to 2 per cent, of the 
 power fails to be transmitted, it must reappear as heat, so that for each horse- 
 power there will be 0-848 B.T.U. per minute or 50-88 per hour. If the feed 
 water per S.H.P. is 12 lbs., this heat would raise its temperature by 4-24° F. ; 
 therefore not so valuable an asset as Dr. Fottinger claims for his system. 
 It is sufficient, however, to demand some method of cooling the lubricating 
 oil beyond atmospheric contact. If the lubricant may be raised by, say, 
 60° F., and its specific heat is 0-35, it will require 3i lbs. of water for 10 lbs. of 
 oil passed ; also that, per horse-power-hour, 0-848 lb. of water will be neces- 
 sary to keep the oil cool. Under these circumstances engines of 10,000 S.H.P. 
 will require circulating water to the amount of 8,480 lbs. or 848 gallons per 
 hour. It is evident, therefore, that in practice a mere oil bath would fail, 
 hence the squirting of cold oil along the angle made by the wheel and 
 pinion faces, and allowing it to be thrown off into a chamber at the 
 bottom of which it runs into a cooling receiver, from which it is pumped 
 into a tank overhead, and so keeps up the supply, is found quite successful 
 and satisfactory as well as economic of oil. In the tropics the cooling water 
 for the oil will, of course, increase in quantity, as does that of the condenser, 
 so that if sea water is at 80° F. and the limit of heat for satisfactory lubrica- 
 tion is 110° F., then the cooling water will be double that named above, 
 and so ships which pass through or trade in tropic seas must have this pro- 
 vision. It need hardly be said, however, that the viscosity of oil at 110° is 
 very low, and, therefore, so high a temperature is to be avoided if possible 
 for continuous running. The extra quantity of cooling water will be cheaply 
 ' bought by the saving in power of transmission effected by the better viscosity 
 of the oil. 
 
SUPERHEATED STEAM AND SUPERHEATERS. 897 
 
 APPENDIX E. 
 
 SUPERHEATED STEAM AND SUPERHEATERS. 
 
 Before the compound engine was generally adopted for ship propulsion 
 superheaters were in general use to dry the steam and fortify it with heat 
 sufficient to carry it' dry at least to the inside of the cylinders. The limit of 
 temperature was then 400° F., so that the amount added was about 125° F. 
 with box boilers and with cylindrical 100° at most ; usually, however, it was 
 not much more than half these, as the reheating was effected then by means 
 of the waste heat in the uptakes, which was not supposed to be much over 
 400° F. On the advent of the # compound engine superheaters were not so 
 popular as formerly as they had become somewhat troublesome and a cause 
 of interference by the Board of Trade officers. Moreover, the gain in economy 
 by the adoption of compound engines was more than that obtained by super- 
 heating in an ordinary engine, and satisfied the shipowner as relieving him 
 of a troublesome appendage. Altogether, superheating dropped out of the 
 marine engineering practice, and was only revived again after experience 
 with it on land installations had demonstrated the possibilities to be great, 
 and the new means practically free from the old objectionable features since 
 metallic packing for rods, and mineral oils for internal lubrication were 
 used. To-day steam at 200 lbs. pressure is in general use ; its natural heat as 
 saturated is 387° F., or about the limit of the superheated of former days. 
 It is fairly dry in the steam space of the boiler as it comes from the water, 
 but any fall in temperature, either from fall in pressure or in loss on trans- 
 mission by radiation, conduction, etc., it will become foggy and deposit 
 moisture on the surfaces it comes in contact with, even if they are no cooler 
 than itself. If these surfaces are cooler, as are the walls of a cylinder, deposi- 
 tion will take place and its efficiency, as also that of the engine itself, will 
 suffer. If, on the other hand, it is fortified against such contingencies by 
 possessing a reserve of heat, no such ill effects will follow. 
 
 Steam is a bad conductor of heat, and consequently when in a stagnant 
 state will absorb no further heat ; it must be in motion over hot surfaces 
 if it is to b.ecome superheated ; hence its passage through a large number 
 of small tubes is the most effective way of superheating it, inasmuch as it 
 all must then be near a hot surface and pass over it rapidly ; moreover, 
 none of it can possibly be stagnant, as was the case in the old box superheaters 
 of fifty years ago. 
 
 The specific heat of steam at 380° F. is 0-574, increasing with its tempera- 
 ture rise till at 600° F. it is 0-6572. 
 
 The Amount of Heat required to superheat a pound of steam from its 
 
 natural temperature T to a temperature T s can be calculated sufficiently 
 
 near by assuming the specific heat to be a mean of the two values. For 
 
 example, if it is required to raise from 200 lbs. with a natural temperature 
 
 o ^ , -r , , L , 0-574 + 0-6572 
 of 380° to 600 F. the specific heat may be taken as ^ — > or 
 
 0-6156. 
 
89S 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Then amount of heat is = mean specific heat X (T, — T). 
 
 In the above case it will be 0-6156 (600 - 380) = 135-4 B.T.U. for each 
 pound of steam. 
 
 The Total Heat of Superheated Steam, calculated by Prof. Peabody's 
 formula, is as follows : — 
 
 (i.) Total heat = 0-4805 (T a - 10-38 ^P) + 857. 
 
 T a is the absolute temperature Fahr. ; P is the absolute pressure. 
 Prof. Ripper's formula is 
 
 (ii.) Total heat = H + 0-48 it t — t^, 
 
 where H is the total heat of saturated steam at a pressure p and temperature 
 £j and t„ that to which it is to be raised. 
 
 Example. — Steam at 200 lbs. gauge pressure is to be raised to a tempera- 
 ture of 500° F. 
 
 (a) By Peabody's formula, 
 
 Total heat = 0-4805 {(800 + 461) - 10'38 V200 + 15} + 857 
 = 442-7 + 857 = 1,299-7° F. 
 
 (b) By Bipper's formula, 
 
 Total heat = 1,200 + 0-48 (500 - 387-5) 
 
 = 1,254. 
 
 Superheated Steam transmitted through pipes is greater than that of 
 saturated steam. It may be calculated by the following formula, where 
 d is the diameter in inches, L the length in feet ; W the weight of a cubic 
 foot of the steam, whose pressure is P x at entry and P 2 at exit. 
 
 Weight per minute = 87 
 
 = 87 fcPx 
 
 Vl( 
 
 P 2 M 5 
 
 The Maximum Work possible from a pound of steam, whose pressure i& 
 200 lbs. and temperature 600° F., is 453 B.T.U., while at 500° it is 405 B.T.U.. 
 as against 340 as saturated steam. This theoretical gain cannot, of course 
 be maintained in practice, but even then it is very material. 
 
 The Economy in using Superheat in practice is proved by the following 
 figures obtained as the result of carefully made experiments in test houses, 
 is the net consumption of the engines, etc., only. 
 
 TABLE CX. — Steam consumed in lbs. per S.H.P.-Hour, Boiler 
 
 Pressure 180. 
 
 S.H.P. 
 
 Triple Expansion 
 Reciprocator. 
 
 Ordinary Turbine. 
 Vac. 28-5. 
 
 Combined Reciprocator and 
 Turbine. 
 
 Vac 20. 
 
 Saturated 
 
 Steam. 
 
 Vac. 26. 
 Superheat 
 to 572° F. 
 
 Vac. 28-5. 
 
 Saturated 
 
 Steam. 
 
 Vac. 28-5. 
 Superheat 
 to 572° F. 
 
 Vac. 28-5. 
 
 Saturated 
 
 Steam. 
 
 Vac. 28-5. 
 Superheat 
 to 572° F. 
 
 500 
 1,000 
 1,500 
 2,000 
 
 16-50 
 15-25 
 14-75 
 14-50 
 
 11-75 
 11-25 
 11 00 
 11 00 
 
 14-00 
 11-75 
 11-25 
 11 00 
 
 11-75 
 9-75 
 9-50 
 9-25 
 
 18 to 22 
 
 12 to 13 
 
 11-25 
 
 11-00 
 
 13 to 15 
 
 9-5 to 9-9 
 
 9-0 
 
 8-7 
 
SUPERHEATED STEAM AND SUPERHEATERS. 
 
 899 
 
 The Heating Surface of a Superheater per pound of steam heated depends 
 largely on the difference in temperature between the hot gases available 
 and the steam as delivered. If it is considerable, as much as 6 B.T.U. per 
 square foot can be transmitted for each degree of difference Fahr., when 
 both surfaces are clean, so that if steam of 200 lbs. pressure is to be raised 
 to 500° with hot gases of 600° temperature, each square foot should transmit 
 6 (600—500) or 600 B.T.U. per hour. 
 
 Each pound of such steam requires (500 — 387) X 0-599, or 67-7 B.T.U. 
 If then the consumption is 12 lbs. of steam per H.P.-hour, the surface will 
 be 67-7 X 12-r 600, or 1-354 square feet per horse-power. 
 
 It will be seen that if high superheat is required the heating gas 
 temperature must be much higher or the surface of the heater larger. 
 
 In ordinary practice the temperature in the uptake should not exceed 
 600° F., and when at a speed of ship lower than given by the full effort of the 
 boilers the temperature will be lower and may be only the old 400° F. of 
 former days. With the Howden air heating for the furnaces the 600° F. 
 may not be wasteful, and even a higher temperature for the sake of super- 
 heating be warranted, but to ensure such it is better not to trust to the 
 smoke-box contents, but seek a source of heat elsewhere. For this reason 
 the modern superheater is inserted in the boiler tubes themselves of a tank 
 boiler, and in the tube chamber of a water-tube boiler where the temperature 
 may be 1,000° F. 
 
 Taking the previous example, but raising the superheat to 600°. then 
 Each pound of steam requires (600 — 387) 0-599, or 127-6 B.T.U., 
 and each square foot passes 6 (1,000 — 600), or 2,400 B.T.U. 
 
 Then the heating surface per H.P. = 127-6 X 12 -e- 2,400, or 0-638 square foot 
 
 It will be seen by the above that for variable service an independently 
 fired superheater possesses many advantages, and when oil-firing is the vogue 
 this could be done, as the waste heat from it could be utilised for feed-heating 
 or other equally useful purposes. 
 
 The claims for superheating in practice on shipboard have been set out 
 by several engineers at the Institution of Naval Architects, in the Transactions- 
 of which they may be read with advantage. 
 
 Mr. H. Gray produced in 1914 the following, as accomplished with a little 
 more than 1 square foot of superheating surface per I.H.P. : — 
 
 Mr. H. Gray's Experiments with Superheated Steam on Shipboard, 1914. 
 
 
 S.S." 
 
 P. A." Triple Expansion 
 
 S.S. "P 
 
 . L." Quadruple Expansion 
 
 Steam Chests of 
 
 
 Et% r ines. 
 
 
 Engines. 
 
 
 Gauge 
 
 Steam 
 
 Pressure. 
 
 Temp, if 
 Satur- 
 ated 
 Steam. 
 
 Actual 
 
 Temp. 
 
 as 
 
 Recorded 
 
 Amount 
 
 of 
 Super- 
 heat. 
 
 Gauge 
 
 Steam 
 
 Pressure. 
 
 Temp, if 
 Satur- 
 ated 
 Steam. 
 
 Actual 
 
 Temp. 
 
 as 
 
 Recorded 
 
 Amount 
 
 of 
 Super- 
 heat. 
 
 
 Lbs. 
 
 o Jj> 
 
 °F. 
 
 o jr 
 
 Lbs. 
 
 o p_ 
 
 o y 
 
 o p 
 
 High-pressure cylr., 
 
 160 
 
 370° 
 
 550 ° 
 
 180° 
 
 206 
 
 389° 
 
 600 : 
 
 211° 
 
 1st intrmdt. cylr., 
 
 53 
 
 300 ° 
 
 400 ° 
 
 100 D 
 
 97 
 
 335° 
 
 460° 
 
 125° 
 
 2nd intrmdt cylr., 
 
 
 , , 
 
 t t 
 
 
 38 
 
 284 D 
 
 290° 
 
 6° 
 
 Low-pressure cylr., 
 
 10 
 
 240° 
 
 
 
 9 
 
 237 D 
 
 220 3 
 
 -17° 
 
 S.B.— Cuts-off— H.P. cylinder, 73 per cent. ; 1 M.P., 70 per cent. : 2 M.P., 70, and L.P., 70 per cent 
 
900 
 
 MANUAL OF MARINE ENGINEERING. 
 
 fefit 
 
 o 
 
 O 
 
 e3 
 
 S-i 
 
 © 
 
 02 
 
 o 
 
 e3 
 
 - 
 
 CO 
 SO 
 
SUPERHEATED STEAM AND SUPERHEATERS. 901 
 
 The s.s. 'Port Augusta," with triple - expansion engines, consumed 
 1-6 lbs. of coal without superheat and 14 lbs. with superheat on an Australian 
 voyage, while a similar ship with quadruple-expansion engines and steam 
 superheated to 600° consumed only 1-29, the fuel being in each case the 
 ordinary coal procurable at the ports of call. 
 
 On one experimental voyage with the latter ship using only good English 
 coal, the consumption was as low as 1-043 lbs., and the average of five similar 
 voyages with such coal showed it to be 145 lbs. 
 
 Mr. Gray claims as the result of his observations that with triple-expansion 
 engines the gain by superheating to about 550° F. is 0-2 lb. per I.H.P., or 
 12| per cent., and with quadruple-expansion engines it is 049 lb., or 14-2 per 
 cent. 
 
 The N. D. Lloyd steamer, "Columbus," of 15,000 I.H.P., with four- 
 crank triple-expansion engines using superheated steam, had a consumption 
 of 1-05 lbs. of coal per I.H.P.-hour for main engines and their own auxiliaries, 
 and 1*20 lbs. consumption for all purposes throughout the ship. 
 
 In another large ship with quadruple engines of 21,650 I.H.P. the con- 
 sumption of saturated steam was 12-71 lbs., and of coal 147 lbs., while with 
 superheating by 200° F. it was reduced to 10-4 lbs. and 1-257 lbs., or 18-2 per 
 cent. less. In a sister ship driven by turbines the consumption without 
 superheat was 11-86 lbs. of steam and 1-385 lbs. of coal, as against 10-9 lbs. 
 and 1-292 with 100° of superheat, or a gain of 8-1 per cent. 
 
 The s.s. " Jupiter " with triple-expansion engines indicating 2,600 H.P. 
 supplied with steam at 185 lbs., having 280° F. of superheat, so that its tem- 
 perature was 661° F., the consumption is only 1-0912 lbs. of Newcastle coal 
 per I.H.P.-hour, which is exceedingly low for such engines. 
 
 The Robinson Superheater (Fig. 334), which is now largely used on ship- 
 board, consists of a series of quite small diameter tubes in pairs joined by a 
 special junction piece at their inner ends, so as to lie side by side for insertion 
 into the boiler tubes ; their outer ends are expanded into holes in special 
 headers (Fig. 335), so that the steam flows from the boiler through them 
 to the main steam pipe leading to the engine. The headers are grouped and 
 secured near the front tube plate, so as not to obstruct gas flow from the tubes 
 more than necessary. In this way the necessary high temperature for super- 
 heat is obtained without so high a temperature in the smoke-box as would 
 otherwise be imperative. With such a large number of these small tubes the 
 velocity of flow through them is comparatively small, so that the drop in 
 pressure is not great. 
 
 It was common practice, and will be so now that for the superheater 
 there is a separate circuit of pipes and valves, so that it may be shut out of 
 action altogether or the portion of steam passing through it regulated so that 
 the temperature may not, at any time as passed to the engines, be above 
 that prescribed. 
 
 The Cylinder Lubricant, if any is used, must be a mineral oil having a 
 high boiling point, and although there is not much risk of internal com- 
 bustion, such a thing has happened with disastrous results, it should have, 
 therefore, a high flash point too. In a general way, of course, there is no 
 free oxygen present in a steam cylinder, but decomposition of water can 
 take place and a supply of oxygen thereby provided. A very soft, greasy 
 graphite has been used as a lubricant for internal purposes, and in moderation 
 
902 
 
 MANUAL OF MARINE ENGINEERING. 
 
 Fig. 335. — Sectional Perspective View of " Robinson " Superheater Header. 
 
SUPERHEATED STEAM AND SUPERHEATERS. 
 
 DOS 
 
 is said to be satisfactory with superheated steam ; so much so that in some 
 foreign ships superheat is much higher than is usual in British practice, 
 the steam temperature being nearly 700° F., or over the melting point of 
 lead, and not much below that of zinc ; consequently no white metal can be 
 used in the stuffing-boxes, and only special bronzes employed with safety 
 there. Probably very soft grey cast iron would be best for neck rings, 
 packing rings, and glands, as the graphite in it would provide the lubricant 
 necessary. 
 
 TABLE CXI. — Weight of Superheated Steam in Lbs. per Minute 
 through Smooth Pipes 240 Diameters Long, with a Drop in 
 Pressure op 1 Lb. 
 
 ] 
 
 1 
 
 
 
 Diameter of Bore 
 
 of Pipe in Inches 
 
 
 
 
 Gauge 
 
 
 
 
 
 
 
 
 
 
 
 Pressure. 
 
 
 
 
 
 
 
 
 
 
 
 
 1 
 
 2 
 
 3 
 
 4 
 
 5 
 
 6 
 
 7 
 
 8 
 
 9 
 
 10 
 
 100 
 
 525 
 
 26-0 
 
 642 
 
 1185 
 
 196 
 
 293 
 
 400 
 
 535 
 
 690 
 
 863 
 
 120 
 
 563 
 
 27-9 
 
 699 
 
 1272 
 
 210 
 
 315 
 
 433 
 
 574 
 
 740 
 
 925 
 
 140 
 
 6-00 
 
 297 
 
 732 
 
 1350 
 
 223 
 
 334 
 
 458 
 
 609 
 
 786 
 
 980 
 
 160 
 
 631 
 
 313 
 
 77-4 
 
 1420 
 
 235 
 
 352 
 
 483 
 
 640 
 
 825 
 
 1033 
 
 180 
 
 664 
 
 330 
 
 815 
 
 1489 
 
 245 
 
 368 
 
 503 
 
 670 
 
 865 
 
 1080 
 
 200 
 
 695 
 
 345 
 
 85-7 
 
 1558 
 
 256 
 
 383 
 
 525 
 
 698 
 
 903 
 
 1125 
 
 220 
 
 7-25 
 
 360 
 
 89-8 
 
 1623 
 
 266 
 
 398 
 
 545 
 
 726 
 
 940 
 
 1172 
 
 240 
 
 7'55 
 
 37'5 
 
 938 
 
 1684 
 
 275 
 
 412 
 
 563 
 
 752 
 
 983 
 
 1214 
 
9Q4 MANUAL OF MARINE ENGINEERING. 
 
 APPENDIX F 
 
 Board of Trade Rules for Boilers. 
 
 The Surveyor's Duty and Responsibility in Fixing Pressures. — The Surveyor 
 is required by the Act to fix the limits of weight to be placed on the safety- 
 valves of passenger steamships. The Surveyor having himself fixed the 
 limits of the weight, is then required to declare, that in his judgment the 
 boiler and machinery are sufficient for the service intended, and in good 
 condition, and that they will be sufficient for twelve months, or such other 
 period as he may, in his judgment, determine. For his guidance the 
 following suggestions are given : — 
 
 Working Pressure to be Fixed by Calculation, &c. — The Surveyor should 
 fix the working pressure for boilers by a series of calculations of the 
 strength of the various parts, and according to the workmanship and 
 material. The senior Engineer Surveyors should receive and report on 
 any plans of boilers intended for passenger vessels that may be submitted 
 in due course with the Form Surveys 6. They are not to report on any 
 tracing or plan that is not accompanied by that form. When the Surveyor 
 has received plans and tracings of new boilers, or of alterations of boilers, 
 and has approved of them, he will of course be careful in making his 
 examinations from time to time to see that they are followed in construc- 
 tion. When he has not had the plans submitted, but is called in to survey 
 a boiler, he will then measure the parts, note the details of construction, 
 and if necessary bore the plates to ascertain their thickness, &c, before he 
 gives his declaration. And in the event of any novelty in construction, or 
 of any departure from the practice of staying and strengthening noted in 
 these regulations, he should report full particulars to the Board of Trade 
 before fixing the working pressure. 
 
 The Surveyor cannot declare a boiler to be safe of whose construction, 
 material, and workmanship he is not fully informed. He should, there- 
 fore, be very careful how he ventures to give a declaration for a boiler 
 that he is not called in to survey until after it is completed and fixed 
 in the ship. 
 
 Girders for Flat Surfaces. — When the tops of combustion boxes, or other 
 parts of a boiler, are supported by solid rectangular girders, the following 
 formula, which is used by the Board of Trade, will be useful for finding the 
 working pressure to be allowed on the girders, assuming that they are not 
 subjected to a greater temperature than the ordinary heat of steam, and in 
 the case of combustion chambers, that the ends are properly bedded to the 
 edges of the tube plate and the back plate of the combustion box : — 
 
 C x d 2 x T „ T . . - 
 
 (W - P) D x L = Workln S P^ssure. 
 
 W = Width of combustion box in inches. 
 P = Pitch of supporting bolts in inches. 
 
 D = Distance between the girders from centre to centre in inches. 
 L = Length of girder in feet. 
 d = Depth of girder in inches. 
 f = Thickness of girder in inches. 
 
BOARD OF TRADE RULES FOR BOILERS. 
 
 905 
 
 N = Number of supporting bolts. 
 
 ~ N x 1,320 . 
 
 yj — — ^ r— when the number of bolts is odd. 
 
 C = - — " J "* when the number of bolts is even. 
 
 N + 2 
 
 The working pressure for the supporting bolts, and for the plate 
 between them, shall be determined by the rule for ordinary stays and 
 plates. 
 
 Plates for Flat Surfaces. — The pressure on plates forming flat surfaces 
 will be easily found by the following formula : — 
 
 = Working pressure. 
 
 o — O 
 
 T = Thickness of the plate in sixteenths of an inch. 
 
 S = Surface supported in square inches. 
 
 C = Constant according to the following circumstances : — 
 
 C = 240 when the plates are not exposed to the impact of heat or flame, 
 and the stays are fitted with nuts on both sides of plates and out- 
 side doubling-strips of the same thickness as the plates they cover, 
 and not less in width than two-thirds of the pitch of the stays. 
 The strips to be riveted on the plate. 
 
 Note. — When doubling-plates cover the whole of the flat surface, the case should 
 be submitted for the consideration of the Board. 
 
 C = 210 when the plates are not exposed to the impact of heat or flame, 
 and the stays are fitted with nuts on both sides of plates and 
 washers, the latter on the outside of boiler, being at least two- 
 thirds the pitch of the stays in diameter, and the same thickness as 
 the plates they cover. These washers to be riveted on the plate. 
 
 C = 165 when the plates are not exposed to the impact of heat or flame, 
 and the stays are fitted with nuts on both sides of plates and out- 
 side washers, the latter being at least three times the diameter of 
 the stay and two-thirds the thickness of the plates they cover. 
 
 C = 150 when the plates are not exposed to the impact of heat or flame, 
 and the stays are fitted with nuts only on both sides of plates. 
 
 C = 1125 when tube plates are not exposed to the direct impact of heat or 
 flame, and the stays are fitted with nuts. 
 
 C = 77 when the tube plates are not exposed to the direct impact of heat 
 or flame, and the stay tubes are screwed and expanded. 
 
 C = 77 when the plates are not exposed to the impact of heat or flame, and 
 the stays are screwed into the plates and riveted over. 
 
 C = 75 when the plates are exposed to the impact of heat or flame, and 
 steam in contact with the plates, and the stays fitted with nuts on 
 both sides of plates, and outside washers, the latter being at least 
 three times the diameter of the stay and two-thirds the thickness 
 of the plate they cover. 
 
 C = 67-5 when the plates are exposed to the impact of heat or flame, and 
 steam in contact with the plate, and the stays fitted with nuts only. 
 
 C = 100 when the plates are exposed to the impact of heat or flame, with 
 water in contact with the plates, and the stays screwed into the 
 plate and fitted with nuts. 
 
906 MANUAL OF MARINE ENGINEERING. 
 
 C = 66 when the plates are exposed to the impact of heat or flame, with 
 water in contact with the plate, and the stays screwed into the 
 plate, having the ends riveted over to form substantial heads. 
 
 C = 39 6 when the plates are exposed to the impact of heat or flame, and 
 steam in contact with the plates, with the stays screwed into the 
 plate, and having the ends riveted over to form substantial heads. 
 
 When a circular flat end is bolted or riveted to a cylindrical shell, S in 
 the formula may be taken as the area of the square inscribed in the circle 
 passing through the centres of the bolts or rivets securing the end, provided 
 the angle ring or flange is of sufficient thickness. 
 
 Compressive Stress on Steel Tube Plates. — 
 
 (D - d) T x 28,000 „ 7 . . 
 
 i -^ — ^ Working pressure. 
 
 D = Least horizontal distance between centres of tubes in inches. 
 
 d = I nside diameter of ordinary tube in inches. 
 
 T = Thickness of tube plate in inches. 
 
 W = Width of combustion box in inches between tube plate and back of 
 fire box, or distance between combustion box tuba plates when 
 boiler is double-ended and the box common to the furnaces at 
 both ends. 
 
 Cylindrical Boilers. — When cylindrical boilers are made of the best 
 material, with all the rivet holes drilled in place, and all the seams fitted 
 with double butt straps, each of at least five-eighths the thickness of the 
 plates they cover, and all the seams at least double-riveted with rivets 
 having an allowance of not more than 87*5 per cent, over the single shear, 
 and provided that the boilers have been open to inspection during the 
 wi.ole period of construction, then 4*5 may be used as the factor of safety. 
 The tensile strength of the material is to be taken as equal to 27 tons per 
 square inch when mild steel is used, tested in accordance with the Rules 
 (p. 442). When plates of higher tensile strength than 27 tons are used, 
 the actual strength, as found on test, may be substituted in the following 
 Rule. The boilers must be tested by hydraulic pressure to twice the 
 working pressure in the presence and to the satisfaction of the Surveyor : — 
 
 To ascertain the strength of shell, the relative sectional areas of plate 
 and rivet must first be determined by the following formulae : — 
 
 (Pitch - dia. of rivet) x 100 _ f Percentage of strength of plate at joint 
 Pitch \ as compared with solid plate. 
 
 For maximum permissible pitch of rivets, &c, see formulae and sketches 
 at end of this section. (Pp. 277 to 281, Seaton and Eounthwaite's Pocket- 
 Book of Marine Engineering.) 
 
 (Area of rivet x No. of rows of rivets) x 100 j Percentage of strength 
 
 ==p — = r-r-i . . ^ = < or rivets as compared 
 
 Pitch x thickness of plate | with golid pkte 
 
 If the rivets are in double shear, multiply the percentage thus found 
 by 1-875 
 
 In consequence of the low shearing strength of steel rivets, the Board 
 require that in all types of joint the nominal rivet section shall be not less 
 
EOARD OF TRADE RULES FOR BOILERS. 907 
 
 than ff of the net plate section; thus, in order that rivet section may be con- 
 sidered to have the same strength as plate section their relation must be : — 
 
 In lap joints — 
 
 (Area of rivet x No. of rows of rivets) x 23 
 = (Pitch - diameter of rivet) x thickness of plate x 27. 
 
 And in butt joints — 
 
 (Area of rivet x No. of rows of rivets x 1875 x 23 
 = (Pitch - diameter of rivets) x thickness of plate x 27. 
 
 But when plate of a higher minimum tensile strength than 27 tons is 
 used, the actual strength in tons should be used instead of the multiplier 27. 
 
 The working pressure per square inch that may be allowed on the safety 
 valves is then given by — 
 
 TTT , • S x 7 x 2 T 
 Working pressure = ==p — ^ — , 
 
 Where S = Tensile strength of material in lbs. per square inch. 
 
 °/ Q = One of the two percentages, found by Rules above divided by 100. 
 
 T = Thickness of plate in inches. 
 
 D = Inside diameter of boiler in inches (inside diameter of outer" 
 
 strake, if any). 
 F = Factor of safety from following Table if % refers to plate 
 
 section. 
 F = 4*5 if % refers to rivet section. 
 Tlie smaller of the two results to be taken. 
 
 Various penal additions are provided in the rules for cases in which the 
 workmanship is of inferior character. 
 
 If it is proposed to use steel for superheaters, particulars should be 
 submitted to the Board for consideration. 
 
 TABLE CIX. — Board of Trade Factors of Safety. 
 
 When all rivet holes are drilled in place after bending ; all seams 
 fitted with double butt-straps, each at least five-eighths the 
 thickness of the plates they cover ; all seams at least double 
 riveted ; and boilers open to inspection during construction, - F = 4'5 
 To be added when circumferential seams are lap and double riveted, - '1 
 
 To be added when longitudinal seams are lap and double riveted, - "2 
 
 To be added when longitudinal seams are lap and treble riveted, - '1 
 
 To be added when boiler is of such length as to fire from both ends, 
 
 unless middle circumferential seams are lap and treble riveted, - - 3 
 
 Compensating rings should be fitted around all manholes and openings, 
 of at least the same effective sectional area as the plate cut out, and in no 
 case should rings be of less thickness than plate to which they are attached. 
 Manholes in shells of cylindrical boilers should have their shorter axis 
 placed longitudinally. 
 
 Tt is very desirable that compensating rings round openings in fiat 
 surfaces be of L or T iron. 
 
 The neutral parts of shells under steam domes must be efficiently 
 stiffened and stayed. 
 
908 MANUAL OF MARINE ENGINEERING. 
 
 Board op Trade Rules for Cylindrical Boiler Shells, 
 joints with drilled holes. 
 Formulae for ordinary chain riveted and ordinary zig-zag riveted joints, 
 and for joints of these descriptions, when every alternate rivet in the outer 
 or in the outer and inner rows have been omitted : — 
 
 Let E = distance from edge of plate to centre of rivet in inches. 
 V = distance between rows of rivets in inches. 
 
 Vj = distance between inner and middle row of rivets in inches, for 
 chain and zig-zag treble riveted butt joints with alternate 
 rivets omitted in outer rows. 
 B — boiler pressure in lbs. per square inch. 
 C = 1 for lap or single butt joints. 
 C = 1-875 for double butt joints. 
 d = diameter of rivets in inches. 
 D = inside diameter of boiler in inches. 
 F = factor of safety for shell plates. 
 n = number of rivets in one pitch. 
 p D = diagonal pitch in inches. 
 
 P D = diagonal pitch in inches between inner and middle rows of 
 rivets in inches, for zig-zag treble riveted butt joints with 
 alternate rivets omitted in outer rows 
 p = greatest pitch of rivets in inches. 
 
 r = percentage of plate left between holes in greatest pitch. 
 R = percentage of rivet section. 
 
 Rj = percentage of combined plate and rivet section. 
 S = tensile strength of material in lbs. per square inch of section. 
 Sj = tensile strength of plates in tons. 
 T = thickness of plate in inches. 
 T a = thickness of each butt strap in inches. 
 % = least value of r, R, R„ as the case may be, divided by 100. 
 
 ORDINARY CHAIN AND ZIG-ZAG RIVETED JOINTS. 
 
 Iron plates and iron rivets, or steel plates and steel rivets : — 
 
 loo ( P - d) _ r 
 
 Steel plates and steel rivets : — * 
 
 100 x 23 x d- x -7854 x n x C x F 
 4 5 x Sj x p x T 
 
 Given C, d, F, n, T, to find p, so that r and R are equal. 
 
 Steel plates and steel rivets : — 
 
 23 x & x -7854 x n x C x F 
 
 _ _ _ + d = p. 
 
 4-o x Sj x T 
 
 Given C, F, n, T, r, to find p and d. 
 Steel plates and steel rivets : — 
 
 4-5 x S x x r x T 
 
 23 x (100 - r) x -7854 xwxCxF 
 100 x 4-5 x SL x r x T 
 
 R. 
 
 23 x (100 - r) 2 x -7854 x n x C x F 
 
 P 
 
BOARD OF TRADE RULES FOR BOILERS. 909 
 
 Steel plates and steel rivets when d is found first, then : — 
 
 100 d 
 100 - r " P ' 
 Steel plates and steel butt straps. 
 Double butt straps : — 
 
 8 " ** 
 Single butt straps : — 
 
 9 x T 
 
 — g- - i v 
 
 For Distance between Rows of Rivets, &c, for Double, Treble, and Quadruple 
 Riveted Lap Joints; and Double and Treble Riveted Butt Joints with 
 same Number of Rivets in Inner and in Outer Rows. 
 
 Steel plates and steel rivets : — 
 
 3 x d v 
 
 Chain-riveted joints, not less than : — 
 
 2 x d = V. 
 Zig-zag riveted joints : — 
 
 s/(Up + 4rf)(p + 4dj = y 
 10 
 Diagonal pitch for zig-zag riveted joints : — ■ 
 
 6 p + 4 d _ 
 10 Pd ' 
 
 To determine the Working Pressure. 
 
 S * % * 2T 
 F x D 
 
 Chain and Zig-zag Riveted Joints in which every alternate Rivet has beeD 
 omitted in the Outer Row, or in the Outer and the Inner Rows. 
 
 Steel plates and steel rivets : — 
 
 100 (p-d) _ r 
 
 V 
 
 Steel plates and steel rivets : — 
 
 100 x 23 x d°- x -7854 x n x C x F 
 
 4-5 x Sj x p x T. ~ 
 
 Steel plates and steel rivets : — 
 
 100 (p- 2d) } R R 
 p n v 
 
 When the Joints are fitted with Single or Double Butt Straps and the Number 
 of Rivets in the Inner Row is double the Number in the Outer Row. 
 
 Steel plates and steel butt straps. 
 Double butt straps : — 
 
 5 x Tjp - d) T 
 8 x (p - 2 d) V 
 
 58 
 
910 MANUAL OF MARINE ENGINEERING. 
 
 Single butt straps : — 
 
 9 x TQ? - d) T 
 
 8x(p-2 d) v 
 
 "When the number of rivets in the inner row is the same as in the outer row. 
 Double butt straps : — 
 
 5 x T 
 
 — g— - J-r 
 
 Single butt straps : — 
 
 9 x T 
 
 For Distance between Rows of Rivets, &c, for Double and Treble Riveted 
 Butt Joints with Alternate Rivets omitted in Outer Rows; and Treble 
 Riveted Lap Joints and Butt Joints with alternate Rivets omitted in 
 Outer and in Inner Rows. 
 
 Steel plates and steel rivets : — 
 
 3 x d 
 
 Chain-riveted joints : — 
 
 J{u P + ±d){ P + id) = V) \ The greatest of these two 
 
 ™ > values of V to be used. 
 
 Dr ,, 1 . See Note below. 
 
 2 x d = V. ) 
 
 For treble chain-riveted joint with alternate rivets omitted in outer row : — 
 
 2 x d = Y v 
 
 Zig-zag riveted joints : — 
 
 Diagonal pitch : — 
 
 tbP + d = p D . 
 For treble zig-zag riveted joint with alternate rivets omitted in outer row : — 
 
 J( 1 1 ~p~T8 ~d)~(p~~+ 8d) v 
 20 i- 
 
 Diagonal pitch : — 
 
 3jp + id 
 
 10 " " D> 
 
 To determine the Working Pressure. 
 
 S x 7 x 2T 
 
 F x D 
 
 = B. 
 
 Note. — The minimum value of V or V 1 for chain-riveted joints is given 
 as 2 d, but = — is more desirable. 
 
 Maximum Pitches for Riveted Joints. 
 
 T = thickness of plate in inches. 
 
 p = maximum pitch of rivets in inches. 
 
 C = constant applicable from the following table : — 
 
BOARD OF TRADE RULES FOR BOILERS. 
 
 TABLE CX. 
 
 911 
 
 Number of Rivets 
 iu One Pitch. 
 
 Constants for Lap 
 Joints. 
 
 Constants for Donltle- 
 iiutt Strap Joints. 
 
 1 
 2 
 3 
 4 
 5 
 
 1 31 
 
 2 62 
 347 
 414 
 
 175 
 3 50 
 463 
 
 5-52 
 6 00 
 
 (C x T) + 1§ = p. 
 
 When the work is first-class, such pitches may be adopted so far as 
 safety is concerned ; yet, in some cases, it may be well not to adopt the 
 greatest pitch found by the formula. The maximum should not, however, 
 exceed 10J inches with the thickest plates for boiler shells. 
 
 If, in any case, the pitch is found to exceed that arrived at by the 
 foregoing formula, for the particular description of joint and thickness of 
 plate, such pitches should not be passed, but, in all cases, reported. 
 
 Note.— If the edge of the plate or butt strap is cut away or vandyked so as to follow 
 the line of rivets, the pitch of rivets in the outer rows may exceed 10| inches. 
 
912 MANUAL OF MARINE ENGINEERING. 
 
 APPENDIX G. 
 
 Lloyd's Rules for Determining the Working Pressure to be 
 Allowed in New Boilers. 
 
 cylindrical shells of steel boilers. 
 
 The strength of cylindrical shells of steel boilers is to be calculated from 
 the following formula : — 
 
 C x (T 2) X B 
 
 1 — =5—^ = working pressure in lbs. per square inch. 
 
 Where D = mean diameter of shell in inches. 
 
 T = thickness of plate in sixteenths of an inch. 
 
 C = 22, when the longitudinal seams are fitted with double butt 
 
 straps of equal width. 
 C = 21*25, when they are fitted with double butt straps of unequal 
 
 width, only covering on one side the reduced section of 
 
 plate at the outer lines of rivets. 
 C = 20*5, when the longitudinal seams are lap joints. 
 
 If the minimum tensile strength of shell plates is other 
 
 than 28 tons per square inch, these values of C may be 
 
 correspondingly increased. 
 B = the least percentage of strength of longitudinal joint,* found 
 
 as follows :— «■ 
 
 For plate at joint, B = ^-^ — x 100. 
 
 „ rivets „ B = — x 85, where steel rivets are used. 
 
 p x t 
 
 „ n x a 
 
 B = ^Tx-7 x '°' " iron 
 
 Where p = pitch of rivets in inches. 
 
 t = thickness of plate in inches. 
 d = diameter of rivet holes in inches. 
 
 n = number of rivets used per pitch in the longitudinal joint. 
 a — sectional area of rivet in square inches. In case of rivets in 
 double shear 1*75 a is to be used instead of a. 
 
 Note. — For the shell plates of superheaters or steam chests enclosed in 
 the uptakes or exposed to the direct action of the flame, the coefficients 
 should be § of those given in the preceding tables. 
 
 Proper deductions are to be made for openings in shell. 
 
 All manholes in circular shells to be stiffened with compensating rings. 
 
 The shell plates under domes in boilers so fitted to be stayed from the 
 top of the dome or otherwise stiffened. 
 
 * The inside butt strap to be at least f of the strength of the longitudinal joint. 
 
LLOYD'S RULES FOR WORKING PRESSURE. 913 
 
 STAYS. 
 
 The strength of stays supporting flat surfaces is to be calculated from 
 the smallest part of the stay or fastening, and the strain upon them is not 
 to exceed the following limits, namely : — 
 
 Iron Stays. — For stays not exceeding 1| inches smallest diameter, and 
 for all stays which are welded, 6000 lbs. per square inch ; for un welded 
 stays above H inches smallest diameter, 7500 lbs. per square inch. 
 
 Steel Stays.— For screw stays not exceeding 1£ inches smallest diameter, 
 8000 lbs. per square inch; for screw stays above 1£ inches smallest diameter, 
 9000 lbs. per square inch. For other stays not exceeding 1£ inches smallest 
 diameter, 9000 lbs. per square inch, and for stays exceeding 1| inches 
 smallest diameter, 10,000 lbs. per square inch. No steel stays are to be 
 welded. 
 
 Stay Tubes. — The stress is not to exceed 7500 lbs per square inch. 
 
 FLAT PLATES. 
 
 The strength of flat plates supported by stays is to be taken from the 
 following formula : — 
 
 O x T 2 
 
 — — - — = working pressure in lbs. per square inch. 
 
 Where T = thickness of plate in sixteenths of an inch. 
 
 P 2 = square of pitch in inches. If the pitch in the rows is not equal 
 
 to that between the rows, then the mean of the squares of 
 
 the two pitches is to be taken. 
 O = 90 for iron or steel plates T 7 g- thick and under, fitted with screw 
 
 stays with riveted heads. 
 C = 100 for iron or steel plates above -1$ thick, fitted with screw 
 
 stays with riveted heads. 
 C = 110 for iron or steel plates T 7 T thick and under, fitted with 
 
 stays and nuts. 
 O = 120 for iron plates above ^ thick, and for steel plates above T 7 V 
 
 and under T 9 g thick, fitted with screw stays and nuts. 
 C = 135 for steel plates T \ thick and above, fitted with screw stays 
 
 and nuts. 
 C = 140 for iron plates, fitted with stays with double nuts. 
 C = 150 for iron plates, fitted with stays with double nuts and 
 
 washers outside the plates, of at least ^ of the pitch in 
 
 diameter and \ the thickness of the plates. 
 O = 160 for iron plates, fitted with stays with double nuts and 
 
 washers riveted to the outside of the plates, of at least ■§ of 
 
 the pitch in diameter and \ the thickness of the plates. 
 O = 175 for iron plates, fitted with stays with double nuts and 
 
 washers riveted to the outside of the plates, when the 
 
 washers are at least § of the pitch in diameter and of the 
 
 same thickness as the plates. 
 
 For iron plates fitted with stays with double nuts and doubling strips 
 riveted to the outside of the plates, of the same thickness as the plates, and 
 of a width equal to f the distance between the rows of stays, C may be 
 
914 MANUAL OF MARINE ENGINEERING. 
 
 taken as 175, if P is taken to be the distance between the rows, and 190 
 when P is taken to be the pitch between the stays in the rows. 
 
 For steel plates, other than those in combustion chambers, the values of 
 may be increased as follows : — 
 
 C - 140 increased to 175. 
 C - 150 „ 185. 
 
 C = 160 „ 200. 
 
 O = 175 „ 220. 
 
 C = 190 „ 240. 
 
 If flat plates are strengthened with doubling plates securely riveted to 
 them, having a thickness of not less than § of that of the plates, the 
 strength to be taken from 
 
 *s« 
 
 * ( T + I)' 
 
 ^— = working pressure in lbs. per square inch. 
 
 Where t = thickness of doubling plates in sixteenths, and 0, T, and P are 
 as above. 
 
 Note. — In the case of front plates of boilers in the steam space, these 
 numbers should be reduced 20 per cent., unless the plates are guarded from 
 the direct action of the heat. 
 
 For steel tube plates in the nest of tubes the strength to be taken from 
 
 140 x T 2 • 
 — = working pressure in lbs. per square inch. 
 
 Where T = the thickness of the plates in sixteenths of an inch. 
 P = the mean pitch of stay tubes from centre to centre. 
 
 For the wide water spaces between the nests of tubes, the strength to- 
 be taken from 
 
 x T 2 
 
 — P2 — = working pressure in lbs. per square inch. 
 
 Where P = the horizontal distance from centre to centre of the bounding 
 rows of tubes ; and 
 
 C = 120, where the stay tubes are pitched with two plain tubes 
 between them and are not fitted with nuts outside the plates. 
 
 C = 130, if they are fitted with nuts outside the plates. 
 
 = 140, if each alternate tube is a stay tube not fitted with nuts. 
 
 C = 150, if they are fitted with nuts outside the plates. 
 
 = 1 60, if every tube in these rows is a stay tube and not fitted 
 with nuts. 
 
 = 170, if every tube in these rows is a stay tube and each alter- 
 nate stay tube is fitted with nuts outside the plates. 
 
 The thickness of tube plates of Combustion Chambers, in cases where 
 the pressure on the top of the chambers is borne by these plates, is not to 
 be less than that given by the following rule : — 
 
 P x W x D 
 1750 xTD - d)' 
 
LLOYDS RULES EOR WORKING TKKSSURE. 915 
 
 Where P = working pressure in lbs. per square inch. 
 
 W = width of Combustion Chamber between plates in inches. 
 D = horizontal pitch of tubes in inches. 
 d = inside diameter of plain tubes in inches. 
 T = thickness of tube plates in sixteenths of an inch. 
 
 GIRDERS. 
 
 The strength of girders supporting the tops of combustion chambers and 
 other flat surfaces is to be taken from the following formula : — 
 
 C x d 2 x T 
 _ : _ — — = working pressure in lbs. per square inch. 
 
 Where L = width between tube plates, or tube plate and back plate of 
 chamber. 
 P = pitch of stays in girders. 
 D = distance from centre to centre of girders. 
 d = depth of girder at centre. 
 T = thickness of girder at centre. 
 
 All these dimensions to be taken in inches. 
 
 Wrought Iron. 
 
 C = 6,000, if there is one stay to each girder 
 
 C = 9,000, if there are two or three stays to each girder. 
 
 C = 10,000, if there are four or five stays to each girder. 
 
 C = 10,500, if there are six or seven stays to each girder. 
 
 = 10,800, if there are eight stays or above to each girder. 
 
 Wrought Steel. 
 
 C = 7,110, if there is one stay to each girder. 
 C = 10,660. if there are two or three stays to each girder. 
 C = 11,850, if there are four or five stays to each girder. 
 C = 12,440, if there are six or seven stays to each girder. 
 C = 12,800, if there are eight stays or above to each girder. 
 
 CIRCULAR FURNACES. 
 
 The strength of plain furnaces to resist collapsing to be calculated as 
 follows : — 
 
 Where the length of the plain cylindrical part of the furnace exceeds 
 120 times the thickness of the plate, the working pressure is to be calculated 
 by the following formula : — 
 
 1,075,200 x T 2 . . . „ • . 
 
 -1 1 _ = working pressure in lbs. per square men. 
 
 Where the length of the plain cylindrical part of the furnace is less than 
 120 times the thickness of the plate, the working pressure is to be calculated 
 by the following formula : — 
 
 ^ - — '- = working pressure in lbs. per square inch. 
 
916 MANUAL OF MARINE ENGINEERING. 
 
 Where D = outside diameter of furnace in inches. 
 T = thickness of plates in inches. 
 
 L = length of plain cylindrical part in inches, measured from the 
 centres of the rivets connecting the furnaces to the flanges of 
 the end and tube plates, or from the commencement of the 
 curvature of the flanges of the furnace where it is flanged or 
 fitted with Adanison rings. 
 
 In the furnaces referred to below, the formulae given are applicable if 
 the steel used has a tensile strength of not less than 26 nor more than 30 
 tons per square inch. If the material of furnaces has a less tensile strength 
 than 26 tons per square inch, then, for each ton per square inch which the 
 minimum tensile strength falls below 26, the coefficient is to be correspond- 
 ingly decreased by ^ part. 
 
 The strength of corrugated furnaces made of steel, on Fox's, Morison's, 
 Deighton's, or Beardmore's plan, and of the Leeds Forge bulb furnace, to be 
 calculated from 
 
 1259 x (T — 2) 
 
 ^ ' = working pressure in lbs. per square inch. 
 
 The strength of Brown's cambered, and improved Purves' furnaces (with 
 ribs 9 inches apart) to be calculated from the following formula : — 
 
 1160 x (T - 2) 
 
 D 
 
 = working pressure in lbs. per square inch. 
 
 The strength of spirally corrugated furnaces is to be calculated from the 
 following formu.a : — 
 
 912 x (T — 2) 
 
 ■ ^r '- = working pressure in lbs. per square inch. 
 
 Where T = thickness of plate in sixteenths of an inch; and 
 
 D = outside diameter of corrugated furnaces, in inches; or smallest 
 outside diameter, in inches, of Brown's cambered, improved 
 Purves', and Leeds Forge bulb furnaces. 
 
 The strength of Holmes' patent furnaces, in which the corrugations are 
 not more than 16 iuches apart from ct ntre to centre, and not less than 
 2 inches high, to be calculated from the following formula: — 
 
 945 x CT — 2) 
 Working pressure in lbs. per square inch = ^- '. 
 
 Where T = thickness of plain portions of furnace in sixteenths of an inch. 
 D = outside diameter of plain parts of the furnace in inches. 
 
BRITISH CORPORATION RULES FOR WORKING PRESSURE, ETC. 917 
 
 APPENDIX H. 
 
 Rules of the British Corporation for the Working Pressure, 
 Thicknesses of Plates, and Sizes of Stays for Steel Boilers. 
 
 The rules following are intended to apply to the construction of steel boilers ; ■ 
 where boilers are to be made of iron, they will be specially considered by 
 the committee. 
 
 Strength Calculations. — The sizes and arrangement of the different parts 
 for a given working pressure, or the working pressure suitable to a given 
 size and arrangement of material, may be found from the following formulae 
 and rules ; — 
 
 Cylindrical Shells— 
 
 C x (T-l) x E 
 
 D ' 
 
 Where C = 20 - 4, when the longitudinal seams are fitted with double butt 
 straps of equal width, and of thickness at least equal to that 
 obtained from the formula for butt straps. 
 
 C = 19 "7, when the double butt straps are of unequal width — i.e., 
 one strap not covering the outer row of rivets, with the 
 thickness as before. 
 
 C = 19*0, when the longitudinal seams are lap joints. 
 
 T — thickness of shell plate in sixteenths of an inch. 
 
 E = the least percentage of strength of longitudinal joints, found 
 as follows : — 
 
 For the plate at the joint, E = *- X 100. 
 
 For the rivets at the joint, E = X - X 100. 
 
 J p X t s 
 
 Where p — pitch of rivets in inches. 
 
 d = diameter of rivet holes in inches. 
 
 n = number of rivets used per pitch. 
 
 a = sectional area of rivets in square inches. 
 
 t = thickness of plate in inches. 
 
 r = 23 for steel. 
 
 s = the minimum tensile strength of plate. 
 
 Where rivets are in double shear, 1 *875 a is used instead of a. 
 D = mean diameter of shell in inches. 
 W = working pressure in lbs. per square inch. 
 
 Nom — The constant, C, is to be Used with steel of the minimum strength 
 of 28 tons per square inch, as required by the rules (see p. 810) ; when a 
 
9 IB MANUAL OF MARINE ENGINEERING. 
 
 higher minimum strength is guaranteed, the constant may be proportionately 
 increased. 
 
 Double Butt Straps.— When all the rows of rivets in the butt are of the 
 same pitch, each strap must be at least five-eighths of the thickness of the 
 shell plate, and, when the pitch of the outer row of rivets is twice that of 
 the centre row, the thickness of each plate is to be in accordance with the 
 following formula : — 
 
 T T 5 ( p ~ ^ 
 
 1 " 8(p- kd) 
 
 Where T 1 = thickness of strap in sixteenths of an inch. 
 T = thickness of shell in sixteenths of an inch. 
 f = pitch of rivets in inches. 
 d = diameter of rivet holes in inches. 
 k = ratio of pitch of rivets in outer to that in inner .row. 
 
 Flat Surfaces, supported by Stays— 
 
 C XT 2 
 P 2 +p 2 
 
 = W. 
 
 Where C = 195 for plates fitted with screwed stays having riveted heads. 
 
 C = 265 for plates fitted with screwed stays and nuts. 
 
 C = 345 for plates fitted with stays and double nuts. 
 
 C = 370 for plates with stays, double nuts, and washers outside. 
 The washers to be at least half the thickness of the plate, 
 and in diameter equal to one-third the pitch of stays. 
 
 C = 400 for plates fitted with stays, double nuts, and washers out- 
 side riveted to the plate. The washers to be at least half 
 the thickness of the plate, and in diameter equal to two- 
 fifths the pitch of the stays. 
 
 C = 450 for plates fitted with stays, double nuts, and washers out- 
 side riveted to the plate. The washers to be the same 
 • thickness as the plate, and in diameter equal to two-thirds 
 the pitch of the stays. 
 
 C = 480 for plates fitted with stays, double nuts, and doubling strips 
 outside riveted to the plates. The strips to be the same 
 thickness as the plate, and in width equal to two-thirds the 
 pitch of the stays. 
 
 T = thickness of plate in sixteenths of an inch. 
 
 P = greatest pitch of stays. 
 
 f = least pitch of stays. 
 
 W = working pressure in lbs. per square inch. 
 
 Flat plates having doublings, at least two-thirds of their thickness, riveted 
 to them. 
 
 CX ( T+ -^ ff 
 
 Where t = thickness of doubling plates in sixteenths of an inch. 
 C, T, P, f, and W, as before. 
 
BRITISH CORPORATION RULES FOR WORKING PRESSURE. ETC 919 
 
 Note. — For front plates in the steam space, which are not protected 
 against the direct action of the flame, the constants given above are to be 
 reduced 20 per cent. 
 
 Tube Plates, with Tubes in Nests.— 
 
 275 x T 2 
 
 Where T and W are as before. 
 
 P = greatest pitch of stay tubes from centre to centre in inches. 
 f = least pitch of stay tubes from centre to centre in inches. 
 For wider spaces, v. p. 904, ante). 
 
 When girders are fitted to the tops of combustion chambers, the thick- 
 ness of the tube plates must be found from the formula — 
 
 W XL xP 
 (P-d) X 1,800 == 
 
 Where W = working pressure in lbs. per square inch. 
 
 L = width of combustion chamber over the plates in inches. 
 
 P = horizontal pitch of tubes in inches. 
 
 d = inside diameter of plain tube in inches. 
 
 T = thickness of tube plate in sixteenths of an inch. 
 
 Stays supporting Flat Surfaces— 
 
 V : 
 
 S ^+i = D. 
 
 Where S = surface in square inches, supported by the stay. 
 W = working pressure in lbs. per square inch. 
 D = effective diameter of stay in inches. 
 C = 8,000 for steel screwed stays and iron if tested to 21 \ tons with 
 
 27 per cent, elongation. 
 C = 6,500 for iron screwed stays. 
 C = 8,900 for longitudinal steel stays. 
 C = 7,000 for longitudinal iron stays. 
 C = 5,000 for welded iron stays. 
 
 Stay Tubes are not to be subjected to a greater stress than 7,500 lbs. 
 per square inch. 
 
 Circular Furnaces. — Thickness of plain furnaces and of furnaces with 
 Adamson rings pitched more than 20 inches apart — 
 
 -^ {18-75 T - (L x 1-03)} = W. 
 
 Thickness of bulb furnaces made by the Leeds Forge Company (Suspension)— 
 
 1,250 X (T - 2) 
 
 D W 
 
 Thickness of furnaces with Adamson rings pitched not more than 20 inches 
 apart, or corrugated, ribbed, and suspension furnaces (Fox, Purves, Morison, 
 Deighton, and Brown's cambered) — : 
 
 1,160 X ( T-2) _ w 
 D 
 
920 MANUAL OF MARINE ENGINEERING. 
 
 Thickness of Farnley spiral furnace and Holmes' furnace— 
 
 950 x (T - 2) 
 D 
 
 Where T = thickness of plate in sixteenths of an inch. 
 
 L = length of furnace or pitch of stiffening rings in inches. 
 D — smallest outside diameter of furnace in inches. 
 \V = working pressure in lbs. per square inch. 
 
 Girders for Combustion Chamber Tops — 
 
 C x d* x T (n + l) 2 
 
 D x L 2 n (w + 2) 
 
 G x d 2 x T 
 
 W, when n = 2, 4, or 6. 
 
 D x L.2 
 
 . = W, when n — 1, 3, or 5. 
 
 Where = 14,500 for steel and 13,180 for iron. 
 
 d = depth of girder at the centre in inches. 
 
 T = thickness of girder in inches. 
 
 D = distance from centre to centre of girders in inches. 
 
 L = length from tube plate to tube plate, or from tube plate to back 
 
 of combustion chamber, in inches. 
 n = number of stays fitted with each girder. 
 
 General Construction. 
 
 1. Each boiler must have at least one glass water gauge, two test cocks, 
 and one steam pressure gauge. Double-ended boilers are to have these 
 fittings at each end. One salinometer cock must also be fitted to each 
 boiler. Where the water gauge pillar is attached by pipes to the steam and 
 water spaces, cocks or valves should be fitted to the shell of the boiler at 
 the ends of these pipes. 
 
 2. A stop- valve is to be fitted to each boiler, so that any one of a series 
 of boilers may be worked independently if required. The neck of the 
 stop-valve to be as short as possible. All boiler and engine stop-valves to 
 be tested to at least twice the working pressure. 
 
 3. Two safety-valves will be required for each main boiler, and they 
 must be tested under steam to the satisfaction of the surveyors, and set to 
 a pressure not more than 3 per cent, in excess of the intended working 
 pressure. The combined area of the valves is to be sufficient to prevent 
 the steam accumulating to more than 10 per cent, of the working pressure 
 during fifteen minutes full firing, with main engines stopped. If the 
 boilers be supplied with forced draught, the valve area must be increased so 
 that the same conditions may be met. 
 
 4. Easing gear is to be provided, and so arranged that the safety-valves 
 on any one boiler may be lifted simultaneously, without interfering with 
 those on any other boiler. 
 
 5. Surface and bottom blow-off cocks should be fitted to the boiler, in 
 addition to the cock on the hull of the vessel, and all cocks must have 
 
•BRITISH CORPORATION RULES FOR WORKING PRESSURE, ETC 921 
 
 spigots extending through the hull plating, with a plate flange round same 
 on the outside. 
 
 6. It is recommended that manhole doors in boilers be not less than 
 16 inches by 12 inches. 
 
 7. Upon completion, the boilers and superheaters are to be tested by 
 hydraulic pressure to twice the intended working pressure, and, after being 
 placed in position in the vessel, they must be efficiently secured by brackets 
 and stays to prevent any fore-and-aft or athwartship movement. It is 
 strongly recommended, because of the rapid corrosion which takes place 
 in material near them, that the boilers be kept as high as possible above the 
 floors or tank top, and that the under side of the boilers be efficiently insulated. 
 
 8. Donkey boilers need not have more than one safety-valve, provided 
 the valve area be not less than \ square inch for each foot of grate surface. 
 In other respects, the requirements for donkey boilers are the same as for 
 main boilers. 
 
 Tubes for water-tube boilers and superheaters to be not less in thickness 
 than required for steam pipes, and tested to 1,000 lbs. per square inch (hy- 
 draulic), or to three times the working pressure, whichever is greater. Headers 
 also to be tested to three times the W.P. All other parts to double W.P. 
 
922 MANUAL OF MARINE ENGINEERING. 
 
 APPENDIX I. 
 
 Bureau Veritas Rules for Boiler Shells, Working Pressure, or 
 Thickness of Plates, and for Sizes of Stays.* 
 
 circular shells and steam-holders with internal pressure. 
 
 A riveted joint may fail through the tearing of the plate or butt strap 
 between the rivets, the shearing of all the rivets, or by a combination of 
 the two. The following formulae apply to these several cases. The plate 
 thickness and the diameter of rivets to be applied to have the highest 
 values which each formula would give separately. 
 
 I. Rupture through Plate. 
 The formulae for working pressure and plate thickness are in this case: — 
 
 2aR(*-Q-Q4) \ 
 
 \ D D - - " (L) 
 
 and t = - — =- + 0-04 inch I 
 
 la ii J 
 
 Where P = allowed working pressure, above atmosphere, in lbs. per 
 square inch. 
 
 D = greatest inside diameter of boiler shell or steam-holder in inches. 
 
 t = thickness of shell plates in inches, t- 0-04 inch represents the 
 
 thickness left after a reduction of 0*04 inch through corrosion. 
 
 R = the tensile stress in lbs. per square inch, which will be allowed 
 in the plate. The value of R will be the breaking strength 
 divided by 4, the latter figure representing the factor of 
 safety for the plate after it has been corroded away by 
 0-04 inch. 
 
 If the actual breaking strength happens to be known by tests carried 
 out to the Administration's satisfaction, it may be applied for finding R; 
 but when, as usual, it is not known, the value of R will be : — 
 
 For Steel. — The fourth part of the lower limit of tensile strength chosen by 
 the designer, which, in such case, is to be stated when a boiler design is 
 submitted for approval. 
 
 For Iron. — 11,200 lbs. per square inch, corresponding with a tensile 
 strength of 20 tons. A table annexed shows the values of 2 R for various 
 tensile strengths. 
 
 a = ratio of the resistance of the plate left between the holes to that of 
 the full plates. It will be determined from the following expression : — 
 
 p - d 
 
 a = 
 
 V 
 
 * X.B. — If a boiler is intended for a vessel belonging to a country where the law 
 prescribes heavier scantlings than those required by the following paragraphs, the 
 builders have, of course, to comply with the legal requirements. 
 
BUREAU VERITAS KULKS FOR BOILER SHELLS, ETC, 
 
 923 
 
 Where p = pitch of rivets in outer row in inches. 
 
 d = diameter of rivet holes in inches, either the real diameter or a 
 corrected one, according to the following clauses : — 
 
 1. When the rivet holes are drilled, or when, having been punched, they 
 are afterwards drilled or rimered out so that the injured metal around is 
 completely removed, the real diameter may be taken. 
 
 2. When the holes are simply punched, they will be considered as being 
 ^ inch larger in diameter than as punched. 
 
 TABLE CXI. showing the Values of 2R. 
 In Formulas (I.) and (IV.) for Various Tensile Strengths of the Material. 
 
 Tensile Strength 
 
 
 Tensile Strength 
 
 
 . of Plates in 
 
 Value of 2R. 
 
 of Plates in 
 
 Value of 2R. 
 
 Tons per Sq. Inch. 
 
 
 Tons per Sq. Inch. 
 
 
 32 
 
 35,800 
 
 24J 
 
 27,400 
 
 31 
 
 34,700 
 
 24 
 
 26,900 
 
 30 
 
 33,600 
 
 23^ 
 
 26,300 
 
 29 
 
 32,500 
 
 23 
 
 25,800 
 
 28 
 
 31,400 
 
 224 
 
 25,200 
 
 27 
 
 30,200 
 
 22 
 
 24,600 
 
 26J 
 
 29,700 
 
 214 
 
 24,100 
 
 26 
 
 29,100 
 
 21 
 
 23,500 
 
 254 
 
 28,600 
 
 204 
 
 22,900 
 
 25 
 
 28,000 
 
 20 
 
 22,400 
 
 II. Rupture through Rivets. 
 In this case the following are the formulae for finding the allowed 
 
 working pressure or required rivet section : — 
 
 2As 
 
 P = 
 
 and 
 
 (IT.) 
 
 Where P and D have the same meaning as before, and 
 
 I - the length in inches of the identical parts into which a riveted 
 joint can be subdivided. In most cases, I is the pitch of the 
 rivets in the outer rows (figs. 333 and 334). In general, it 
 depends upon the system of joint adopted. 
 * = the maximum shearing stress in lbs. per square inch which will 
 be allowed on the rivets. It will be the fourth part of the 
 actual shearing resistance of the material, if known, from 
 tests. If the actual shearing resistance of the rivet bars is 
 not known, it will be assumed to amount to - 8 of their 
 tensile strength, and the value of s will be one-fifth of the 
 lower tensile limit adopted by the designer, who is free to 
 adopt from 24 to 29 tons steel. 
 
924 
 
 MANUAL OF MARINE ENGINEERING. 
 
 For Iron. — 9,000 lbs. per square inch, corresponding with a tensile strength 
 of about 20 tons per square inch. 
 
 A = the total shearing surface in square inches of the rivets (that is, 
 twice the area of the rivet hole when a rivet is in double 
 shear). Only fifteen-sixteenths of the full area are to be 
 taken when the riveting is done by hand. 
 
 III. Combined Rupture through Plate and Rivets. 
 
 This case is only to be examined when the outer row has a wider pitch 
 than the inner ones. 
 
 The formula to be applied in this case is : — 
 
 2 (B X R + C x S) 
 
 P = 
 
 (III.) 
 
 D xl ' ' 
 
 Where P, R, S, D, and I have the same meaning as before {v. figs. 333 and 334). 
 
 B = the sectional area in square inches of the plate on the portion. I, 
 
 of the joint along the line of its supposed rupture, assuming 
 
 that, in case the plate is liable to corrosion, its thickness has 
 
 been reduced by 0'04 inch. 
 
 C = the total area of the rivets which are supposed to shear on the 
 
 length, I, corrected, if required, in the same way as prescribed 
 
 above. 
 
 ....j!:::.?.;: 
 
 -4- P —\ 
 ... r 
 
 JL 
 
 
 
 Fig. 336. 
 
 Fig. 337. 
 
 For a rivet in double shear, the resistance will be coll idered as being 
 twice that of one in single shear. 
 
 IV. Rupture through Butt Straps. 
 
 Rupture may take place along one of the inner rows of rivets (see P, Q, 
 figs. 333 and 334). The formula for this case, based on the same principle 
 as (I.), are : — 
 
 2oE(«- 0-04) 
 
 P = 
 
 D 
 
 and 
 
 P x D 
 ' = -7^r + °'°- im ' h - 
 
 2 a n 
 
 . (IV.) 
 
BUREAU VERITAS RULES FOR BOILER SHELLS ETC. 925 
 
 Where P, D, and R have the same meaning as before. 
 
 t — thickness in inches of butt strap, or sum of thicknesses, if there 
 are two straps. (The thickness, of course, not to be less than 
 required for caulking.) 
 
 q — d 
 
 a = 3 . 
 
 S 
 
 Where q — pitch of rivets in the inner row in inches. 
 
 d = diameter in inches of the rivet holes in the inner row. 
 
 V. Combined Rupture through Butt Straps and Rivets. 
 
 Formula (III.) applies to this case, B being the section of the butt strap, 
 or straps, along which rupture would take place. 
 
 Remarks. — No rivet holes to be nearer the edge of any plate than the 
 diameter of the rivet. In zigzag riveting, the distance between the rows is 
 to be such that no rupture through plate or butt strap is to be feared along 
 the zigzag line. 
 
 When stays are bolted through the shell, they should be so arranged 
 that they do not weaken the shell plates more than the riveted joints. If 
 the resistance at the stay bolts is the smaller of the two, the plate's thickness 
 shall be determined by it. It will be found from a formula the same as 
 (I.), p and d applying to the stay bolts. 
 
 Shells of Superheaters. — The same mode of determining the working 
 pressure or the thickness of shall plates will be followed for circular cylin- 
 drical superheaters, and the same f ormulae may be used, but with the following 
 alterations : — 
 
 1. When the plates are exposed to the direct action of the products of 
 combustion, the values of R and S, as given for Cases I. and II., will have 
 to be multiplied by - 8 and the addition to the thickness of plates, on account 
 of corrosion, will be increased from 0'04 to -^ of an inch, to compensate for 
 the corrosive action of the gases. 
 
 Formulae (I.) become, therefore — 
 
 (HaR<-M 
 D ' 
 
 and ^pg-^+^inch. 
 
 2. When the plates are protected from the direct action of the products 
 of combustion, R and S will be multiplied by 9, and the additional thickness 
 for burning away will be J inch. 
 
 Formulae (I.) become in this case — 
 
 P = S H, and t = j^g-H £ inch. 
 
 59 
 
926 MANUAL OF MARINE ENGINEERING. 
 
 FLAT PLATES. 
 
 The allowed working pressure or the thickness of flat plates is to be 
 determined by the following formulae : — 
 
 p = ( ^rl x c' and ' = * + V (a2+ b2) ¥• 
 
 Where P = allowed working pressure above atmosphere in lbs. per sq. inch. 
 t = thickness of plate in sixteenths of an inch. 
 a = pitch of stays in inches, in one row. 
 b = distance in inches between two rows of stay. 
 
 Note. — When plates are effectively stiffened by doubling plates well 
 
 riveted thereto, and having a thickness, t, in sixteenths, the value, t-\- - 
 may be substituted for t in the formula. 
 
 • In case ot irregular staying, such as in the annexed 
 sketch, (fig. 335)— 
 
 Q 
 
 , 
 
 
 
 
 Fig. 
 
 338. 
 
 I (Pj + P 2 ) 2 shall be taken instead of (a 2 + b 2 ). 
 
 T = tensile strength of the material in tons per 
 square inch of the original section. 
 
 It is to be determined in the same way as for shell plates, that is — 
 
 For steel it will be equal to the lower limit of the tensile stress, which 
 is to be stated on the drawing. 
 
 For iron, 21 tons per square inch. 
 
 C = a constant, the value of which depends upon the mode of staying, 
 as follows ; — 
 
 C = 0*104 when the stays are screwed into the plates and riveted over. 
 
 C — 0*079 when the stays are screwed into the plates and fitted with 
 outside nuts at both ends. 
 
 C = 0*062 when the stays are fitted with inside and outside nuts 
 and washers, provided the diameter of the outside washers 
 be at least 0*4 of the pitch between the rows of stays. The 
 thickness of the washer to be at least § that of the plate. 
 
 C = 0*055 when the stays are fitted with inside and outside nuts 
 and washers, the outside washer being riveted to the plate 
 and having f of the plate's thickness and a diameter equal 
 to 0*6 of the pitch between the rows of stays. 
 
 C = 0*050 when the outside washers are replaced by strips of plate, 
 
 having a width of at least 0*6 of the distance between the 
 
 rows of stays, with a thickness not less than f that of the 
 
 plate : the strip being well riveted to the plate. 
 
 T 
 For the values of ~, see Table cxii. 
 
 When the plates are in contact with steam on one side and flame or hot 
 gases on the other, the thickness is to be increased. 
 
BUREAU VERITAS RULES FOR BOILER SHELLS, ETC. 
 
 927 
 
 1)1 
 
 For instance, when, in return tube boilers, the top front plates are L 
 no way protected from the hot gases, the working pressure or thickness 
 will, in such a case, be determined by the formula; : — 
 
 p _ (t :i 2g 0-9 T 
 02+62 * - c ' 
 
 and 
 
 * = 2 + */(a 2 +& 2 ) 
 
 PC 
 
 ()■'.) T 
 
 When the said front plates are protected by a flame plate, no increase 
 of thickness will be required. 
 
 When front plates are in two pieces, the lap should be double-riveted 
 if the thicker plate is \ inch or above. 
 
 TABLE CXII. — Values of ~ in the Formulae for Flat Plates. 
 
 Tensile Strength 
 
 
 
 
 
 
 of Plate in 
 
 C = 0-104. 
 
 C = 0-079.* 
 
 C = 0-062. 
 
 C = 0-055. 
 
 C = 0-050. 
 
 Tons per Sq. Inch. 
 
 
 
 
 
 
 20 
 
 T 
 
 £ = 192-3 
 
 T 
 
 £ = 253-1 
 
 T 
 
 i = 322-6 
 
 T 
 
 c = 363-6 
 
 T 
 
 G = 400 
 
 21 
 
 202-0 
 
 265-8 
 
 338-6 
 
 381-8 
 
 420 
 
 22 
 
 211-5 
 
 278-4 
 
 354-8 
 
 400-0 
 
 440 
 
 23 
 
 221-1 
 
 291-1 
 
 371-0 
 
 418-0 
 
 460 
 
 24 
 
 230-8 
 
 303-8 
 
 387-0 
 
 436-2 
 
 480 
 
 25 
 
 240-4 
 
 316-4 
 
 403-2 
 
 454-4 
 
 500 
 
 26 
 
 250 
 
 329-1 
 
 419-4 
 
 472-6 
 
 520 
 
 27 
 
 259-6 
 
 341-8 
 
 435-4 
 
 490-8 
 
 540 
 
 28 
 
 269-2 
 
 354-4 
 
 451-6 
 
 509-0 
 
 560 
 
 29 
 
 278-8 
 
 367-1 
 
 467-8 
 
 527-2 
 
 580 
 
 30 
 
 288-4 
 
 379-7 
 
 483-8 
 
 545-4 
 
 600 
 
 *N B. — When plates are not subjected to flame or hot gases, C may be reduced 
 
 T 
 from 0-079 to 0066; or the values of ~ in this column by 1-2. 
 
 STAYS. 
 
 The diameter of stays supporting flat surfaces is to be determined by 
 the following formula : — 
 
 d = | inch + J 
 
 Where d = effective diameter in inches (for instance, the diameter at bottom 
 of thread in screw stays). 
 Q = total load on stay in lbs. 
 T = tensile strength of the material in tons per square inch. 
 
 For steel, this tensile strength will be the lower limit chosen by the boiler 
 designer ; for iron, it will be taken at 22 tons. In both cases the actual 
 -strength may be applied, if it is known from tests. 
 
 If the stays are not round, their cross-section must be such that tbe 
 
928 MANUAL OF MARINE ENGINEERING. 
 
 stress per square inch, caused by the load, Q, nowhere exceeds 5 "75th part 
 of the tensile strength, after deducting ^ of an inch all round as an allowance 
 for corrosion or wear. 
 
 In welded stays, the stress, as just described, will be reduced by 20 per 
 cent. Welding of steel stays is only allowed for very mild qualities. 
 
 For high working pressures, such as used in triple-expansion engines, it 
 is recommended to screw all stays into the plates they support, in addition 
 to fitting them with nuts. 
 
 This also applies to stay tubes, with the exception that it is recommended 
 that nuts should not be fitted in combustion chambers. 
 
 CIRCULAR FURNACES. 
 
 Plain Cylindrical Furnaces. 
 
 When made truly circular as possible, and of steel, not less than 26 tons 
 tensile, the working pressure and the thickness of the plates may be calculated 
 from the following formula? : — 
 
 16,000 t — 60 L 
 
 P = 
 
 and t = 
 
 D 
 
 P X D + 60 L 
 
 16,000 
 
 Where t = the required thickness of plates in inches. 
 D = outside diameter of furnace in inches. 
 
 P = working pressure in lbs. per square inch (above atmosphere). 
 L = length of furnace in inches ; or, if made or fitted with efficient 
 rings, the length between the rings. 
 
 For iron or low tensile steel, constant must be reduced to 14,400, 
 Furnace plates should not exceed |f inch in thickness. 
 
 DOMED FURNACES IN VERTICAL BOILERS. 
 
 When the tops of furnaces in vertical boilers are portions of spheres 
 the thickness must not be less than : — 
 
 Pr Pr 
 
 t = ;™q + -15 for iron plates, or t = ^— : + "15 for steel plates. 
 
 CORRUGATED AND RIBBED FURNACES. 
 
 The plate thickness is to be found by the following formulae :-— 
 1. For corrugated furnaces — 
 
 T- ™+2 
 
 1,260^ 
 
 Where P = working pressure in lbs. per square inch. 
 T = thickness of plate in sixteenths of an inch. 
 
 D = outside diameter in inches measured on the top of the corru- 
 gations, 
 
 ■ The formula applies to corrugations 6 inches long and U inches deep. 
 
BUREAU VERITAS RULES FOR BOILER SHELLS, ETC. 929 
 
 2. For ribbed furnaces, when manufactured to the satisfaction of the 
 Administration — 
 
 T - ™ + 2 
 
 1 " 1,160 + L 
 
 Where P and T are as above. 
 
 D = outside diameter of plain parts between the ribs in inches. The 
 formula applies to ribs spaced 9 inches and projecting If 
 inches, the difference between the greatest and the smallest 
 diameters in any part of the furnace not exceeding i^jj. 
 
 3. For bulb furnaces- 
 
 T- PP I 2 
 " 1,260 " r 
 
 Where D is the outside diameter in inches measured between the bulbs. 
 
 The coefficients in the above formula apply to the case where the tensile 
 strength of the material is 26 tons or above per square inch. When it is 
 below 26 tons the coefficient is to be reduced ^ for each ton below 26. 
 
 Combustion Chamber Girders. 
 
 The strength of girders on the tops of combustion chambers shall be 
 determined as follows : — 
 
 CdH 
 
 (W-p)DL' 
 
 Where P = working pressure. 
 
 C = a constant found as under. 
 N = number of bolts in each girder. 
 cl = depth of girder in inches. 
 t = thickness of girder in inches. 
 p = pitch of bolts in girder in inches. 
 L = length of girder between supports in inches. 
 D = distance between girders, centre to centre, in inches. 
 W = width of firebox, from tube plate to back plate, in inches. 
 
 C = ~ =— for odd numbers of bolts. 
 
 N+ 1 
 
 C = — ' XT — - for even numbers of bolts, 
 
 GENERAL AS TO BOILER ARRANGEMENTS AND FITTINGS. 
 
 When two or more boilers are fitted, they should be arranged to work 
 separately or independently of each other, either by stop-valves between 
 the boilers and the common superheater, or by stop-valves between the 
 separate superheaters and the main steam pipes. 
 
 Each boiler must be fitted with a separate pressure gauge. Double- 
 ended boilers to have a pressure gauge at each end. 
 
930 MANUAL OF MARINE ENGINEERING. 
 
 Steam pipes of donkey engines must be independent of the main pipes, 
 so as to keep the steam off the main engines when only winches or other 
 auxiliary engines are working. 
 
 The bottom blow-off to be arranged with one cock directly attached to 
 the shell of the boiler, and another directly attached to the skin of the vessel. 
 The surface blow-off must be similarly arranged. The main stop-valve to 
 be so situated and fitted that it can be worked from the starting platform 
 or the stokehole floor. 
 
 To protect the plating, the cocks on the ship's bottom should be fitted 
 with spigots passing through the plating and through a flange on the outside. 
 If this flange is of iron it must be galvanised. 
 
 Steam domes or superheaters, when placed in the uptake and exposed to 
 the direct impact of the flame, will only be allowed as an exception, and 
 must be efficiently protected by flame plates. In all cases it must be possible 
 to examine efficiently the interior and the exterior of the domes or super- 
 heaters. 
 
 To prevent the boilers shifting in a transverse direction through the 
 rolling of the vessel, or longitudinally in case of collision, they must be 
 properly secured in their seats. 
 
 All manholes to be fitted with compensating rings. 
 
 At least two safety-valves of an approved design must be fitted to each 
 main boiler. 
 
 Their total area is given by the following formula : — 
 
 A- 8 ' 7 
 
 V(P - 18) 3 
 
 Where A = sectional area of safety-valve in square inches per square foot 
 of grate surface. 
 P = working pressure of boiler in lbs. per square inch. 
 
 When forced draught is provided for, the area is to be increased in pro- 
 portion to the increased evaporative power of the boilers. 
 
 Suitable arrangements and gear to be fitted in connection with the safety- 
 valves, whereby they may be lifted from the deck as well as from the stoke- 
 hole floor. 
 
 If it be practicable to isolate a superheater communicating simultaneously 
 with two or more boilers, it will be necessary to furnish it with a safety- 
 valve of suitable dimensions. 
 
ADMIRALTY RULES FOR STEEL BOILERS. 931 
 
 APPENDIX J. 
 
 Admiralty Rules for Steel Boilers and Materials and 
 
 their Tests. 
 
 Admiralty rule for working pressure provides that the water-pressure 
 test shall not exceed £ of the ultimate strength of the shell, or a factor of 
 safety of 2£, when subjected to water test pressure, and the working steam 
 pressure is fixed at 90 lbs. below the test pressure, which is called their 
 "constant margin" of safety for all pressures. 
 
 SUPERVISION OP BOILER WORK. 
 
 The following instructions to boiler-maker overseers are those usually 
 given in Admiralty specification : — 
 
 The boilers will be subject to the supervision of an overseer, who will 
 be directed to attend on the premises of the contractors during the progress 
 of the work on the boilers, to examine the material and workmanship used 
 in their construction, to witness the prescribed tests, and to see that this 
 specification, as regards the boilers and work in connection, is conformed to 
 in all respects by the contractors. The extent of supervision is described in 
 the following paragraphs extracted from Admiralty instructions to over- 
 seers, and the contractors are to afford him every facility for their proper 
 execution. 
 
 The plates and other material used in the construction of the boilers to 
 be subjected to such tests as may be directed in the specification. Every 
 plate used is to be carefully examined by the overseer for laminations, 
 blisters, veins, and other defects, and to ensure that it is of the proper 
 thickness and brand. No plate, angle, &c, which from any cause is con- 
 sidered by the overseer to be unfit for the intended use is to be fitted. 
 
 During the construction of the various parts of the boilers, the overseer 
 is to satisfy himself that the dimensions as shown on the approved drawings 
 are being adhered to by the contractors. 
 
 Whenever plates are flanged or welded, or in any case where iron or 
 steel is worked in such manner that it is particularly liable to suffer in 
 strength unless carefully handled, the overseer is to be present if possible 
 on all occasions during the time the work on each article is in progress, and 
 he is to fully satisfy himself that it is sound before he allows any part to be 
 put in the boilers. 
 
 Samples of the rivets being used for the boilers are to be taken by the 
 overseer during the progress of the work and tested as specified hereafter, 
 and any batches of rivets found defective are to be rejected. Before rivets 
 are put in, the overseer is to see that the plates are brought properly 
 together, and that the holes are fair with one another. He is not to allow 
 drifting on any account, but he is to see that they are carefully rimmed fair 
 where necessary. He is also to make sure during the progress of the work 
 that the rivets fill the holes completely, and that the heads are properly set 
 up, well formed, and finished. 
 
932 MANUAL OF MARINE ENGINEERING. 
 
 The overseer is to see that all internal parts of the boiler are riveted with 
 rivets having heads and points of approved shape, and that any seams he 
 considers necessary are riveted on the fire side. , No snap heads are to be 
 allowed in the internal parts. Any proposal for hydraulic riveting the 
 internal parts is to be submitted to the Admiralty, with sketch of the 
 proposed heads and points. In all parts where the rivets are not closed by 
 hydraulic riveting machinery he is to see that the rivet holes are counter- 
 sunk and that coned rivets are used. All holes in the plates, angles, &c, 
 are to be drilled, and not punched, and are to be drilled in place after 
 bending. The clearance between rivet hole and rivet before closing is not 
 to be greater than approved by the overseer. 
 
 The overseer will see that the particulars of the form, dimensions, and 
 pitch of the various stays shown on the drawings are adhered to, and 
 samples of them are to be tested as directed in this specification ; and he 
 will be guided by his experience as a workman in testing and judging of the 
 soundness of the forging and construction of the various stays. 
 
 He is to see that palm stays, if fitted, are forged from the solid and not 
 welded, that all short stays are nutted on all flat surfaces except where 
 otherwise approved and screwed to a pitch of twelve threads per inch for 
 stays of 1 inch diameter and above, that the holes for the screwed stays in 
 the water spaces are drilled and tapped together after the furnaces and 
 combustion chambers have been riveted in place in the boiler, that the 
 combustion chamber stays are drilled square to the bevel of the combustion 
 chamber 'plates, and that no bevel washers are inside the chamber. Any 
 girder stays used for combustion chambers are to be well bedded on to the 
 tube plates to the satisfaction of the overseer. 
 
 The overseer is to see that the arrangement of the zinc plates shown on 
 the approved drawings is adhered to, that the metallic surfaces in contact 
 are filed bright, and that means are adopted to secure a firm grip of the 
 clips by which the plates are attached. 
 
 The overseer is to witness the testing, in all cases, of the boiler tubes, 
 in accordance with this specification, before they are put in the boiler. 
 
 When the boilers are reported to the overseer by the contractors as 
 being completed, ready for testing by water pressure, the overseer is to 
 witness a preliminary test of them in accordance with the specification, 
 carefully observing with the assistance of gauges, and -straight edges 
 whether any bulging or deflection of the plates has taken place. 
 
 The official test will be conducted on all occasions in the presence of an 
 inspecting officer. A test pressure gauge is supplied to the overseer from 
 the Admiralty, and the official test is to be made with this gauge. 
 
 After the boilers have been tested by water pressure the overseer is to 
 see that they are properly cleaned inside and outside, and then well painted 
 with red lead. It is important that the whole surface of the boilers 
 should be thoroughly cleansed of scale formed in manufacture before 
 any paint is put on them. The boilers are not to be exposed to the 
 weather till they are so painted, and properly cleaned and closed up to his 
 satisfaction. 
 
 The overseer is to make himself fully acquainted with the progress of 
 the whole of the work in its various stages, to satisfy himself that every 
 part is sound before it is allowed to be put in the boilers, and to see that 
 the following instructions for the treatment of mild steel are strictly com- 
 plied with. 
 
ADMIRALTY RULES FOR STEEL BOILERS. 933 
 
 QUALITY OP AND TESTS FOR MATERIALS. 
 
 (a) Steel generally. 
 
 Quality of. All steel used throughout the whole of the work is to be of 
 
 quality to be approved by the Admiralty. The names of the firms 
 from whom it is proposed to ohtain this material are to be sub- 
 mitted for approval before the order is placed. The quality and 
 makers' names of all steel are to be marked on all drawings sent to 
 the Admiralty. 
 Solid ingots All steel forgings are to be made from solid ingots, and test 
 for forgings. pi eces f r0 m each important completed forging will be subject to 
 the tests for steel enumerated below, and facilities are to be 
 afforded, by a proper system of numbering and registering the 
 ingots, to enable the origin of any forging to be ascertained. 
 Testing in The whole of the temper and bending tests for furnace and 
 
 presence of ther plates exposed to flame, also those for the important castings 
 officers, &c. and forgings, and not less than 10 per cent, of all the other tests 
 enumerated, are to be made in the presence and to the satisfaction 
 of the Admiralty officers. These tests are intended to facilitate 
 supply, but are not to preclude examination at the machinery con- 
 tractors' works, and rejection of any plates or castings found defective 
 after delivery. The tensile strength and elongation per cent, are 
 to be stamped on each plate. 
 Copies of Copies of the invoices upon which the plates, bars, angles, rivets, 
 
 be V su CeS iied s ^ a y s j * UDes j & c -j f° r the boilers have been received, with the names 
 &c. ' of the manufacturers of the material and their tests, are to be sup- 
 
 plied to the boiler overseer for his use in accepting them. When 
 any orders are placed with steel makers, copies of the orders show- 
 ing the particular purpose of the materials should be sent to the 
 Admiralty overseer, with a copy of the specified tests. 
 
 (b) Plate, Angle, and Bar Steel, for Boilers and any Steel 
 
 Steam Pipes. 
 
 Tests for Every plate, &c, used is to be tested as follows: — Strips cut 
 
 stee pa s, i en gthwise or crosswise are to have an ultimate tensile strength 
 
 Tensile test. p er square inch of section as follows : — (1) Plates, &c, not exposed 
 
 to flame, not less than 27 tons, and not exceeding 30 tons ; (2) 
 
 lower front plates not less than 25 tons, and not exceeding 28 tons; 
 
 (3) furnace plates not exceeding 25 tons, and not less than 23 tons; 
 
 (4) tire-box and other plates exposed to flame not exceeding 26 
 tons, and not less than 24 tons. The elongation taken on a length 
 of 8 inches, must be at least 20 per cent, for strips from shell and 
 other plates which are not exposed to flame and which will not be 
 flanged, and at least 25 per cent, for plates which are not exposed 
 to flame and which will be flanged, at least 27 per cent, for strips 
 from the furnaces, and at least 26 per cent, for strips from fire-box 
 and other plates exposed to flame. Edge shearings, to bring plates 
 to proper dimensions, are to be equal on opposite sides. 
 
 Temper Strips cut lengthwise or crosswise, 1J inches wide, heated 
 
 tests. uniformly to a low cherry red, and cooled in water of 82° Fahren- 
 
934 MANUAL OF MARINE ENGINEERING. 
 
 heit, must stand bending double in a press to a curve of which the 
 inner radius is one and a-half times the thickness of the steel 
 tested. 
 
 The plates for furnaces and other parts exposed to flame are, in Welding and 
 addition, to be tested l>y welding and forging, and some of the opiates 65 ' 
 welded pieces are to be broken in the testing machine to ascertain exposed to 
 the efficiency ot the welding. 
 
 The pieces cut out for testing are all to be cut in a planing Test pieces, 
 machine, to have the sharp edges removed, and are to be of parallel 
 width for at least 8 inches of length. 
 
 The angle, tee, and bar steel are to stand such other forge tests, Forged tests 
 both hot and cold, as may be sufficient, in the opinion of the ° t f eei gle,,t0- ' 
 overseer, to prove the soundness of the material and fitness for the 
 service intended. 
 
 The furnaces are to be subjected to a hammering test if required Hammering 
 before the machinery is accepted. The test is to be carried out in 
 accordance with Admiralty practice. 
 
 (c) Steel Rivets for Boilers. 
 
 The whole of the rivets are to be properly heated in making, Manufac- 
 and care is to be taken that the finished rivets cool gradually. 
 
 The rivets are to be made from steel bars, the ultimate tensile Tensile test 
 strength of which should not be less than 25 tons, and should not 
 exceed 27 tons per square inch, with a minimum elongation of 25 
 per cent, in a length of 8 inches. 
 
 Pieces cut from the bars, heated uniformly to a low cherrv red, Temper 
 and cooled in water of 82° Fahrenheit, must stand bending double tests 
 in a press to a curve of which the inner diameter is equal to the 
 diameter of the bar tested. 
 
 Samples from each batch of rivets are to be selected and are to Forge tests, 
 stand the following forge tests satisfactorily without fracture : — 
 The rivet to be bent double, cold, to a curve of which the inner 
 diameter is equal to the diameter of the rivet tested ; the rivet to 
 be bent double when hot. and hammered till the two parts of the 
 shank touch ; the head to be flattened when hot, without cracking 
 at the edges, till its diameter is two and a-half times the diameter 
 of the shank. Finally, the shank of a rivet is to be nicked on one 
 side, and bent over to show the quality of the metal. 
 
 (d) Boiler Tubes. 
 
 The tubes are to be made from acid or basic " open hearth " Boiler 
 steel. Strips cut from the tubes must have a tensile strength not tubes - 
 exceeding 25 tons and not less than 21 tons per square inch, with 
 an extension of at least 25 per cent, in a length of 8 inches. Strips 
 cut from the tubes flattened, heated to a blood heat and plunged 
 into water 82° Fahrenheit temperature, should be capable of being 
 doubled over a radius of \ inch without fracture. Pieces 2 inches 
 long, cut from the ends of tubes under -^ inch thick are to be 
 capable, when cold, of being hammered down endwise until their 
 length is reduced to 1^ inch, and of being flattened until the sidea 
 
ADMIRALTY RULES FOR STKEL BOILERS. 035 
 
 are close together in each case without fracture. The ends of the 
 tubes under T ^- inch thick are to admit of being expanded cold by 
 a roller expander, worked in a tube hole to an increase of diameter 
 of 15 per cent., and hot by a solid drift to an increase of diameter 
 of 20 per cent. Tubes -j\ inch and over in thickness are to admit 
 of being expanded hot and cold to half the increases of diameter 
 required for tubes under ^ inch thick. The above tests are to be 
 applied to 2 per cent, of the tubes, to be selected by the examining 
 officers. The failure of the tubes selected to stand the specified 
 tests in a satisfactory manner may render the whole of any delivery 
 liable to rejection. 
 
936 MANUAL OF MARINE ENGINEERING. 
 
 APPENDIX K. 
 German Government Rules for Boilers, 
 regulations respecting steam boilers. 
 
 1. Plating of Boilers. — When the smallest dimensions of a cylindrical 
 boiler exceeds 25 centimetres, or of a spherical-shaped boiler 30 centimetres, 
 the portions of the boiler exposed to the flame, furnaces, and tubes shall not 
 be made of cast iron. 
 
 Brass (copper) plates are only to be used for furnaces where their 
 smallest dimension does not exceed 10 centimetres. 
 
 2. Furnaces, Flues, and Combustion Chambers. — Flues running round or 
 through steam boilers must, at their highest point, be at least 10 centimetres 
 below the minimum water level. This minimum level must, in lake and 
 river vessels, be maintained when the vessel is inclined at an angle of 4°, 
 and for sea-going ships at an angle of 8°. In boilers with a breadth of 1 to 
 2 metres, the distance to the water level must be at least 15 centimetres, 
 and in boilers of greater breadth at least 25 centimetres. 
 
 These conditions do not apply to boilers in which the tubes are less than 
 10 centimetres diameter, nor to those wherein the portions of the plates in 
 contact with steam are not liable to become red hot. The risk of plates 
 becoming red hot may be assumed to be done away with if the heating sur- 
 face exposed to the flame, before the flame reaches the point in question, is, 
 with natural draught 20 times and with forced draught 40 times, greater 
 than the area of the grate. The highest point in marine boiler flues, &c, is 
 the upper surface of the plates in question. 
 
 3. Feed Apparatus. — Each boiler must have a feed valve which will close 
 by boiler pressure when the feed is off. Where only one feed check-valve 
 is provided shut-off" valves must be fitted in the feed delivery pipes. 
 
 4. Each boiler must be fitted with two independent feed apparatus, each 
 worked separately, and each to be capable, by itself, of keeping the boiler 
 supplied. Several boilers connected together and used for the same purpose 
 may, in this respect, be regarded as one boiler. Each feed apparatus to be 
 capable of delivering 30 litres of water per square metre of heating surface 
 per hour. Apparatus for boilers worked under forced draught, or having 
 heating surface less than 35 times the grate area, must be proportionately 
 larger. Each feed apparatus must have an independent suction pipe, or 
 else a suitable arrangement of switch shut-off valves must be fitted. An 
 auxiliary feed apparatus for a main boiler may be used for feeding a donkey 
 boiler if there are independent delivery pipes, or if suitable switch shut-off 
 valves are provided. If an injector is used for auxiliary feed purposes it 
 must be capable of working satisfactorily with both high and low pressure 
 steam ; if it will not so work, some other provision must be made for use 
 when pressure is low. Where a main boiler is used for driving winches in 
 port, the main engine feed pumps will not be reckoned as a secondary feed 
 apparatus. If supplementary feed arrangements, beyond those required by 
 
GERMAN GOVERNMENT RULES FOR BOILERS. 937 
 
 the regulations, are provided, the piping for them must either be inde- 
 pendent of that belonging to the regulation apparatus, or must be capable 
 of being shut off from the latter by valves. Each main feed-pump must be 
 fitted with an escape valve that cannot be shut off from the pump. 
 
 5. Water Gauges. — Each boiler must be fitted, with a gauge glass, and 
 also with a second means of ascertaining the height of the water. Each of 
 these fittings must have a separate connection to the boiler, or if they have 
 a common connection, the stand pipe must be at least 60 square centimetres 
 in area, or 88 millimetres in diameter. Pipes for connecting water gauges 
 to boilers must not be less than 16 centimetres in area, or, say, 45 millimetres 
 in diameter. Long or much bent pipes must be larger in proportion. 
 Internal connecting pipes must not be used, but two pipes may be led to 
 one opening in the plating, provided the area of opening be made equal to 
 the combined areas of the pipes. Gauge glasses are to be placed so as to be 
 easily visible. The mark indicating lowest water level should not be higher 
 than half-glass, and the mark showing highest point of heating surface 
 should be above upper edge of lower nut which secures glass. The cocks 
 must be easy to open and shut, and it must be possible to replace a broken 
 glass without risk to the operator. 
 
 6. Water test cocks to be fitted : the lowest one must be placed at the 
 level of the regulation water-line. All gauge cocks must be so fitted as to 
 have a straight way through for cleaning them from scale or salt. Spindle 
 valves must not be used. 
 
 7. Water Level. — The regulation lowest water level is to be distinctly 
 marked on the gauge glass, and also in a prominent place on the boiler 
 shell. The level of the flue, at its highest point, is also to be plainly and 
 permanently marked on the shell of the boiler in way of the ship's beam. 
 
 Two water-gauge glasses are to be fitted on the boiler at the standard 
 level of the water when the vessel is inclined as above, one on each side of 
 the boiler, as far apart as possible, and symmetrically placed right and left 
 of centre line of boiler. 
 
 When these two gauge glasses are fitted, the additional means named in 
 Rule 5, for ascertaining the water level, may be dispensed with. 
 
 The marks indicating lowest water level are to be not less than 5 centi- 
 metres long, and are to be cut into the plating near to the water gauges. 
 They are also to be distinguished by having the letters N. W. stamped on 
 them. They must be on before the inspection of the boiler is made, and 
 must be visible after it is lagged. Where water gauges of main boilers of 
 river steamers and barges and of donkey boilers are direct on the boiler 
 shell, a plate must be fitted at the lowest water - level mark with 
 " Niedrigster Wasserstand " on it, and an arrow pointing to the glass. 
 Where water gauges are connected to boilers by pipes (as must be the 
 arrangements in all sea-going steamships) a similarly marked plate is to be 
 fixed to the stand-pipe. The highest point of the heating surface is to be 
 indicated as follows: — Where the water gauge is fitted direct to the boiler 
 shell a plate indicating the point in question, and marked " Hochster 
 Feuer beriihrter Punkt," is to be fixed near the gauge. Where the gauge is 
 connected to the boiler by pipes a similar plate is to be affixed to the gauge 
 standard. These plates are to be permanently secured — not with removable 
 screws. 
 
 In cases where the fore and aft trim of the vessel is subject to consider- 
 able variation, the minimum water level (paragraph 2) must either be raised 
 
938 
 
 MANUAL OF MARINE ENGINEERING. 
 
 10 centimetres, or the mark indicating highest point of heating surface must 
 be raised a similar amount. The permanent mark showing regulation 
 lowest water level must, however, correspond with the particulars given 
 on the maker's name-plate, as hereafter specified. 
 
 \ third water gauge must be fitted on every double-ended boiler. 
 
 8. Safety -Valves. — Each boiler must be fitted with, at the least, one 
 reliable safety- vajye. 
 
 When severar boilers have a common steam chest, from which they 
 cannot be separately disconnected, two safety-valves will be sufficient. 
 
 Boilers of steamships, locomotives, and portable engines must have at 
 least two safety-valves. 
 
 In steamships, except those that are sea-going, one valve is to be placed 
 in such a position that the load on it can easily be ascertained from the 
 deck. 
 
 The valves must be so arranged that they can be readily lifted at any 
 time. They are to be loaded so that they will blow off immediately the 
 working pressure is reached. 
 
 Safety-valve areas must be such that after firing up for 15 minutes the 
 pressure does not accumulate more than 10 per cent. 
 
 Provided that the springs are satisfactorily fitted and loaded, the area of 
 sifety-valves. in square millimetres, may be as follows for each square 
 metre of heating surface : — 
 
 Pressure above atmosphere, 
 Area in square millimetres, 
 
 5 
 131 
 
 6 
 112 
 
 7 
 98 
 
 8 
 86 
 
 9 
 
 79 
 
 10 
 
 72 
 
 11 
 66 
 
 12 
 60 
 
 13 
 
 56 
 
 14 
 54 
 
 15 
 52 
 
 ■el 
 
 61 1 
 
 For boilers worked under forced draught, or in which the heating surface 
 is less than 35 times the grate surface, the safety-valve areas must be 
 correspondingly increased. Safety-valve boxes must be fitted with drain 
 pipes that cannot be shut off. The waste steam pipes from all safety-valves 
 must blow off into the open air. Blowing off into a tank, or into the 
 chimney, will not be permitted. The thickness of the washers which 
 determine the compressions of the safety-valve springs will be fixed by the 
 Official Surveyor and recorded in his book, and no alterations must be made 
 except by his authority. Exception is made, however (by the regulations 
 of 19th December, 1883), in the case of long voyages, and the chief engineer 
 of the vessel is permitted to re-adjust the washers in case the valves blow off 
 too freely. To meet this case the washers may be fitted in halves. 
 
 9. Pressure Gauge. — Each boiler must have a reliable pressure gauge on 
 which the highest regulation working pressure is to be plainly marked. 
 
 Boilers of steamships must have two pressure gauges — one clearly 
 visible to the fireman, the other, except when the vessel is sea-going, in 
 a, convenient position for being observed from the deck. If a vessel has 
 several boilers connected to a common steam chest, one pressure gauge on 
 ■deck will be sufficient, in addition to the one on each boiler. 
 
 Gauges to be placed where they can easily be seen. Dials to be marked 
 in kilogiammes per square centimetre (or in atmospheres, — 1 atmosphere 
 equals 1 kilogramme per square centimetre), and marking must go up to 
 50 per cent, above the regulation working pressure. In vessels having 
 more than one boiler, one pressure gauge (in prominent position in the 
 engine room), in addition to those on the boilers, will suffice. The pipe to 
 
GERMAN GOVERNMENT RULES FOR BOILEHS. 
 
 939 
 
 this gausje must be connected to all boilers, but must be capable of being 
 shut off from each. 
 
 10. Name Plate on Boiler. — (1) Each boiler must have the highest 
 working pressure, the maker's name, the shop number of the boiler, and 
 the date of completion, plainly and permanently stamped on it. Steam- 
 ships, moreover, must also Lave the standard lowest water level marked on. 
 
 (2) Boilers that have already been constructed, at the date of these 
 regulations, are not required to be altered to comply with them. 
 
 (3) The regulations for boilers of steamships apply in all cases in which 
 boilers are permanently fixed in, or connected with, the vessel. 
 
 The address of the maker must be given as well as his name. 
 
 10a. These particulars must be stamped on a metal plate, which is to be 
 fixed on the boiler by copper rivets, the heads of which are to be not less 
 than 12 millimetres in diameter. The plate to be in such a position that it 
 is easily visible when there are casings on the boiler. 
 
 11. Testing of Boilers. — Every new boiler must, on completion, be tested 
 by water pressure. 
 
 The test pressure for boilers intended for a working pressure of not 
 more than 5 atmospheres to be double the working pressure, and for boilers 
 
 intended for a working pressure of more than 5 
 atmospheres, the test pressure is to be 5 atmos- 
 pheres above the working pressure. 
 
 For boilers where the working .pressure is 
 under atmospheric pressure, the test pressure is 
 to be 1 kilogramme per squai-e centimetre. The 
 plates of the boiler must withstand the test 
 pressure without showing any permanent change 
 of form, and must also remain tight under the 
 pressure. 
 
 The boiler is to be regarded as leaky, if the 
 water at the highest pressure leaks through in 
 other forms than dewy moisture or pearly drops. 
 11a. When the boilers have been tested and 
 found satisfactory, the copper rivets which fix 
 the brass plate are to be officially stamped, and 
 an impression of this mark is to be inserted on 
 the certificate of test. 
 
 12. When boilers have undergone a thorough 
 repair in the boiler shop, or when they have had 
 to be stripped and extensively repaired on board, 
 they must be tested by hydraulic pressure the same as new boilers. 
 
 When in boilers with internal furnace, the furnace has been taken out 
 to repair or renew, and when in locomotives the fire-box has been taken 
 out to repair or renew, or when in cylindrical boilers one or more plates 
 have been renewed, the boilers must, on completion of such repairs, be 
 tested by hydraulic pressure as above. 
 
 When a boiler has undergone considerable repairs the steam pressure to 
 be allowed must be determined afresh. 
 
 13. Test Gauge. — The test pressure is to be registered by an open 
 mercurial gauge, or by the standard pressure gauge provided by the 
 inspecting official. 
 
 Each boiler to be tested must be fitted with a connection to suit the 
 
9 tO MANUAL OF MARINE ENGINEERING. 
 
 standard gauge, and by which the gauge may be applied by the inspecting 
 officer. 
 
 To facilitate application of official test gauge, a flanged branch is to be 
 provided immediately below the usual pressure gauge. The opening in the 
 flange and cock must not be less than 7 millimetres, the length of the 
 branch must be 60 millimetres, and its breadth 25 millimetres. Form to 
 be as shown in sketch (fig. 336). 
 
 RULES FOR STRENGTH OF BOILERS. 
 
 5th August, 1890. 
 The quantities are expressed in centimetres and kilogrammes : — 
 
 (1 kilo, per sq. cm. = 1 atm. = 14*223 lbs. per sq. in English). 
 
 Formulae for thickness of shell plate and working pressure : — 
 
 TTT , . S x R x 2T 
 Working pressure = =r = — • 
 
 T .., f , " P x D x F 
 lhickness of plate = -^ ^ „• 
 
 O X XV X L 
 
 Where P = highest steam pressure (kilos, per sq. cm.). 
 S = lowest tensile strength of material. 
 T = thickness of shell plate. 
 D = greatest internal diameter of boiler. 
 
 4*75 when rivet holes in longitudinal seams are drilled 
 
 and rivets closed by hydraulic pressure. 
 4*5 when longitudinal seams are double strapped and have 
 
 holes drilled and rivets closed by hydraulic pressure. 
 Pitch - Diameter of rivet 
 
 F = 
 
 R = 
 
 Pitch 
 
 ~ T , Area of rivet x No. of rivets in pitch x a n _ S, 
 
 Or, R= : — = z-7-, r i ,. x0'9x-J. 
 
 pitch x thickness ot plate b 
 
 w , _ (1 for lap and single riveted butt joints. 
 
 vvnere a = } 2 for double riveted, &c, butt joints. 
 
 Sj = Tensile strength of rivet material. Unless actual test has 
 been made, this is to be taken as 3300 for iron, and 
 3800 for steel ; but it must not exceed 4800 kilos, per 
 sq. cm. in any case. 
 
 The lower of these two values of R must be used. 
 
 S is to be taken as 3300 for iron; for steel the lowest value given by 
 the tests is to be used, but this must not exceed 5000 kilos, per sq. cm., 
 and the material must show an elongation of at least 20 per cent, in a 
 length of 20 centimetres. Steel of higher tensile strength may only be 
 used under specially arranged conditions as to tests, &c. 
 
 In riveted joints the shearing strength of the rivets must not be less 
 than the tensile strength of the plate left between holes. 
 
 Butt-straps are to be of the same quality as the shell plates, and of not 
 less than 75 per cent, of their thickness. 
 
GERMAN GOVERNMENT RULES FOR BOILERS. 
 
 941 
 
 In boilers without a central circumferential seam the outside butt-strap 
 must not be thinner than the shell plate. If the shell plates are thicker 
 than 1*25 centimetres, the central circumferential seam must be double 
 riveted, and if they are 2 -5 centimetres or more it must be treble riveted. 
 
 Rivet holes must be at least one diameter of rivet from edge of plate. 
 Diameter of rivet hole must not be less than thickness of plate, or more 
 than 2 x thickness of plate, the former limit to apply to thick plates and 
 the latter to thin ones. 
 
 Diagonal pitch must not be less than 2 - l x diameter of rivet. 
 
 Section 2. — Cylindrical Furnaces. 
 The thickness of plate for cylindrical furnaces is given by the formula — 
 
 Thickness = K N /P xLx D + 0-2cm. 
 
 Where P = working pressure (kilos, per sq. cm.). 
 
 L = length between stiffeners (centimetres). 
 
 D = external diameter of furnace (centimetres). 
 
 K = O0033 when the longitudinal joints are welded and fitted 
 
 with double butt-straps, and the furnace is truly 
 
 cylindrical. 
 K = 00035 when the longitudinal joints are lapped only. 
 
 The thickness of plate must not, in any case, be less than that given by 
 the formula — 
 
 P x D 
 
 Thickness = — ~ — + 0-2 cm. 
 \j 
 
 Wherein = 670 for plain furnaces without Adamson rings or other 
 stiffeners. 
 
 C = 820 for furnaces with one Adamson ring, when the un- 
 supported distance between the ring and the ends is 
 not more than 122 cm. 
 
 C = 920 for furnaces with two Adamson rings, when the un- 
 supported distance between the stiffeners is not more 
 than 79 cm. 
 
 C = 1160 for Fox or Morison corrugated, or Purves ribbed 
 furnaces. In corrugated furnaces D is the external 
 diameter of the depressed corrugation, and in ribbed 
 furnaces the external diameter of the plain portion 
 between the ribs is to be taken. 
 
 The thickness proposed for any type of furnace, such as Farnley, 
 Deighton, «kc, must be referred for special consideration. 
 
 For the boilers of river steamers the constant addition may be 0*1 cm. 
 in place of 0*2 cm. 
 
 Section 3. — Screw and Through Stays. 
 
 The strength of screw and through stays must not be taken higher 
 
 than — 
 
 500 kilos, per sq. cm. for iron. 
 
 600 „ .. „ steel. 
 
 oO 
 
912 MANUAL OF MARINE ENGINEERING. 
 
 Steel stays must not be welded. 
 
 In welded iron stays the section must be increased 20 per cent. 
 The pitch of screw stays in combustion chamber backs must not exceed 
 20 cm., and that of through stays in the steam space 45 cm. 
 
 Section 4. — Flat Surfaces. 
 
 The thicknesses of stayed flat surfaces are to be those given by the 
 formula — 
 
 Thickness = 0T5 cm. + P*J^- 
 
 Wherein p = greatest pitch of stays (in centimetres). 
 
 P = working pressure (kilos, per sq. cm.). 
 
 K = 2200 for stays screwed in and riveted over. 
 
 K = 2800 for stays screwed in and fitted with nuts. 
 
 K = 3000 for stays fitted with inside and outside nuts and 
 washers, the outer washers being three times the 
 diameter of the stay end and half the thickness of 
 the plate. 
 
 K = 3400 for stays fitted with inside and outside nuts and 
 washers, the outer washers having a diameter of 
 •6 x pitch, and a thickness of "66 x thickness of plate, 
 and being riveted to the plate. 
 
 If the combustion chamber top is supported by bridge stays only, which 
 rest on the tube-plate, and are not assisted by any sling rods from the boiler 
 shell, the thickness of the tube-plate must not be less than that given by 
 the formula — 
 
 ™, . , P x W x p 
 
 Thlckne8S = 1800 (p - dy 
 
 Wherein P = working pressure (kilos per square cm.). 
 
 W = width of chamber (tube-plate to back), in cm. 
 p = pitch of tubes, in cm. 
 d = internal diameter of plain tube, in cm. 
 
 If the tube-plates are of steel an 1 are not exposed to the direct action of 
 flame, the thickness may be reduced by 12£ per cent. 
 
 • Section 5. — Combustion Chamber Bridge Stays. 
 
 The strength of these stays is to be that given by the following 
 loi mula : — 
 
 ml . , P (L - p) S x L 
 
 Thickness = — v ^ *' . 
 
 K x d 
 
 Wherein P = working pressure (kilos, per square cm.). 
 L = length of bridge stay, in cm. 
 p = pitch of stay bolts in bridge stay, in cm. 
 S = distance apart of bridge stays, in cm. 
 d = depth of bridge stay at centre, in cm. 
 K = 420 for one stay bolt in each bridge stay. 
 K = 630 for two or three stay bolts 
 K = 720 for four stay bolts. 
 
GERMAN GOVERNMENT RULES FOR BOILERS. 943 
 
 " Thickness " stands for the collective thickness of the two plates of the 
 bridge stay if constructed in that way. 
 
 If the plates are of steel they may be 10 per cent, thinner; and a further 
 reduction may be made if they are assisted by slings from above. 
 
 GENERAL. 
 
 Manholes, <fec, must be fitted with stiffening rings to compensate for 
 plate cut away ; they should not, as a rule, be less than 30 cm. x 40 cm., 
 but in special cases 26 cm. x 38 cm. may be accepted. Cast-iron manhole 
 covers cannot be accepted. It is strongly recommended that the studs of 
 manhole covers be both screwed into the covers and riveted over on the 
 inside. 
 
 Boilers are to be properly secured against movement both fore and aft 
 and athwartships. 
 
 When there are several boilers they must be provided with proper shut- 
 off valves to enable them to be worked independently. 
 
 Where the donkey boiler works at a lower pressure than the main 
 boilers, satisfactory arrangements must be made to prevent steam passing 
 from the latter into the former. 
 
 The space between the combustion chamber back and the back plate of 
 the boiler should not be less than 12 cm., to allow for proper cleaning. In 
 the cases of extra large boilers, and of boilers worked under forced draught, 
 a greater space should be allowed. 
 
 Every marine boiler must be fitted with a blow-off valve and pipe, and 
 also with a scum valve to draw the water from the level of the lower end of 
 the water gauge. The blow-off pipe must be fitted with one cock on the 
 boiler and a second on the ship's skin. 
 
 Wood must not be used either for lagging boilers or bulkheads in the 
 neighbourhood of boilers. 
 
 All the larger mountings are to be secured to strong flanges riveted to 
 the shell of the boiler ; and the studs securing the mountings are not to 
 penetrate the shell. 
 
 TESTS OP MATERIAL FOR STEEL BOILERS. 
 
 Regulations of 1st July, 1899. 
 
 All material to be of good and tough quality ; any that proves brittle in 
 working to be rejected. 
 
 A range of 5 kilos, will be allowed where ultimate tensile strength is 
 42 kilos, or under, aud 6 kilos, where ultimate tensile is above 42 kilos. 
 
 Shell plates and other material not exposed to flame must show an 
 elongation of at least 20 per cent, in a length of 200 millimetres ; and 
 material exposed to flame must show at least 22-5 per cent, on same length. 
 
 Material of less than 38 kilos, ultimate strength must elongate at least 
 25 per cent. 
 
 Bending tests are to be made on strips 25 millimetres wide, that have 
 been heated to a dark red and quenched in water at 28° C. 
 
 They must stand bending without cracking to an angle of 180° over a 
 radius, for shell plates, &c, of four times the thickness of the plate ; and for 
 material exposed to flame, of twice the thickness of the plate. 
 
 A tensile and a bending test is to be made from each shell plate. As 
 
94 4 MANUAL OF MARINE ENGINEERING. 
 
 regards other plates, a tensile test is to be made from every fourth plate"; 
 but, as far as possible, these test pieces should be so selected that there may 
 be one from each furnace charge. 
 
 As regards plates other than shell plates and not exposed to flame, a 
 bending test is to be made from each furnace charge. A bending test must 
 be made from each plate that is exposed to flame. 
 
 Plates of over 41 kilos, tensile must be annealed if they have been 
 •worked in the fire. 
 
 Plates that are to be subjected to the direct action of flame should no$ 
 have a higher ultimate tensile than 42 kilos, per square millimetre. 
 
 Steel of more than 46 kilos, ultimate tensile must not be used. 
 
Lloyd's rules fob electrtc light installations. 945 
 
 APPENDIX L, 
 
 Lloyd's Rules for Electric Light Installations on Board Vessels. 
 
 The following requirements as to the sizes, positions, and protection 
 of the cables, and to the fitting oi the cut-outs are now embodied as rules : — 
 
 leads or circuits. 
 
 1. The sectional area of the copper wires in the cables should be at least 
 in the proportion of 1 square inch per 1,000 amperes carried. 
 
 2. No single wire of greater size than 14 or of less than 18 standard wire 
 gauge should be used. For portable leads, cables composed of stranded 
 wires should be used having sufficient conductivity and flexibility for the 
 purposes intended. 
 
 3. The copper used in all wires or cables should have a conductivity of 
 at least 98 per cent, that of pure copper. 
 
 4. The insulation resistance of all wires, including portable leads, should 
 be not less than 600 megohms per statute mile, after 24 hours' immersion in 
 sea-water. 
 
 5. The insulating material used must not appreciably soften if subjected 
 to a temperature of 180° F. If india-rubber insulation is used, the wires 
 should be first covered with a layer of pure rubber, then with a separator, 
 then with a layer of india-rubber coated tape. The whole should then be 
 vulcanised together. The cable should afterwards be satisfactorily pro- 
 tected, preferably with a braided covering of waterproof fibre. 
 
 6. Wires which are insulated with any other material than india-rubber 
 should fulfil the same conditions as to insulation resistance, and should be 
 of equal durability with those above specified. 
 
 JOINTS. 
 
 7. Joints in branches, or of branches with leads of small circuits, must 
 be made in properly constructed water-tight junction boxes, or should have 
 the copper wires thoroughly soldered and the insulation carefully carried 
 out, all the joints being made watertight. Joints in flow and return wires 
 should not be made opposite one another. All joints should be in accessible 
 positions, none being made in bunkers, cargo spaces, or spaces which may 
 at any time be used for carrying cargo, stores, or baggage. 
 
 8. For soldering wires, resin only should be used as a flux. 
 
 9. Where practicable, the leads should be placed where they can be 
 always accessible ; if they are laid in wood battens the covers should be 
 screwed on, not nailed, and care should be taken that the casings are so 
 arranged that water will not lodge in them. Cables which are properly 
 
946 MANUAL OF MARINE ENGINEERING. 
 
 covered with protective metal sheathing, or which are protected by galvanised 
 wire armouring may be unencased. They should, however, be secured by 
 screwed clips, not by staples. All sharp bends in cables should be avoided. 
 
 10. All cables which are liable to be exposed to the weather or moisture 
 should be lead covered, or be otherwise specially protected. Where great 
 heat is experienced, no wood casing should be used, but the 'cables should 
 be protected by iron casings, or if they are not exposed to mechanical injury, 
 they may be armoured with galvanised wire and fastened to decks or bulk- 
 heads with screwed clips spaced not more than 12 inches apart. 
 
 11. If cables are led through cargo spaces, coal bunkers, or spaces which 
 may at any time be used for carrying cargo, stores, or baggage, or which 
 are not at all times accessible, they should be strongly protected against 
 damage, preferably by iron casings. If they are led through metal tubes, 
 these must be strongly secured, and should be fitted so that the water cannot 
 lodge in them. 
 
 Armoured cables may be used without casings or tubes provided they 
 are strongly secured to the under side of decks or to bulkheads by screwed 
 clips, and provided they are armoured in conformity with the standard of 
 the Engineering Standards Committee, viz. : — 
 
 For cables below \ inch diameter over lead by galvanised steel wires 
 •072 diameter, for cables \ inch to 1 inch over lead by two layers of steel 
 tape, each *03 inch thick, for cables above 1 inch to 2 inches diameter by 
 two layers of steel tape, each "04 inch thick, and for larger cables by two 
 layers of steel tape *06 thick. 
 
 12. Where cables pass through beams, bulkheads, or other iron work, 
 they should be led through special fittings of sheet lead, hard wood, or vul- 
 canised fibre to prevent their being chafed, and where they pass through 
 decks they should be led through metal tubes lined with wood, or vulcanised 
 fibre, and securely fastened to the decks, standing at such a height above 
 the deck level that water cannot stand above them. Where cables pass 
 through water-tight bulkheads the fittings should be provided with brass 
 water-tight screwed glands. 
 
 13. In vessels having spaces allotted alternately for passengers and 
 cargo, the lamp fittings in these spaces should be removable, and the ter- 
 minals so arranged that they can be properly covered up with strong metal 
 covers, or the whole of the fittings should be similarly provided with strong 
 metal covers. The main switches and cut-outs should be outside these 
 spaces, or if placed inside, they should be in strong iron boxes provided 
 with iron covers, or otherwise securely arranged to prevent the fittings being 
 tampered with. 
 
 DISTRIBUTION. 
 
 14. A main switchboard should be fitted in the dynamo-room, to which 
 all the main circuits throughout the ship should be brought, a switch and 
 cut-out being fitted thereon for each circuit. The auxiliary switchboards 
 for further sub-division of the current should be placed in conveniently 
 accessible positions, and each such switchboard should be similarly fitted 
 with a separate switch and cut-out for each sub-circuit. Cut-outs should 
 be fitted to each lamp circuit where these are made with reduced size of wire. 
 
LLOYD S RULES FOR ELECTRIC LIGHT INSTALLATIONS. 947 
 
 If vessels are wired on the double-wire system cut-outs should be fitted to 
 each cable of these circuits. 
 
 15. In cases where electric lights are used for the mast-head light and 
 side lights, the switches controlling these lights should be placed in a position 
 where they can be controlled by the officer of the watch, or other responsible 
 person, and cannot be tampered with by other members of the crew, or by 
 passengers, etc. 
 
 16. The switchboards should be of slate or other incombustible material. 
 The switches should be on the quick-break principle, and should be so con- 
 structed that they must be either full " on " or completely " off " — that is, 
 they must not be able to remain in an intermediate position. They should 
 have ample rubbing surfaces and their conductivity should not be less than 
 that of the wires connected to them. 
 
 17. Cut-outs should be fitted to each main or auxiliary circuit.'on the 
 switchboards, as near as possible to the switches of these circuits. If the 
 switchboard is not fitted near the dynamo, or if more than one dynamo 
 may be used on any one circuit, then cut-outs should also be fitted to the 
 main cable as near as possible to each of the dynamo terminals. 
 
 18. All other cut-outs should also be in easily accessible places, and as 
 near as possible to the commencement of the cables or wires they protect. 
 They should be mounted on slate or other incombustible bases, and be arranged 
 so that the fused metal may not be a source of danger, and where fitted 
 with covers these should be incombustible. 
 
 19. All fuses should be of easily fusible and non-oxidisable metal, and 
 should be so proportioned as to melt with a current 100 per cent, in excess 
 of the normal current ; that is, they should melt with a current in the pro- 
 portion of 2,000 amperes per scpiare inch of section of the wires they protect. 
 The fuses for branch wires to single lights should, however, if of tin wire, 
 be of not greater size than 22 S.W.G. 
 
 20. The fuses for each cable should be made of standard dimensions, 
 so that a large fuse cannot be used for a small cable by mistake, or, if wire 
 fuses are used, permanent instructions should be fitted on or near each switch- 
 board, giving particulars of the proper size of fuse for each circuit. 
 
 21. In shaft passages and in damp places, all lamp switches and cut-outs 
 should be of a strong water-tight pattern, or should be placed in water-tight 
 boxes having hinged or portable water-tight covers. No switches or cut- 
 outs are to be placed in bunkers. 
 
 22. There should be no joints in the "cables leading from the dynamo 
 to the main switchboard, nor in those leading from the main to auxiliary 
 switchboards, nor should branches to single lamps be taken off these cables. 
 
 23. A voltmeter should be supplied with each installation. If more than 
 one dynamo is fitted, neither being capable of the whole of the output, an 
 ampere meter should be supplied with each dynamo. 
 
 JOINTS W TH HULL. 
 
 24. In vessels fitted on the single-wire system, all the joints with the 
 hull should be placed in accessible positions. Those for single lamps or for 
 small cables should be made with brass screws not less than three-eighth 
 of an inch in diameter, carefully tapped into the iron or steel, having white 
 
948 MANUAL OF MARINE ENGINEERING. 
 
 brass washers, between the wires and the vessel, or the wires should 
 be soldered to brass-faced washers. For larger cables and for the pole of 
 dynamo the cable wires should be properly sweated into brass or copper 
 shoes, which should be bolted to the vessel. The iron or steel where contact 
 is made should be filed bright, and the area of contact should not be less 
 than eight times the section of the copper of the cable. 
 
 IN VESSELS CARRYING PETROLEUM. 
 
 25. The single-wire system must not be adopted for any part of the 
 installation. Switches and cut-outs must not be fitted in places liable to 
 the accumulation of petroleum vapour or gas, and all lamps in places where 
 it is possible for gas to accumulate must be made with an outer glass globe 
 made air-tight. All wires in such places are to be lead covered, or the 
 insulation of the cables employed is to be of such a nature as not to be 
 affected by petroleum. No joints of cables, switches, or cut-outs should 
 be fitted in the pump-room, but the wires for each lamp therein should be 
 carried to the lamp from a distributing junction box placed outside the 
 pump-room or companion. 
 
 The following paragraphs referring to the effect of the electric light 
 installations upon the compasses are issued as suggestions, not as rules : — 
 
 POSITION OF DYNAMOS AND OF ELECTRIC MOTORS. 
 
 26. The position and type of dynamos and electric motors should be such 
 that the compasses will not be affected. Dynamos and large motors should 
 be at least 30 feet from the standard compass. 
 
 CABLES. 
 
 27. In vessels fitted with continuous-current dynamos, and wired on 
 the single- wire system, no single cable should be carried within 15 feet of 
 any compass, and cables conveying heavy currents should be fixed at still 
 greater distance. If it is necessary to fix any cables within this distance, 
 then for all parts of the vessel lighted from this cable the concentric or double 
 wire system should be adopted, the return wire being carried as near the 
 flow as possible in the vicinity of the compasses. 
 
 London, 11th June % 1909. 
 
Lloyd's registers' rules tor refrigerating machinery, etc. 949 
 
 APPENDIX M. 
 
 Refrigerating Machinery and Appliances. 
 
 Lloyd's registers' rules. 
 
 On the application of the owners of vessels titted for carrying refrigerated 
 cargoes, the Committee will authorise their surveyors to survey the refri- 
 gerating machinery and appliances, and in those cases where the following 
 conditions are complied with and a satisfactory report is received from the 
 surveyor, certificates of these surveys will be issued and the notation " R.M.C." 
 (in red) (i.e., Refrigerating Machinery Certificate) will be made against the 
 vessel's name in the Society's Register Book, and in the special list of vessels 
 fitted with refrigerating appliances. In cases in which the refrigerating 
 machinery and appliances are constructed under the special survey of the 
 Society's Surveyors and to their entire satisfaction, the notation " -f- R.M.C." 
 (in red) will be made in the Register Book. The name of the maker and 
 description and number of refrigerating machines, whether single or duplex, 
 and the refrigerating power of the machines will be recorded in the special 
 list in the Register Book, as will also the number and capacity of insulated 
 cargo chambers and the nature of insulation and the method employed for 
 cooling the holds. 
 
 1. The insulation must be sound and in good order and of efficient con- 
 struction. The details of construction showing the amount and nature of 
 the insulating material employed in the various parts are to be reported to 
 the Committee. 
 
 Bilge suction and sounding pipes and ballast tank air and sounding pipes, 
 passing through insulated spaces, should be well insulated to prevent their 
 being frozen up. No sluice valves, scuppers, or drain pipes are to be fitted, 
 which will permit drainage from spaces outside of the insulated chambers 
 into the bilges of the insulated holds. 
 
 It is recommended that the woodwork of the insulation over tunnel tops 
 be fastened with screws to facilitate the examination of this part, and that 
 extra strong battens of American elm be fitted upon it under the hatches. 
 Insulated removable portions are to be arranged in the bulkhead insulation, 
 where required, to give easy access to sluice valves and bilge suction roses. 
 The bottoms, sides, and coamings of all insulated hatches and limbers should 
 be painted to prevent decay. 
 
 Thermometer tube flanges and covers should be arranged so that water 
 does not run down and freeze in them when taking the temperature. 
 
 Cargo battens should be provided for the floor or deck and the sides of 
 the chambers previous to loading the homeward cargo. Those for the sides 
 of the chambers should be fastened, and should be at least \\ inches in depth, 
 and 2 inches wide, one batten being placed over each frame or ground, the 
 
950 MANUAL OF MARINE ENGINEERING. 
 
 others being intermediately arranged. The battens for the floor and decks 
 should be at least 2 inches X 2 inches. 
 
 Where the brine system of refrigerating is employed, the brine circulating 
 pipes and tanks should not be galvanised on the inside. 
 
 In cases where internally galvanised tanks and cooling pipes have been 
 fitted, the brine cooling and return tanks, if closed, should be provided with 
 two ventilating pipes communicating with the atmosphere. If the tanks 
 are not closed, the cooling room should be efficiently ventilated. 
 
 2. The refrigerating machinery is to be of approved construction, and 
 of sufficient power to maintain the necessary low temperature in the cargo 
 chambers in tropical climates when running 18 hours per day. For cargo 
 capacities of above 70,000 cubic feet the machinery is to be either duplex 
 or in duplicate. 
 
 3. A sufficient amount of spare gear is to be supplied and stowed where 
 it is readily accessible. 
 
 No spare gear will, however, be required in cases where two complete 
 sets of refrigerating machines are fitted, each being of sufficient power to 
 maintain the necessary low temperature in the cargo chambers in tropical 
 climates when running 18 hours per day, provided all the working parts of 
 these machines are interchangeable. 
 
 When two similar machines are fitted, each connected to different cargo 
 compartments, one set of spare gear suitable for either machine will suffice. 
 
 Where one single dry air machine is fitted to each compartment, the 
 following will be required : — 
 
 1 crank-shaft with eccentric sheaves, complete, or one half shaft if the 
 
 halves are interchangeable. 
 1 piston-rod and nuts for steam and air cylinders. 
 1 set of piston-rod and connecting-rod brasses. 
 1 piston, complete, for each steam and air cylinder. 
 1 cylinder cover for each pattern used in steam and air cylinders. 
 1 air-pump bucket and rod. 
 1 circulating-pump bucket and rod. 
 1 pair main bearing brasses, complete. 
 Main and cut-off valves for each steam cylinder. 
 Balance springs and rings for steam and air slide valves. 
 False valve face for each pattern fitted in steam cylinders, with screws. 
 1 eccentric rod for each pattern used. 
 1 eccentric strap for each pattern used. 
 
 1 slide valve spindle and nuts for steam and air cylinders, for each pattern 
 
 used. 
 
 2 main bearing bolts. 
 
 1 set of connecting-rod and piston-rod bolts. 
 Full set of air valves and seats for air compressor. 
 
 1 set of inlet and outlet valves and 1 set valve faces (if fitted) for air- 
 expansion cylinders, with screws. 
 1 set of valves for air, circulating, and feed pumps. 
 1 set of escape valve springs. 
 50 suction springs. 
 50 delivery springs. 
 
Lloyd's registers' rules for refrigerating machinery, etc. 951 
 
 50 buffer springs. 
 
 6 tubes and 24 ferrules for condenser. 
 
 6 tubes for cooler. 
 
 6 tubes for air-drying chamber. 
 
 Assorted bolts, studs, and nuts. 
 
 1 set of lead-lined nuts for air-expansion cylinder cover. 
 
 A quantity of packings and joint rings. 
 
 Where one duplex or two single dry air machines are fitted to each com- 
 partment the following will be required : — 
 
 1 crank-shaft with eccentric sheaves complete, or one half shaft if the 
 halves are interchangeable. 
 
 1 piston-rod and nuts for steam and air cylinders. 
 
 1 set of connecting-rod and crosshead brasses. 
 
 1 piston for H.P. steam cylinder. 
 
 1 piston, complete, for air compressor ; and 1 for air-expansion cylinder. 
 
 1 set of piston springs for each steam cylinder. 
 
 1 cylinder cover for each pattern used in air compression and expansion- 
 cylinders. 
 
 1 air pump bucket and rod. 
 
 1 circulating pump bucket and rod. 
 
 Main and cut-off slide valves and spindles with nuts complete for H.P. 
 steam cylinder. 
 
 Balance springs and rings for steam and air slide valves. 
 
 1 H.P. steam cylinder valve and valve face with screws. 
 
 1 eccentric sheave, strap, and rod for each pattern used. 
 
 1 slide valve spindle and nuts for steam and air cylinders for each pattern 
 
 used. 
 
 2 main bearing bolts. 
 
 1 set of connecting-rod and piston-rod bolts. 
 
 Half set of air valves and seats for air compressor. 
 
 1 inlet and 1 outlet valve, and half set of valve faces (if fitted) for air 
 
 expansion cylinder, with screws. 
 1 set of valves for air, circulating, and feed pumps. 
 1 set of escape valve springs. 
 20 suction springs. 
 40 delivery springs. 
 40 buffer springs. 
 
 6 tubes and. 24 ferrules for condenser. 
 6 tubes for cooler and 6 for air-drying chamber. 
 Assorted bolts, studs, and nuts. 
 
 \ set of lead-lined nuts for air-expansion cylinder cover. 
 A. quantity of packings and joint rings. 
 
 Where one single ammonia or carbonic anhydride comoression machine 
 is fitted : — 
 
 1 crank-shaft with eccentric sheaves, complete, or one half shaft if the 
 
 halves are interchangeable. 
 Piston and rods complete with nuts for each steam cylinder and gas 
 
 compressor. 
 
952 MANUAL OF MARINE ENGINEERING. 
 
 1 air pump bucket and rod. 
 
 1 circulating pump bucket and rod. 
 
 1 pair main bearing brasses, complete. 
 
 1 set of connecting-rod and crosshead brasses. 
 
 Main and cut-off valves for steam cylinders. 
 
 1 valve spindle for each pattern used and nuts complete. 
 
 1 eccentric strap and rod for each pattern used. 
 
 1 brine pump complete. 
 
 1 cover for each pattern used. 
 
 2 main bearing bolts. 
 
 1 set of connecting-rod and piston-rod bolts. 
 
 1 set compressor suction and delivery valves with springs and boxes, 
 
 complete. 
 1 set of valves for air, circulating, feed, and brine pumps. 
 Crank-shaft for fan engine. 
 
 1 steam piston and rod, etc., for fan engine, complete. 
 1 pair of connecting-rod brasses for fan engines, with bolts, etc., complete. 
 1 set of blocks for making all leather packings used. 
 6 tubes and 24 ferrules for condenser. 
 Lengths and bends of piping of each size used, together with flanges, 
 
 couplings, and screwing apparatus for effecting repairs. 
 1 gas regulating valve. 
 1 distributing and 1 collecting piece with multiple branches for coils 
 
 for each pattern used. If these pieces are made of forged steel, 
 
 no spare pieces are required. 
 Sundry valves, cocks, flanges, and fittings. 
 Assorted bolts, studs, and nuts. 
 Quantity of leather packings and joint rings. 
 
 For ammonia and carbonic anhydride compression machines, the following 
 spare gear will be required, where one duplex or two single machines are 
 .fitted to each compartment : — 
 
 1 crank-shaft, or one half shaft if the halves are interchangeable. 
 
 1 steam piston-rod and nut for each pattern used. 
 
 1 piston for H.P. steam cylinder, with springs, complete. 
 
 1 set of piston rings for each steam cylinder. 
 
 1 set of piston rings for each size of compressor. 
 
 1 compressor piston-rod and nuts, complete, for each pattern used. 
 
 1 air-pump bucket and rod. 
 
 1 circulating pump bucket and rod. 
 
 Main and cut-off slide valves for H.P. steam cylinder. 
 
 Main and cut-off valve spindles and nuts for H.P. steam cylinder. 
 
 1 eccentric sheave, strap, and rod, for each pattern used. 
 
 1 brine pump complete. 
 
 1 cover for each end of gas compressor, except where screwed plugs are 
 
 used. 
 ■2 main bearing bolts. 
 
 | set of connecting-rod and piston-rod bolts. 
 | set compressor suction and 1 delivery valve with springs and box, 
 
 complete. 
 
Lloyd's registers' rules for refrigerating machinery, etc. 953 
 
 1 set of valves for air, circulating, feed, and brine pumps. 
 
 1 steam piston and rod, etc., for fan engine, complete. 
 
 1 pair of connecting-rod brasses for fan engines, with bolts, etc.. complete. 
 
 1 set of blocks and leather for making all leather packings used. 
 
 6 tubes and 24 ferrules for condenser. 
 
 Lengths and bends of piping of each size used, together with Hanges, 
 
 couplings, and screwing apparatus for effecting repairs. 
 1 gas regulating valve. 
 1 distributing and 1 collecting piece with multiple branches for coils 
 
 for each pattern used. If these pieces are made of forged steel, 
 
 no spare pieces are required. 
 Sundry valves, cocks, flanges, and fittings. 
 Assorted bolts, studs, and nuts. 
 A quantity of joint rings. 
 
 In cases where an independent circulating water pump is used, and its 
 work cannot be performed by the main or auxiliary engines, a duplicate 
 pump complete should be fitted. 
 
 In cases where an independent circulating water pump is used, and its 
 work can be performed by the main or auxiliary engines, a pump bucket 
 and rod should be carried, and h set of valves for water end. 
 
 In cases where an independent surface condenser with air, circulating 
 and feed pumps combined is fitted, and its work cannot be performed by the 
 main engines : — 
 
 1 crank-shaft with eccentric sheaves complete. 
 1 piston and rod complete for each pattern used. 
 1 eccentric strap and rod complete for each pattern used. 
 1 slide valve and spindle complete for each pattern used. 
 1 pump bucket and rod complete for each pattern used. 
 1 set of connecting-rod and piston-rod bolts and nuts. 
 1 set of valves for air, circulating and feed pumps. 
 6 condenser tubes and 12 ferrules. 
 
 In cases where an independent surface condenser with air, circulating, 
 and feed pumps combined is fitted, and its work can be performed by the 
 main engines : — 
 
 1 pump bucket and rod complete for each pattern used. 
 1 set of connecting-rod and piston-rod bolts and nuts. 
 •| set of valves for air, circulating, and feed pumps. 
 6 condenser tubes and 12 ferrules. 
 
 PERIODICAL SURVEYS. 
 
 9. The complete periodical survey required in par. 4 will consist of the 
 following : — 
 
 The insulation throughout the holds is to be carefully examined and 
 tested for dryness and fulness by sounding with a hammer and by boring. 
 The test holes are to be afterwards efficiently closed. Special attention is 
 to be paid to the spaces under the snow boxes, trunks, and hatches where 
 dampness may accumulate, to the sides under stringers and under decks 
 
954 1IANUAL OF MARINE ENGINEERING. 
 
 and to the tunnel tops. All limber hatches are to be removed, the limbers 
 cleared, and the suction pipes and roses, sluices, and sounding pipes are to 
 be examined. Hatches, air trunk-ways, and thermometer tubes with their 
 connections and fastenings are to be examined, and where trunk- ways pass 
 through water-tight bulkheads, the water-tight doors are to be examined 
 and worked. 
 
 The trunk-ways should be as air-tight as practicable, and their fastenings 
 should be secure. 
 
 The steam pipes, water pipes, and connections, the crank-shaft and 
 bearings, connecting-rods, steam and air cylinders, pistons, slides and valves, 
 compressors and pistons, compressor rods and glands, surface condenser, and 
 air or gas coolers, circulating, air, feed, and bilge pumps, are to be carefully 
 examined, and the condensers and coolers tested if deemed necessary. 
 
 The auxiliary machinery, where fitted, is also to be examined. 
 
 The spare gear is to be examined. - 
 
 In dry air machines special attention is to be given to the condition of 
 the air expansion cylinders, their pistons and valves. In other machines 
 special attention is to be given to the condition of the compressors, including 
 the pistons, rods, and glands, and to the expansion valve. 
 
 The refrigerator coils and their connections and the brine pipes and tanks, 
 where fitted, are to be carefully examined and tested if deemed necessary. 
 
 Where the brine may escape to the bilges, the cement is to be examined. 
 
 The machinery is to be examined under working conditions and tested 
 on the snow box or refrigerators, the time and fall of temperature being 
 noted. 
 
LLOYDS RULES FOR BURNING AND CARRYING LIQUID FUEL. 955 
 
 APPENDIX N. 
 
 Lloyd's Eules for Vessels Fitted for Burning and Carrying 
 
 Liquid Fuel. 
 
 (Section 49, Steel Ship Rules.) 
 
 1. In vessels fitted for burning liquid fuel, the record " fitted for liquid 
 fuel " will be made in the Register Book. 
 
 2. The compartments for carrying oil fuel must be strengthened to 
 withstand efficiently the pressure of the oil when only partly filled and in a 
 seaway. . They must be tested by a head of water extending to the highest 
 point of the filling pipes or 12 feet above the load line, or 12 feet above the 
 highest point of the compartment, whichever of these is the greater. 
 
 3. If peak tanks or other deep tanks are used for carrying liquid fuel 
 the riveting of these should be as required in the case of vessels carrying 
 petroleum in bulk. The strengthening of these compartments must be to 
 the committee's satisfaction. 
 
 4. Each compartment must be fitted with an air pipe to be always 
 open discharging above the upper deck. 
 
 5. Efficient means must be provided by wells and sparring or lining to 
 prevent any leakage from any of the oil compartments from coming into, 
 contact with cargo or into the ordinary engine-room bilges. 
 
 6. If double bottoms under holds are used for carrying liquid fuel, the 
 ceiling must be laid on transverse battens, leading at least two inches air 
 space between the ceiling and tank top and permitting free drainage from 
 the tank top into the limbers. 
 
 7. The pumping arrangements of the oil fuel compartments and their wells 
 must be absolutely distinct from those of other parts of the vessel, and must 
 be submitted for approval. 
 
 If it is intended to carry sometimes oil and sometimes water ballast in 
 the various compartments of the double bottom, the valves controlling the 
 connections between these compartments and the ballast donkey pump, 
 and also those controlling the suctions of the special oil pump, must be so 
 arranged that the suction for each separate compartment cannot be con- 
 nected at the same time to both pumps. 
 
 8. No wood fittings or bearers are to be fitted in the stokehold spaces. 
 
 9. Where oil fuel compartments are at the sides of, or above, or below 
 the boilers, special insulation is to be fitted where necessary to protect them 
 from the heat from the boilers, their smoke boxes, casings, etc. 
 
 10. If the fuel is sprayed by steam, means are to be provided to make 
 up for the fresh water used for this purpose. 
 
 11. If the oil fuel is heated by a steam coil the condensed water should 
 not be taken directly to the condenser, but should be led into a tank or an 
 open funnel mouth, and thence led to the hot-well or feed tank. 
 
 12. The above arrangements are applicable only to the case of oil fuel, 
 the flash point of which, as determined by. Abel's close test, does not fall 
 below 150° Fahrenheit. 
 
 London, 17th June, 1909. 
 
956 
 
 MANUAL OF MARINE ENGINEERING. 
 
 APPENDIX O. 
 
 Board of Trade Rules for Safety-valves. 
 
 Provisions of the Act as regards Safety-valves. — Every steamship of which 
 a survey is required by the Act shall be provided with a safety-valve upon 
 each boiler, so constructed as to be out of the control of the engineer when 
 the steam is up, and, if such valve is in addition to the ordinary valve, it 
 shall be so constructed as to have an area not less, and a pressure not 
 greater, than the area of and pressure on that valve. 
 
 The surveyors are instructed that, in all new boilers and whenever 
 alterations can be easily made, the valve chest should be placed directly on 
 the boiler ; and the neck, or part between the chest and the flange which is 
 bolted on to the boiler, should be as short as possible and be cast in one 
 with the chest. 
 
 If any person place an undue weight on the safety-valve of any steam- 
 ship, or, in the case of steamships surveyed under the Act, increase such 
 weight beyond the limits fixed by the engineer-surveyor, he shall, in 
 addition to any other liability he may incur by so doing, be liable for each 
 offence to a fine not exceeding one hundred pounds. 
 
 The engineer-surveyor shall declare, amongst other things, the limits of 
 the weight to be placed on the safety-valves ; that the safety-valves are 
 such, and in such condition, as required by the Act ; and that the 
 machinery is sufficient for the service for the time he fixes, and is in good 
 condition. 
 
 Area of Safety-valves. — When natural draught is used, the area per 
 square foot of fire-grate surface of the locked-up safety-valves should not be 
 less than that given in the following table opposite the boiler pressure 
 intended, but in no case should the valves be less than 2 inches in diameter. 
 This applies to new vessels and to vessels which have not previously 
 received a passenger certificate. 
 
 When, however, the valves are of the common description, and are 
 made in accordance with the table, it will be necessary to fit them with 
 springs having great elasticity, or to provide other means to keep the 
 accumulation within moderate limits. 
 
 When forced draught is used, the area of the safety-valves should not be 
 less than that found by the following formula : — 
 
 — ^r — = area of valves required. 
 
 Where A = area of valves as found from the table. 
 
 C = estimated coal consumption in lbs. per sq. ft. of grate per hour. 
 
BOARD OF TRADE RULES FOR SAFETY-VALVES. 
 
 TABLE CXIIL— Safety-valve Areas, 
 (natural draught.) 
 
 957 
 
 
 Area of 
 
 
 Area of 
 
 
 Area of 
 
 A 
 
 rea of 
 
 Boiler 
 
 Valve per 
 
 Boiler 
 
 Valve per 
 
 Boiler 
 
 Valve per 
 
 Boiler Val 
 
 ve per 
 
 Pressure. 
 
 Square Foot 
 
 Pressure. 
 
 Square Foot 
 
 Pressure. 
 
 Square Foot 
 
 Pressure. Squs 
 
 ire Foot 
 
 
 of Fire-grate. 
 
 
 of Fire-grate. 
 
 
 of Fire-grate. 
 
 of Fi 
 
 re-grate. 
 
 Lbs. 
 
 Sq. inch. 
 
 Lbs. 
 
 Sq. inch. 
 
 Lbs. 
 
 Sq. inch. 
 
 Lbs. Sq 
 
 inch. 
 
 15 
 
 1-250 
 
 67 
 
 •457 
 
 119 
 
 •279 
 
 170 
 
 202 
 
 16 
 
 1-209 
 
 68 
 
 •451 
 
 120 
 
 •277 
 
 171 
 
 201 
 
 17 
 
 1171 
 
 69 
 
 •446 
 
 121 
 
 •275 
 
 172 
 
 200 
 
 18 
 
 1-136 
 
 70 
 
 •441 
 
 122 
 
 •273 
 
 173 
 
 199 
 
 19 
 
 1*102 
 
 71 
 
 •436 
 
 123 
 
 •271 
 
 174 
 
 19S 
 
 20 
 
 1-071 
 
 72 
 
 •431 
 
 124 
 
 •269 
 
 175 
 
 197 
 
 21 
 
 1-041 
 
 73 
 
 •426 
 
 125 
 
 •267 
 
 176 
 
 196 
 
 22 
 
 1013 
 
 74 
 
 •421 
 
 126 
 
 •265 
 
 177 
 
 195 
 
 23 
 
 •986 
 
 75 
 
 •416 
 
 127 
 
 264 
 
 178 
 
 194 
 
 24 
 
 •961 
 
 76 
 
 •412 
 
 128 
 
 •262 
 
 179 
 
 193 
 
 25 
 
 •937 
 
 77 
 
 •407 
 
 129 
 
 •260 
 
 180 
 
 192 
 
 26 
 
 •914 
 
 78 
 
 •403 
 
 130 
 
 •258 
 
 181 
 
 191 
 
 27 
 
 •892 
 
 79 
 
 •398 
 
 131 
 
 •256 
 
 182 
 
 190 
 
 28 
 
 •872 
 
 80 
 
 •394 
 
 132 
 
 •255 
 
 183 
 
 189 
 
 29 
 
 •852 
 
 81 
 
 •390 
 
 133 
 
 •253 
 
 184 
 
 188 
 
 30 
 
 •833 
 
 82 
 
 •386 
 
 134 
 
 •251 
 
 185 
 
 187 
 
 31 
 
 •815 
 
 83 
 
 •382 
 
 135 
 
 •250 
 
 186 
 
 186 
 
 32 
 
 •797 
 
 84 
 
 •378 
 
 136 
 
 •248 
 
 187 
 
 185 
 
 33 
 
 •781 
 
 85 
 
 •375 
 
 137 
 
 •246 
 
 180 
 
 184 
 
 34 
 
 •765 
 
 86 
 
 •371 
 
 138 
 
 •245 
 
 189 
 
 183 
 
 35 
 
 •750 
 
 87 
 
 •367 
 
 139 
 
 •243 
 
 190 
 
 182 
 
 36 
 
 •735 
 
 88 
 
 •364 
 
 140 
 
 •241 
 
 191 
 
 181 
 
 37 
 
 •721 
 
 89 
 
 •360 
 
 141 
 
 •240 
 
 192 
 
 181 
 
 38 
 
 •707 
 
 90 
 
 •357 
 
 142 
 
 •238 
 
 193 
 
 180 
 
 39 
 
 •694 
 
 91 
 
 •353 
 
 143 
 
 •237 
 
 194 
 
 179 
 
 40 
 
 •681 
 
 92 
 
 •350 
 
 144 
 
 •235 
 
 195 
 
 178 
 
 41 
 
 •669 
 
 93 
 
 •347 
 
 145 
 
 •234 
 
 196 
 
 177 
 
 42 
 
 •657 
 
 94 
 
 •344 
 
 146 
 
 •232 
 
 197 
 
 176 
 
 43 
 
 •646 
 
 95 
 
 •340 
 
 147 
 
 •231 
 
 198 
 
 176 
 
 44 
 
 •635 
 
 96 
 
 •337 
 
 148 
 
 •230 
 
 199 
 
 175 
 
 45 
 
 •625 
 
 97 
 
 •334 
 
 149 
 
 •228 
 
 200 
 
 174 
 
 46 
 
 •614 
 
 98 
 
 •331 
 
 150 
 
 •227 
 
 201 
 
 173 
 
 47 
 
 •604 
 
 99 
 
 •328 
 
 151 
 
 •225 
 
 202 
 
 173 
 
 48 
 
 •595 
 
 100 
 
 •326 
 
 152 
 
 •224 
 
 203 
 
 172 
 
 49 
 
 •585 
 
 101 
 
 •323 
 
 153 
 
 •223 
 
 204 
 
 171 
 
 50 
 
 •576 
 
 102 
 
 •320 
 
 154 
 
 •221 
 
 205 
 
 170 
 
 51 
 
 •568 
 
 103 
 
 •317 
 
 155 
 
 •220 
 
 206 
 
 169 
 
 52 
 
 •559 
 
 104 
 
 •315 
 
 156 
 
 •219 
 
 207 
 
 169 
 
 53 
 
 •551 
 
 105 
 
 •312 
 
 157 
 
 •218 
 
 208 
 
 168 
 
 54 
 
 •543 
 
 106 
 
 •309 
 
 158 
 
 •216 
 
 209 
 
 167 
 
 55 
 
 •535 
 
 107 
 
 •307 
 
 159 
 
 •215 
 
 210 
 
 166 
 
 56 
 
 •528 
 
 108 
 
 •304 
 
 160 
 
 •214 
 
 211 
 
 166 
 
 57 
 
 •520 
 
 109 
 
 •302 
 
 161 
 
 •213 
 
 212 
 
 •165 
 
 58 
 
 •513 
 
 110 
 
 •300 
 
 162 
 
 •211 
 
 213 
 
 •164 
 
 59 
 
 •506 
 
 111 
 
 •297 
 
 163 
 
 •210 
 
 214 
 
 164 
 
 60 
 
 •500 
 
 112 
 
 •295 
 
 164 
 
 •209 
 
 215 
 
 •163 
 
 61 
 
 •493 
 
 113 
 
 •292 
 
 165 
 
 •208 
 
 216 
 
 •162 
 
 62 
 
 •487 
 
 114 
 
 •290 
 
 166 
 
 •207 
 
 217 
 
 •161 
 
 63 
 
 •480 
 
 115 
 
 •288 
 
 167 
 
 •206 
 
 218 
 
 •161 
 
 64 
 
 •474 
 
 116 
 
 •286 
 
 168 
 
 •204 
 
 219 
 
 •160 
 
 65 
 
 •468 
 
 117 
 
 •284 
 
 169 
 
 •203 
 
 220 
 
 159 
 
 66 
 
 •462 
 
 118 
 
 ■281 
 
 
 
 
 
 H.B.— When oil fired or with forced draught, take an assumed grate arta equal to the total heating 
 
 surface divided by 30. 
 
 61 
 
1 
 
 95 8 MANUAL O* MAR1NK ENGINEERING. 
 
 When the pressure exceeds 180 lbs. per square inch, the accumulation 
 of pressure at the steam test will probably be exceptionally high, unless the 
 area of the branch leading from the valve chest is in excess of the area of 
 the valves, and the area of the main waste steam pipe is correspondingly in 
 excess of the gross area of the valves. 
 
 In ascertaining the fire-grate area, the length of the grate should be 
 measured from the inner edge of the dead plate to the front of the bridge, 
 and the width from side to side of the furnace on the top of the bars at the 
 middle of their length. 
 
 Examination of Safety-valves. — The surveyor, in his examination of the 
 machinery and boilers, is particularly to direct his attention to the safety- 
 valves, and, whenever he considers it necessary, he is to satisfy himself as 
 to the pressure on the boiler by actual trial. 
 
 The surveyor is to examine the whole of the valves, weights, and springs 
 at every survey. 
 
 The responsibility of seeing to the efficiency of the mode by which the 
 valves are fitted, so as to be out of the control of the engineer when steam 
 is up, rests with the surveyor, who should see that the method adopted is 
 efficient and approved of by the Board of Trade. 
 
 The safety-valves should be fitted with lifting gear, so arranged that the 
 two or more valves on any one boiler can at all times be eased together, 
 without interfering with the valves on any other boiler. The lifting gear 
 should, in all cases, be arranged so that it can be worked by hand either 
 from the engine-room or stokehole. 
 
 Care should be taken that the safety-valves have a lift equal to at least 
 one-fourth their diameter; that the openings for the passage of steam to and 
 from the valves, including the waste steam pipe, should each have an area 
 not less than the area of valves required bv Table cxiii. ; that each valve- 
 box has a drain-pipe fitted at its lowest part. In the case of lever valves, 
 if the holes in the lever are not bushed with brass, the pins must be of 
 brass; iron and iron working together must not be passed. The valve seats 
 should be secured by studs and nuts. 
 
 Surveyor to See Valves Weighted. — When the surveyor has determined 
 the amount of pressure, he is to see the valves weighted accordingly, and 
 the weights or springs fixed in such a manner as to preclude the possibility 
 of their shifting or in any way increasing the pressure. The limit of the 
 weight on the valves is to be inserted in the declaration, and should it at 
 any time come to a surveyor's knowledge that the weights have been shifted 
 or the loading of the valves otherwise altered, or that the valves have been 
 in any way interfered with, so as to increase the pressure, without the 
 sanction of the Board of Trade, he is at once to report the facts to the 
 Board of Trade. 
 
 Spring Safety-valves. — If the following conditions are complied with, 
 the surveyor need raise no question as to the substitution of spring loaded 
 valves for dead-weighted valves : — 
 
 (1) That at least two valves are fitted to each boiler. 
 
 (2) That the valves are of the proper size, as required by Table lxxvi. 
 
 (3) That the springs and valves be so cased in that they cannot be 
 
 tampered with. 
 
 (4) That provision be made to prevent the valves flying off in case of 
 
 the springs breaking. 
 
BOARD OF TRADE RULES FOR SAFETY-VALVES. 959 
 
 (5) That the requisite safety-valve area is cased and locked up in the 
 
 usual manner of the Government valves. 
 
 (6) That screw lifting gear be provided to meet all the valves. 
 
 (7) That the size of the steel of which the springs are made is to be 
 
 found by the following formula : — 
 
 y 
 
 s x D , 
 — = a. 
 
 8 — the load on the spring in lbs. 
 D = the diameter of the spring (from centre to centre of wire) in 
 
 inches. 
 d — the diameter, or side of square, of the wire in inches, 
 c =» 8000 for round steel. 
 c = 11,000 for square steeL 
 
 (8) That the springs be protected from the steam and impurities issuing 
 
 from the valves. 
 
 (9) That, when valves are loaded by direct springs, the compressing 
 
 screws abut against metal stops or washers when the loads 
 sanctioned by the surveyor are on the valves. 
 (10) That the springs have a sufficient number of coils to allow a com- 
 pression under the working load of at least one-quarter the 
 diameter of the valve. 
 The size of steel of springs of safety-valves should not, as a rule, be 
 less than \ inch. 
 
 In no case is the surveyor to give a declaration for spring-loaded 
 valves unless he has tried them under full steam and full firing for at 
 least twenty minutes with the feed-water shut off and stop-valve closed, 
 and is fully satisfied with the result of the test. In special cases, or 
 when the valves are of novel design, the results of the test under full 
 steam should be reported to the Board, but, if the surveyor is fully satis- 
 fied with them, he need not delay the granting of the declaration for the 
 vessel pending the approval of the Board. If, however, the accumulation 
 of pressure exceed 10 per cent, of the loaded pressure, he should withhold 
 his declaration, and report the case to the Board of Trade, accompanied 
 by a sketch, if necessary, and stating the strength pressure and the 
 working pressure of the boilers. 
 
 In the case of safety-valves, of which the principle and details have 
 already been passed by the Board of Trade, the surveyor need not require 
 plans to be submitted, so long as the details are unaltered, of which he must 
 fully satisfy himself; but in any new arrangement of valves, or in any case 
 in which any detail of approved valves is altered, he should, before 
 assuming the responsibility of passing them, report particulars, with a 
 drawing to scale, to the Board of Trade. 
 
 The tracings of new safety-valve designs should, if possible, be trans- 
 mitted to the Board of Trade for consideration before the construction of 
 the safety-valves is commenced. 
 
 Owners, Masters, and Engineers to see that Valves are kept in Proper 
 Order. — It is clearly the duty of the masters and engineers of vessels to 
 see, in the intervals between the surveys, that the locked-up safety-valves, 
 as well as the other safety-valves and the rest of the machinery, are in 
 proper working order. There is no provision in the Merchant Shipping 
 
960 MANUAL OF MARINE ENGINEERING. 
 
 Act, 1854, exempting the owner of any vessel, on the ground that she 
 has been surveyed by the Board of Trade surveyors, from any liability, 
 civil or criminal, to which he would otherwise be subject. The Act of 
 Parliament requires the Government safety - valve to be out of the 
 control of the engineer when the steam is up ; this enactment, far from 
 implying that he is not to have access to it, and to see to its working, at 
 proper intervals when the vessel is in port, rather implies the contrary; 
 and the master should take care that the engineer has access to them 
 for that purpose. Substantial locks that cannot be easily tampered with, 
 and as far as possible weather-proof, should be used for locking up the 
 safety-valve boxes. 
 
 All Tests for Pressure and Accumulation are to be made with Board of 
 Trade Gauges. — In witnessing the hydraulic tests of boilers, &c, and in 
 witnessing all safety-valve tests for accumulation of pressure, the surveyors 
 are to use the pressure-gauges supplied by the Board of Trade for the 
 purpose. No steam gauge should be used without having a syphon filled 
 with water between it and the boiler. 
 
BUREAU VERITAS RULES FOR BOILER SAFETY-VALVES, ETC. 961 
 
 APPENDIX P. 
 
 Bureau Veritas Rules for Boiler Safety-Valves and 
 Metal Tests, etc. 
 
 Area of safety-valves = - 4 - , = square inches. 
 
 P is the working pressure in pounds per square inch. 
 No valve should be less than 1| inches diameter. 
 
 Thickness of shell plate is deduced from the following : — 
 Working pressure = (t — O04). 
 
 PD 
 
 Therefore, t = —^= + 0-04 inch. 
 
 D is the greatest inside diameter in inches. 
 
 t is the thickness in inches. 
 
 04 the allowance for corrosion. 
 
 R is the tensile stress in lbs. per square inch (which is the ultimate 4- 4) : 
 
 and must be stated by the maker. 
 a is the ratio of the plate between the holes to the full plate, or pitch — 
 
 diameter of rivet -f- pitch. 
 
 Tests of Materials. 
 
 Steel Plates. — Tensile in tons plus percentage of elongation in 8 inches not 
 less than 50. 
 
 Iron Plates. — Tensile with 23 and 18 per cent, in 8 inches ; across, 21£ 
 and 12 per cent, in 8 inches. 
 
 Bend cold, if 32nds, 120° with 100° across. 
 1? 32nds 90° 70° 
 
 Nickel Steel, 3 25 NiO-35 C— f| ton and 20 per cent, in 4. 
 
 Copper Plates. — Tensile, 12| and 22 per cent, in 8 inches. Bend double 
 at dull red heat. 
 
 Copper Tubes. — Tensile, 13 and 30 per cent, in 4 inches. 
 
 Brass Plates. — Tensile, 19 with x 12 per cent. 
 
 Steel Pin Rivets. — 1| ton over §§, 25 per cent, under £f nds, 20 per cent. 
 in 8 inches. 
 
 Brass Tubes. — Internal pressure, 430 lbs. per square inch. Condenser 
 tubes, 570 lbs. per square inch. 
 
 Brass Castings. — 11 J tons, 8 per cent, in 4 inches. 
 
 Iron Castings. — Minimum pressure, If square, 6J in supports, series 
 shocks. Pressure of 27 lbs. weight falling 12 inches up to 18 inches by 
 2 inches increase. 
 
 Common castings, 1\ tons, and fall 10 to 16 inches. 
 
962 MANUAL OF MARINE ENGINEERING. 
 
 APPENDIX Q. 
 
 Admiralty Rules for Testing Materials for Machinery. 
 (e) Steel Castings for Machinery. 
 
 steel Steel castings for the machinery are to satisfy the following 
 Twto 1 8S ' conditions : — Tensile strength, not less than 28 tons per square 
 inch, with an extension of 2 inches of length of at least 23 per 
 cent. Bars of the same metal, 1 inch square, should be capable of 
 bending cold, without fracture, over a radius not greater than 
 If inches, through an angle depending on the ultimate tensile 
 strength, this angle to be not less than 90° at 28 tons ultimate 
 strength, and not less than 60° at 35 tons ultimate strength, and 
 in proportion for strengths between these limits. For intricate 
 thin castings the extension in 2 inches of length is to be at hast 
 10 per cent., and the bending angle is to be not less than 20° at 
 28 tons ultimate strength, and 15° at 35 tons ultimate strength, 
 and in proportion for strengths between these limits. Test pieces 
 are to be taken from each casting. All steel castings are also to 
 satisfactorily stand a falling test, the articles being dropped from 
 a height of 12 feet (or as may be approved) on a hard macadamised 
 road or a floor of equivalent hardness. 
 
 It is to be distinctly understood that contractions or defects in 
 steel castings are not to be made good by patching, burning, or by 
 electric welding, without the sanction of the Admiralty overseers. 
 
 (f) Steel Forgings jor Machinery. 
 
 Bteel All steel forgings are to satisfy the following conditions : — 
 
 forgings. Ultimate tensile strength, not less than 28 tons per square inch, 
 with an extension in 2 inches of length of at least 30 per cent. 
 Bars of the same metal, 1 inch square, should be capable of being 
 bent cold, without fracture, through an angle of 180° over a radius 
 not greater than \ inch. Test pieces are to be taken from each 
 forging. Crank and propeller shafts are to have test pieces taken 
 from each end, and the ultimate tensile strength of the material of 
 these shafts must not exceed 32 tons per square inch. 
 
 For all important forgings, such as crank and propeller shafts, 
 connecting- and piston-rods, the forgings are to be gradually and 
 uniformly forged from solid ingots, from which at least 30 per cent, 
 of the top end of the ingot has been removed before forging, and 
 at least 3 per cent, of the total weight of the ingot from the bottom 
 end after forging. The sectional area of the body of the finished 
 forging is to be not more than \ the original sectional area of 
 the ingot. 
 
ADMIRALTY RULES FOR TESTING MATERIALS. 963 
 
 (g) Cast Iron. 
 
 Test pieces to be taken from such castings as may be considered Cast > r0 °- 
 necessary by the inspecting officer. The minimum tensile strength 
 to be 9 tons per square inch, taken on a length of not les3 than 
 2 inches. The transverse breaking load for a bar 1 inch square, 
 loaded at the middle between supports 1 foot apart, is not to be 
 less than 2000 lbs. 
 
 (h) Gun-metal, Naval Brass, end White Metal. 
 
 The gun-metal used for all castings throughout the whole of composi- 
 the work supplied by the contractors is, unless otherwise specified, ti0I V . 
 to contain not less than 8 per cent, of tin, and not more than and test's oi 
 5 per cent, of zinc, the remainder to be of approved quality copper, fnjj'lllfvai 1 
 The exact proportion of tin above 8 per cent, being arranged as brass, &c. 
 may be required, depending on the use for which the gun-iuetal 
 is intended. The ultimate tensile strength of gun-metal is to be 
 not less than 14 tons per square inch, with an extension in 
 2 inches of length of at least 1\ per cent. The composition of any 
 naval brass used is to be : — Copper, 62 per cent. ; zinc, 37 per 
 cent. ; and tin, 1 per cent. All naval brass bars are to be cleaned 
 and straightened. They are to be capable of (1) being hammered 
 hot to a fine point, (2) being bent cold through an angle of 75° over 
 a radius equal to the diameter or thickness of the bars. The 
 ultimate tensile strength of naval brass bars \ inch diameter and 
 under is not to be less than 26 tons per square inch, and for round 
 bars above \ inch diameter, and square bars not less than 22 tons 
 per square inch, whether turned down in the middle or not. The 
 extension in 2 or 4 inches of length is to be at least 10 per cent. 
 Breaks within less than \ inch of the grip are not to count. 
 Cuttings from the propellers and other important gun-metal cast- 
 ings and naval brass work will be sent to Portsmouth dockyard for 
 analysis. The white metal used for bearing surfaces is to contain 
 at least 85 per cent, of tin, not less than 8 per cent, of antimony, 
 and about 5 per cent, of copper ; zinc or lead should not be used. 
 The brasses are to be carefully tinned before filling with white 
 metal. 
 
 The gun-metal castings are to be perfectly sound, clean, and castings 
 free from blowholes. The steel castings are required to be clean, »«d . 
 sound, and to be out of twist, and as free as possible from blow- orging8 
 holes ; the steel forgings are required to be quite sound, clean, and 
 free from all flaws. All castings and forgings must admit of being 
 machined, planed, and bored to the required dimensions ; and no 
 piecing, patching, bushing, stopping, or lining, will be permitted, 
 nor will any manufactures in which these conditions have been 
 infringed be accepted. In cases of doubt as to the suitability of 
 castings or other materials, for the purpose intended, early refer- 
 ence should be made to the inspecting officer to avoid subsequent 
 delay by the rejection of such parts after delivery. The whole of 
 the steel plates, angles, and rivets used in the construction of any 
 part of the work supplied is to be manufactured by the Siemens- 
 
96 4 MANUAL OF MARINE ENGINEERING. 
 
 Martin process. All castings are to be of steel or gun-metal, 
 except where otherwise specified ; and all forgings, plates, bolts, 
 &c, are to be of steel. No steel is to be toughened without 
 sanction from the Admiralty. 
 
 Copper for Pipes. 
 
 Copper Strips cut from the steam and other pipes, either longitudinally 
 
 or transversely, are to have an ultimate tensile strength of not less 
 than 13 tons per square inch when annealed in water, with an 
 elongation in length of 2 inches or 4 inches of not less than 35 and 
 30 per cent, respectively. Such strips are also to stand bending 
 through 180° cold until the two sides meet, and of hammering to a 
 fine edge without cracking. 
 
 Treatment of Mild Steel. 
 
 Hammer All plates and furnaces for boilers and steam pipes are to be 
 
 removed, treated as follows, with a view of removing the black oxide or scale 
 formed during manufacture :-^The plates and furnaces, previous to 
 their being taken in hand for working, are to stand for not less 
 than 8 hours in a liquid consisting of 19 parts of water and 1 of 
 hydrochloric acid. The plates should be placed in the bath on 
 edge, and not laid flat. When the plates and furnaces are removed 
 from the dilute acid both the surfaces are to be well brushed and 
 washed to remove any scale which may still adhere to them. They 
 ghould then be placed in another similar bath filled and kept well 
 supplied with fresh water, or be thoroughly washed with a hose, as 
 may be found necessary. The plates on removal from the fresh 
 water should be placed on edge to dry. This treatment is to be 
 carried out on the premises where the boilers and pipes are made. 
 Cold bend- All plates or bars which can be bent cold are to be so treated ; 
 ,ng - and if the whole length cannot be bent cold, heating is to be had 
 
 recourse to over as little length as possible. 
 Hydraulic The front and end plates of the boilers, and also all other plates, 
 flanging, including those for the combustion chambers, are to be flanged by 
 
 hydraulic pressure, and in as few heats as possible. 
 Dangerous In cases where plates or bars have to be heated, the greatest 
 ture P tobe care snou ld be taken to prevent any work being done upon the 
 avoided, material after it has fallen to the dangerous limit of temperature 
 known as a " blue heat " — say from 600° to 400° Fahrenheit. 
 Should this limit be reached during working, the plates or bars 
 should be reheated. 
 Annealing. Plates or bars which have been worked while hot are to be 
 subsequently annealed simultaneously over the whole of each plate 
 or bar. 
 Failure of ^ n ca8es where any bar or plate shows signs of failure or 
 metal. fracture in working it is to be rejected. Any doubtful cases 
 are to be referred to the Admiralty. 
 
RECOMMENDATIONS FOR MAIN CYLINDRICAL BOILERS. 965 
 
 APPENDIX K. 
 
 RECOMMENDATIONS OF THE BRITISH MARINE ENGINEERING DESIGN 
 AND CONSTRUCTION COMMITTEE FOR MAIN CYLINDRICAL 
 BOILERS SUMMARISED. 
 
 The Steel used must be made in an open-hearth furnace, either by the acid 
 or basic process. 
 
 Ordinary shell plates and longitudinal stays shall have a tensile strength 
 from 28 to 32 tons per square inch. 
 
 Special steel plates and stays may be as high as 35 tons tensile. If higher 
 tensile than 35 tons is desired, special arrangements must be made with the 
 Registering Body. 
 
 Steel plates for welding and flanging, and the bars for screw stays and for 
 rivets, 26 to 30 tons tensile. 
 
 Steel strips for welded tubes must not have a higher tensile than 28 tons. 
 
 Wrought-iron bars may be used for screwed stays and allowed the same 
 working stress as for steel, if it is tested and stands 21"5 tons tensile with an 
 extension of 25 per cent, in 8 inches. 
 
 The methods of testing and expected results are to be as prescribed by the 
 British Engineering Standards Committee. 
 
 The hydraulic test to be applied to new boilers shall be twice the working 
 pressure when that does not exceed 100 lbs. per square inch ; when it does, 
 then the test pressure is to be 50 lbs. in excess of one and a half times 
 the W.P. 
 
 For boilers which have been on service, the test must not be more than 
 1-5 the W.P. 
 
 Cylindrical shells.— The W.P. shall equal ( * ~ 2 J * ^ X J - 
 
 \j X iJ 
 
 Where t is the thickness of the shell plate in 32nds of an inch. 
 
 S is the minimum tensile strength of the plates in tons per square inch. 
 J is the percentage of strength of joint as found in the usual way 
 
 {v. p. 906). 
 C is a coefficient which is 2*67 when the longitudinal seams are with 
 
 double butt straps, and 2 - 87 when they are lapped. 
 D is the inside diameter of the outer drum of plating. 
 
 The maximum pitch of rivets in inches = (K x -f 1'625, where K is as 
 given on p. 911, but with no restriction to 10 inches. 
 For furnaces the following are the rules : — 
 
 r ' M . wp 1,450 (*-l) 2 
 
 (i.) Pto. w.p. = (L+24)xD . 
 
 (L+24) xD t™ , , , , 
 
 >The minimum to be taken. 
 
 W.P. = ~ { 
 
966 MANUAL OF MARINE ENGINEERING. 
 
 Where D is the external diameter in inches. 
 t is the thickness in 32nds of an inch. 
 
 L is the length in inches between the lines of substantial support 
 from other parts. 
 
 (ii.) Corrugated. W.P. = C (t — 1), -*- D. 
 
 Where D is the least external diameter as taken at the bottom of corrugations. 
 t is the thickness in 32nds of an inch. 
 
 C is 480 for the Fox, Morison, Purves & Deighton, and similar 
 furnaces ; 510 for Leeds Forge suspension furnace. 
 
 No furnace should be more than §4 inch thick. 
 Flat plates supported by screwed stays. 
 
 (a) W.P. = ((-l) 2 xCv a 2 + b 2 . 
 
 Where t is the thickness in 32nds of an inch: 
 
 a is the distance apart of the rows of stays in inches. 
 
 b is the pitch of the stays in the rows in inches. 
 
 C = 50 when the stays are riveted over only. 
 
 C = 55 for screwed stay tubes when expanded in three holes. 
 
 C = 75 when screwed stays are fitted with nuts on the outer ends and 
 
 exposed to flame, and 88 when not exposed. 
 C = 100 when the stay simply passes through the plate and fitted 
 
 with nuts inside and outside the plate. 
 
 When the stays pass through the plates and have nuts inside and outside 
 with substantial washers under the outer nuts, 3'5 times the stay in diameter, 
 and the thickness t w is not less than 0"2 the diameter of the stay, then 
 
 (b) W.P. = 100 {(t-l) 2 + 0-15 £} -a 2 +6 2 , 
 
 when the washers are in diameter at least 067 the pitch of the stays, and in 
 thickness t w is not less than 0'6 the thickness of the plate and riveted to it. 
 
 (c) W.P. = 100 {(/ - l) 2 + 0-354} + « 2 + o 2 , 
 
 when the plate is stiffened by strips riveted to it whose breadth is at least 
 0'67 the pitch, and the thickness t w is not less than 0*6 the thickness of the 
 plate to which it is efficiently riveted. 
 
 (d) W.P. = 100 {(t-l) 2 -\- 0-55 4} + a 2 + b\ 
 
 when there is a doubling plate at least 0'6 its thickness efficiently riveted 
 to it. Then 
 
 (e) W.P. = 100 {(t - l) 2 + 0-85 %,} -*■ a 2 + I 2 . 
 
 For front tube plates having spaces between the nests of tubes, and 
 between them and the boiler shell, and between them and the furnaces and 
 longitudinal stays. t w is the thickness when a doubling plate is riveted to 
 the tube plate. 
 
 (a) W.P. = 55 {{t - l) 2 + 0-55 *„} 4- a 2 + b 2 , 
 
 when all the stay tubes at margins are screwed into plates, but without nuts. 
 
RECOMMENDATIONS FOR MAIN CYLINDRICAL BOILERS. 967 
 
 (6) W.P. = 72 {(t - l) 2 -f 0-55 £} + a 2 + /, 2 , when all have nuts. 
 
 (c) When only alternate stay tubes have nuts, the factor is 63 instead of 72 
 For back tube plates, where there must be no nuts, 
 
 (d) W.P.= 38(f-l) 2 -=- 7> 2 , 
 
 where p is the mean pitch got by adding together the four sides of the 
 quadrilateral from tube stay centre to tube stay centre and dividing by 4. 
 
 (e) For the W.P. for the front tube plate for stays other than the margin 
 ones the same rule applies, and if the stay tubes have nuts the factor is 49. 
 
 Screw stays to combustion chambers must have nine threads per inch, and 
 whether of steel or tested wrought iron 
 
 W.P. = {d - 0267) 2 x 8,250 -*- a X b, 
 
 where d is the diameter over thread, a and b are the vertical and horizontal 
 pitches in inches. 
 
 Longitudinal stays must have six threads per inch. 
 
 a X o 28 
 
 where d is the diameter over thread and S the minimum tensile strength of 
 bars in tons. 
 
 Stay tubes, whether of lapwelded wrought iron or steel, shall be allowed a 
 working stress of 7,500 lbs. per square inch at bottom of threads. 
 
 Dished ends shall have a radius of curvature not exceeding the diameter of 
 the shell if without a stay. The inside radius of curvature at flange shall be 
 not less than four times the thickness. 
 
 W.P. = 15xS(J-l)-=-R, 
 
 where t is the thickness in 32nds of an inch, R the radius of curvature, and S 
 the minimum tensile strength of steel in tons. 
 
 Mountings and Fittings. — Water-gauge glasses may be | or § inch in outside 
 diameter, and 20, 18, 16, or 14 inches long. Stand pillars for water gauges 
 2 - 5 inches diameter inside. Single-ended boilers over 15'5 feet diameter must 
 have two, and double-ended over 16 feet three such gauges. Two independent 
 means of feeding and origin of feed. 
 
 Safety valves to be such that there is not on any one valve a greater load 
 than 2,600 lbs. The aggregate area on any one boiler shall be found thus — 
 
 Rule for area in sq. inches = total heating surface in sq. feet xf — ), 
 
 f being the working or gauge pressure in lbs. per square inch. 
 
 The valves may be loaded to 2| per cent., but not to more than 5 lbs. in 
 excess of the working pressure. 
 
 The waste steam pipe and passages from the valves must have a sectional 
 area equal to O'Ol square inch for each square foot of total heating surface. 
 
969 
 
 INDEX. 
 
 Acttng surface of screws, 485. 
 Action of steam in turbines, 109, 110. 
 Actual mean pressure in practice, 123. 
 Adiabatic expansion, 112. 
 Admiralty brass, mixture, 963. 
 
 bronze, Properties of, 820. 
 
 nominal horse-power, 196. 
 
 old rules for estimating Ind. H.P. for 
 
 speed, 33. 
 
 rules for boilers and tests, 931. 
 
 for testing machinery materials, 
 
 962. 
 
 for tests of machinery, 849. 
 
 steel, 809. 
 
 tests of copper pipes, 964. 
 
 of materials, 931, 962. 
 
 trials of machinery under steam, 855. 
 
 white metal, 963. 
 
 Advantages of compound cylinders, 192. 
 Air and water vapour, 356, 357. 
 
 compressors, Navy type, 525, 527. 
 
 in condensers, Effect of, 355. 
 
 in sea water, 373. 
 
 leaks to condenser, 355. 
 
 pump buckets, Scantlings of, 391. 
 
 clearance, 381. 
 
 foot valves, 381. 
 
 ■ indicator diagrams, 378. 
 
 kinetic system, 386, 387. 
 
 piston packings, 391. 
 
 Size of, 383, 384, 385. 
 
 suction pipes, 395. 
 
 pumps, Capacity of, 375. 
 
 Edwards', 376. 
 
 Efficiency of, 381. 
 
 Function of, 373. 
 
 Horizontal, 375. 
 
 operated by independent means, 
 
 385. 
 
 Vertical, 376. 
 
 Weir's, 377. 
 
 — Wet and dry, 378. 
 
 valve on boilers, 687. 
 
 Alloys, Composition of, 823, 824. 
 of copper differentiated, 819. 
 
 Alternating stresses, 725. 
 
 Effect of, 289. 
 
 on shafts, 288. 
 
 Aluminium, Properties of, 818. 
 
 Use of, to marine engineer, 226. 
 
 American lake paddle steamer, 8. 
 
 walking beam engines, 65. 
 
 Amsler's torsion meter, 145. 
 
 Angle of entrance, suitable for speed, 33. 
 
 Antimony, Use of, 819. 
 
 Area through boiler tubes, 581. 
 
 " Ariel," Trials of s.s., 189. 
 
 Arrangement of cylinders of marine screw 
 
 engines, 78. 
 Ash ejectors, 704, 705. 
 Assistant cylinder, Joy's patent, 324. 
 
 cylinders for slide valves, Indicator 
 
 diagram of, 324. 
 Atlantic steamer service, 732-740. 
 Augmentor of vacuum, 389. 
 Automatic drains, 710. 
 
 feed, arrangement of pumps, 520. 
 
 Auxiliary condensers, 535, 536. 
 
 exhaust, 514. 
 
 — — machinery, 511, 729. 
 
 Effect of, on vacuum, 355. 
 
 generating electricity, division 
 
 of units, 516. 
 
 in older ships, 210. 
 
 Steam used by, 512. 
 
 - Vibration of, 225. 
 worked by electricity, 516. 
 
 valves on cylinders, 282. 
 
 B 
 
 Back pressure due to lead, 124. 
 in L.P. cylinder, 233. 
 
 Balance of engine, Necessity for, 225. 
 
 piston for slide valves, 322. 
 
 - weights, Design of, 799, 800. 
 Materials of, 799. 
 
 bob- 
 
 Balancing a single-crank engine with 
 
 weights," 769. 
 an engine of any number of cranks, 
 
 770. 
 
970 
 
 INDEX. 
 
 Balancing engines, 222. 
 
 Modified principles of, 791. 
 
 ■ weights of moving parts, 779. 
 
 of an ordinary 4-crank engine, 785. 
 
 of a 4-crank engine ( Yarrow-Schlick- 
 
 Tweedy), 785. 
 
 • of inertia effects, 755. 
 
 of moving parts of an engine, 750, 752. 
 
 Bar links, Size of, 432. 
 
 Battle cruisers, Modern, 733. 
 
 Battleship, Modern, 733. 
 
 Beam engines, 65. 
 
 Beardmore's oil engines, 872. 
 
 Bearings of shafting, Limits of pitch of, 288. 
 
 Bed plates and foundations, Arrangement 
 
 of, 334. 
 Belliss and Morcom auxiliaries, 523. 
 
 — ■ experiments for efficiency, 
 
 157, 158. 
 
 oil engine, 59. 
 
 Bending moments on shafts, 293. 
 Bevis-Gibson torsion meter, 148. 
 Bilge pipes, 506. 
 
 pumps, 408. 
 
 valves and boxes, 506. 
 
 Blade tips, Path of, 460. 
 Blades of screws, 485, 491. 
 
 Breadth of, 492. 
 
 Numbers of, 492. 
 
 Blow-off cocks and valves, 685. 
 Board of Trade, factors of safety, 907. 
 influence of design, 221. 
 requirements for boiler 
 
 shells, 638. 
 rules and regulations for 
 safety valves, 956, 960. 
 - for boiler shell, 906. 
 for boiler shell joints, 
 908. 
 
 for flat plates, 905. 
 
 for girders, 904. 
 
 for pipes, 508. 
 
 ■ steel, 809. 
 
 Boiler clothing or lagging, 689, 690. 
 
 efficiency, 555. 
 
 — ends, method of working, 652. 
 
 making, Continental practice different 
 
 from British, 665, 666. 
 
 pressures and mean pressures in 
 
 cylinders, 721. 
 — in early days, 10. 
 
 seatings and bearers, 693, 695. 
 
 shell, Capacity of, 581. 
 
 - Cylindrical, 637. 
 
 - joints, Board of Trade rules for 
 
 908. 
 
 shells, Board of Trade rules for, 906, 
 
 907. 
 — — - British Corporation rules for, 
 917. 
 Bureau Veritas rules for, 922. 
 Lloyd's rules for, 912. 
 
 German rules for, 936. 
 
 Boiler stays, German rules for, 942. 
 
 tubes, Area through, 581. 
 
 work, Supervision of, by Admiraltv. 
 
 931. ' 
 
 Boilers, Admiralty rules for, 931. 
 
 Babcock & Wilcox, 595. 
 
 Belleville, 613. 
 
 Blechynden, 607. 
 
 Combination, 634. 
 
 Cylindrical, 558-561. 
 
 Details of, 584-585. 
 
 Double-ended, 562, 564, 567. 
 
 Du Temple, 609. 
 
 Efficiency of, 581. 
 
 Ferguson & Fleming, 601. 
 
 General dimensions of, 574. 
 
 German rules for, 936. 
 
 gunboat type, 565, 571. 
 
 Herreshoff, 592. 
 
 Hohenstein, 621. 
 
 Holt's design, 569. 
 
 Leading features of, 538. 
 
 Locomotive, 572, 573. 
 
 Manholes of, 666-668. 
 
 Modern examples of, 667. 
 
 Mumford, 609. 
 
 Myabara, 622. 
 
 Niclausse, 617. 
 
 Normand, 600. 
 
 Oval, 558, 569. 
 
 Reed, 611. 
 
 Seaton, 603. 
 
 small size, 567. 
 
 Stirling, 620. 
 
 Stresses on, 728. 
 
 Thornycroft, 609. 
 
 Thornycroft-Marshall, 623. 
 
 — - Various types of, 557. 
 
 Vertical cylindrical, 571. 
 
 Ward's coil, 593. 
 
 Water space of, 666. 
 
 , Weight of, 668. 
 
 White-Foster, 606. 
 
 with dry combustion chambers, 569. 
 
 Yarrow, 604. 
 
 Bolts of column flanges, 254. 
 
 of connecting-rods, 282. 
 
 of cylinder covers, 255. 
 
 of piston-rods, 280. 
 
 of screw propeller blades, 497. 
 
 Stress on, 289. 
 
 Boring holes, 252. 
 
 the sternpost, 691. 
 
 Boss of screw propeller, 490. 
 
 Boyle and Marriott's law of expansion, 1 10, 
 
 in. 
 
 Brake horse-power, 203. 
 
 " Brandon," s.s., first ship with compound 
 
 engines, 10. 
 Brass, Mixtures of, 819. 
 
 tube metal, Properties of, 819. 
 
 Brasses of connecting-rods, 282. 
 of main bearings, 335, 338. 
 
INDEX. 
 
 971 
 
 British Corporation rules for boiler shells, 
 
 917. 
 
 for flat surfaces, 918. 
 
 for general construction and 
 
 fitting of boilers, 920, 
 
 921. 
 
 for girders, 920. 
 
 for stays, 919. 
 
 Bronze, Admiralty, 82, 963. 
 
 Effect of heat on, 833, 835. 
 
 Special, for high temperatures, 821. 
 
 for screw, 821. 
 
 Bronzes for high temperatures, 833. 
 
 Various, 822. 
 
 Buckley's springs, 265. 
 
 Built crank-shafts, Sizes of, 307. 
 
 Bull's metal, 822. 
 
 Bulkhead stop valve, 676. 
 
 Bureau Veritas, general rules for boilers and 
 
 fittings, 929. 
 rules and regulations for safety 
 
 valves, 930, 961. 
 
 for boiler shells, 922. 
 
 for flat plates, 926. 
 
 for furnaces, 928. 
 
 for oil and other internal 
 
 combustion engines, 874, 
 
 875. 
 
 for pipes, 510. 
 
 . for stays, 927. 
 
 tests of materials, 933, 961. 
 
 Burners for oil fuel, 540-543, 674. 
 Butt joints, double-riveting, 647. 
 
 quadruple-riveting, 647, 649. 
 
 treble-riveting, 647, 648. 
 
 various riveting, 641, 644. 
 
 Cameron's patent springs, 265. 
 
 Capacity of air pumps, Calculations of, 374. 
 
 of circulating pumps, 39o. 
 
 of funnels, 583. 
 
 Cargo ships, effects of differing drafts, 230. 
 
 steam, Trials of, 232, 744a. 
 
 Cast iron, Effects of temperature on, 833. 
 
 Kinds of, 802. 
 
 Modern, 215. 
 
 Scotch pig, 803. 
 
 Causes of mechanical loss, 163. 
 
 of superiority of compound systems, 
 
 192. 
 Cavitation causing vibration, 225. 
 Centrifugal force on shafts, 281. 
 
 pumps, 399, 400,, 410. 
 
 of s.3. " Mauretania," 402. 
 
 Channel service steamers, Steam pressure 
 
 in, 177. 
 '" Charlotte Dundas " and her machinery, 6. 
 Check feed valves on boilers, 681, 682. 
 Chimney draught, 539. 
 Chrome steel, 814. 
 
 Circumferential seams, 650, 651. 
 Circulating pump rods, 397. 
 
 pumps, 396. 
 
 Power required to drive, 409. 
 
 ratio of H.P. cylinder, 397. 
 
 Size of, 370. 
 
 - suction and delivery pipes, 398. 
 
 of 
 
 water, Velocity of, 370. 
 
 " City of Detroit," p.s., 7,600 H.P., 8. 
 
 Clearance of pistons, 256. 
 
 (stroke), 113. 
 
 volume, 113. 
 
 Clyde turbine steamer, " Duchess 
 Argyll," 9. 
 
 Coal, consumption per I.H.P., 577. 
 
 Varieties of, 539, 544. 
 
 Coefficients for computing power, 26, 27, 29. 
 
 of fineness of ships, 28. 
 
 of friction in practice, 843, 844. 
 
 or factors for Admiralty speed for- 
 mulae, 34. 
 
 Collie's torsion meter, 149. 
 
 " Columbia " (tj.s.s.), 16. 
 
 Columns, 334, 340. 
 
 Sizes of, 342. 
 
 Combination of turbines with reciprocators, 
 93, 94. 
 
 indicator diagrams — Quadruple sys- 
 tem, 193. 
 
 turbines, 136. 
 
 Combustion, Air required for, 547. 
 
 chambers, 659, 660, 663. 
 
 of fuels, how effected, 553. 
 
 " Comet," p.s., and her machinery, 7. 
 
 Comparative theoretical efficiency of marine 
 engines, 181, 184. 
 
 Comparison of trials of turbines and re- 
 ciprocators, 92. 
 
 Competitive trials of compound engines, 
 183, 185. 
 
 with triple-expansion 
 
 engines, 188, 189. 
 
 Compound engine, Direct-expansion, 179. 
 
 Inventors and history of, 10. 
 
 systems of cylinders, 190, 192. 
 
 turbines, 138, 139. 
 
 ■ with receiver, 177. 
 
 with two L.P. cylinders and one 
 
 H.P., 184-186. 
 
 Compression of steam, 114. 
 
 Condenser efficiency, 350. 
 
 for winches, etc., 536. 
 
 Jet, 344. 
 
 tests (Admiralty), 372. 
 
 tube, ferrules, 365. 
 
 — ■ packings, 364. 
 
 tubes, Admiralty, 349. 
 
 Faults of, 363. 
 
 Kinds of, 348. 
 
 Protection of, 349, 363. 
 
 Spacing of, 365. 
 
 Tests of, 363. 
 
 ^onfbmsers, 344. 
 
972 
 
 INDEX. 
 
 Condensers, Auxiliary, 535, 536. 
 
 Method of testing, 844. 
 
 Shape of, 352, 353, 367. 
 
 Surface, 347. 
 
 Conditions imposed on the marine engine 
 
 designer, 56. 
 Connecting-rod bolts, 282. 
 
 brasses, 283 
 
 caps, 285. 
 
 Diameter of, 280, 281. 
 
 jaws, 285. 
 
 Connecting-rods, various kinds, 279-284. 
 Considerations influencing efficiency of 
 
 machinery, 152, 153. 
 Contraflo condensers, 352-354. 
 Construction of boilers, 634. 
 Consumption of fuel, 576. 
 
 per I. HP., 19. 
 
 per I.H.P. in various ships, 
 
 730. 
 
 of steam, 512, 513. 
 
 Controlling feed supply automatically, 632, 
 
 683. 
 Cooling surface, Allowances of, 359. 
 
 Basis of, 360, 361. 
 
 how determined, 353, 354. 
 
 per I.H.P., 359. 
 
 Rankine's formula for, 360. 
 
 water, Flow of, 361, 362. 
 
 Passage of, 370. 
 
 Quantity of, 369. 
 
 Ratio of, to steam, 370. 
 
 Copper, 815, 816. 
 
 Analysis and strength of, 816. 
 
 and steel piping, Thickness of, 696. 
 
 Effect of temperature on, 833. 
 
 High price of, 214. 
 
 pipes, 695, 696. 
 
 Admiralty tests, 964. 
 
 Thickness of, 509. 
 
 Corrugated and ribbed furnaces, Rules for, 
 
 657. 
 
 furnaces, 656. 
 
 Cost of machinery affected by weight, 728. 
 Coupling bolts, 308. 
 Couplings with cross keys, 310. 
 Coutt's and Adamson's governor, 708. 
 Covers of steam cylinders, 255. 
 
 of valve boxes, 256. 
 
 Crank arms, Dimensions of, 300. 
 
 efforts, Diagram of, 297. 
 
 of paddle-wheel engines, 299. 
 
 of screw engines, Stresses on, 304, 
 
 305, 306. 
 
 overhung, 299. 
 
 pin, Diameter of, 299. 
 
 ■ pins, Surface of, 309. 
 
 - shaft arm, Stresses on, 289. 
 bearings, 336. 
 
 shafts, Built-up, 307. 
 Details of, 331. 
 for naval ships, 
 Hollow, 308. 
 
 Westgarth-Carel 
 
 oil 
 
 engine, 
 
 308. 
 
 Crank shafts of screw engines, 303. 
 
 screw engines, Rules for, 332. 
 
 Criterion of quality of metals, 845. 
 
 Crotarite or bronze for stays, 821. 
 
 Crude oil engines, 85, 859. 
 
 Cruisers, 733. 
 
 Cylindrical condensers, 353. 
 
 covers, 255. 
 
 Bolts of, 255. 
 
 drain locks, 253. 
 
 - — flanges, 256. 
 
 liners, 248. 
 
 ratio, 241. 
 
 relief valves, 253. 
 
 ■ Scantlings of, 257, 258. 
 
 section, 
 
 870. 
 
 Cylinders and liners, Iron for, 804. 
 
 Arrangement of, 235. 
 
 comp. system, 190, 192. 
 
 diameter of, To ascertain, 231. 
 
 Horizontal, 254. 
 
 in war ships, 229. 
 
 Large versus small, 229. 
 
 Number of, 235. 
 
 Object of testing, 848. 
 
 of naval engines, 244. 
 
 Oscillating, 254. 
 
 quadruple-expansion engines, Arrange- 
 ments of, 237, 239, 240. 
 
 reasons for small ones, 229. 
 
 Requirements in, 228. 
 
 Size of, 228. 
 
 triple -expansion engines, 
 
 ments of, 236, 237, 238. 
 
 Curtis turbine, 141. 
 
 turbines in U.S. cruiser, " Salem," 93. 
 
 Curves of revolution, slip, etc., Exami- 
 nation of, 38, 39. 
 
 of twisting moments, 295. 
 
 Cushioning, Effect of, in cylinder, 114, 115. 
 
 Cut-off with slide valve, To find, 446. 
 
 Defects of indicator readings, 200. 
 Deighton's furnace, 656. 
 Delta metal, 822. 
 
 Denny's estimate of efficiency, 153, 156. 
 Denny, Johnson, torsion meter, 150. 
 Deposits in boilers, Composition of, 832. 
 Design influenced by Board of Trade, 221. 
 by Admiralty, etc., 222, 722. 
 
 Arrange - 
 
 of marine oil engines, 101. 
 
 Destroyers, 720, 727. 
 
 Determining nominal working stress, 727. 
 
 Development of existing single-screw ships, 
 
 97, 100. 
 Dewrance's bronze for high temperatures, 
 
 833. 
 Diagonal compound paddle engine, 61, 62, 
 
 63 
 
INDEX. 
 
 973 
 
 Diagonal triple-compound paddle engine, 
 
 64. 
 Diagrams of a quadruple-expansion engine, 
 
 193. 
 Diameter of L.P. cylinder, how found, 233. 
 " Diana," h.m.s., Trials of, 628. 
 Diesel engine, 87, 101, 859. 
 
 applied to marine propulsion, 5, 
 
 859. 
 
 Efficiency of, 88, 859. 
 
 fuel consumption, 88, 105. 
 
 type of oil engines, 859. 
 
 Direct-acting engine on American paddle 
 
 steamers, 65. 
 Direct-expansion compound engine, 179. 
 Directing boxes, 409. 
 Discharge pipes for cooling water, 371. 
 Division of work in compound cylinders, 
 
 179. 
 Double-acting Diesel engine, 103. 
 
 oil engine, 87. 
 
 reversing oil engine, 104. 
 
 " Draco," Trials of s.s., 188, 189. 
 Drain cocks and pipes, 711. 
 Draught in funnels, 539. 
 
 Artificial, 547. 
 
 Howden's, 548, 549. 
 
 Ellis, Eaves, 549. 
 
 Drop, Effect of, in condenser, 178. 
 
 valves, 251. 
 
 in marine work, 243. 
 
 Du Temple water-tube boilers, 609. 
 Dual air pumps, Weir's, 377. 
 Dunlop's governor, 706. 
 Duplex pumps, 518. 
 Duralumin, Advantages of, 226. 
 
 Properties of, 818, 819. 
 
 Durham and Churchill's governor, 706. 
 
 E 
 
 Early marine engines, 10, 11, 12. 
 Eccentric rods, 442. 
 
 straps, 442. 
 
 Eccentrics, 440. 
 
 Economy due to superheat, 170, 897. 
 Edwards' patent air pump, 376. 
 Effect of clearance, 113, 124. 
 
 ■ of design on weight, 722. 
 
 of drop in boiler pressure on turbines, 
 
 143. 
 
 of drop in receiver, 178. 
 
 of drop in vacuum on turbine, 143. 
 
 of end pressure on torque, 151. 
 
 of high vacuum on L.P. cylinder, 358. 
 
 of inertia of various parts of an engine, 
 
 746-749. 
 
 of momentum on loads, 298. 
 
 of priming on turbines, 143. 
 
 of steam jackets on efficiency, 169. 
 
 of superheat on efficiency, 169, 170. 
 
 — — of temperature on metals, 83.'!, 
 
 Effect of vacuum on steam consumption, 
 
 362. 
 of weight of various parts of an 
 
 engine, 745. 
 Effective H.P., Constants for, 26. 
 
 mean pressure, 116. 
 
 Effects generally of increase of pressure, 
 
 175. 
 Efficiency, Comparative theoretical, of 
 
 marine engines, 181, 184. 
 
 of air pumps, 381. 
 
 of an engine, 152. 
 
 of boilers, 581. 
 
 of boilers, How to ascertain, 555. 
 
 of condenser, 355. 
 
 of Diesel engines, 859. 
 
 of furnace, 538. 
 
 of German engines, 156. 
 
 of heating surface, 580. 
 
 of modern engines, 153. 
 
 of old marine engines, 153. 
 
 r of propellers, 457. 
 
 of quadruple engines, 161, 162. 
 
 of triple-compound engines, 161, 162. 
 
 of tube surface, 554. 
 
 of turbine, 142. 
 
 Eight-crank engines, 85. 
 
 Ejectors, 409. 
 
 Elastic limit v. yield point as criterions, 
 
 725. 
 Elder's and Dr. Kirk's productions, 10, 11. 
 " Eldorado," s.s., Trials of, 190. 
 Electric light engines, 515. 
 
 installations, Lloyd's rules for 
 
 945. 
 power generating by turbine, 525, 
 
 892. 
 
 transmission, 892. 
 
 Elementary steam engine, 174. 
 " Empress Queen," p.s., 14. 
 Endurance of metals, 289. 
 
 trials, 857. 
 
 Engine overload, 229. 
 
 seatings and beds, 691, 692. 
 
 Engines, Balanced four-crank, 220. 
 
 for generating electricity on ship 
 
 board, 515, 521-523. 
 
 Internal combustion, 85, 859. 
 
 of H.M.S. " Rattler " and " Alecto," 
 
 4. 
 
 " Comet," 1812, p.s., 66. 
 
 " Kaiser Wilhelm II.," s.s., 86. 
 
 Short and long, 216. 
 
 with special valve gear, 219, 220. 
 
 Entablatures of paddle engines, 342. 
 Equivalent twisting moments, 294. 
 Estimated horse-power, 197. 
 
 — Rule for, 198. 
 
 Estimating I.H.P. for speed by wettcd-skin 
 
 resistance, 35, 36. 
 Evaporation, Methods of, 552. 
 Evaporative power of boilers, 555. 
 Evaporators and distillers, 713, 714. 
 
 62 
 
974 
 
 INDEX 
 
 Evaporators, Use of, 372. 
 
 " Eveston," s.s., Engines of, on Westgarth- 
 Carel design for oil, 868, 869. 
 
 General arrangement of machin- 
 ery in, 871. 
 
 Examination doors on cylinders, 256. 
 
 Examples of engines compared, 176. 
 
 showing effects of different margins, 
 
 37. 
 
 Exhaust passages and pipes, Size of, 247. 
 
 steam from auxiliaries, how dealt 
 
 with, 513. 
 
 Expansion diagram showing stages in pro- 
 gress, 194. 
 
 joints, 507. 
 
 of moist steam, 110. 
 
 of steam by stages, 109. 
 
 Rates of, in marine engines, 230. 
 
 valve on independent face, Diagram 
 
 of, 451. 
 
 valves, 437, 438, 439. 
 
 on back of main valves, diagram 
 
 of motion, 452, 453, 454. 
 
 on compound engines, 440. 
 
 piston type, 439. 
 
 Expansive engine, 175. 
 
 compared with compound, 178, 
 
 191. 
 
 with two cylinders, 182. 
 
 Experiments with metal by Prof. Arnold, 
 289. 
 
 with metal by Wohler, 289. 
 
 with superheated steam, 71, 897. 
 
 with surface condensers, 350. 
 
 Express and cross-channel steamers, 734. 
 
 steamer and naval engines, Stress on, 
 
 727. 
 
 steamers, Service speed of, 230. 
 
 Extra supply cock for feed water, 371. 
 
 Extruded material, Dick's, 215. 
 
 Facings and feet on cylinders, 253. 
 Factors for determining size of steam 
 exhaust pipes, 247. 
 
 of safety, 846. 
 
 Board of Trade, 907. 
 
 to determine mean pressure, 125. 
 
 and 
 
 683, 
 
 False valve faces, 280. 
 
 Fans, Ventilating, etc., 521, 522. 
 
 Feathering floats, Gear for restraining, 464. 
 
 screws (Bevis), 499. 
 
 Feed check valves, Size of, 682. 
 
 filters, 717. 
 
 heaters, 711, 712. 
 
 pipes, Sizes of, 406. 
 
 pump rods, 407. 
 
 valves, 405. 
 
 pumps, 402. 
 
 Main and auxiliary, 518, 519. 
 tanks, 407, 
 
 Feed temperature regulators, 356. 
 
 valves on boilers, Automatic, 
 
 684. 
 
 water, Extra supply of, 371. 
 
 Net, 402. 
 
 Ferrules for fire end of boiler tubes (Ad- 
 miralty), 662. 
 Firebars and doors (Henderson's), 673. 
 - of boiler furnaces, 672, 673. 
 Martin's, 673. 
 
 Fire grate, Area of, 574. 
 
 Fireplaces in water-tube boilers, 632. 
 
 Fitting machinery into a ship, 691, 694. 
 
 Five-crank engines, 85. 
 
 Flanges to pipes, 508. 
 
 Flat plates, Board of Trade rules for, 905. 
 
 British Corporation rules for, 
 
 919. 
 Bureau Veritas rules for, 926, 
 
 927. 
 
 Lloyd's rules, 913. 
 
 surfaces, German Government rules, 
 
 942. 
 Stiffening of, 258. 
 
 Flow of cooling water, Rate of, 361. 
 
 Foot valves, Effect of, on air pump action, 
 
 381. 
 Form of ship suitable for speed, 24. 
 Formula for efficiency of marine engines. 
 
 154, 155. 
 Fottinger's torsion meter, 144, 145. 
 Foundations, 334. 
 Four-crank balanced engines, 220. 
 
 engines, 84. 
 
 Framing of engines, 341. 
 
 of horizontal engines, 341. 
 
 French cruiser, " Dupuy de Lome," 16. 
 Fresh water distilling installation, 715. 
 Friction, Coefficients of, for various sub- 
 stances, 843, 844. 
 
 Horse-power, 203. 
 
 in ports and passages, 124. 
 
 Laws of, 165. 
 
 of diagonal engines, 163. 
 
 of guides, 164. 
 
 of journals, 165. 
 
 of pistons, 163. 
 
 of stuffing-boxes, 164. 
 
 of valve gear, 167. . 
 
 Frictional losses, 165. 
 
 resistance of clean surfaces, 49. 
 
 of steam 123. 
 
 Froude's experiments, 29. 
 
 Fuel, Combustion of, effected, 553. 
 
 consumed per square foot grate, 549. 
 
 consumption, 730. 
 
 of oil engines compared with 
 
 turbines, 105. 
 
 " Lusitania," Ml. 
 
 per I.H.P., 676. 
 
 Fuels. Calorific value of. 5 '4. 545. 
 
 to calculate, 546. 
 
 ■- ... Chemical composition of, 544, 545, 
 
iNDEX. 
 
 975 
 
 Fuels, rate of combustion, 547. 
 
 solid varieties, 539, 544. 
 
 Fundamental principles of marine pro- 
 pulsion, 1. 
 Funnel and uptake sections, 582. 
 
 Size of, 550, 552. 
 
 Funnels, diameter of, Rules for, 583. 
 
 of steam ships, 670, 671. 
 
 Furance, Efficiency of, 538. 
 
 fronts and fittings, 672. 
 
 Furnaces, British Corporation rules, 919. 
 
 ■ Bureau Veritas rules, 928. 
 
 • corrugated, ribbed, etc., Rules for, 657. 
 
 German Government rules, 941. 
 
 Lloyd's rules, 915. 
 
 methods of connecting to ends and 
 
 chambers, 658, 659. 
 
 methods of stiffening, 655. 
 
 special designs for withdrawals, 658. 
 
 Thickness of, 654. 
 
 Gain in efficiency by superheating steam, 
 
 170. 
 Gas engines on board ship, 108. 
 Gauges for steam and other pressures, 707. 
 Geared paddle engines, 5. 
 
 turbines, 887. 
 
 Gearing by wheels, Lubrication of, 895. 
 General arrangement of machinery, " Lusi- 
 
 tania," 90. 
 design of marine engine, Influences 
 
 affecting, 210, 211, 212. 
 Genesis of the compound engine, 174. 
 German Government experiments with 
 turbines, 95. 
 
 rules for boilers, 936. 
 
 1 for boiler shells, 940. 
 
 for flat surfaces, 942. 
 
 for furnaces, 941. 
 
 ■ for general for boilers, 943. 
 
 . for stays, 941. 
 
 ■ for tests of material, 943. 
 
 silver, 822. 
 
 Girders, Board of Trade rules, 904. 
 
 British Corporation rules, 920. 
 
 Lloyd's rules, 915. 
 
 Gland packings, 259. 
 
 Governors, 705, 706, 707. 
 
 Graphic method of ascertaining mean pres- 
 sure, 112. 
 
 of finding mean pressure of a 
 
 compound system, 128. 
 
 Graphite and other solid lubricants, 842. 
 
 Grate area, Proportions of, 577. 
 
 of water-tube boilers, 590. 
 
 Gratings and platforms, 716. 
 
 Grease, 842. 
 
 " Great Britain," S.S., Shaft of, 214. 
 
 " Great Eastern," Particulars of paddle 
 engines, 70. 
 
 '' Greyhound," Experiments with, 29. 
 
 Griffin's engine for heavy oils, 59. 
 
 Gross horse-power, 203. 
 
 Guide plates, 340. 
 
 Guides for valve-rods, 325, 326. 
 
 Gun-metal or bronze for valve-rods, 820. 
 
 Gurney's water-tube boiler, 587. 
 
 H 
 
 Harmonic motion, 758. 
 
 Heat, Effect of, on metals, 833, 834. 
 
 Heating surface, 553. 
 
 Effective, 577. 
 
 Efficiency of, 580. 
 
 for superheat, 898. 
 
 in practice, 579. 
 
 of water-tube boilers, 590. 
 
 Heavy oil engine, 400 B.H.P., 107. 
 
 or crude oil engine, 85. 
 
 Herreshoff, water-tube boiler, 592. 
 
 High-pressure steam, Advantages of, 162. 
 
 Hirsch's screw, 495. 
 
 Hirn's estimation of net power, 200. 
 
 Hohenstein water-tube boiler, 621. 
 
 Holding down bolts, 254, 694. 
 
 Hollow shafts, 332. 
 
 compared with solid, 333. 
 
 Hopkinson-Thring torsion meter, 146, 147. 
 Hopkinson, Prof., on shaft horse-power, 
 
 202. 
 Horizontal cylinders, Feet of, 254. 
 
 direct-acting engine, 72. 
 
 Hornblowers' patent of 17, 81, 174. 
 Horse-power, Brake, 203. 
 
 Friction, 203. 
 
 — — Gross, 203. 
 
 Indicated, 198. 
 
 Net, 204. 
 
 Nominal, 195, 196, 197. 
 
 Shaft, 150,201. 
 
 Thrust, 203. 
 
 Tow-rope, 204. 
 
 Watts' nominal, 195. 
 
 Hot surfaces, Covering of, 257. 
 Howden's system of forced draught, 548. 
 Hydraulic installations, 526. 
 
 transmission, 893. 
 
 Hydrokineters, Weir's, 687. 
 Hydromotors of kinds, 3. 
 
 Immadium, a bronze for boiler mountings, 
 
 821. 
 Impermeators, Mechanical, 710. 
 Impulse turbine, 136. 
 
 compounded for pressure and 
 
 velocity, 137, 138. 
 Increased pressure of steam, 187. 
 Indicated horse-power. 198, 230. 
 
976 
 
 1NUEX. 
 
 Indicated thrust, 204. 
 
 (Fronde), 475, 47i'>, 477 
 
 Indicator diagram, 198, 199. 
 
 s.s. " Vasari," 193. 
 
 diagrams showing effect of notching 
 
 up, 441. 
 
 readings, Defects of, 200. 
 
 Inertia and weight, Effect of, general case, 
 793. 
 
 of the connecting-rod, 762. 
 
 of valve gears, 765. 
 
 of various parts of an engine, 7(30. 
 
 Stresses due to, 797. 
 
 Influence of auxiliary machinery on con- 
 sumption, steam, and fuel, 152. 
 
 of Board of Trade on steel, 221. 
 
 of surface on friction, 165, 166. 
 
 of temperature on friction, 1C6. 
 
 of tonnage laws, 215. 
 
 of velocity on friction, 166. 
 
 Injection orifice, Size of, 346. 
 
 water, Amount of, 345. 
 
 Inlet and discharge valves, 502, 504. 
 
 pipes for cooling water, 371. 
 
 valves, Admiralty, Kingston's, 503. 
 
 Intermediate shafts, Loads on, 287. 
 
 Intermittent stresses, 289, 723. 
 
 Internal combustion engines, 85, 859. 
 
 pipes in boilers, 681. 
 
 Inventors of the compound engine, 174. 
 
 " Ireland," p.s., large oscillating paddle 
 engines, 70. 
 
 Iron bars, 805. 
 
 Best Yorkshire, 806. 
 
 cast, Strength of, 805. 
 
 Cleveland, 803. 
 
 Cold blast, 804. 
 
 Cumberland haematite, 804. 
 
 forgings, 807. 
 
 ■ Lincolnshire, 803. 
 
 ■ mixtures, 804. 
 
 — Shropshire, 803. 
 
 Specific gravity of cast, 804. 
 
 ■ Staffordshire, 803, 806. 
 
 Welsh, 803. 
 
 wrought, Kinds of, 805. 
 
 Jacket drains, 711. 
 
 John's coefficients for E.H.P., 26. 
 
 Junk ring, 262. 
 
 bolts, 269. 
 
 " Juno," Trials of, 190. 
 
 K 
 
 Kinetic system of air pump, 386. 
 Kingston valves, 502. 
 Kirk's analysis of a ship's form, 24. 32. 
 " Kovno," Trials of, 188, 189. 
 
 Ladders, 710. 
 
 Lagging and clothing of hot surfaces, 
 relative value of materials, 689. 
 
 ■ of cylinders, 257. 
 
 Lake paddle steamer, S. 
 
 Lap joints with various riveting, G40, 041 
 
 of valve, Effect of, 448. 
 
 Large bolts, Stresses on, 289. 
 
 Lead, Properties of, 818. 
 
 " Lead," The effect of, 447. 
 
 Length of stroke of piston, 209. 
 
 Life of a water-tube boiler, 633. 
 
 Limits to superheating, 169. 
 
 Limitations of speed, Method of determin- 
 ing, 25. 
 
 to piston speeds, 206. 
 
 Liners of steam cylinders, 248, 249. 
 
 Advantage of, 250. 
 
 Link motion, 428. 
 
 double bars, 431, 432. 
 
 position of suspension pin, 430. 
 
 single bars, 431. 
 
 Liquefaction during expansion of steam 
 124. 
 
 Ljungstrom turbine, 140. 
 
 Lloyd's rules for boiler furnaces, 915* 
 
 shells, 912. 
 
 stays, 913. 
 
 electric lighting, 945. 
 
 flat plates, 913. 
 
 girders, 915. 
 
 nominal horse-power, 197. 
 
 oil and other internal com- 
 bustion engines, 872, 873. 
 
 refrigerating machinery,949. 
 
 Load on piston-rods, 273. 
 
 Loads and stresses, Effect of, 723. 
 
 Locomotive boilers, 572, 573. 
 
 Locus of a piston, Diagram of, 444. 
 
 Loss by conductivity, 173. 
 
 by friction of steam in pipes and 
 
 passages, 172. 
 
 by radiation, 169. 
 
 due to air pump, 167, 168. 
 
 due to circulating pumps, 168. 
 
 due to variation in velocity of steam, 
 
 172. 
 
 from driving pumps of an engine, 167. 
 
 Losses due to heat leaks, 169. 
 
 due to liquefaction, 169. 
 
 due to mechanical defects, 169. 
 
 in engine due to physical causes, 169. 
 
 Low -pressure steam. Characteristics of, 363. 
 
 Lubricants, limits of pressure per square inch 
 for good working, 843. 
 
 Solid,- 842, 901. 
 
 Lubricating arrangements, 212, 895. 
 
 Lubricators and impermeators, 709. 
 
 Centrifugal. 709. 
 
 Sight feed, Tin 
 
 " Lusitania," Steam consumption of, 513. 
 
INDKX. 
 
 977 
 
 M 
 
 MacAlpine's quadruple self-balanced en- 
 gine, 752, 753. 
 
 Machinery in Atlantic steamers, 732. 
 
 of battleships, 719. 
 
 of cruisers, 719, 720. 
 
 of destroyers, 719. 
 
 of s.s. " Olympic," 733. 
 
 of various cargo and passenger steam- 
 ers, 741 . 
 
 general small passenger 
 
 steamers, 742. 
 
 naval ships, 736, 739. 
 
 ocean express steamers, 740. 
 
 paddle-wheel steamers, 743. 
 
 steamers, 735. 
 
 MacLaine's piston, 266. 
 
 Mail and passenger steamers, 733. 
 
 Main bearing, bolts, 337. 
 
 brasses, Sizes of, 339. 
 
 caps, 337. 
 
 frames, 338. 
 
 steel, 338. 
 
 bearings, Loads on, 335. 
 
 steam pipes, Size of, 243, 246. 
 
 Manganese bronze, 820. 
 
 steel, 815. 
 
 Manholes in boilers, 666-668. 
 Manoeuvring turbines, 143, 700. 
 Margins of power discussed, 36, 37. 
 
 of safety, 846. 
 
 of speed in ships, 230. 
 
 Marine engine, compared with ones of simi- 
 lar design on shore, 56, 57. 
 
 requisites, 180. 
 
 engines differ from land ones, 228. 
 
 machinery, weight of modern instal- 
 lations, 731. 
 
 motors, Functions of, 4. 
 
 propellers, Forms of, 1 . 
 
 Fundamental principle of, 457. 
 
 steam engines, Kinds of, 5, 11, 12. 
 
 Material, Influence of, 214. 
 
 Supply of, 214. 
 
 Materials used by marine engineers, 
 
 802, 828. 
 Mean pressure, 112, 721. 
 
 in compound engines, 117, 
 
 119. 
 
 in Diesel oil engine, 82. 
 
 in a steam cylinder, 199. 
 
 in method of finishing, 199. 
 
 Mechanical efficiency of engines, 163. 
 
 of marine engine, 153. 
 
 of reciprocating engine, 154. 
 
 of turbines, 143, 154. 
 
 losses in marine engines, 154. 
 
 722, 
 
 118, 
 
 Medium-pressure cylinders, Size of, 234. 
 Melloid, a tin bronze, 822. 
 Metallic packings, 261. 
 Metals, effect of heat, 833. 
 modern effect, etc., 722. 
 
 Metals, Prices of, 828. 
 
 Properties of, 826, 827. 
 
 Methods of boiler work, 650. 
 
 of gauging efficiency, 152. 
 
 of working air and circulating pumps, 
 
 390. 
 Miyabara water-tube boiler, 622. 
 Modern direct-acting paddle engine, 60, 61. 
 
 experiments with superheating, 170. 
 
 of M. Felix Goddard, 170. 
 
 high -pressure boilers, Scantlings of, 
 
 . 667. 
 
 nominal horse-power, 196. 
 
 screw blades, 493. 
 
 Momentum of moving parts, 298. ' 
 
 Morison's condenser, 362. 
 
 furnace, 656. 
 
 Motion, Harmonic, 758. 
 
 of a piston, 444. 
 
 Mudd's experiments for efficiency, 153. 
 Multiple screws, Russian " Popoffskas," 
 
 " Ermack," " Lubeck," 16. 
 Mumford's water-tube boiler, 609, 610. 
 Muntz metal, Properties of, 819. 
 Murdoch's governor, 708. 
 
 N 
 
 Naval brass, Admiralty mixture, 963. 
 
 Properties of, 819. 
 
 Net horse-power, 204. 
 
 Nickel steel, 813. 
 
 Niclausse water-tube boiler, 617, 618. 
 
 Nominal horse-power, Value of, 195. 
 
 Lloyd's rule for, 197. 
 
 Modern rule for, 196. 
 
 North Atlantic steamer service, 732. 
 
 " Northern," Trials of, 188, 189. 
 
 Notching up diagram, showing how effected, 
 450, 456. 
 
 Effect of, 441, 449. 
 
 Number of cylinders of compound engine, 
 82-83. 
 
 of triple- and quadruple- 
 compound engines, 83. 
 
 Objection to high pressures in engines, 
 
 162. 
 Oil, Consumption of, on ships, 836. 
 
 engine, Double-acting, 862, 866. 
 
 Kind & Co. twin, 863. 
 
 Pumps of, 860. 
 
 Junkers, design by Weser Co., 
 
 864. 
 section of a Westgarth-Carel 
 
 cylinder, 870. 
 
 engines, Beardmore's semi-Diesel, 872. 
 
 Classes of, 85, 859. 
 
 Diesel type, 859. 
 
978 
 
 INDEX. 
 
 Nuremburg, 
 
 8151, 
 
 Oil engines, History of, 6. 
 
 Lloyd's rules for, 872, 873 
 
 M. A. N. Co., 
 
 862. 
 Manoeuvring with, 865. 
 rate of revolution, 860. 
 Reversal of, 88. 
 Reversing of, 865. 
 Shafts of, Rules for, 322, 329. 
 Weight of, S59. 
 
 Westgarth-Carel, in s.s. " Eves- 
 ton," 867, 868. 
 fuel, Burners for, 540, 543. 
 
 burning and carrying, Lloyd's 
 
 rules for, 955. 
 
 Methods of burning, 674. 
 
 fuels, Furnace fittings for, 542. 
 
 Oils and lubricants, 836 
 
 Animal, 836. 
 
 Boiling, setting, and flash points of, 
 
 840. 
 
 Compound, 840. 
 
 crude, Various, 543, 545. 
 
 Density of, 838. 
 
 Mineral, 838. 
 
 Mixing of, 841. 
 
 Prices of, 840. 
 
 Specific gravity of, 838. 
 
 Vegetable, 837. 
 
 Viscosity of, 838. 
 
 " Olympic," Particulars of, 97, 98, 99. 
 
 Particulars of machinery, 733. 
 
 triple screw and combination reci- 
 
 procator and turbine, 16. 
 Oscillating cylinders, Trunnions of, 254. 
 
 engines, Framing of, 342. 
 
 paddle engines, 68, 69. 
 
 " Otaki," s.s., Trials of, 97. 
 Oval boilers, 558, 569. 
 
 Packing rings of pistons, 262. 
 Packings of pistons, 163. 
 Paddle engines, 727. 
 
 Modern forms of, 60, 73. 
 
 float, Immersion of, 471. 
 
 - floats. Area of, 466, 469. 
 Number of, 467. 
 - Wooden, 469. 
 shafts, Diameter of, 302. 
 
 Rules for, 331. 
 
 Stresses on, 301. 
 
 wheel, Applications of, 1, 2, 60. 
 areas, 471. 
 bosses, 471. 
 
 common with fixed floats, 461. 
 Diameter of, 464,466. 
 floats, Locus of, 463. 
 machinery, Features of, 2. 
 propulsion problems, 460, 7446. 
 rims, 472. 
 
 Paddle wheel, Shafts of, 473. 
 
 steamships, " Ch. Dundas," 
 
 " Comet," " Claremont," 
 " Bessemer," 13, 48. 
 
 with feathering floats, 461, 4G5. 
 
 with steel floats, 468. 
 
 wheels, feathering gear, 469, 470, 472, 
 
 473. 
 Paraffin engine, 85. 
 Parsons' proposals for combinations, 94, 9r, 
 
 96. 
 
 geared turbines, 5. 
 
 turbine, 140. 
 
 Passage of steam through pipes, 248. 
 Passages in turbines, Shape of, 138. 
 Patent fuels, 541, 544. 
 Perfect air pump, Functions of, 386. 
 Performance of s.s. " Otaki " and " Orari," 
 
 96, 97. 
 Period of steam exhaust, 124. 
 Petrol engine, 85. 
 
 Phosphor-bronze, Properties of, 820. 
 Pipes, Board of Trade Rules for, 508. 
 
 ■ Flow of steam through, 248. 
 
 in bilges, 506. 
 
 Inlet and discharge, 371. 
 
 through bulkheads, how fitted, 507. 
 
 Piston, Locks for bolts of, 269. 
 
 packings of air pumps, 391. 
 
 Piston-rod bolts, Sizes of, 280. 
 
 crosshead gudgeons, 277. 
 
 crossheads, 274, 275, 277, 278. 
 
 ends, 270. 
 
 how fitted, 273. 
 
 guide slippers, 277. 
 
 guides and "udgeons, Size of, 278. 
 
 Surface of, 276. 
 
 Thrust on, 275. 
 
 Piston-rods, 271. 
 
 Loads on, 273. 
 
 Sizes of, 272. 
 
 Piston speeds, 204, 720. 
 
 springs, 263. 
 
 valves, 251. 
 
 Packings of, 416. 
 
 - — Thorn's patent, 418. 
 
 Pistons, 262. 
 
 Cast steel, 267. 
 
 — Clearance of, 256. 
 Construction of, 268. 
 
 Pitch of screw, 483. 
 
 - of shaft bearings, 288. 
 ratio, 485. 
 
 Planimetcr, Use of, 200. 
 
 Position of condenser, 211 212. 
 
 Power developed by turbines, 143. 
 
 required to drive a ship. Rules for, 25, 
 
 26, 31. 
 Practical method of estimating M.P., 128. 
 Present-day practice and design, 77. 
 Pressure in cylinder and condenser, Ratio 
 
 of, 233. 
 limit for lubricants, 843. 
 
INDEX. 
 
 979 
 
 Prices of machinery, 720. 
 
 Prismatic coefficients of ships, 28. 
 
 Progress in naval engineering, List of ships, 
 17. 
 
 in mercantile engineering, List of 
 
 ship's, 18. 
 
 made by early engineers, 176. 
 
 Progressive trials, 855, 856. 
 
 ■ — of steamships, methods of plot- 
 ting results, 37, 38. 
 
 of T.s.s. " Lusitania " shown by 
 
 curves, 50. 
 
 Propeller shafts, screw shafts, 311. 
 
 Stresses on, 286. 
 
 Propellers, Best, for actual service, 461. 
 
 — — Dimensions of, 461. 
 
 — — Efficiency of, 457. 
 
 ' Propontis " s.s., first ship with triple- 
 compound engine, 11. 
 
 Pump barrels and liners, 391. 
 
 buckets, 391. 
 
 Circulating, 396, 397, 398, 399. 
 
 ■ rods, 389. 
 
 valves, 392. 
 
 Pumps, feed, Double-ram flywheel, 517. 
 
 etc., Duplex, 518. 
 
 Weir's, 519. 
 
 for cooling water, Size of, 370. 
 
 Main feed, 402, 404. 
 
 Methods of working, 390. 
 
 Q 
 
 Quadruple engines, Cylinders of, 231, 239, 
 
 240. 
 Quadruple-expansion engine diagrams, 193. 
 Quadruple-riveting, 649. 
 
 R 
 
 Ramming chocks and stays, 694. 
 
 Ramsbottom rings, 262. 
 
 Rankine's rule for ascertaining effective 
 
 horse-power, 31. 
 Rate of expansion in earlv marine engine, 
 
 10. 
 
 of revolution in war ships, 208. 
 
 in mercantile ships, 208. 
 
 of marine turbines, 89. 
 
 of revolutions of marine engines, 5, 
 
 207, 208. 
 Ratio of expansion in a turbine, 109. 
 
 of maximum to mean torque, 291. 
 
 of mean pressure to maximum and 
 
 minimum pressures, 110. 
 Ratios of cylinders in compound engines,241 . 
 Reaction turbines, 136. 
 Receiver, Compound engine, 177. 
 
 space of cylinders, 253. 
 
 Reciprocators in U.S. cruiser " Birming- 
 ham," 93. 
 
 Reed's water-tube boiler, 611. 
 Refrigerating machines, 533, 534. 
 
 machinery, Lloyd's rules for, 935. 
 
 Regulating and stop valves, 701, 703. 
 
 valve, 703. 
 
 Relation of frictional to residual resistance, 
 
 49. 
 
 of power to displacements, 49. 
 
 Relief frame for flat valves, Dawe & Holt's, 
 418. 
 
 ■ Martin and Andrews', 420. 
 
 Relief frames for flat valves, 41(>. 
 
 Common Admiralty, 419, 420. 
 
 Church's, 421. 
 
 Requisites in the marine engine, 180. 
 Resistance of pipes and valves, 123. 
 
 of shaft to bending, 294. 
 
 of ships, wave formation, 20, 22. 
 
 to twisting, 290. 
 
 Restrained packings, 266. 
 
 Results of trials of merchant ships, 131. 
 
 of naval engines, 130. 
 
 of various steamers, 41-48. 
 
 Return connecting-rod engines, 74, 75. 
 Reversal of oil engines, 88. 
 
 of propellers of oil engines, 88. 
 
 of stress, influence of numbers on 
 
 materials, 289. 
 Reversing and manoeuvring gear for very 
 
 large turbines, 701, 702. 
 
 engines, 696. 
 
 gear for turbines, 699. 
 
 gears, 442, 521. 
 
 all-round type, Particulars of, 
 
 of, 699. 
 
 ■ Brown's specials, 697, 698. 
 
 Common, 697. 
 
 of oil engines, 865. 
 
 Revolutions of screws, 205, 206. 
 
 Ribbed furnaces (Purves), 656. 
 
 River steamers, Special features of, 2. 
 
 Riveted joints, Strength of, 638. 
 
 Riveting of funnels, smoke boxes, and 
 
 castings, 671. 
 Rods for slide valves, 325. 
 Rotary air pumps, 388. 
 Rule for thickness of a boiler shell, 652. 
 
 Safe working stresses, 289. 
 
 on various materials and differing 
 
 conditions, 724, 726. 
 Safety collars to expansion joints, 508. 
 
 locks of piston bolts, 269. 
 
 Margins and factors of, 846. 
 
 valves, B. of T. rules, 956. 
 
 (Cockburn's), 677, 679. 
 
 French, 930. 
 
 German regulations, 938. 
 
 of boilers, 676. 
 
 ■ " » — — rules of authorities, 679, 680, 
 
980 
 
 INDEX. 
 
 458, 
 
 Safety valves, Size of, 678. 
 
 Sanitary pump, 409. 
 
 Scantlings of engine seats, 692. 
 
 Schlick's inventions, 222. 
 
 " Scotia," Engines of, particulars, 69. 
 
 Scouts and destroyers, 734. 
 
 Screw blades, Bolts of, 497. 
 
 Material of, 490, 498. 
 
 Modern, 493. 
 
 of different pitch, Effect of, 223 
 
 Sections of, 494, 497. 
 
 Shapes of, 495. 
 
 Surface of, 485. 
 
 Thickness of, 488. 
 
 bosses, 490. 
 
 Power of sucking in air, 224. 
 
 propeller, Action of, in water, 
 
 459. 
 
 - Diameter of, 474. 
 Features and applications of, 3. 
 
 - for shallow draft, Thornycroft 
 and Yarrow's, 3. 
 
 original forms, 474. 
 
 propellers, Diameter of, 480, 481. 
 
 Early history of (" Scotia," 
 
 " Priscilla," " Empress Queen," 
 " La Marguerite," Woodcroft, 
 F. P. Smith, Ericsson), 13, 15. 
 
 Forms of (Griffiths), 15. 
 
 Feathering, 499, 500. 
 
 Leading features of, 480. 
 
 Loads on, 286. 
 
 Method of lifting on board, 501. 
 
 ■ Multiple, 16. 
 
 Particulars of, 478. 
 
 Pitch of, 483. 
 
 Pressure on, 478, 479. 
 
 Rules for, 482. 
 
 Surface and surface ratios of, 
 
 485. 
 
 Weight of, 499. 
 
 shafts, Castings of, 312, 313. 
 
 Stresses on, 311. 
 
 Thrust of, 475, 487. 
 
 working in disturbed water, 224. 
 
 Screwed stays in boilers, 663. 
 
 Screws affected by the hull, 224. 
 
 ■ damaged, 498. 
 
 of turbine steamers, 89. 
 
 Scum cocks on boilers, 685. 
 
 Sea cocks and valves, 502. 
 
 experiences with turbines in U.S. 
 
 cruiser " Chester," by Lieut. Yates, 
 143. 
 
 — — performance of steamers and steam- 
 ships on service, 39, 40. 
 
 water, Air contained in, 373. 
 
 Solid matter in, 402. 
 
 Securing engines and boilers in the ship, 
 695. 
 
 "Selandia," S.O., Particulars of, 866, 867. 
 
 Self-balanced flat valves, 417. 
 
 Semi-Diesel engines, 87, 872. 
 
 Sentinel steam valve, 687. 
 Separators for water deposit, 689. 
 Sequence of cranks of compound engines, 
 
 Effect of, on diagrams, 455, 456. 
 Shaft coupling bolts, 308. 
 
 couplings, 308. 
 
 ■ with drivers, 310. 
 
 ■ with taper bolts, 310. 
 
 ends, Outer bearings of, 312. 
 
 horse-power, 701. 
 
 Formula for, 150. 
 
 journals, Surface of, 309. 
 
 Shafting, hollow, Strength of, 286. 
 
 of marine engines, 286. 
 
 Shafts, Diameter of, to resist torsion, 291. 
 
 Effects of centrifugal force on, 287. 
 
 for oil engines, 287, 322-329. 
 
 for turbines, 287, 328, 329, 332. 
 
 Rules for, 329. 
 
 of Board of Trade, 311. 
 
 British Corporation, 323. 
 
 Bureau Veritas, 325. 
 
 Lloyd's, 322. 
 
 Stiffness of, 292. 
 
 Stresses on, 288. 
 
 Shallow water ships, 734. 
 S.H.P., Calculations of, 201. 
 Ship, Form of, for speed, 24. 
 Side lever engines, 66, 67. 
 Single-acting Mirrlees' Diesel, 105. 
 
 oil engine, 102, 103. 
 
 Single-crank engine, 83. 
 
 engine and valve gear balanced by a 
 
 rotating weight, 766. 
 engines, Ward's, 217, 218. 
 
 Six-crank engine, 85. 
 
 Size of intermediate cylinders, 234. 
 
 Skin friction, residual resistance, Rules for 
 
 Taylor's, 21. 
 resistance of ships, Calculations of, 
 
 23. 
 Slide valves, Locomotive, 412. 
 
 Proportions of, 327, 328. 
 
 valve-rods, Guides of, 315. 
 
 Slip of propeller, Apparent, 458, 460. 
 
 Real, 458, 460. 
 
 Ratio of, turbine steamers' screws, 
 
 89. 
 Slot link, Proportions of, 431. 
 links, 429. 
 
 Smoke box and fittings, 669. 
 Snifling valve, Uses of, 347. 
 Soaps, 841. 
 
 Solid bronze screw of s.s. 
 496. 
 matter in sea water, 832. 
 
 Mauretania,' 
 
 Space occupied by machinery, 215. 
 Spare gear, 730. 
 
 Special modern metals, Effect on design of, 
 722. 
 
 pistons, 264. 
 
 Speed margin in ships, 230. 
 of piston, 204-205. 
 
INDEX. 
 
 981 
 
 Speed ot ship (suitable), 24. 
 
 of torpedo boats and destroyers, 37. 
 
 Stages in use of steam shown in an expansion 
 
 diagram, 194. 
 Standard design of engine, 210. 
 Starting turbines, 143. 
 Stay bars, Board of Trade rule for, 660. 
 
 — — of boilers, stresses allowed by the 
 
 authorities, 663, 664. 
 
 Particulars of, 660. 
 
 Pitch of, by rules of various 
 
 authorities, 664. 
 
 tubes for boilers, 662. 
 
 Staying engines in the ship, 694. 
 Stays, British Corporation rules, 919. 
 
 ■ Bureau Veritas rules, 927. 
 
 German rules, 941. 
 
 Steam ash hoists, 705. 
 
 consumption, H.M.S. " Achilles," 89. 
 
 of turbine, Rateau's formula, 
 
 143. 
 consumptions at various speeds of 
 
 H.M.S. " Amethyst " and " Topaze," 
 
 93. 
 
 Effect of increased pressure, 187. 
 
 efficiency, 162. 
 
 of reciprocators, 161. 
 
 — ■ — of turbines, 161. 
 
 — engine indicator, Advantage of, 151. 
 
 — gauges, 687. 
 
 — hammer, Introduction of, 214. 
 
 — jackets, 252. 
 
 — ports and passages, Size of 246. 
 of L.P. cylinders, 82. 
 
 Size of, 235. 
 
 Width of, 251. 
 
 space in boilers, 582. 
 
 — - steering gears, 528. 
 
 Superheated, 170, 897. 
 
 turbines, 136. 
 
 turning gear, 704. 
 
 ■ used expansively, 109. 
 
 whistles, 688. 
 
 and syrens, Waste by, 517. 
 
 Steamers for shallow waters, 734. 
 
 on Clyde Estuary, 734. 
 
 on N. Atlantic, 732. 
 
 Various, 735. 
 
 Steel bars and plates, 807. 
 
 Basic, 808. 
 
 Bessemer, 807. 
 
 B. of T. tests, 810. 
 
 boiler plates, extras of cost, 811, 812. 
 
 Sizes of, 811. 
 
 castings, 812. 
 
 • Introduction of, 725. 
 
 chrome alloys, 814. 
 
 vanadium alloys, 814. 
 
 — effect of heat treatment, 
 
 815. 
 
 electrical, Use of, 214. 
 
 for boilers, 808. 
 
 for British Corporation, 810 
 
 Steel for solid-drawn tubes, 815. 
 
 forgings, 812. 
 
 Physical properties of. 830, 831. 
 
 Lloyd's tests, 810. 
 
 nickel alloys, 813. 
 
 - pipes, Rules for thickness of, 509. 
 Sizes of, 510. 
 
 pistons, 267. 
 
 platesand forgings, Introduction of, 214. 
 
 Silicon high-tensile, 811. 
 
 Size of, 214. 
 
 Self -hardening, 214. 
 
 Siemens-Martin, 808. 
 
 springs for safety valves, 681. 
 
 Tests for, 809. 
 
 Steeple engines for paddle steamers, 70, 71. 
 Steering gears, Essential features of, 530. 
 
 Steam, as fitted in warships, 529. 
 
 Brown's, 530, 531. 
 
 Stern bush, 312, 313, 314. 
 — tube, 314. 
 
 Cedervall's, 316. 
 
 wheeler, engines and wheels, 1 1 . 
 
 Stirling water-tube boiler, 620, 621. 
 Stokehole ventilators, Rules for, 717. 
 Stone's bronze, 822. 
 Stop and throttle valves, Use of, 246. 
 — valve on boilers, Rules for, 674. 
 valves (Admiralty), 656. 
 
 Stresses, Alternating, 725. 
 
 due to inertia, 797. 
 
 Intermittent, 723. 
 
 on rod bolts, 289. 
 
 on the arm of a crank shaft, 289. 
 
 safe-working, Limits of, 289. 
 
 Stretch of metals under tension, 726. 
 Stroke of piston, Length of, 209. 
 Studs and bolts, Pitch of, 258. 
 Stuffing-boxes, 259. 
 Sizes of, 260. 
 
 Superheated steam, 170, 897. 
 
 Economy of, 170, 898. 
 
 Superheaters, Surface of, 899. 
 Supply of materials, 214. 
 Surface condensation, Economy of, 349. 
 — inside v. outside of tubes, 366. 
 Necessity for, 349. 
 
 condenser, 347. 
 
 as fitted by Hall, 348. 
 
 Efficiency of, 355. 
 
 - History of, 347. 
 Inventors of, 347. 
 
 condensers, Construction of, 367. 
 
 Supporting flat surfaces of, 368. 
 
 Trials of, in battleship " Ibuki,' 
 
 368. 
 
 of boiler tubes, 554, 578. 
 
 — ■ — ■ of screw propellers, 473, 474. 
 — - — ratio of screws, 485. 
 
 Total heating, 579. 
 
 Suspension pin of link motions, 430. 
 Symington's, William, inventions, 6. 
 
 Syrens, 688. 
 
 63 
 
982 
 
 INDEX. 
 
 Tallow, 842. 
 
 Tank experiment with ship's models, 29, 
 
 30. 
 Taylor's formula for depth of water, 55. 
 Temperature, Effect of, on metals, 833. 
 Tension on metals, Effect of, 726. 
 Test and trials, Object of, 845. 
 Testing boilers (Admiralty rule), 63b\ 
 
 Bureau Veritas, 636. 
 
 German Government, 636. 
 
 regulations, 939. 
 
 by water, 634, 636, 849. 
 
 machinery (Admiralty), 849. 
 
 — — materials by Admiralty, 962. 
 
 of boilers by water, 634, 636. 
 
 Tests and trials of machinery, 845. 
 
 of boiler material by various authori- 
 ties, 635. 
 
 of condensers (Admiralty), 372. 
 
 of machinery before and after fitting 
 
 on board, 849. 
 
 of material, criterion, 845. 
 
 of materials, Bureau Veritas, 961. 
 
 for boilers by Admiralty, 
 
 931. 
 
 ■ German rules, 943. 
 
 of steel, American standards, 829, 
 
 830. 
 
 of turbines, 848, 849. 
 
 particulars for various parts (Ad- 
 miralty), 850, 851. 
 
 - (Board of Trade), 853. 
 
 — : (British Corporation), 852. 
 
 (Germanischer Lloyd's), 853. 
 
 854. 
 — ■ (Italian Naval), 852. 
 
 - (Lloyd's Register), 852. 
 Theoretical indicator diagrams for three- 
 stage expansion, 129. 
 
 Thickness of copper pipes for various 
 
 services, 509. 
 Thornycroft's oil engines for boats, 58. 
 
 ■ water-tube boiler, 609. 
 
 Thornycroft-Marshall water-tube boiler, 
 
 623, 624. 
 Three-crank engines, 84. 
 Thrust, 316, 457, 475. 
 — — bearing, Pressure on, 317. 
 
 block, Friction of, 317. 
 
 seating, 692. 
 
 blocks, 319. 
 
 collars, Details of, 318. 
 
 ■ Diameter of, 317. 
 
 horse-power, 203. 
 
 H.P., Heck's method of determining, 
 
 203. 
 — — ■ Indicated, 475. 
 
 Rankine's rule for, 477. 
 
 Tin, Properties of, 817. 
 
 Tonnage, Gross and net, 215. 
 
 laws. Influence of. on design, 215. 
 
 Tonpie, Effect of, on shafts, 292. 
 
 on shafts, 290. 
 
 Torsion and shearing of metals, 725. 
 
 meter diagrams, 144, 151. 
 
 meters, 145. 
 
 Torsional stiffness of shafts, 292. 
 
 Total heating surface, Allowance of, 578. 
 
 Method of estimating, 580. 
 
 resistance of destroyers, 59. 
 
 Tow rope horse-power, 204. 
 
 Tramp steamers' machinery, Stresses on 
 
 728. 
 Transmission of power, 887. 
 Treatment of hot bearings of turbine, 143. 
 Trial trip conditions considered, 36. 
 
 trips of the mercantile marine, 858. 
 
 hips 
 
 Trials at sea, Progressive, 856. 
 
 for endurance, 857. 
 
 of Belleville boiler in H.M.S. " Sharp- 
 shooter," 616. 
 
 of compound engines in naval 
 
 183. 
 
 of four-stage quadruple engines, 135. 
 
 of s.s. " Garonne " and " Ranee," 
 
 170. 
 
 of H.M.S. " Amethyst " and sister 
 
 ships, 91. 
 
 of H.M.S. " Cossack " in water of 
 
 different depths, 51. 
 
 of H.M.S. "Hermes" for water con- 
 sumption, 627. 
 
 of modern steamships, 45, 47, 855. 
 
 of paddle-wheel steamers, 48. 
 
 of screw steamers, 46. 
 
 of three-stage expansion four-crank 
 
 engines, 134. 
 
 of three-stage three-crank, 133 
 
 of triple-compound engines, 188. 
 
 of turbine steamers, 47. 
 
 of two-stage compound engines, 132. 
 
 — — Progressive, 855. 
 
 under steam, Admiralty, 855. 
 
 Purposes of, 853. 
 
 with superheaters, 170, 897. 
 
 Triple-expansion compound engine, 185. 
 — ■ — engines, Cylinders of, 236, 237, 238. 
 
 engines for fishing craft, etc., 81. 
 
 Triple screws, introduction and history, 
 15, 16. 
 compound engines, Trials of, 188. 
 
 Trunk engines by John Penn & Son, 72, 73. 
 Trunnions of oscillating cylinders, 254. 
 Tube plates of condensers, 364. 
 - surface of boilers, 554. 
 
 Allowance of, 578. 
 
 Tubes. for boilers, 660, 661, 662. 
 of water-tube boilers, 591. 
 
 Tug boats, Special features of, 1. 
 Tunnel-shaft bearings, 320. 
 Tunnel shafting. Loads on, 287. 
 
 pedestals, 693. 
 
 shafts, Diameter of, 292. 
 
 Turbadium or bronze for propellers, 821. 
 
INDEX. 
 
 983 
 
 Turbine and reciprocators compared, 89. 
 
 arrangement on ships, 139, 887. 
 
 cruiser " Lubeck," 89, 93. 
 
 driven generating set, 525. 
 
 Efficiency of, 88, 887. 
 
 experiments in U.S. America, 91. 
 
 gearing, Varieties of, 890. 
 
 installations, Weight of, 893. 
 
 Ljungstrom, 140. 
 
 machinery, 88, 887. 
 
 screw steamers, Particulars of. 744c. 
 
 shafts, Rules for, 328, 329, 332. 
 
 — steamers, Trials of, 46. 
 Zoelly, 142. 
 
 Turbines, Application of, as additions to 
 reciprocators, 12, 13, 890. 
 
 Geared, 887-893. 
 
 of s.s. " Lusitania," 89. 
 
 of two elementary kinds, 136. 
 
 Turbo reciprocators, Trials of, 97. 
 Twin-cylinder engines of H.M.S. " Rattler," 
 
 70. 
 Twin screws, Introduction of, 15. 
 Twisting moment, 290. 
 
 moments, Curves of, 296, 297. 
 
 Two-crank engines, 83. 
 
 Two-cycle Diesel engine, 103, 106, 107. 
 
 oil engine, modus operandi, 108. 
 
 Types of marine boilers, 557. 
 
 U 
 
 Uniflux condenser, 351. 
 Uptake and funnel section, 582. 
 
 Vacuum augmentor, Parson's, 389. 
 
 blower, Roger's, 389. 
 
 Effect of, 12, 362. 
 
 Highest possible, 370 
 
 Value of trial trips, 19. 
 Valve box covers, 256. 
 
 Common locomotive, 413 
 
 diagram, 445. 
 
 - — — Double-ported, 251. 
 — - — faces, False, 250. 
 
 gear, Bremme's investigations of, 876. 
 
 — Hackworth's, 433. 
 
 — Joy's, 436. 
 
 Marshall's, 434, 435. 
 
 — Martin's arrangement, 433. 
 
 Morton's investigations of, 880. 
 
 with single eccentric, 433. 
 
 gears, 428. 
 
 - rod bolts, 325. 
 
 spindles, 325. 
 
 scats, 395. 
 
 travel, 415. 
 
 Valves and boxes for bilge pipes, 506. 
 Double-ported slide, 413, 414. 
 
 Valves for distributing steam and their 
 
 gear, 412. 
 
 of air pumps, 392, 393, 394. 
 
 of oscillating cylinders, 254. 
 
 Piston, 415. 
 
 Treble-ported, 414. 
 
 trick kinds, 413, 414. 
 
 with exhaust port through the back, 
 
 423. 
 Vanadium steel, 814. 
 Variation in value of fuel, 152. 
 Various experiments with ships and models 
 
 in water of different depth, 52, 53, 
 
 54. 
 
 types and designs of engine, 57. 
 
 Varying thickness of shell plates for different 
 
 tensile strengths, 650. 
 ' Vasari," s.s., Indicator diagrams of, 193. 
 Velocity flow in turbines, 109 
 Ventilating and force-draught fans, 521, 
 
 522. 
 Ventilators for stokeholds, 717. 
 Vertical air pumps, 376. 
 Diesel acting inverted screw engines, 
 
 75, 76, 77. 
 " Vespasian," s.s., altered by Parsons to 
 turbine driven, 100. 
 
 Consumption of steam on, 231. 
 
 Vibration, Cause of cavitation, 225. 
 
 caused by screws, 223. 
 
 in ships, 222. 
 
 of auxiliary machinery, 225. 
 
 " Violet," p.s., new triple-steeple engines, 
 70. 
 
 W 
 
 Wake currents, Effect of, 459. 
 Ward's single-crank engines, 217, 218. 
 Warships, consumption of steam and fuel, 
 
 232. 
 - — Trials of, 232. 
 Waste of steam by whistles, 517. 
 Water consumed per S.H.P., 153. 
 
 consumption in main and auxiliary 
 
 machinery, H.M.S. "Diana," 628. 
 
 consumption in main and auxiliary 
 
 machinery, H.M.S. " Hermes," 627. 
 
 gauges on boilers, 686. 
 
 service, 507. 
 
 spaces in boilers, 666. 
 
 Water-tube boiler, Du Temple, 609. 
 
 Elements of, 588. 
 
 Essential features of, 588. 
 
 prate area, 590. 
 
 Hohenstein, 621. 
 
 installations compared with tank 
 boiler ones, 630. 
 
 installations, Weight of, 625. 
 
 Miyabara, 622. 
 
 Mumford's, 609, 610. 
 
 Niclausse, 617, 618. 
 
984 
 
 INDEX. 
 
 Water-tube boiler, Reed's, 611. 
 
 Stirling, 620, 621. 
 
 Thomycroft's, 609. 
 
 Thomcroft -Marshall, 623, 624. 
 
 boilers, Drums of, 592. 
 
 Fireplaces of, 632. 
 
 Heating surface of, 590. 
 
 History of, 589. 
 
 - Life of, 633. 
 
 - Tubes of, 591. 
 Types of, 592. 
 
 Watt's nominal horse-power, 195. 
 
 Weigh shafts, Size of, 443. 
 
 Weight in relation to tonnage, 732. 
 
 of engines and boilers, 718, 895. 
 
 of machinery in relation to tonnage, 
 
 732. 
 
 of modern marine machinery in- 
 stallations, Rules for, 731. 
 
 of screw propellers, Rule for, 499. 
 
 of water-tube boiler installations, 625. 
 
 Weights of moving parts of an engine, 779. 
 
 Weighton's experiments on condensation, 
 350. 
 
 Weir's condenser, 351. 
 
 dual air pumps, 377. 
 
 pumps, 519. 
 
 Welded joints of boiler shells, 650. 
 
 Westinghouse governor, 707. 
 
 Wetted skin estimated, Kirk's method, 33. 
 
 Mumford's method, 24. 
 
 Seaton's method, 24. 
 
 Wheel gearing for screw engines, 71, 887. 
 
 Lubrication of, 895. 
 
 — ratios, 887. 
 
 White brass, Parsons', 822. 
 
 metal, Admiralty, 822. 
 
 mixture, 963. 
 
 Babbit's, 822. 
 
 White metal, Fenton's, 822. 
 
 — ■ in bearings, 339. 
 
 — ■ Magnolia, 822. 
 
 Plumtin, 822. 
 
 — ■ Richards' plastic, 822. 
 
 Stone's, 822. 
 
 metals, Composition of, 825. 
 
 Wigzell's triple self-balanced engine, 754. 
 
 755. 
 " Wilberforce," p.s. (1841), 348. 
 Winch condenser, 536. 
 Wire-drawing of steam, 123. 
 Woolf's patent compound engine (1804). 
 
 174. 
 Work, Division of, in compound cylinder, 
 179. 
 
 done theoretically bv a pound of 
 
 steam, 161, 897. 
 Working stress, To determine, 727. 
 
 Yarrow boiler, 604. 
 
 Yarrow's experiments for efficiency, 159. 
 
 Yarrow-Schlick balanced engines, 220. 
 
 Tweedy method of balancing, 754. 
 
 system of balancing an engine, 
 
 788. 
 
 Zetjner's diagram for common valve 
 
 motion, 445. 
 Zinc or spelter, Properties of, 817. 
 Zoelly turbine, 142. 
 
 BBLL ANl> BAlIi, LIMITED, PRIMERS, GLASGOW 
 


 VD 07592 
 
 
 405145 
 
 UNIVERSITY OF CALIFORNIA LIBRARY