WORKS OF 
 PROFESSOR A. M. GREENE, 
 
 PUBLISHED BY 
 
 JOHN WILEY & SONS, Inc. 
 
 Elements of Heating and Ventilation. 
 
 A Text-book for Technical Students and a 
 Reference Book for Engineers. 8vo, vi +324 
 pages, 223 figures. Cloth, $2.50 net. 
 
 Pumping Machinery. 
 
 A Treatise on the History, Design, Construction, 
 and Operation of Various Forms of Pumps. 
 8vo, vi +703 pages, 504 figures. Cloth, $4.00 net. 
 
 The Elements of Refrigeration. 
 
 A Text-book for Students, Engineers and Ware- 
 housemen, vi +472 pages. 6 by 9. 192 figures. 
 Cloth, $4.00 net. 
 
 BY SPANGLER, GREENE, AND MARSHALL: 
 
 Elements of Steam Engineering. 
 
 Third Edition, Revised. 8vo, v+296 pages, 
 284 figures. Cloth, $3.00. 
 
PUMPING MACHINERY 
 
 A TREATISE ON THE 
 
 HISTORY, DESIGN, CONSTRUCTION 
 
 AND OPERATION OF VARIOUS 
 
 FORMS OF PUMPS 
 
 BY 
 
 ARTHUR M. GREENE, JR. 
 
 Professor of Mechanical Engineering, Russell Sage Foundation, Rensselaer Polytechnic 
 Institute; Sometime Junior Dean, School of Engineering , University of Missouri 
 
 SECOND EDITION, REVISED 
 
 NEW YORK 
 
 JOHN WILEY & SONS, INC. 
 
 LONDON: CHAPMAN & HALL, LIMITED 
 
 1919 
 
Engineering 
 Library 
 
 Copyright, 1911, 1918, 1919, 
 
 BY 
 
 ARTHUR M. GREENE, JR. 
 
 PRESS OF 
 
 BRAUNWORTH & CO. 
 
 BOOK MANUFACTURER* 
 
 BROOKLYN. N. Y. 
 
PREFACE TO SECOND EDITION 
 
 IN the second edition of Pumping Machinery it has been 
 thought advisable to rearrange the treatment of water valve 
 design and the dynamics of the water end. The air lift pump 
 has been analyzed in a different manner to agree more closely 
 with the true variation of pressure. Other small minor changes 
 have been made. 
 
 A. M. G., Jr. 
 
 Sunnyslope, Troy, N. Y. 
 May, 1918. 
 
 lii 
 
PREFACE 
 
 SEVEN YEARS ago the course of lectures which forms the 
 basis of this book was given as a required course to the mechan- 
 ical engineering students at the University of Missouri. As a 
 number of civil engineering students elected this, it occurred 
 to the author that there was a demand for a book to cover 
 the field as he had planned the course. 
 
 The general plan of the work is to give a brief historical 
 outline of the development of pumping machinery which is 
 closely connected with the history of the steam engine; then 
 to examine and describe the action of a number of common 
 forms of pumps, following this by the methods of design of 
 pumping apparatus. 
 
 The book is intended for the use of those who have studied 
 hydraulics, mechanics, and the strength of materials. It is 
 intended to develop certain general principles of mechanics 
 which are applicable to pumping machinery as well as to train 
 the student in application of certain of the theoretical portions 
 of the studies of an engineering course. 
 
 The first two chapters are devoted to the development of 
 pumping machinery, because the author feels the need, of all 
 engineers, of a more intimate knowledge of the great works of 
 the early members of the profession. The pump is so closely 
 allied with the steam engine in early times that this history 
 traces the early forms of that machine. The historical works 
 mentioned in the list of text-books in the bibliography as well as 
 the historical articles given in the same catalogue have been 
 indispensable in the preparation of these chapters. 
 
 The descriptive chapters on modern pumps have been 
 prepared from the catalogues and bulletins of manufacturers 
 and from current technical literature. The figures have been 
 redrawn from these sources to fit them for this work. The 
 half-tones have been made from photographs or cuts of publi- 
 cations kindly furnished by the various manufacturers. The 
 
VI 
 
 PREFACE 
 
 thanks of the author to these manufacturers are here publicly 
 expressed. The various articles in the technical press which 
 have been of assistance in the preparation of the lectures and 
 this work are listed as bibliography. The author is indebted 
 to the writers of these articles and texts and to their publishers 
 for the aid he has received. These chapters have been incor- 
 porated into the book with certain tables from the catalogues, 
 so that the student may have in his library a reference book 
 which will not only give him the peculiarities of the different 
 forms of special pumps, but the dimensions or other data may 
 be found so that he may know what size of pump may be had 
 for a certain purpose. 
 
 In the preparation of the theoretical part of the work of 
 design the author has used the methods employed by him in 
 courses in steam-engine design and hydraulic motors, adding 
 thereto methods developed from "Die Pumpen," by Hartmann- 
 Knoke and Berg, and " Die Zentrifugal Pumpen," by Fritz 
 Neumann. These two books have been of great assistance in 
 the preparation of the work. The theory relating to air-lift 
 pumps differs from that given by Harris in his articles in the 
 Proceedings of the American Society of Civil Engineers, which 
 were originally considered. 
 
 The author especially wishes to express his thanks in this 
 place, to his wife, Mary E. Lewis Greene, for her aid in the 
 preparation of the manuscript and in proof-reading, as well as 
 for her advice during the progress of the work. 
 
 Grateful acknowledgments are made to the following: 
 International Steam Pump Company and to Mr. William 
 Schwanhauser, Henry R. Worthington Company, the Allis- 
 Chalmers Company, and Mr. Win Sando, the Snow Pump Works, 
 The Holly Pump Works, and Mr. Decrow, The American Steam 
 Pump Company, the Alberger Company, the Buffalo Pump 
 Company, the Prescott Pump Company, Deane Brothers Pump 
 Company, the National Board of Fire Underwriters, Mr. Francis 
 Head, Prof. Lewis F. Moody, Mr. F. C. Dunlap, Mr. Lester 
 French, and Mr. A. F. Rolf for their aid in furnishing data., 
 illustrations and other materials. 
 
 THE AUTHOR 
 
 TROY, March 15, 1911. 
 
TABLE OF CONTENTS 
 
 CHAPTER I 
 
 PAGB 
 
 HISTORICAL DEVELOPMENT i 
 
 CHAPTER II 
 RECENT HISTORY 56 
 
 CHAPTER III 
 MODERN RECIPROCATING PUMPS : 130 
 
 CHAPTER IV 
 SIMPLEX PUMPS 166 
 
 CHAPTER V 
 DYNAMICS OF WATER END , 190 
 
 CHAPTER VI 
 DESIGN OF PARTS 260 
 
 CHAPTER VII. 
 DYNAMICS OF STEAM END 324 
 
 CHAPTER VIII 
 STEAM END DETAILS . 344 
 
 CHAPTER IX 
 TEST OF PUMPING ENGINES T 410 
 
 CHAPTER X 
 
 HIGH DUTY PUMPS AND WATER WORKS STATIONS 428 
 
 vii 
 
viii TABLE OF CONTENTS 
 
 CHAPTER XI 
 
 PAGE 
 
 SPECIAL PUMPING MACHINERY 456 
 
 CHAPTER XII 
 INJECTOR AND PULSOMETER 503 
 
 CHAPTER XIII 
 AIR-LIFT PUMPS AND PNEUMATIC PUMPS 512 
 
 CHAPTER XIV 
 CENTRIFUGAL PUMPS 535 
 
 CHAPTER XV 
 MINE PUMPS 659 
 
PUMPING MACHINERY 
 
 CHAPTER I 
 
 HISTORICAL DEVELOPMENT 
 
 THE pump, considered as any apparatus used to raise water, 
 is one of the oldest machines known to man. Long before 
 the Christian Era, when man emerged from the age of the hunter 
 into that of the shepherd, he found it necessary to raise water from 
 wells for his flocks in places where there were no pure streams. 
 Indeed, even before this age there must have been devices 
 to raise water from low levels tc supply the personal needs 
 of the hunter and those dependent on him. 
 
 Following the age of the shepherd came that of the farmer, 
 when the demand for an apparatus to raise water was greater 
 than before. The shepherd no longer wandered over the 
 country in search of pasture, but now he cared for a definite 
 tract of land to furnish the food supply for his flocks, his herds, 
 and his family. Unfortunately it became necessary for him 
 to use lands on which there was not a sufficient natural supply 
 of water for irrigation, and he was compelled to lift water from 
 low streams to these fields, so as to increase the yield from his 
 land. 
 
 The next age was that of townsman and manufacturer. 
 For protection, mutual aid, and comfort, man began to live 
 in towns and cities. This necessitated supplying water for 
 the common use by gravity from some higher source to foun- 
 tains, by means of hand pumps from wells, or by utilizing 
 natural springs. These town fountains, pumps, or springs are 
 still prominent objects in the cities of the Old World, and also 
 
2 FUMPIXG ^lACHINERY 
 
 in our own colonial towns. As the town developed the supply 
 was carried into some of the houses. Finally, as the burgher 
 became a manufacturer and was compelled to dig into the 
 earth for his raw materials, it became necessary to clear his 
 excavations of the surface and subsurface waters which filled 
 them. 
 
 Modern civilization has demanded more and more as pump- 
 ing machinery has been perfected, until to-day running water 
 in unlimited supply is found not only in the houses of cities, 
 large and small, but even the isolated farm often has its own 
 water works, giving ample water to each room of its house, 
 barn, and dairy. The latest demand is the supply, at a moment's 
 notice, of large quantities of water, under great pressure, to 
 the congested districts of trade in our large cities for fire 
 protection. 
 
 Thus from earliest times may be traced a demand for some 
 means of raising water for man; for his herds and lands; for 
 the purpose of clearing his mines, and finally for his own per- 
 sonal convenience as well as the protection of his property. 
 The pumping machine has developed from a very crude origin, 
 it is true, yet its earliest types were so effective that they may 
 be found in use to-day although in modified form. 
 
 The present work will not consider the sources of water 
 supply, the gauging of the flow, the distribution of water, its 
 analysis or purification, or any of the problems of hydraulics 
 save those which are concerned with the raising of water from 
 one level to another. 
 
 Two of the earliest forms of pumps are the Shadoof and the 
 Noria, the former being common in Egypt, while the latter is 
 found in China and along the Euphrates as well as on the Nile. 
 The first of these is the ordinary well sweep seen on many old 
 farms in this country. A leather, earthen, or woven bucket 
 is attached to an arm by means of ropes or tree branches and 
 ropes. This arm is tied to a crossbeam supported in crotches 
 of tree trunks planted in the ground at the edge of some river 
 or well. The arm supporting the bucket is counterweighted 
 by a stone or a mud ball, so that there will be practically no 
 
HISTORICAL DEVELOPMENT 
 
 FIG. i. Shadoof. 
 
 FIG. 2. Shadoofs in Series. 
 
PUMPING MACHINERY 
 
 weight to lift. A man then pulls on the bucket support, 
 putting the bucket beneath the water, and then allows the 
 counterweight to lift it to the proper level, where he empties 
 
 the water into the canal or 
 basin. From the canal it 
 flows to the land which is 
 to be irrigated. At times 
 a series of these shadoofs is 
 placed in line (Fig. 2), each 
 shadoof raising the water a 
 portion of the total lift which 
 would be too great for any 
 one machine. The shadoof 
 is probably the oldest ap- 
 paratus for raising water, 
 although the simple bucket 
 attached to a rope (Fig. 3) 
 must have been used in 
 early times. Wilson states 
 that in India one or two 
 men operate a Paecottah (the 
 southern Indian name for 
 the Egyptian shadoof), and 
 lift from 1000 to 3000 cubic 
 feet in from six to eight 
 hours. The lift varies from 
 
 FIG. 3. Bucket. 
 
 5 to 12 feet. This apparatus is known also as the Swape 
 and in upper India it is called the Ldt. Buckley states that 
 the experiments of Wilson showed that two men could lift 
 57,600 cubic feet through a height of one foot in ten hours, 
 while one man could lift 33,000 cubic feet. 
 
 The Noria (Fig. 4) is a machine by which the water of a 
 stream is raised in buckets attached to a wheel, the wheel being 
 moved by the stream or in some cases by animal power. The 
 Chinese claim to have used these as early as one thousand years 
 before the Christian Era. One of them is described as con- 
 sisting of eighteen or twenty arms with paddles, to the periph- 
 
HISTORICAL DEVELOPMENT 5 
 
 ery of which is attached a number of buckets. At the lowest 
 limit -of the motion of the wheel these buckets dip below the 
 water and are filled. The motion of the wheel thus raises 
 the water to a higher level. In some wheels the impact of the 
 stream on the buckets is sufficient to drive the wheel, while 
 in other cases the wheel is provided with additional vanes to 
 drive it. A simple Chinese wheel (Fig. 4) is formed of bamboo. 
 The spokes of this wheel are attached to a central shaft and 
 cross each other at two-thirds the distance to the periphery, 
 
 FIG. 4. Noria. 
 
 where they are firmly lashed together, while at the end they 
 are fastened to end pieces. These act as the buckets, the 
 natural joint in the bamboo forming the bottom of the 
 bucket. The triangle formed by the spokes and the end 
 piece is filled with bamboo basket work, thus forming a 
 paddle to drive the noria. The end tubes are cut and 
 fastened at such an angle that when the vane is in a hori- 
 zontal plane the end tube is inclined upward at an angle of 
 twenty degrees so that the buckets will not discharge their 
 contents until they pass the level of the axle. The buckets 
 
6 PUMPING MACHINERY 
 
 finally discharge into a trough which conducts the water to a 
 canal or reservoir from which it can be used. The wheels are 
 from 20 to 40 feet in diameter. In the case of a wheel 20 feet 
 in diameter, containing twenty buckets 2 inches in diameter 
 and 4 feet long, 70,000 gallons were raised in twenty-four 
 hours when the wheel made four revolutions per minute. 
 
 These wheels were used throughout the East in Asia and 
 Egypt. Colonel Chesney of the British army reports several 
 of them in operation along the Euphrates as motors and pumps. 
 Some of these were arranged so that their axles could be raised 
 by means of stones in order that their depths of immersion 
 
 FIG. 5. Modern Nona. 
 
 would be the same at different stages of the river. Along the 
 river aqueducts are seen which carried the water away, and 
 walls or dams are found which served to raise the water at the 
 season of small flow and also served to direct the flow to 
 the openings in the walls where the norias were placed. At 
 the center of the river openings were left through which boats 
 could pass. 
 
 An American wheel of modern application (Fig. 5) shows 
 that the idea' of centuries ago is still of value as the basis of an 
 irrigating apparatus; as arranged here the full diameter is 
 almost available for the lift. Such a machine is simple and 
 very effective. 
 
HISTORICAL DEVELOPMENT 7 
 
 The name "noria" is also applied to a development of the 
 older ancient machine, in which the buckets are attached to a 
 chain or rope which, with its series of buckets or pitchers, is 
 placed over the wheel and extends down to the water at con- 
 siderable distance below. In this case, however, the wheel is 
 driven by animal power. This form is known as the Persian 
 Wheel (Fig. 6), while according to Buckley, the term Sakias 
 is used for it in certain other places. To operate the Persian 
 wheel a buffalo or camel is driven around a vertical axis, to an 
 arm of which it is yoked. The axle contains a cog wheel 
 (Figs. 6, 7, and 8) of crude form which engages with a second 
 cog wheel mounted on a horizontal axis ori which is placed the 
 
 FIG. 6. Sakias. 
 
 band or chain wheel for the support of the bucket chain. The 
 water is discharged from the buckets into a trough and flows 
 to whatever fields are to be irrigated. It is stated that two 
 bullocks will lift 2000 'cubic feet of water per day and the 
 heights of the lift will vary from 25 to 100 feet. The apparatus 
 is common in the East, and in 1890 the American consul at 
 Cairo reported 20,000 of these in use in the upper and lower 
 Nile valleys. 
 
 A form in which the buckets are attached to the wheel as 
 in the old noria, but in which the wheel is driven by animal 
 power as in the Persian wheel, is still used at times, and is 
 known as the Sakias. The term Taboot is applied to such an 
 apparatus when the buckets are replaced by bags made of 
 animal skins, 
 
PUMPING MACHINERY 
 
 FIG. 7. Persian Wheel. 
 
 FIG. 8. Persian Wheel. 
 
HISTORICAL DEVELOPMENT 
 
 I 
 
 ' 
 
 I 
 
 I 
 
 1 
 
 I 
 
 I 
 
 s 
 
 FIG. 9. Joseph's Well. 
 
 FIG. 10. Chain Pump. 
 
 I, . "i||IK,n Illlll' .ill. "illllli t-nlllh ,i!i.ol 
 
 FIG. n. American Form of Persian Wheel. 
 
10 
 
 PUMPING MACHINERY 
 
 - An old Persian wheel whose date of construction is not known 
 is located at Joseph's well at Cairo. This is shown in Fig. 9. 
 The total depth is almost 300 feet, and this depth is divided 
 into two lifts. The inclined passageway around the upper 
 shaft enables the animals used for driving to be taken down 
 to the lower driving wheel. The Romans built similar appar- 
 atus, calling them Roman Buckets. They were similar to 
 our coal and grain elevators. Another development of this 
 same apparatus consisted of a number of square discs mounted 
 on a chain and fitted into a square pipe like small pistons. 
 These chain discs, which have become the common chain 
 
 FIG. 12. Mental. 
 
 pump of to-day, have been used in China from time imme- 
 morial. Fig. 10 shows the construction of this. The pallets or 
 chaplets A A attached to the chain B are really loose-fitting 
 pistons and in the application shown there is no suction. 
 
 The Persian wheel has been applied in America (Fig. n); 
 with it a horse is used and the crude gearing and buckets are 
 replaced by metal parts. The principle, however, is the same 
 as that of the old wheel. With such a machine a horse is said 
 to lift 500 cubic feet per hour through a height of 20 feet. 
 
 The Mental, Katweh or Latha is a form of apparatus 
 used in Egypt and India to-day (Fig. 12). It consists of a 
 basket attached to two ropes in such a manner that two men 
 may swing it into a stream when swinging it in one direction, 
 
HISTORICAL DEVELOPMENT 
 
 11 
 
 while on the return swing the basket by a dextrous twist is 
 discharged. In this manner two men may raise 20,000 cubic 
 feet of water one foot in a day of ten hours. 
 
 The Doon (Fig. 13) consists of a long trough pivoted near 
 one enoT and balanced by a stone attached to an overhead lever. 
 The operator stands on the long counterbalanced end of the 
 trough, overcoming the counterbalance and causing the outer 
 end of . the trough to dip into the water. On stepping from 
 this, the weight lifts the trough and discharges the water into 
 the ditch above the stream or pond. 
 
 Another primitive apparatus was the Mot, which consisted 
 
 FIG. 13. Doon. 
 
 of a bag of skin attached to a rope. The bag was raised to 
 the surface by oxen, discharged, and then dropped back for 
 another supply. This apparatus is effective, as two bullocks 
 and a man can lift 79,000 cubic feet one foot in ten hours. 
 
 The double Zig-Zag Balance (Fig. 14) is shown by Mr. 
 H. M. Wilson, as an apparatus used in Asia Minor and Egypt. 
 Two men oscillate the frame and thus cause the water to flow 
 past wooden valves at the intersection of the steps, and the 
 water gradually passes from one step to the next. On each 
 swing, water is lifted and gradually travels upward to the dis- 
 cliarging trough. 
 
12 
 
 PUMPING MACHINERY 
 
 Fig. 15 illustrates a method shown on certain Egyptian 
 monuments. This should hardly be called a pumping appli- 
 ance, but it serves to illustrate the great importance of irriga- 
 tion when in ancient times such expensive methods were 
 employed. 
 
 This brings one near the beginning of the Christian Era, 
 and at this time there were several important applications of 
 
 FIG. 14. Zig-zag Balance. 
 
 mechanical principles to the raising of water. The use of 
 suction to raise water was applied and valves were added to 
 tubes carrying water while the movable diaphragm, partition 
 or piston closely fitting the water vessel or tube was employed. 
 The principle of the syphon was known at this time, and 
 the syringe, which employed these various principles, was in 
 use. 
 
 The principle of atmospheric pressure was not understood, 
 although it was employed in the suction of water in early 
 
HISTORICAL DEVELOPMENT 
 
 13 
 
 machines. It was not until the time of Torricelli, 1644, that 
 this was fully comprehended. 
 
 Fig. 16 is known as a Tympanum. It was employed 
 
 FIG. 15. Egyptian Irrigation. 
 
 in Egypt. A spiral tube is attached to the face of the wheel, 
 which is driven by the current. The end of the tube dips 
 
 FIG. 16. Tympanum. 
 
 beneath the stream, and as the wheel turns part of this water 
 is caught within the tube and is gradually lifted to the hub 
 of the wheel, where it is discharged into an irrigation flume. 
 
14 
 
 PUMPIXG MACHINERY 
 
 The Archimedean Screw (Fig. 17) is one of the first com- 
 binations of a movable diaphragm in a tube. By turning the 
 screw it is seen that the water is compelled to travel upward. 
 The helical surface A on the axis B is rotated only, but the 
 effect of it is that of axial movement. 
 
 The application of the principle of suction and one of the 
 first uses of heat to lift water is described by Hero of Alexandria 
 (dr. 120 B.C.) in his "Pneumatics." In this apparatus the 
 heat from the burning sacrifice on the altar A (Fig. 18) is used 
 to open the doors B of the temple. This was accomplished 
 
 FIG. 17. Archimedean Screw. 
 
 by the pressure produced on heating the air contained within 
 the hollow altar A. The air was led into the top of the sphere 
 C and its pressure drove the water into the bucket Z), the weight 
 of which acted on ropes attached to the trunnions E of the 
 doors. When the sacrifice was burned the air contracted, 
 producing a vacuum in C; the water was sucked back from D 
 and the doors were closed by the weight F. In this apparatus 
 it is important to notice the use of pressure to force water 
 from one vessel to another, and of a vacuum or suction to 
 draw water into a vessel. 
 
 From the description of Hero it is not possible to know 
 
HISTORICAL DEVELOPMENT 
 
 15 
 
 whether he invented any of the devices described, but it is 
 reasonable to suppose that many of them were the inventions 
 of Ctesibius, who made several mechanical inventions. The 
 
 FIG. 18. Temple Pump. 
 
 FIG. 19. Force Pump. 
 
 FIG. 20. Sun Fountain. 
 
 Force Pump is ascribed to him and to Hero. This pump 
 (Fig. 19) may have been suggested by the noria, in which 
 valves were used at the bottom of the buckets to facilitate 
 
16 
 
 PUMPING MACHINERY 
 
 filling. There were two single-acting pistons, moved up and 
 down by a cross rod or beam, and although the pumps were 
 single acting the stream was continuous, owing to the arrange- 
 ment of the pumps. It is to be noted that this was intended 
 for a fire pump, and is the earliest fire engine of which there is 
 record. 
 
 Another machine (Fig. 20) called a Fountain uses the 
 expansive force of heated air to operate it Water is driven 
 
 FIG. 21. Steam Pump. 
 
 FIG. 22. Fountain of de Caus. 
 
 from the vessel A through the syphon B by the expansion of 
 the air in the top of A when exposed to the sun's rays. This 
 water then falls into the box D and when the vessel A cools 
 water is drawn up from D through the pipe E. Although 
 Hero describes it in this manner, it is well to note that some 
 form of valve must be placed in C and E to have the action 
 proceed as described. 
 
 The use of steam pressure was suggested by Giovanni 
 Baptista della Porta in 1601 in an apparatus ' (Fig. 21) in which 
 
HISTORICAL DEVELOPMENT 
 
 17 
 
 steam is generated in a boiler A , from which it is discharged 
 into the top of a vessel B filled with water. A pipe C reaches 
 to the bottom of B, and when the steam pressure is exerted 
 on the water in B it is forced from the discharge pipe C. The 
 apparatus is the same as that of Hero, and della Porta suggests 
 in his book, published in 1601, that the vacuum caused by the 
 condensation of the steam be used for filling the vessel with 
 water. Della Porta should have shown another pipe leading 
 from the vessel B to the source of water supply. It is interest- 
 ing that della Porta suggested the separation of the boiler and 
 
 FIG. 23. Pump of da Vinci. 
 
 the pump cylinder, a very valuable point, although in 1615 
 Solomon de Caus proposed to combine the two by placing' the 
 equivalent of the vessel B on the fire and heating the water in 
 it for the purpose of driving the same by the pressure of 
 the steam. (Fig. 22.) 
 
 Before this time, in the fifteenth century, Leonardo da Vinci 
 suggested the pump shown in Fig. 23. The lantern wheel A 
 was turned by two cranks and this motion was transmitted 
 to a cylinder B through the toothed wheel C, and a helical 
 groove in the cylinder B caused a piston rod D to travel back 
 and forth by means of a pin on the rod which fitted into 
 
18 PUMPING MACHINERY 
 
 the helical groove. The movement of the solid piston of the 
 cylinder E would then suck water from the well through the 
 suction pipe F and discharge it through the discharge pipe 
 G. Valves must have been employed here, although not shown, 
 and the machine at once suggests the complete understanding 
 of the action of the piston within a cylinder. 
 
 In 1630 a patent was granted to David Ramsey e by the 
 English king which covered two points: " to raise water from 
 low pitts by fire," and " to raise water from low places and 
 mynes and coal pits, by a new waie never yet in use." 
 
 There is no record of what he did, but this seems to be one 
 of the first useful applications of heat to the raising of water. 
 
 While the use of heat for pumping was progressing, the 
 application of water power and animal power was being devel- 
 oped extensively. In England there was the installation of 
 the London Bridge Water Works in I58i ; by which the current 
 of the Thames was used, while in France Agostino Ramelli 
 published a book describing his inventions, including pumps. 
 This book was published in 1588, and the pumps described 
 were operated by men and animals as well as by the currents 
 of streams. According to an old account the application of 
 the lift pump was made in 1581, " when Peter Morrys was 
 given a grant by the Lord Mayor and Commonalty of the city 
 of London for the term of 500 years for supplying and con- 
 veyance of water into houses by pipes from an artificial force 
 from London Bridge on condition that he pay ten shillings 
 annually into the chamber of London." He was authorized 
 to use the first arch of London Bridge for this purpose. 
 
 In a paper before the American Water W T orks Association, 
 Mr. T. W. Yardley quotes from a description published in 1633 
 as follows: " The present supply of good water for London is 
 like to be very much enlarged by the great improvement of 
 the water works of Peter Morrys before mentioned, who, being 
 a Dutchman, in the twenty- third year of Queen Elizabeth, 
 first gave assurance of his skill in raising Thames water so 
 high as should supply the upper parts of London; for the Mayor 
 and Aldermen came down to observe the experiment^ and 
 
HISTORICAL DEVELAPMENT 
 
 19 
 
 they saw him throw water over St. Magnus steeple, before which 
 time no such thing was known in England as the raising of 
 water." The success of this invention was such that Morrys 
 obtained additional grants for two other arches under London 
 Bridge. * The second grant was for 2000 years and was finally 
 secured by the New River Water Company. 
 
 In 1731 Henry Height on, the engineer, described the London 
 Bridge water works in the " Philosophical Transactions," and 
 accompanied his minute account by an engraving. It may be 
 that the pumping apparatus had changed 1 from the time of its 
 first installation, but that is not known. 
 
 There were three water wheels (Fig. 24.) at the time of this 
 
 FIG. 24. London Bridge Water Works. 
 
 description, each about 20 feet in diameter. These wheels A 
 were carried on heavy frames BB and were raised and lowered 
 with the tide. The frames B supporting the wheel A were 
 pivoted at the axles of the lantern wheels CC, so that although 
 the axle of the wheel A was raised and lowered the pin wheels 
 DD attached to A on its axle were always in contact with 
 the lantern wheels CC. The beams BB were supported at 
 their outer ends by chains, which were raised or lowered by the 
 cranks EE acting on a series of crown wheels and lantern 
 wheels. 
 
 The axles of the lantern wheels CC were connected by a 
 
20 PUMPING MACHINERY 
 
 coupling to crank shafts FF, each provided with four cranks, 
 which in turn were connected to a series of vibrating levers. 
 These levers were connected to pump rods at each end, although 
 in the figure one set of pumps is omitted for clearness. There 
 were sixteen pumps to the wheel with cranks arranged so that 
 four of them would work alternately. The pump cylinders 
 were 4 feet 9 inches long and 7 inches in diameter; they were 
 fixed to the top of an iron cistern which contained the proper 
 foot valves, while the discharge valves were placed in another 
 box. 
 
 According to the description of this pump the gears were 
 so arranged that the crank turned 2,\ times per turn of the 
 main wheel. With the several wheels used in 1733, containing 
 52 forcers or cylinders in all, 1954 hogsheads were pumped 
 per hour while the main wheels made six turns per minute. 
 This would occur with full tide, and at that rate 46,896 hogs- 
 heads would be pumped per twenty-four hours. 
 
 Beighton states that the amount of leakage or " slip " in 
 the pump amounted to from 20 to 25 per cent of its displace- 
 ment. The reasons for this he gives: ist, the amount of 
 valve lift will cause leakage; 2d, no leather can be made 
 strong enough for pistons. He states that loose leathers cause 
 leakage while tight leathers cause excessive friction. 
 
 Such pumps excite admiration when the amount of experi- 
 ence possessed at that day and the state of the art of both 
 machinists and millwrights is considered. 
 
 While the works of Peter Morrys were being constructed 
 in England, a book appeared in Paris, 1588, by Captain Agos- 
 tino Ramelli, an Italian engineer. An account of this book on 
 the various machines of Ramelli is given by Mr. W. F. Durfee 
 in Gassier' s Magazine for June, 1895. Among the machines 
 for pumping water given by Mr. Durfee three have been taken 
 to illustrate methods described by Ramelli, and also to illus- 
 trate the state of the art at this time. A number of the machines 
 show that Ramelli was familiar with the action of the piston 
 and the rotary engine as well as with the principles of gearing 
 to transmit power between shafts at an angle and not in the 
 
HISTORICAL DEVELOPMENT 21 
 
 same plane; and for the purpose of obtaining reciprocating 
 motion from continuous rotary motion. 
 
 In Fig. 25 a method is shown for draining an excavation 
 or site beside a flowing stream. The wheel A is driven by the 
 stream. Attached to its axle or shaft B is a series of cranks 
 moving a set of levers CC, which cause the two shafts D to 
 
 FIG. 25. Ramelli's Pump. 
 
 oscillate. Each of the shafts DD has four arms which serve 
 to drive the pump rods of four submerged pumps. 
 
 It is to be noted that in both this machine and that of Peter 
 Morrys the crank and connecting rod are in evidence long 
 before the time of James Watt. Here are also seen successful 
 methods of securing reciprocating motion from rotary motion. 
 
 The wheel A of Fig. 26 was turned by a man walking on the 
 inside, and in this the crank B gave a reciprocating motion to 
 
22 
 
 P UM PI NG MA CHI N ER Y 
 
 FIG. 27. Ramelli's Rotary Pump. 
 
HISTORICAL DEVELOPMENT 
 
 23 
 
 the chain CC, which pulled in each direction on the arm E, 
 driving the arms DD up and down and pumping on each stroke, 
 one pump F forcing water on the up stroke of E and the other 
 on the down stroke. 
 
 This not only shows the application of the crank and con- 
 necting rod, but there is much ingenuity displayed in arranging 
 a chain to operate on both strokes. 
 
 The rotary pump (Fig. 27 ) of Ramelli is deserving of partic- 
 ular notice as a pump and also for the ingenious arrangement 
 of its gears. For some reason the chain was not carried down 
 to the axle of the pump. It may have been to keep this out 
 of the water or for the purpose of getting a higher speed. The 
 pin and lantern wheels at A and the crown and lantern wheels 
 at B serve to increase the number of 
 rotations, while the spiral gears at C 
 serve only to change the direction of 
 motion. 
 
 In 1628, two years before Ramseye 
 secured his patent from King Charles 
 for the use of fire, Edward Somerset, 
 Marquis of Worcester, is thought to 
 have installed an apparatus for raising 
 water at Raglan Castle. He did not 
 secure a patent on this until 1663, how- 
 ever, and there were no drawings nor 
 even a mcdel with his patent. From 
 the description of the apparatus in his 
 patent and from the grooves in the 
 walls of Raglan Castle, an idea of the 
 construction and operation of the ma- 
 chine may be formed, although it is not FlG 28 ._ WorCester s Pump . 
 certain that this is correct in every detail. 
 
 Two vessels A A (Fig. 28) are connected to a boiler through 
 a valve B. In the figure the connection to the boiler is made 
 to the right-hand vessel and the steam from the boiler presses 
 on top of the water in A and forces it out through the valve C. 
 During this action the right-hand foot valv D is held down on 
 
24 PUMPING MACHINERY 
 
 its seat. The water is driven to a height corresponding to the 
 steam pressure, and for that reason high steam pressure was 
 required for great heads. While the right-hand vessel A is 
 discharging, the condensation of steam in the left-hand vessel 
 produces a vacuum in the vessel and water is drawn up through 
 the foot valve D. 
 
 On the reversal of the valves B and C the operation just 
 described is repeated with opposite vessels and so the operation 
 became continuous. 
 
 Worcester introduced this for the water supply at Vauxhall 
 near the old city of London. Upon examination it is seen that 
 apart from the useful application of the invention there is 
 nothing new in this apparatus, as the elements are all found in 
 della Porta's work. 
 
 In 1683 Sir Samuel Moreland, the master mechanic of the 
 laboratory of Charles II, published a book from Paris on " The 
 Elevation of Water by All Sorts of Machines." He had been 
 sent to Paris by Charles on business relating to the water works 
 which the king had erected. In his book Moreland speaks 
 of the expansion of steam and its pressure, showing that a good 
 idea of the pressure and volume of saturated steam was common 
 in that day. He also refers to the duty of his pumping engines, 
 using the term as it is used to-day the amount of work per 
 hundred weight of coal. The question of the application of 
 steam to the raising of water was one which not only occupied 
 Moreland, but many other mechanicians, on account of the 
 difficulty which was then experienced in clearing the shafts 
 of the English mines from the vast quantities of water which 
 collected therein from the underground flow. Many mines 
 had to be abandoned because of the cost of carrying the mines 
 deeper when this expense of draining was too great. The 
 work was mostly done by animal power and the cost was rather 
 startling when compared with the steam pump of even the 
 early days. 
 
 It is to be noted that Moreland introduced and invented 
 the plunger type of pump in 1675. This type of pump was 
 one in which an enlarged end of a rod was forced into a cham- 
 
HISTORICAL DEVELOPMENT 
 
 25 
 
 her, displacing the water. This was followed shortly afterward 
 by a bucket pump, in which the water passed through valves 
 in the piston on the down stroke, as was the case in the later 
 engines of Simpson used at Thames-Ditton. 
 
 In 1698 Thomas Savery patented the design of an engine 
 
 FIG. 29. Savery 's Pump of 1702. 
 
 for freeing the mines of Cornwall from water. It was the first 
 steam apparatus applied to this kind of work. In 1699 he 
 submitted a model to the Royal Society of London and success- 
 ful experiments were made with it. A model fire engine was 
 exbihited before King William III at Hampton Court in 1698, 
 and the success of this led to the granting of the patent which 
 
26 PUMPING MACHINERY 
 
 read: " A grant to -Thomas Savery, Gent'L, of the sole exercise 
 of a new invention by him invented for raising of water and 
 occasioning motion to all sorts of mill works by the impellant 
 force of fire, which will be of great use for draining mines, 
 serving towTis with water and for the working of all sorts of 
 mills, when they have not. the benefit of water nor constant 
 winds; to hold for 14 years with usual clauses." 
 
 The apparatus is almost identical with that of Worcester, 
 and it is not known whether or not Savery knew of the earlier 
 work. The form of his pump of 1702, which is an improvement 
 on that of 1698, is shown in Fig. 29. It consists of a main 
 boiler A and the pumping chambers B and C. The water from 
 the tank G discharges through the valve / on one of the pumping 
 vessels, which condenses the steam in that vessel and the vacuum 
 produced thereby draws water through the suction pipe L 
 and foot valve E. When steam is then admitted from A through 
 the steam valve, the water is forced out through the valve D 
 into the discharge pipe. \Vhen the water is low in the boiler A, 
 the auxiliary boiler is filled from the main K through the pipe 
 H and valve /. This water is driven into the main boiler by 
 raising steam in M. The pipe connecting the two boilers is 
 carried to the bottom of the auxiliary boiler M and the steam 
 pressure on the water drives it over into the main boiler A, 
 provided the pressure in M is higher than that in A . 
 
 Savery seems to have been a wide-awake promoter and 
 advertiser, for he began a systematic scheme for making his 
 invention known. He explained it to the Royal Society and 
 presented them with a drawing and specifications which ap- 
 peared in their " Transactions," and he published a prospectus, 
 called " The Miner's Friend; or a Description of an Engine 
 to raise water by fire described and the manner of fixing it in 
 mines, with an account of several uses it is applicable to, and 
 an answer to the objections against it. London, 1702." This 
 invention of Savery was intended to do away with the great 
 expense in- the use of animal power to operate the pumps of 
 'the mines, which not only were expensive, but reached their 
 limit of capacity, so that workings could not be carried much 
 
HISTORICAL DEVELOPMENT 
 
 27 
 
 farther. In one mine 500 horses were employed in handling 
 the water. 
 
 Savery's improvements were the addition of surface con- 
 densation, the secondary boiler, and the use of water cocks 
 It must be remembered that this machine could not be used in 
 deep mines, as sufficient steam pressure could not be carried. 
 The joints of the sheets forming the boilers, pump chambers, 
 and other parts were fastened by solder, and at high tempera- 
 
 FIG. 30. Piston Pump of Papin. 
 
 tures these joints would not hold. Desaguliers reports such 
 accidents. 
 
 The Savery engine was built by others after his time by 
 such engineers as Desaguliers- and Gravesande. 
 
 While Savery was using the direct action of steam for raising 
 water, Papin, in Germany, was proposing the use of the piston 
 to separate the steam and the water, thus leading to the further 
 application of two pistons of different sizes, so that the steam 
 pressure did not have to be equal to the water pressure. In 
 1707 Papin proposed a water pump shown in Fig. 30. This 
 was brought out in Cassel, Germany. In this machine steam was 
 generated in the boiler A and was conducted through the cock 
 B to the cylinder C. It acted on the piston H and forced water 
 
28 PUMPING MACHINERY 
 
 through the discharge valve G into an air chamber, from which 
 it was delivered. Water was admitted through the valve K 
 after the piston had been driven to the end of its stroke. Papin 
 realized the cooling effect of the cylinder walls and suggested 
 the introduction of a piece of heated iron E to warm up the 
 piston before the introduction of the steam. The weighted 
 cover D, similar to Papin's safety valve, served for the intro- 
 duction of the iron. 
 
 Although Papin had in his pump the possibility of making 
 the steam pressure different from that of the water, it was 
 not to him 'that the honor for the first use of pistons of various 
 sizes is .due. Pistons for the forcing of water were explained by 
 Ramelli and Leonardo da Vinci, but these pistons were driven 
 through some mechanical medium from water power or man 
 power, and it was Thomas Newcomen, a blacksmith of Dart- 
 mouth, England, who first employed steam for the operation 
 of a water piston of different size from the steam piston. This 
 at once enabled mines to be sunk to a greater depth, as pump- 
 ing could now be done with greater efficiency. The Savery 
 engine was used to lift water 350 feet, but with the engine 
 invented by Newcomen the height was limited only by the 
 strength of the materials employed. 
 
 The engine of 1705 is shown in Fig. 31. Steam is generated 
 in boiler A at about atmospheric pressure, and as the piston of 
 the cylinder B is drawn up by the unbalanced weight of the 
 pump rod C steam is drawn into the cylinder from the boiler 
 through a valve at the top. When the piston reaches the top 
 of its stroke the valve is closed and water is sprayed into the 
 steam space from reservoir D, thus condensing the steam in 
 the cylinder and producing a vacuum. The air pressure on 
 top of the piston is then sufficient to force the piston down, 
 raising the pump rod C, and with it the pump piston at its 
 lower end. The water of condensation then falls through a 
 pipe G into a hot well, the height of the water in the drain being 
 fixed by the vacuum in the cylinder. The walking-beam E 
 served to connect the two piston rods by chains and sectors, 
 and the rod C could be of any length, as the greater part of its 
 
HISTORICAL DEVELOPMENT 
 
 29 
 
 weight and that of the column of water in the discharge pipe 
 could be balanced by counterweights, on a rod such as F on 
 either side of the center. Sufficient weight was left unbalanced 
 to cause the piston of cylinder B to rise when low-pressure 
 steam was admitted. 
 
 Since the engine was really driven by atmospheric pressure 
 and was operated by steam at practically no pressure above the 
 atmosphere, it was known as "the atmospheric engine." 
 
 FIG. 31. The Atmospheric Engine of Newcomen. 
 
 Indeed, so low was the steam pressure that the weight of the 
 safety valve H was sufficient to keep it closed. 
 
 The try cocks K served to indicate the quantity of water 
 in the boiler, and the cock above the cylinder B was used to 
 introduce water on the top of the piston to keep the edge of it 
 airtight. The use of a spray inside the cylinder to condense 
 the steam resulted from the discovery that leakage of water 
 around the piston condensed the steam more quickly and 
 
30 ' PUMPING MACHINERY 
 
 better than the original method of cooling the cylinder wall 
 by spraying it with water. 
 
 It is remarkable that a man of such little training as New- 
 comen appears to have had should have been able to combine 
 the necessary elements to make such a great machine. He did 
 not occupy a very high position in the town; however, he was a 
 good workman. When he and his colleague, John Galley, 
 wrote to Dr. Hooke, the famous physicist, in regard to their 
 use of a steam cylinder and piston to drive a separate pump, 
 lie advised against their plan. They were not to be put down, 
 however, and in 1705 they secured a patent. 
 
 The engine was then applied to drain mines and pump 
 water, but Newcomen and Galley did not have sufficient mathe- 
 matical knowledge to properly design their machines and they 
 had many failures, while their successes were accidents. 
 
 In 1713 Humphrey Potter, a boy who operated valves by 
 hand, arranged an automatic method of shutting off the steam 
 and water by the use of beams, strings, and catches, and made 
 the machine independent of an attendant. Henry Beighton 
 improved this in 1718 by substituting a vertical beam with 
 pins which struck the valve handle as it was raised and low- 
 ered by the walking beam. The vertical beam was known as 
 a " plug rod," " plug tree," or " plug frame." 
 
 These pumps were built for various purposes and were of 
 different dimensions; one in 1714 for Ansthorpe in Yorkshire 
 vras 23 inches in diameter and of a 6-foot stroke. It made 15 
 strokes per minute. One, described by Farey, was 8 inches in 
 diameter on the water side and 24 inches on the steam side. 
 The stroke was 60 inches and there were 15 strokes per minute. 
 This' pump lifted water 162 feet and the water column on the 
 piston weighed over 3500 pounds, which with 660 pounds of 
 unbalanced weight of the pump rod necessitated almost 5000 
 pounds on the piston. Such a pressure could be obtained 
 wiVh a vacuum of 21 inches of mercury. The pump devel- 
 oped 8 horse power. Another engine at Griff in Warwick- 
 shire cost ^"150 per year to operate it and displaced 500 horses 
 at an expense of 900 per year. The first Newcomen engine 
 
HISTORICAL DEVELOPMENT 31 
 
 was introduced on the continent in 1723 at Konigsberg, Hun- 
 gary. In 1735 cast iron was used in place of wrought iron for 
 the parts of the engine. 
 
 The engine was improved by many engineers. About 1769 
 John Smeaton, one of the most distinguished engineers of his 
 day, built several engines with greater stiokes than those usu- 
 ally employed. By using the proper diameters for his pistons 
 he was enabled to get much higher speed. Before building 
 pumps he experimentally determined the proper proportions 
 of the engine and so improved its construction. 
 
 Before the last quarter of the century these engines were 
 introduced to such an extent that the coal mines of Coventry 
 and Newcastle, the tin and copper mines of Cornwall, blowing 
 engines of the English and Scotch furnaces, the docks of Cron- 
 stadt in Russia, the lowlands of Holland and the salt mines 
 of Hungary bore testimony to the success of this invention. 
 The mines were carried to greater depths, the cost of pumping 
 water and air was reduced, and the supply of water to towns 
 was more certain. 
 
 One of Smeaton's Newcomen engines is shown in Fig. 32. 
 The figure shows the cylinder A connected with the boiler by 
 means of the steam pipe B. The boiler is placed in another 
 building. The valve C admits steam to the cylinder through 
 the admission pipe, which is carried above the bottom of the 
 cylinder so as to keep the injection water from entering it. 
 When the steam is admitted, the piston is driven or pulled 
 upward and when the top of its stroke is reached the upward 
 movement of the plug tree or working plug D acts on the 
 handles E through pins, turning the axle F, and the Y or 
 " tumbling bob " G is thus moved, shifting the rod H and 
 handle / and thus shutting off the steam. At the same time 
 the handle K opens the valve M , allowing water from the cis- 
 tern N to enter the cylinder through the spray head 0. This 
 immediately condenses the steam in the cylinder and the 
 vacuum produced permits the atmospheric pressure to drive 
 down the piston. The pins P and the springs Q stop the down- 
 ward motion at the proper point. At this lowest point the 
 
32 
 
 PUMPING MACHINERY 
 
 FIG. 32. Newcomen Engine of Smeaton. 
 
HISTORICAL DEVELOPMENT 33 
 
 drain pipe R is opened, allowing the condensed steam and con- 
 densing water to discharge into the hot well 5. The sniffing 
 valve T is then opened, allowing any air to escape. The feed 
 water of the boiler is taken from the hot well. The cistern 
 N is supplied by a jack-head pump 7, driven from a small 
 beam. Both beams get their motion through chains and 
 sectors, so that there is always a straight pull on the piston or 
 pump rods. The cock V admits water around the piston so 
 that the oakum packing of the piston is kept in proper con- 
 dition. The excess of this water is carried off through the drain 
 pipe to the hot well. The main beam is carried on sectors so 
 as to- reduce the friction. 
 
 To improve the efficiency, Smeaton covered the steam 
 side of the piston with planks and when the injection water 
 contained salts which would form a scale, the water in the , 
 hot well was not used for boiler feed, but the clear feed 
 water was passed through a coil of pipe immersed in the hot 
 well. 
 
 The next important step in the improvement of the steam 
 engine and pump was that of James Watt. The invention of 
 this man was one of the greatest events in the history of civili- 
 zation, as it not only improved the existing machines, but in 
 his specifications are contained the fundamental ideas of all 
 modern improvements to the steam engine. This event clearly 
 demonstrates what may be done by a careful and detailed 
 study of existing conditions. 
 
 The early history of Watt, who was born in 1736, is one with 
 which every engineer should be familiar. After many trials 
 and successes in the south of England he came back to Glasgow 
 and was employed to repair some of the apparatus belonging to 
 the university. In 1763 he repaired its model of the New- 
 comen engine, and this led to his making a study of the history 
 of the steam engine. He read the treatise of Desaguliers and 
 the works of others. In this study he learned of the accom- 
 plishment of Savery, Newcomen, and those who had preceded 
 them. Watt now began a series of experiments on the action 
 of the engine, determining quantitative relations between the 
 
34 PUMPING MACHINERY 
 
 amounts of steam, cooling water, and the heat of steam and 
 water. He discovered the great loss by radiation and absorption 
 and was led to use non-conducting materials for his vessels as well 
 as for the coverings of them. Not having experimental data 
 for this work, he made a series of original experiments on the 
 temperature and pressure of steam at whatever points he could 
 observe these and constructed a curve to give values at other 
 points. He determined the amount of steam used by the 
 Newcomen engine and the amount which should have been 
 used had none of the steam condensed; in addition he com- 
 pared the amount of injection water used in the engine with 
 the amount which should have been used. These experiments 
 showed him that three-fourths of the steam taken into the 
 cylinder was wasted and that the engine used four times as 
 much injection water as it should have used. His calculation 
 showed him at once that the method of producing the vacuum 
 was a poor one, as the cylinder had to be heated by steam at 
 each stroke so that.it could be filled, and then this heat was 
 removed again on the condensation of the steam. He then 
 tried to keep the cylinder hot, which necessitated that the 
 steam be taken from it for condensation. He invented the 
 independent condenser, which made his improvement complete. 
 After constructing a number of experimental machines and 
 after many vicissitudes he took out his patent in 1769 in con- 
 nection with Dr. Roebuck. 
 
 The patent of 1769 gave the following description: 
 " My method of lessening the consumption of steam, and 
 consequently fuel, in fire engines, consists in the following 
 principles: 
 
 " ist. That the vessel in which the powers of steam are to 
 be employed to work the engine, which is called ' the cylinder ' 
 in common fire engines, and which I call ' the steam vessel ' 
 must, during the whole time that the engine is at work, be 
 kept as hot as the steam which enters it; first, by inclosing 
 it in a case of wood, or any other materials that transmit heat 
 slowly; secondly, by surrounding it with steam or other heated 
 bodies; and thirdly, by suffering neither water nor other sub- 
 
HISTORICAL DEVELOPMENT 35 
 
 stances colder than the steam to enter or touch it during that 
 time. 
 
 " 2dly. In engines that are to be worked, wholly or partially 
 by condensation of steam, the steam is to be condensed in 
 vessels distinct from the steam vessel or cylinder, though 
 occasionally communicating with them. These vessels I call 
 condensers; and while the engines are working, those condensers 
 ought at least to be kept as cold as the air in the neighbor- 
 hood of the engines, by application of water or other cold 
 bodies. 
 
 " 3dly. Whatever air or other elastic vapor is not condensed 
 by the cold of the condenser, and may impede the working of 
 the engine, is to be drawn out of the steam vessels or condensers 
 by means of pumps, wrought by engines themselves, or other- 
 wise. 
 
 " 4thly. I intend in many cases to employ the expansive 
 force of steam to press on the pistons or whatever may be used 
 instead of them, in the same manner as the pressure of the 
 atmosphere is now employed in common fire engines. In cases 
 where cold water cannot be had in plenty, the engines may 
 be wrought by this force of steam only, by discharging the 
 steam into the open air after it has done its office. 
 
 " 5thly. Where motions round an axis are required, I make 
 the steam vessels in form of hollow rings or circular channels, 
 with proper inlets and outlets for the steam, mounted on hori- 
 zontal axles like the wheels of a water mill. Within them are 
 placed a number of valves that suffer any body to go round the 
 channel in one direction only. In these steam vessels are 
 placed weights, so fitted to them as to fill up a part or portion 
 of their channels, yet rendered capable of moving freely in 
 them by the means hereinafter mentioned or specified. When 
 the steam is admitted in these engines between these weights 
 and valves, it acts equally on both, so as to raise the weight on 
 one side of the wheel, and by the reaction of the valves suc- 
 cessively, to give a circular motion to the wheel, the valves 
 opening in the direction in which the weights are pressed, but 
 not in the contrary. As the vessel moves round, it is supplied 
 
36 PUMPING MACHINERY 
 
 with steam from the boiler, and that which has performed ats 
 office may either be discharged by means of condensers, or into 
 the open air. 
 
 " 6thly. I intend in some cases to apply a degree of cold 
 not capable of reducing the steam to water, but of contracting 
 it considerably, so that the engines shall be worked by the 
 alternate expansion and contraction of the steam. 
 
 " Lastly, instead of using water to render the piston or 
 other parts of the engine air- or steam-tight, I employ oils, 
 wax, resinous bodies, fat of animals, quicksilver, and other 
 metals in their fluid state." 
 
 It is to be noted that these claims covered the following 
 points : 
 
 ist. Lagging and jackets. 
 
 2d. Condensers. 
 
 3d. Air pumps. 
 
 4th. Expansive use of steam and the non-condensing engine. 
 
 5th. A rotary engine. 
 
 6th. Packings. 
 
 Mathew Boulton became the partner of James Watt, and 
 it is to him that much of the credit of the actual engine is due. 
 He was the owner of one of the most famous manufactories 
 of the day at Soho, near Birmingham, England. Here he 
 manufactured ornamental metal ware, gold- and silver-plated 
 ware and works of art, such as vases, statues, and bronzes. 
 His factories were noted for the good work done, and for the 
 broad policy of management. 
 
 Although the arrangement of the partnership was agreed 
 on in 1769, it was not until the spring of 1774 that Watt could 
 go to Birmingham. By November of that year their first 
 engine was built. The form of this is shown in Fig. 33. 
 
 Steam enters from the boiler by the pipe A and the valve 
 B passing to the steam jacket C. The condenser D is then 
 connected to the cylinder by the valve E, and the vacuum 
 thus produced in the space F causes the piston G to move 
 downward, and steam flows in above .the piston. When the 
 piston reaches the lower end of the stroke, the valves B and E 
 
HISTORICAL DEVELOPMENT 
 
 37 
 
 are closed while the valve H opens. This connects the spaces 
 on each side of the piston, and the weights of the pump rods 
 7 and / on the outer end of the beam K overbalance the weight 
 of the piston G and its rod, and so the piston is pulled rapidly 
 
 FIG. 33. Watt's Engine. 
 
 upward to the top of the cylinder, the steam above the piston 
 passing over to the lower side. 
 
 After closing H and opening B and E the operation is 
 repeated, and the air pump L removes the condensed steam 
 and the air from the surface condenser. The pump M supplies 
 the cooling water, and the pump N takes water from the hot 
 well and feeds the boiler. The pump rod of the air pump 
 contains the pins which operate the handles of the valves. 
 
38 PUMPING MACHINERY 
 
 An outlet P is used when the air is driven out from the cylinder 
 and the air pump before the engine is started. 
 
 There was much trouble in getting men and machines to 
 make these engines, and it may be said that the demand for 
 better work developed the machinist's trade of that day. Much 
 of the development was made by the firm of Boulton & Watt. 
 
 In the building, erection, and operation of their engines, 
 Boulton & Watt were led to take out patents for the following 
 articles: 
 
 A letter copy press. 
 
 A cloth dryer by the use of steam in copper rolls. 
 
 Five devices for getting rotary motion from reciprocating 
 motion without the use of a crank. 
 
 The expansive use of steam. 
 
 The double-acting engine. 
 
 The double-coupled engine. 
 
 A rotary engine. 
 
 A trunk engine. 
 
 A steam hammer. 
 
 Parallel motion. 
 
 The engine governor. 
 
 Mercury steam gauge. 
 
 Water gauge. 
 
 Steam-engine indicator. 
 
 Watt seemed to be one who could always find some 
 means of meeting every need: when it took too much time in 
 copying his reports to Boulton, he invented the copy press; 
 when it was necessary to study the action of steam in the 
 cylinder, he brought out his indicator. These numerous inven- 
 tions do not indicate that the firm was always successful. 
 Many times they were on the verge of bankruptcy, and had 
 their patent not been extended for twenty-four years, when it 
 first expired, their labor would have been in vain, because of 
 financial straits. The extension gave them the needed relief, 
 and at the expiration of the patent the firm was in good 
 condition. The story of the trials and successes of this firm 
 in the development of the engine is given in the biographies 
 
HISTORICAL DEVELOPMENT 
 
 39 
 
 of these two men, and the student is recommended to study 
 these most interesting books. 
 
 To gauge the power of his pumps, Watt introduced the 
 term " horse power," in so common use to-day. This, with the 
 term, " duty," gave those using pumps a method of comparing 
 
 FIG. 34. Hornblower's Pump. 
 
 the operations of various machines. An economical operation 
 was aimed at in all of this work, and in order to interest the 
 engine attendants, monthly prizes were given for the best 
 results in duty during the month. An operative machine had 
 been constructed, and it was now their object to improve the 
 efficiency of it. 
 
 Boulton died in 1809 and Watt in 1819, but before that 
 
40 PUMPING MACHINERY 
 
 time these men had given the business over to their sons. 
 They enjoyed the protection of their fathers' patents until 
 1801, during which time others were at work on the improve- 
 ment of the engine end of the pump. 
 
 Jonathan Hornblower patented a compound. engine in 1781, 
 although Watt claimed this invention. The engine is shown 
 in Fig. 34. Steam is admitted into the cylinder A from the 
 boiler, and from this cylinder it discharges into the cylinder B 
 and thence into the condenser C, shown in section. The plug- 
 tree rod D serves as the rod for the air pumps as well as to 
 operate the valve handles, which are not shown. The operation 
 of the engine is practically the same as that of the Watt engine. 
 To start, all air is driven out by allowing boiler steam to flow 
 through the cylinders and condenser, and thence through the 
 sniffing valve at E. The condenser will now condense steam 
 below the piston in B, and as the piston descends the valve 
 between A and B allows steam below the piston of A to expand 
 and press down on top of the piston of B. The steam from the 
 boiler enters on top of that of A. When the bottom of the 
 stroke is reached the boiler and condenser are cut off and 
 the top of each cylinder is connected to the lower portion. The 
 weight of the main pump rod F and the boiler feed pump G 
 pulls the pistons upward and the operation is repeated. This 
 was declared an infringement on the Watt patent. It did not 
 give a much higher duty than the best single-cylinder Watt 
 engines of the day, although the same idea as applied by Arthur 
 Wolf in 1804 with higher pressure steam gave duties of from 
 40,000,000 to 57,000,000 foot-pounds per bushel of coal, while 
 the Watt engine gave a little over 30,000,000. 
 
 The Bull Cornish pumping engine of 1798 was brought out 
 by William Bull and Richard Trevi thick. This type of engine 
 is the one which remained in use longer than any other, as it 
 was much simpler than that of Watt and had all of the ele- 
 ments of economy. It is shown in Fig. 35. 
 
 The steam cylinder A is carried on the timbers BB, extending 
 from the walls of the pump house in such a manner as to bring 
 the piston rod directly over the pump well. The piston rod 
 
HISTORICAL DEVELOPMENT 
 
 41 
 
 C is connected to the pump rod Z), and this, in turn, to the 
 counter balancing beam E by the rod K' ' . A pump rod G with 
 its pump .F is also shown. The counterweight H is to balance 
 as much of the weight as is thought necessary. The rod / 
 
 FIG. 35. Cornish Pumping Engine. 
 
 actuates the piston of the air pump and is used as a plug rod. 
 The valves of the air pump are in the base and the piston is 
 solid. The tank P surrounded the air pump and the pipe R; 
 it was filled with water. The pipe R acted as a condenser, but 
 
42 
 
 PUMPING MACHINERY 
 
 water was admitted through 0, making it really a jet condenser. 
 The pins 'on the plug rod K operated the rods leading to the 
 valves at L and M. In starting, the valves were operated 
 from the floor 5 and the air was driven out from the cylinder, 
 condenser, and air pump through the sniffing valve N, which 
 was water-sealed. 
 
 This engine was adjudged an infringement on the Watt 
 patents, which prevented its introduction for some time, but 
 it was afterward used exclusively in America and Europe 
 with many improvements. It had many advantages over the 
 other engine in its simplicity, but it was objectionable in that 
 it must be placed directly over the opening of the mines. It 
 was of great value further in that by properly selecting the 
 
 FIG. 36. Action of Cornish Engine. 
 
 mass of the parts the inertia of these could be used to permit 
 the expansion of the steam, although the water pressure or 
 resistance was constant. This may be shown by the diagrams 
 of Fig. 36. The steam pressure in excess of the resistance of 
 the water is used in accelerating the piston, piston rod, pump 
 rod, balancing beam, and counterbalance, and after passing the 
 point at which the steam pressure equals the water pressure 
 the inertia of the parts will supply the deficiency of energy on 
 being brought to rest after the steam pressure falls below the 
 resistance of the water. Neglecting friction the area abc will 
 equal the area bde. These areas will always be a measure 
 of the energy stored up in the moving parts and will be a func- 
 tion of the maximum velocity and the mass moved, so that 
 by changing the amount of mass the speed of the apparatus 
 
HISTORICAL DEVELOPMENT 43 
 
 could be altered. The use of heavy counterweights at times 
 required all of the steam pressure on the up stroke of the weight 
 to move them, while on the down .stroke the excess of counter- 
 weight acting with the steam was used to lift the water from 
 deep mines. 
 
 It is important to note the advantage of use of steam expan- 
 sively, as it is this which has made the modern pumping engine 
 with a fly wheel so economical. These early engines were 
 quite efficient, as is seen from the table below, which demon- 
 strates the great advantage of this Cornish engine. 
 
 DUTIES IN FOOT-POUNDS PER BUSHEL (94 POUNDS) OF 
 
 WELSH COAL 
 
 In 1769, the Newc 
 In 1772, the Newc 
 In 1778 to 1815, "V 
 In 1820, improved 
 In 1826, 
 In 1827, 
 In 18-28, 
 In 1830, 
 In 1839, 
 In 1850, 
 In 1827, highest d 
 In 1832, 
 In 1842. 
 
 omen engine 
 
 5,500,000 ft.-l 
 9,500,000 
 20,000,000 
 28,000,000 
 30,000,000 
 32,000,000 
 37,000,000 
 
 43,350,000 
 
 54,000,000 
 60,000,000 
 67,000,000 
 97,000,000 
 108.000. ooo 
 
 omen engine, improved by Smeaton 
 V^att engine 
 
 Cornish engine 
 
 uty, Consolidat< 
 Fowey Cons* 
 United Mine 
 
 (averaged 
 
 
 
 
 sd Mines 
 
 3 ls 
 
 s. . 
 
 This engine was developed into the beam engine and was 
 used for water works. Fig. 37 is a cut of a Cornish engine of 
 1840 for the East London Water Works, with a capacity of 
 6,000,000 gallons per twenty-four hours. This Tepresented the 
 best engine of the day. 
 
 While the Cornish engine was being used for pumping water 
 from mines and for the water supply of cities, another form of 
 pump was sucessfully operated in 1830 in New York by a Mr. 
 McCarty. This was the centrifugal pump. It was a pump which 
 had been known for a long time, as Euler discussed its theory 
 in a paper in 1754. According to one author the invention of 
 it is due to Denys Papin in 1689, who took his idea from Johann 
 Jordan. Jordan designed a centrifugal pump in 1680. Demour 
 in 1730 invented the equivalent of a centrifugal pump. It 
 
44 
 
 PUMPING MACHINERY 
 
 consisted of a tube, Fig. 38, mounted on a vertical axis so that 
 the lower end entered the water to be raised near the axis. 
 On turning this the centrifugal force overcame the effect 'of 
 gravity and the water rose. In 1818, a few years before 
 
 FIG. 37. East London Water Works. 
 
 McCarty's work, a centrifugal pump was designed in Boston 
 and known as the Massachusetts pump. It was a successful 
 machine. Fig. 39 shows the general arrangement of the Boston 
 pump. The vanes were parallel to radial lines and removed 
 
HISTORICAL DEVELOPMENT 
 
 45 
 
 several inches from them. They were placed on each side 
 of a disc, and this runner revolved within a casing. 
 
 After McCarty the improve- 
 ment of this form of pump was 
 undertaken by Blake and An- 
 drews in this country in 1831 
 and 1839, respectively, and by 
 Appold, Thompson, and Gwynne 
 in England. The original blades 
 of the Massachusetts pump were 
 radial, but those of Andrews in 
 1846 were curved, as shown in 
 
 Fig. 4 0. This pump had the FlG> 38 ._ Dera0 ur's Centrifugal Tump. 
 
 vanes held between two discs. 
 
 This patent was bought by John Gwynne for England, and his 
 firm began the manufacture of these pumps. The development 
 of the centrifugal pump in England is closely connected with 
 
 FIG. 39. The Boston Pump. 
 
 the history of this firm. They were the constructors of the 
 most notable installations for many years. 
 
 In 1848 Lloyd took out a patent for a centrifugal fan, and 
 Appold began the manufacture of it and applied it to the 
 
46 
 
 PUMPING MACHINERY 
 
 lifting of water. In 1851 he exhibited this and showed its 
 practicability. The tests of the pump with the curved vanes 
 showed it to be about three times as efficient as that with 
 straight arms. The advantage of this pump is its ability to 
 lift large quantities of water considering the space occupied 
 by the machine, and its ability to pump small solid particles 
 without clogging. It was originally thought that this was only 
 
 FIG. 40. Andrews Pump. 
 
 FIG. 41. Serviere's Rotary Pump. 
 
 applicable to low lifts, but to-day pumps of this form are used 
 for lifts of several hundred feet. 
 
 Another old form of pump in which rotary motion of the 
 parts is utilized is shown in Fig. 41. This is the rotary pump, 
 and is an old invention found in the form of Fig. 41 among a 
 collection of models made by Serviere, a Frenchman, born in 
 I 593- This is one of the best forms of this type, as will be 
 seen later, when the rotary pump will be examined in detail. 
 
HISTORICAL DEVELOPMENT 47 
 
 Since the action of this pump is positive, it may be used against 
 
 FIG. '42. Ramelli's Rotary Pump. 
 
 high heads, although leakage may be excessive, due to the wear 
 which occurs in the parts. 
 Some claim that this was 
 the invention of Pappen- 
 heim, a German, who lived 
 in the seventeenth century. 
 Ramelli, whose publica- 
 tion of 1588 illustrates a 
 rotary pump shown pre- 
 viously in Fig. 27, uses a 
 slightly different form from 
 that of Serviere. In his 
 pump, Fig. 42, a series of 
 flat pistons are driven by 
 a rotating cylinder which is 
 placed eccentrically within 
 the outer casing. These 
 flat plates are held out by 
 
 FIG. 43. Sixteenth Century Rotary Pumps. 
 
 springs, as shown in the 
 
 figure, and the rotation of the inner cylinder forces the watei 
 
 through the machine, 
 
 
PUMPING MACHINERY 
 
 FIG. 44. Watt's Rotary Pump. 
 
 FIG. 45. Eve's Pump. 
 
HISTORICAL DEVELOPMENT 49 
 
 Another form of rotary pump of the sixteenth century is 
 given in Fig. 43. As seen in the figure the space A is being 
 filled through an opening below the water level while the space 
 B, which is closed by the sliding partition C, is being discharged. 
 The sliding partition C extends from one side of the casing to 
 the other, and slides through the stuffing box. After rising 
 to the highest point it drops by its weight, which is made suffi- 
 cient to overcome the friction of the s tuning box. The stuffing 
 box shows the form used at that day. The friction of this 
 machine was very great. 
 
 A form of rotative pump similar to that patented by Watt 
 in 1782 as a rotative engine is shown in Fig. 44. The operation 
 is clear from the figure; the flap or butment A serves to divide 
 the two sides of the pump and when the projecting piece or 
 piston strikes the butment it swings on its pivot. The move- 
 ment of this butment is controlled by a heavy spring, or by 
 rods and cams, so that it is held against the water pressure from 
 force main, moving only at the proper time. This scheme was 
 altered in 1825 by J. Eve in that he substituted a revolving 
 cylinder for the pivoted butment and inserted three moving 
 pistons for the one. The small drum was driven by gearing 
 from the main shaft at three times the revolutions of the main 
 shaft, as it was one- third the size of the main drum. This is 
 shown in Fig. 45. 
 
 In 1805 John Trotter introduced a different form, Fig. 46, 
 in which a plate was driven in such a manner as to touch two 
 fixed concentric drums, its position being radial. The operation 
 of the machine is evident from the figure. There is some 
 chance for leakage after the piston crosses the discharge pipe 
 and before it crosses the suction pipe, so that it is really neces- 
 sary to have more than one piston. Fig. 47 illustrates another 
 form of this used for water, although it had been used in 1790 
 for a steam engine. 
 
 From these earlier forms a number of new rotaries were 
 developed which finally became a variation of the older form 
 of Serviere, as will be seen in the next chapter. 
 
 Another old type to be mentioned is the reciprocating 
 
50 
 
 PUMPING MACHINERY 
 
 ..FiG. 46. Trotter's Rotary Pump. 
 
 FIG. 47. Four-Bladed Rotary Pump. 
 
HISTORICAL DEVELOPMENT 
 
 51 
 
 rotary form, Fig. 48. The figure shows the operation of the 
 pump, and a further description is unnecessary. The objection 
 to these and the rotary pumps is the fact that it is very difficult 
 to keep their pistons tight. The rotary pumps have the great 
 advantage that the flow 
 of water is always in the 
 same direction through the 
 pump. 
 
 A pump somewhat al- 
 lied to the rotary is the 
 screw pump, Fig. 49. The 
 pump illustrated was the 
 invention of Revillion, and 
 was patented in Paris in 
 1830. It consisted of a 
 right- and a left-handed 
 screw meshing together, 
 being driven in opposite 
 directions at the proper 
 speed by means of gears 
 AB. The point of one screw 
 touches the root of the 
 other, and thus incloses a 
 definite volume of water 
 between the screw and the 
 walls of the pump cham- FlG " 48 -Reciproca^ng Pump. 
 
 ber, which travels upward as the screw rotates. At the end 
 of the travel this water is forced out at the center. 
 
 In passing, it is well to note that the use of pumps for the 
 extinguishing of fires had been common from the earliest times, 
 Fig. 19 being one described by Hero. Until about 1840, how- 
 ever, these were all driven by hand power. Fig. 50 shows a 
 French engine of 1829. This was hauled to the fire by the fire 
 company, which in America was a very important social organ- 
 ization during the first half of the urban history of the last 
 century. 
 
 The first steam-driven fire engine of note in this country 
 
52 
 
 PUMPING MACHINERY 
 
 was one planned by Captain John Ericsson, in a competition 
 for a prize offered by the Mechanics Institute of New York, 
 
 n 
 
 FIG. 49. Screw Ptfmp of Revillion. 
 
 FIG. 50. French Fire Pump of 1829. 
 
 in 1840, although in 1830 Braithwaite and Ericsson brought 
 out a steam fire engine in London. The pump developed 
 
HISTORICAL DEVELOPMENT 53 
 
 6 H.P., and pumped 150 gallons per minute a distance of 80 
 or go feet. It was drawn by horses and practically was the 
 same as that designed by Ericsson for New York. Before this 
 time stationary steam fire" pumps were used. 
 
 Fig. 51 gives a view of the Ericsson engine. The boiler 
 was of the locomotive type, the barrel A being connected to 
 
 FIG. 51. Steam Fire Pump of 1840. 
 
 the firebox B. The steam pipe C supplied the cylinder D 
 with steam. The water cylinder was in line with the steam 
 cylinder and above it was placed an air chamber. The smoke 
 pipe from the boiler was carried around the air cylinder in 
 the form of a serpent. On the front of the engine was a blowing 
 box which could be worked by the cross-head of the steam 
 engine, or by hand or by a crank attached to the wheels of the 
 engine. The latter arrangement served to force the fire within 
 the firebox when the engine was on its way to a conflagration. 
 This is a forerunner of the modern fire engine, and it is note- 
 worthy that the design was most thoughtfully and carefully 
 made. 
 
 The first record of the hydraulic ram was that of Mr. White- 
 hurst of Derby, England. In 1772 he erected a machine shown 
 in Fig. 52. A was a spring or 'reservoir supplying the cock C 
 
54 
 
 PUMPING MACHINERY 
 
 through the pipe B, which was about 600 feet long and H 
 inches in diameter. The cock was 16 feet below the level in 
 A, and on closing this after drawing water the momentum of 
 this long column of water was sufficient to force the water 
 
 FIG. 52. Hydraulic Ram of Whitehursi. 
 
 into air chamber D, which was under the pressure of the higher 
 reservoir E. 
 
 It was Montgolfier in 1796 who independently invented 
 the same scheme, but made it of more value by using a device 
 
 FIG. 5*3. Ram cf Montgclfkr. 
 
 which worked, continuously and automatically, in place of the 
 cock C. His scheme is sho\vn in Fig. 53. The w.ater descends 
 from the source of supply through A, escaping at B. When 
 the water has acquired a certain velocity it raises the ball and 
 closes the opening at B. The momentum of the water causes 
 an increase of pressure, and this is finally sufficient to open the 
 
HISTORICAL DEVELOPMENT 
 
 55 
 
 valve in C against the high pressure of the discharge. The 
 valve at B may be of the disc form, but opening downward; 
 the principle, however, is the same in all cases. 
 
 The development of the ram in the years which follow the 
 work of Montgolfier consists in improvement in details, the 
 
 FIG. 54. Flash Wheels. 
 
 latest forms of the present day being quite similar to this 
 early type. 
 
 Scoop wheels or flash wheels, Fig. 54, were used from 
 early times. They were in reality water wheels turning back- 
 ward. These, as will be seen, have been used to some advan- 
 tage in later times. They were used extensively in Holland 
 about the time of the introduction of the steam engine. 
 
CHAPTER II 
 RECENT HISTORY 
 
 THE year 1840 marks an era in the history of pumping 
 machinery, for it was in that year that Henry R. Worthington 
 began his brilliant inventions, which have led to most of the 
 modern forms of steam pumps. 
 
 Before this lime the boiler-feed pump was usually driven 
 by the main engine through some extension from the piston 
 or pump rod or by an auxiliary rod from the walking beam. 
 Worthington was concerned in the design of a steamboat for 
 canal navigation. It happened that when the boat was stopped 
 at locks or by obstructions the attendants had to resort to the 
 use of hand pumps to keep the boiler properly filled. An 
 independent steam pump was thought necessary, and on Sep- 
 tember 7, 1841, the pump shown in Fig. 55 was patented. 
 
 In the illustration this pump is mounted on a base for 
 exhibition purposes, although it was originally bolted to the 
 side of the boiler setting at the front. Steam enters the cylinder 
 A through the pipe B, the steam being directed to either 
 end by a valve in the steam chest C. In the position shown, 
 the valve has just been moved so as to admit steam to the 
 right-hand end of the cylinder. This drives the piston, piston 
 rod, and plunger to the left. The plunger D drives the water 
 frpm the cylinder G through an ordinary conical valve in the 
 valve box E to the feed pipe F and from there to the boiler. 
 As the rod moves to the left the arm H, attached to the rod, 
 moves the tappet rod / to the left. Finally near the end of the 
 stroke the right tappet / strikes the lever K, pivoted at its 
 center, and forces it against the sloping 'top of the rod M. This 
 forces the rod M down against the pressure of a spring. After 
 the lever K passes over the point, the spring pressure suddenly 
 
 56 
 
RECENT HISTORY 
 
 57 
 
 forces the lever K over to the left away from the tappet. This 
 moves the arm N to the right, which controls the steam valve 
 through the axis of the arm AT so as to admit steam to the left- 
 hand end of the cylinder A. 
 
 The piston, piston rod, and plunger now move to the right, 
 
 FIG. 55. Original Worthington Pump. 
 
 and water is sucked into the cylinder G through a conical valve 
 at the bottom of E and the suction pipe 0. The left-hand tappet 
 drives the lever K to the right and finally the spring forces it 
 over, suddenly reversing the motion. The spring action was 
 necessary to properly reverse the steam valve, as a tappet 
 
 FIG. 56. Spring-Thrown Valve. 
 
 alone would permit the valve to cut off steam from each end 
 when the motion was slow. It is necessary to have the valve 
 reversed positively while the pump has positive motion, however 
 slow. 
 
 The admission of steam to this particular pump was con- 
 trolled by a float in the boiler so that when the water became 
 
58 
 
 PUMPING MACHINERY 
 
 low the pump would start automatically and continue until 
 the float would again cut oft the supply of steam. The pump 
 shown in the figure was in service for twenty-five years and 
 was finally bought back by the Worthingtons as a cherished 
 relic. 
 
 In 1844 Worthington used a helical spring in a casing A, 
 Fig. 56. The arm B pressed against a helical feather or pro- 
 jection on the casing, turning the casing against the action of 
 the spring. When the arm went beyond the end of the feather 
 the spring forced the casing over suddenly, thus moving the 
 valve at C by turning the valve rod, moving the valve across 
 the cylinder, perpendicular to the piston motion. 
 
 The water valves at D, in which E is the discharge space 
 and F the suction, are of the forms used in many pumps, of 
 
 FIG. 57. B Valve. 
 
 that day. The pistons, stuffing boxes and cylindrical parts are 
 quite similar to those of the present. 
 
 The B valve, Fig. 57, was invented for the purpose of 
 admitting steam to the right-hand end of the cylinder by a 
 motion to the right when steam is above the valve. In the 
 figure it is seen that when the valve is moved to the right, 
 the steam above the valve at A will enter the space B through 
 the space C and thus pass to the right-hand end of the cylinder 
 through D. At the same time the left-hand end E is con- 
 nected to the exhaust port G through the cavity F. This is 
 necessary when a slide valve is moved directly by the motion of 
 the piston, since the motion to the right moves the valve to the 
 right and with this movement steam is admitted to the right- 
 hand end, reversing the pump. 
 
RECENT HISTORY 
 
 59 
 
 The next improvement was a steam- thrown valve, Fig. 58. 
 This was in 1849. In this arrangement an auxiliary valve 
 rod not shown moved an auxiliary valve, admitting steam to 
 the right or left of the piston A, which forced the small cylinder 
 to the right or left, thus moving the main valve, which was 
 
 FIG. 58. Steam-Thrown Valve. 
 
 part of the small cylinder. The excessive friction from the steam 
 on top of the auxiliary piston led to the design of a balanced 
 valve which took steam in through the ordinary exhaust pas- 
 sage and used the so-called steam chest as an exhaust chest. . 
 
 FIG. 59. Positive Water and Steam Valve. 
 
 Fig. 59 shows a pump built about 1850 in which both the 
 water and steam valves were controlled by the movement of 
 the valve rod. This was in a measure a water dash pot for 
 the purpose of stopping the motion of the pistons. Worthing- 
 ton used this in some cases, but he did not advocate its general 
 use. 
 
60 , 
 
 PUMPING MACHINERY 
 
 In 1849 Worthington, with his partner, Wm. H. Baker, 
 patented a relief- valve motion. This scheme is shown in 
 Fig. 60, where radial water valves or clack valves are used. 
 The invention consisted of the use of two water passages A A 
 at each end of the water cylinder, so that when the water 
 piston uncovered the inner of these ports the pressure in front 
 of the piston was relieved suddenly and the steam in the steam 
 cylinder drove the piston to the extreme end, moving the main 
 slide valve B over by means of the arm C and the tappets on 
 the valve rod D. This same scheme was applied by cutting 
 grooves in the water-cylinder bore at each end. 
 
 The aim in all of these later pumps was to simplify the 
 
 FIG. 60. Relief Valve Motion. 
 
 mechanism so that the pump would do the work in a 
 positive manner and still be simple to care for and to 
 operate. 
 
 In 1850 Worthington made another improvement by sub- 
 stituting a number of small valves for the four large valves 
 used on pumps heretofore, and also employed a plunger and 
 ring in place of the piston of former pumps. The pump, Fig. 61, 
 was used on the steamer "Washington." There were thirty- 
 six of these small valves arranged on valve decks as indicated 
 in the figure. The large number of -small valves gave the 
 requisite amount of opening with a small lift and hence the 
 amount of leakage passing the valves when the pump was 
 
RECENT HISTORY 
 
 61 
 
62 
 
 PUMPING MACHINERY 
 
 reversed was greatly reduced. The valves were of rubber, 
 one-half inch thick. 
 
 The arrangement of the plunger and ring of Fig. 61 is 
 markedly different from the pistons of the former figures, and 
 Worthington adopted it for several reasons: 
 
 ist. It gives ample space above and below for the valves. 
 
 2d. The ample space around the plunger forms a subsiding 
 chamber where harmful materials may settle out of the way 
 of the plunger packing. 
 
 3d. The constant protrusion of the plunger tends to carry 
 foreign matter away from the packed joint. 
 
 FIG. 62. Steam End of Savannah Pump. 
 
 4th. The construction makes the removal or renovation of 
 the parts an easy matter. 
 
 5th. The deflection of the water currents is less than in 
 ordinary arrangements. The friction is very small. 
 
 It is to be noted in passing that this design simplified the 
 construction of the water end of the pump as far as the foundry- 
 work and the machine-shop work were concerned. The use of 
 a solid metal packing around the plunger as seen at A was an 
 innovation, but it has proved a success. It was made long 
 and with clear water the wear was not sufficient to cause exces- 
 sive leakage for some time. A steam-thrown valve was used 
 in this case. Worthington was not satisfied with this method 
 of valve operation and after his invention of the duplex pump 
 in 1859 ne rarely used it, although inventors such as Knowles, 
 Blake, and others continued to use it. 
 
RECENT HISTORY 63 
 
 In 1854 Worthington erected his first water-works pump for 
 trie city of Savannah. It was a compound pump of peculiar 
 design, in that the high-pressure cylinder was at the center of 
 the annular low-pressure cylinder. Another new feature of the 
 pump was the balancing of the steam valve by carrying part 
 of the pressure on the back of the valve by a piston supported 
 from the steam-chest cover as shown in Fig. 62. This figure 
 illustrates the use of this balancing method on an ordinary 
 cylinder. The Savannah engine was a success and a duplicate 
 of it was built for Cambridge, Mass., in 1856. This engine, 
 and a companion engine, built soon after, were of the following 
 dimensions : 
 
 High-pressure cylinder ....................... 12 inches diameter 
 
 Low-pressure cylinder ....................... 25 
 
 Plunger .................................... 14 " 
 
 Length of stroke .................. .......... 25 inches 
 
 Capacity ................................... 300,000 gals, per 24 hrs. 
 
 This engine was the best of its day, as was shown by tests 
 made for the Brooklyn Water Works in 1857 and 1859, when 
 that city was about to install new pumping engines. The 
 results of these tests as shown in the report of Mr. James P. 
 Kirkwood, chief engineer of the Brooklyn Water Works, were 
 as follows: 
 
 Date of Test. Name of Engine. pe^lb^ 
 
 April, 1857 Worthington engine, at Cambridge .............. 669,411 
 
 June, 1857 ........ . ..... 675,746 
 
 Jan., 1857 Cornish engine, at Jersey City ................. 628,233 
 
 July, 1857 Hartford crank engine ................ , ....... 646,994 
 
 July, 1857 " ...................... .. 614,426 
 
 Jan., 1860 Brooklyn new rotative engine ................. 601,407 
 
 June, 1856 Cornish engine, Philadelphia ................... 589,903 
 
 As a result of this the Worthington pump was selected 
 and a modification of the Cambridge pump design, known as 
 the duplex pump, was accepted; this was, however, abandoned 
 on account of trouble with the contractor and it was not until 
 1863 at Charlestown, Mass., that this important type was 
 installed in a water works. It will be remembered that the 
 term duty as used by Moreland and Watt meant the useful 
 
64 
 
 PUMPING MACHINERY 
 
RECENT HISTORY 65 
 
 work per bushel or hundred weight of coal, but in later years 
 the duty was figured as the useful work per 1000 pounds of 
 steam or per 1,000,000 British thermal units. 
 
 The date of 1859 is that of the invention of the duplex 
 pump, one of the simplest contrivances for operating the valves 
 of a pump, doing away with the complex arrangements used 
 heretofore for this work. This then became for many years the 
 standard type of pumping engine, as the Corliss engine became 
 the standard mill engine. 
 
 To make clear the operation of the duplex pump a modern 
 form of boiler-feed pump, Fig. 63, is illustrated. This con- 
 sists really of two pumps placed side by side. In this figure 
 the piston rod A of the front pump has just reached the right- 
 hand end of its stroke, and the bell-crank lever BC, which 
 extends from the front piston rod to the back valve rod has 
 moved the D slide valve of the back pump 'to the right, uncover- 
 ing the steam port of the left-hand end of the rear pump, causing 
 that pump to move to the right. This motion is transmitted 
 to the front valve stem through the reverse lever DE, which 
 connects the back piston rod to the front valve rod. The D 
 slide valve of the front pump is moved to the left, admitting 
 steam to the right-hand side of the cylinder and thus the piston 
 of the front pump moves to the left. This motion, in turn, 
 causes the back pump to move to the left and this then moves 
 the front valve so that the piston moves to the right, when the 
 operation is repeated. 
 
 The action of the pump may be represented by the diagram, 
 Fig. 64, in which vertical distance represents time and horizontal 
 distance represents motion of the pump. The solid line repre- 
 sents the front pump and the dotted one the rear pump. At 
 the end of each stroke there is a period of rest while the other 
 pump follows the stroke of the first one. This is accomplished, 
 as will be explained later, by the method of moving the valve 
 by the valve rod or its equivalent. 
 
 Not only is the claim .for a simpler valve gearing made 
 
 or this pump, but there should also be a steadier discharge 
 
 of water, because as one pump nears the end of its stroke the 
 
<56 
 
 PUMPING MACHINERY 
 
 other one discharges water to keep up the flow while the first 
 reverses. 
 
 This duplex pump was one in which, by the use of an addi- 
 tional unit, a simple form of valve gear was obtained. While 
 Worthington was inventing steam-thrown valves of various 
 forms previous to his invention of the duplex pump the same 
 
 Stroke 
 FIG. 04. Action of Duplex Pump. 
 
 problems were being investigated by others in this country 
 as well as abroad. It is not the intention of this book to follow 
 all of the different forms of simplex pumps invented from the 
 time of Worthington to the present, as his inventions mark 
 the era. Later a number of modern simplex mechanisms will 
 be examined in detail to illustrate what had been done and 
 
RECENT HISTORY 67 
 
 what forms have survived. It is necessary at this point, 
 however, to mention such names as Blake, Cameron, Knowles, 
 Earle, Cope, Maxwell, Marsh, Silver, Dean, Davison, and 
 Gordon in America; and Moreland, Thompson, Baummans, 
 Tyler, Clarkson, and Davies in Europe, as some of those who 
 had been working on this problem. Their work is of interest 
 in comparing the complicated manner of the earlier forms with 
 the simpler forms of to-day. 
 
 The introduction of the duplex pump, at least while the 
 patent lasted, caused considerable argument as to the relative 
 merits of the simplex and duplex forms, the simplex manufac- 
 turers claiming a lack of positive length of stroke for the duplex 
 against a full stroke of the simplex to offset the greater com- 
 plication. The proper adjustment of each pump, however, 
 will give a proper stroke. 
 
 Pumps of the duplex type were introduced into water 
 works in 1860, and in 1863 one of 5,000,000 gallons per twenty- 
 four hours was installed by Worthington at Charlestown, Mass., 
 and in 1871 a larger one for 19,000,000 gallons was built for 
 the city of Philadelphia. The earlier engines were single 
 expansion on the steam end. 
 
 The pumping engine of Mr. George Shields for the city of 
 Cincinnati, which was built in 1861, was one of the largest of 
 its day; it was to lift 9,600,000 gallons per twenty-four hours 
 against a head of 170 feet. It was direct acting and had no 
 balance beam. The steam cylinder was 100 inches in diameter 
 and had a stroke of 12 feet. The pump cylinder was 45 inches 
 in diameter. This engine was one of the unfortunate structures 
 of American engineering in that it cost the municipality many 
 times its original estimated figure; but to its credit it may be 
 said that twice it was the means of saving the city from a 
 water famine, when the other pumps gave out. The use of 
 vertical engines in which the steam and water cylinders were 
 placed over each other was quite common, as, for example, 
 in the Bull Cornish engines, but when it was desired to eliminate 
 the inertia bob weight to cut down the weight of the engine, 
 fly wheels were introduced. The fly wheels were driven from a 
 
68 
 
 P UMPI XC MA GHINER Y 
 
 beam in most cases, but in 1868 Richard Moreland, Jr., and 
 David Thompson invented the direct-acting engine shown in 
 Fig. 65. In this the cross-head was connected to the pump 
 plunger by two or four rods which spanned the crank shaft. 
 The cylinder was supported by A frames. The pump barrel 
 
 was bolted to the base of 
 the engine. The pump was 
 single acting and the general 
 arrangement was excellent and 
 simple. 
 
 This type o* engine was 
 installed in 1868 and 1876 at 
 the Eastbourne Water Works 
 in England. It proved to be 
 so successful in operation and 
 needed so little repairing after 
 being in service until 1881 
 that new engines for that 
 plant were made of the same 
 type. 
 
 The general arrangement 
 of steam end and pump for 
 this machine will be seen to 
 be quite similar to the modern 
 American pumps. The design 
 was good and one of the 
 simplest for the application of 
 the fly wheel to the pump. 
 
 About this time a unique 
 FIG. 65. Fly Wheel Pump. pum p was introduced by Bird- 
 
 sill Holly to care for his direct-pressure system, which dispensed 
 with a reservoir or standpipe, obtaining pressure direct from 
 the engine. He introduced the system in 1866 in Lockport, 
 N. Y., using a pump driven by a water wheel, but in 1871 he 
 used his quadruplex engine at Dunkirk, N. Y. The four pumps, 
 Fig. 66, were placed in tandem with the steam cylinders, which 
 were arranged in pairs, each acting on a crank. The two 
 
RECENT HISTORY 
 
 69 
 
 I 
 
70 
 
 PUMPING MACHINERY 
 
 cranks were arranged at 135 and the frame of the engine was 
 such that the center lines of the engines intersected at 90. 
 The figure shows the construction of the engine. 
 
 Such an engine was necessary for this system, as the engine 
 had to start from any position as soon as the drop, in pressure 
 in the mains moved the governor. This system saved the 
 expense of a reservoir, but the necessity of keeping the engine 
 under steam continually made the steam use and the labor 
 
 FIG. 67. Leavitt's Lynn Pump. 
 
 expense very high. The engine could be run with any number 
 of steam or water cylinders, and the engines could be used as 
 single-expansion cylinders or by exhausting from one to the 
 other three, as was done in 1874 at Rochester, the advantages 
 of compounding could be had. On account of the intermittent 
 action of these engines their duty was not better than that of 
 other pumps. A duty of 86,176,315 foot-pounds was obtained 
 on the 6,000,000 gallon pump at Buffalo, N. Y., in 1879. 
 
 The first important compound steam end for water works 
 in America was designed in 1872 by Frederick Graff, chief 
 
RECENT HISTORY 7 1 
 
 engineer of the water department of the city of Philadelphia. 
 Compound pumps were built by Simpson & Co. of England, 
 in 1848, although most of the pumping engines were of single- 
 cylinder type. 
 
 This Graff engine was followed in 1873 by an engine designed 
 by Mr. E. D. Leavitt, jr., which was quite similar to the Graff en- 
 gine in its theoretical operation, although different in its arrange- 
 ment. It was intended for the water works of Lynn, Mass. 
 
 As shown in Fig. 67, this had two inclined steam cylinders 
 and a pump which was arranged to discharge on each stroke, 
 although it was single acting on the suction side. This was 
 accomplished by the plunger attached above the pump bucket. 
 The supplementary pipe and valves connecting the two ends 
 of the pump cylinder were for the purpose of reducing the 
 friction. The valves were of large diameter and double beat, 
 that is, each valve had two discharging edges. This type of 
 pump end was first built in 1848 and was known as a Thames- 
 Ditton pump. The figure also shows the application of the 
 fly wheel to the pump for the purpose of using steam expan- 
 sively. The cylinders were steam jacketed and the steam 
 valves were of the gridiron type, driven by cams. The double- 
 acting air pump was driven from the beam. 
 
 This pump gave a duty of almost 104,000,000 foot-pounds 
 per 100 pounds of coal on a 52-hour test, when pumping about 
 5,060,000 gallons per day. The duty for its year's record was 
 about 75,000,000. The dimensions of the pump are given below: 
 
 Diameter of high-pressure cylinder 17 1 ins. 
 
 low-pressure cylinder 36 ' 
 
 high-pressure piston rods 3 ' 
 
 low-pressure piston rods 3 1 ' ' 
 
 air pump *. i T \ " 
 
 pump barrel 26.1 ins. 
 
 plunger 18^ 
 
 Length of stroke of steam and water pistons 7 ft. 
 
 air pump 44^ ins. 
 
 Distances between end centers of the beam 1 1 ft. 
 
 Weight of fly wheel 10.7 tons 
 
 beams 4.2 
 
 moving parts connected with beams 5 
 
 Length from pump to top of vertical pipe reservoir. J 9 4 ft. 
 
 Height of top of vertical pipe above bottom of well 163 .34 ft. 
 
72 
 
 PUMPING MACHINERY 
 
 This pump with its walking beam may be taken as an 
 example of many of the pumping engines to be found in Europe 
 and America, although the arrangement of the cylinders in an 
 inclined position was novel. Practically all of these machines, 
 with the exception of Graff's for the Philadelphia Water Works, 
 
 FIG. 68. English Pump of 1866. 
 
 were with single cylinders, although compound engines had 
 been known. To give an idea of one of the English pumps of 
 this period the West Middlesex pump of 1866 is shown in Fig. 
 68, as this illustrates the general arrangement of such pumps. 
 The steam cylinder was 80 X 120 inches, while the water cylinder 
 was 24^ X 120 inches. The pump was.double acting, although the 
 steam cylinder was single acting. The large acorn-shaped box 
 
kECRNT HISTORY 73 
 
 B on the pump rod directly beneath the parallel motion is the 
 balance box, which was loaded so as to force the water from 
 the pump against a head of 200 feet on the down stroke. During 
 this stroke the two ends of the steam cylinder were connected 
 through a 15-inch equilibrium pipe as in the old type of Watt 
 engine. The weighted box pulled the piston up until the piston 
 reached the top of the zo-foot stroke, when the plug rod R 
 reversed the valves, bringing live steam on. top of the piston 
 and connecting the lower end of the cylinder to the condenser 
 C through a ig-inch exhaust valve. The steam on the upper 
 side then drove the piston downward, the air pump A driven 
 'from the main beam maintaining the vacuum in the condenser. 
 This pump was provided with a safety device so that pressure 
 would be held in the discharge chamber in case the discharge 
 pipe should break. This pump is the equivalent of Watt's 
 original pump, although it was worked with steam at 40 pounds 
 pressure. The water pressure was 200 feet and the pump made 
 1 6^ strokes a minute. 
 
 Several important inventions of Worthington's should be 
 mentioned here: the dash relief valve, the rotary pump valve, 
 and the cross connection for compound pumps. 
 
 The dash relief valve was used on pumps to regulate the 
 length of the stroke and prevent pounding. The motion of the 
 piston could be stopped by cutting off the exhaust steam before 
 the end of the stroke and compressing the steam in the space 
 behind the pistons. This was done by having the piston ride 
 beyond the port to the exhaust passage, but in order to intro- 
 duce live steam into the cylinder in this cushion space it was 
 nedessary to use another passage beyond the one for the exhaust. 
 This made five steam passages in the cylinder, Fig. 69, the two 
 outer for steam and the three inner ones for exhaust. It is 
 seen from the figure that when the piston reaches the position 
 shown, the steam to the left of the piston, retained behind the 
 piston after passing B, has been compressed into the clearance 
 space and into the passage A. This acts as a cushion, and as 
 soon as the valve moves to the right high-pressure steam 
 enters and drives this piston to the right. If the piston is 
 
74 
 
 PUMPING MACHINERY 
 
 brought to rest too far from the end of its stroke, a valve D 
 is opened slightly between the passages A and B and some of 
 the steam compressed in A exhausts into B and then to C, 
 
 FIG. 69. Dash Relief Valve. 
 
 thus allowing the piston to travel farther. If D is opened too 
 much the piston will strike the cylinder head, causing pounding, 
 and it is then necessary to close the valve D slightly. The 
 
 FIG. 70. Rotary Steam Valve. 
 
 FIG. 71. Cross Connections. 
 
 application of the rotating valve was the application of the 
 common valve used by Corliss to the steam cylinder of the pump, 
 Fig. 70. This valve operates in the same manner as a slide 
 
RECENT HISTORY 
 
 75 
 
 valve, the steam or exhaust being conducted by the passage 
 A or the center of the valv6 casting B. The dotted passage 
 leading to the end controlled by a dash relief valve acts as the 
 
.76 
 
 P UMPI NG M A CHI NER Y 
 
 1 
 
 J 
 
 passage for the starting steam. 1 In such an arrangement the 
 piston starts slowly because all of the steam to start it would 
 have to pass the relief valve. 
 
kLCENT HISTORY 77 
 
 The cross connection of the low-pressure cylinders, Fig. 71, 
 is one which is used to supply steam for the demands of eithei 
 cylinders C or D from the exhaust of both of the high-pressure 
 cylinders A or B. Without the cross connection, the steam 
 from A would be used only in C and that of B in D, but by 
 this cross connection of the valve chests a steady motion is 
 obtained. 
 
 The duplex Worthington pump was also made in the com- 
 pound form, and one tested in 1873 at the Belmont Water Works 
 of Philadelphia gave a duty of 54,416,694 foot-pounds on 100 
 pounds of coal. This pump had steam cylinders of 29 and 
 50^ inches in diameter, while the water cylinder was 22^- inches 
 in diameter. The common stroke was 50 inches. The capacity 
 of the pump was 5,000,000. Fig. 72 shows the general form 
 of this pump. A similar one at Newark, N. J., gave 77,358,478 
 foot-pounds duty. 
 
 The difference between the duty of the Leavitt pumping 
 engine and that of the direct-acting duplex Worthington is 
 rather marked, arid where the engine is to run at full capacity 
 for considerable length of time these figures show conclusively 
 that such a high-duty engine, even though its cost is much 
 greater than the simpler pump, would prove to be a paying 
 investment. How r ever this may be, when the use of the 
 pump is quite intermittent, if the pump is too large for the 
 average consumption of water, the loss in starting up, together 
 with interest, depreciation, insurance, taxes, and repairs 
 on the more expensive machine might make the cost of 
 pumping greater than it would be with the less efficient ma- 
 chine. 
 
 This important fact was emphasized by Mr. Worthington 
 and undoubtedly accounts for the greater use of these duplex 
 machines in the years which immediately followed. 
 
 At the Centennial Exhibition of 1876 the Worthington 
 pump was used to supply all of the water. The pump 
 is shown in Figs. 72 and 73. The dimensions were as fol- 
 lows: 
 
78 PUMPING MACHINERY 
 
 Diameter high-pressure cylinder 29 ins. 
 
 low-pressure cylinder 50} 
 
 water plunger 22^ 
 
 Stroke 48 
 
 Air-pump diameter 29! 
 
 stroke 24 
 
 From the sectional view of the pump it is seen that the 
 steam and water cylinders are in tandem on each side of the 
 pump and that the valves are of the swinging-piston balanced 
 type. The cylinders have separate steam and exhaust pas- 
 sages and the outside cut shows the adjusting relief valves. 
 The cylinders were jacketed on the barrel and heads. It will 
 be noted that the valves are driven from bell-crank levers 
 attached to the walking beams of the air pumps by connecting 
 rods. The air-pump beams are worked from cross heads on the 
 piston rods through short connecting rods. It is evident that 
 the connections of the rods, beams, and levers are such that the 
 motion of one engine controls the valve of the other engine. 
 
 The construction of the jet condenser and the air pump is 
 seen in the picture. The water end of the pump shows the 
 small valves advocated by Worthing ton as well as the use of 
 the plunger and ring. The direct path, of the water through 
 the pump is one which diminishes the friction loss in the 
 machine. 
 
 The centennial year 1876 marks the installation of another 
 Leavitt pump at Lawrence, Mass. This pump was built on the 
 same lines as the Lynn pump, but it gave a duty of 117,550,800 
 foot-pounds per 100 pounds of coal, making a new record. 
 
 The Corliss engine of the Centennial Exposition was one of 
 the most remarkable features of the exhibition. In 1878 Mr. 
 George H. Corliss used hir engine in the construction of a pump 
 for Pawtucket, R. I. This pump, Fig. 74, consisted of two 
 steam cylinders, each one in tandem with a water cylinder. 
 The tail rod from the water end was attached by a connecting 
 rod to a vibrating lever, pivoted in the base and joined by a 
 rod to the crank of the engine. By use of the fly wheel in this 
 and the other fly-wheel engines, steam could be used expansively 
 without the use of heavy reciprocating parts. The fly wheel 
 
RECENT HISTORY 
 
 79 
 
 was mounted on the shaft carried by the bearings, which were 
 supported by the air chambers of the pump. A diagonal 
 brace was carried from the bearings to the main center ped- 
 estals. 
 
 The steam cylinders were 15 and 30 J inches in diameter, 
 30 inches in stroke, and were furnished with the Corliss valve 
 gear. They were steam jacketed on the barrels, heads, and 
 
 FIG. 74. Corliss Pump for Pawtucket. 
 
 valve boxes. There was a receiver between the cylinders, the 
 volume of which was equal to that of the low-pressure cylinder. 
 The drips from the jackets were delivered into the boiler feed, 
 heating it, while the drip from the receiver was passed through a 
 coil in the boiler flue and returned to the receiver in a super- 
 heated condition. The steam throttle valves were so connected 
 with the governor that they limited the speed to 52 revolu- 
 tions. 
 
80 
 
 PUMPING MACHINERY 
 
 FIG. 75. Moreland's Compound Steam End. 
 
 The exhaust from the low-pressure engine was carried to a 
 jet condenser placed beside the engine, while the air pump was 
 
RECENT HISTORY 81 
 
 placed below the main pedestal bearing. The pump was 
 10.52 inches in diameter with 2|-inch rods. It was to lift 
 3,000,000 gallons per twenty-four hours against 270 feet head. 
 It was of the inside-packed plunger type with 280 small 
 valves. 
 
 In August, 1878, this engine was tested for ten hours per 
 day for twelve days, with all coal and wood charged, and gave 
 a duty per 100 pounds of coal of 104,357,654 when delivering 
 3,060,000 gallons. On October 3, 1878, a 24-hour test gave 
 133,522,060 foot-pounds. This was a remarkable result, but 
 no more so than the annual duty of 123,656,000 foot-pounds 
 for the year 1888. 
 
 In 1889 Professor Denton tested this engine and obtained 
 a duty of 124,720,000 foot-pounds with hard coal and 127,- 
 350,000 foot-pounds with soft coal. These results were obtained 
 with boilers giving an equivalent evaporation of 8.88 pounds 
 of water per pound of coal for the hard coal and 9.35 
 pounds for the soft coal. A higher duty could be credited 
 to the engine if the usual amount of 10 pounds were evap- 
 orated. 
 
 The next important water-works engine was that used at 
 the Pettaconsett Water Works of Providence, R. I. This was 
 built by Corliss originally for Boston, but was not accepted for 
 some reason. Its test for six days in May, 1882, gave a duty 
 of 113,271,000 foot-pounds. 
 
 The inverted pump of Moreland and Thompson of 1868 
 proved to be so successful that in 1880 Moreland & Son 
 designed a compound pump for the Eastbourne Water Works 
 of England. This pump, Fig. 75, was of the close tandem com- 
 pound steam type with a single-acting piston pump attached 
 below. The plunger on the piston rod as shown in the earlier 
 pump, Fig. 65, was arranged to give a discharge on each stroke. 
 The general arrangement may be seen in the figure and reminds 
 one of the lines of recent pumps. The valve gearing was of the 
 Myer type and the high- and low-pressure cut off could be 
 adjusted separately. 
 
 The general dimensions were as follows: 
 
82 PUMPING MACHINERY 
 
 High-pressure diameter 20 ins. 
 
 Low-pressure diameter 38^ ' 
 
 Pump-piston diameter 20 
 
 Plunger diameter 15 " 
 
 Stroke 40 
 
 Fly-wheel, 15 tons 15 ft. 7 ins. diam. 
 
 This pump was tested on February 18 and 19, 1884, by 
 Messrs. Wallis and Borias, and showed a duty of 124,600,000 
 foot-pounds per 112 pounds of coal or 111,300,000 per 100 
 pounds of coal. It was running under a low head during this 
 test arid developed only 90.7 H.P., while it was designed to 
 develop 160 H.P. The following data are given from the 
 test: 
 
 Total revolutions in twelve hours 16,875 
 
 Average revolution per minute 23.44 
 
 Total amount of water, in gallons 759,456 
 
 Average total lift, in feet 243.94 
 
 steam pressure above atmosphere 69.15 
 
 vacuum 28.58 iris. 
 
 Horse-power water 77-97 
 
 Indicated horse power. ." 90 . 70 
 
 Mechanical efficiency 86% 
 
 Coal consumed in twelve hours 1666 Ibs. 
 
 Coal per horse-power hour of water i . 78 
 
 Coal per indicated horse-power hour i . 53 
 
 Duty per 112 Ibs. of coal in ft. Ibs 124,600,000 
 
 Duty per 100 Ibs. of coal in ft. Ibs 111,300,000 
 
 At this time the Holly Manufacturing Company brought out 
 a high-duty engine to replace the quadruplex engine. This 
 was designed by Mr. Harvey F. Gaskell, their superintendent. 
 It was a simpler machine, and was the first standard design of 
 high duty applicable to all forms of water works, the previous 
 high-duty engines being of special design. 
 
 The first of these engines (Figs. 76, 77, and 78) was installed 
 at Saratoga Springs, N. Y., in 1882. 
 
 From the figures it will be seen that the high-pressure 
 cylinder is directly over the low-pressure cylinder and that 
 the piston rods of these two cylinders are connected to the 
 extreme ends of a massive walking beam by short connecting 
 rods. This connection means that as one piston moves to 
 the right the other moves to the left, each being at the dead 
 
RECENT HISTORY 
 
 83 
 
 i 
 
 point at the same instant, and therefore it was of the Wolf 
 compound type. This arrangement made it possible to dis- 
 
PUMPING MACHINERY 
 
 charge directly from the high-pressure cylinder to the low- 
 pressure through the gridiron valve seen in Fig. 78. 
 
RECENT HISTORY 
 
 85 
 
 A connecting rod from the upper end of the beam joined 
 it to the crank of the fly-wheel shaft and by connecting the 
 two sides or engines by cranks at right angles the motion of 
 
 OUTSIDE. SECTIONAL. 
 
 -STEAM END ELEVATION, GASKILL PUMPING ENGINE. 
 
 FIG. 78. Section of Gaskell Engine. 
 
 the pumps was made steady. The valve gear is driven by 
 two longitudinal revolving shafts driven by bevel wheels, the 
 high-pressure steam valve being of the poppet type with an 
 
86 PUMPING MACHINERY 
 
 automatic relief while the other valves are of the gridiron 
 type. The condenser can be driven from a lever attached to 
 the trunnions of the walking beam, although in the Saratoga 
 engine a Buckley condenser was used. The water end con- 
 sisted of an inside packed plunger, the Worthington system 
 of a large number of small valves being used in it. In this 
 engine there were 672 water valves. There were two tests 
 made on this pumping engine; one in November, 1882, and 
 the other hi June, 1883; both of the results showed a duty of 
 about 113,000,000 foot-pounds, although for a short period a 
 duty of 127,000,000 was obtained. The dimensions of this 
 engine were: 
 
 Diameter high-pressure cylinder 21 ins. 
 
 low-pressure cylinder 42 ' 
 
 pump plunger 20 " 
 
 rods 4 " 
 
 St-x5ke 36 " 
 
 Capacity 5,000,000 gals. 
 
 There were subsequently many changes made in the 
 detaik of this engine, but the general plan was unaltered. 
 Charles T. Porter in his report of 1883 on the Saratoga engine 
 makes the following statement: " In the details through 
 which this general plan has been carried out, there are no 
 features which seem open to criticism, but, on the contrary, 
 all seem entitled to commendation. The construction is 
 thoroughly mechanical in every respect. The forces are trans- 
 mitted and the strains are resisted in the manner theoretically 
 the most correct. ... I do not think that the study of any 
 machine has ever given me a stronger feeling of confidence in 
 its durability." This shows the kind of designing done by 
 Mr. Gaskell and the kind of work done by the Holly com- 
 pany. 
 
 While these pumps were being built in America, the Euro- 
 pean practice seemed to hold to the Cornish steam pump 
 (Fig. 79), although in many cases horizontal fly-wheel engines 
 were connected through other rods or gears to horizontal or 
 vertical pump cylinders (Figs. 80 and 81), and in certain cases 
 the beam engine was built with a fly wheel (Figs. 82 and 83). 
 
RECENT HISTORY 
 
 87 
 
 FIG. 79. Cornish Engine of 1878. 
 
 FIG. 80. Horizontal Fly Wheel Pumping Engine. 
 
88 
 
 PUMPING MACHINERY 
 
 The pump shown in Fig. 83 was one constructed by Simpson 
 & Co., for the Lambeth Water Works of London. This pump 
 was of the type used for many years by this company and gave 
 unusually good results. The steam end consists of a com- 
 pound engine connected to the beam by a Watt parallel 
 motion. A fly wheel was used to replace the heavy bob 
 weight. 
 
 The technical press of that time does not record any remark- 
 
 
 FIG. 81. 1874 Fly Wheel Pump. 
 
 able duties for these machines, and for that reason these are not 
 described here, although the student who cares for examples is 
 referred to the bibliography at the end of the book. 
 
 An exception to the statement above must be mentioned 
 the pump used for the city of Paris at the St. Maur station. 
 Although installed in 1876, it was somewhat similar to one 
 installed in 1871. They were built by Farcot & Sons. The 
 steam cylinder was 39.4X70.8 inches, and the water cylinder 
 was 14,2X70.8 inches. The distance from the center of 
 the shaft to the center of the water cylinder was 50 feet 
 8 inches. The pump had a piston speed of 350 feet per 
 
RECENT HISTORY 
 
 89 
 
90 PUMPING MACHINERY 
 
 minute, which meant about 30 R.P.M., a high speed for that 
 day. At this speed it pumped 3,000,000 gallons in twenty- 
 four hours. To get such a high speed the valves on the suc- 
 
 FIG. 83. Simpson's Pumping Engine. 
 
 tion and discharge side were made so as to open simply and 
 give a free passage to the water. This may be seen in the 
 figure. The passages through the pump were made quite large 
 so that the water would have a free path. 
 
RECENT HISTORY 
 
 91 
 
92 
 
 PUMPING MACHINERY 
 
 The extended tail rod on the pump plunger was used to 
 pump air into the air chambers to keep them charged. 
 
 The pump was driven by a single-cylinder condensing 
 engine, with the air pump driven by a lever connected to the 
 
 FIG. 85. 
 
 cross-head, and so well was the machine designed that the 
 makers reported the development of an indicated horse power 
 on 12. i pounds of steam per hour or 1.54 pounds of ordi- 
 nary coal while 2.03 pounds _of coal were used per pump horse 
 
RECENT HISTORY 
 
 93 
 
 power per hour. This would mean a duty of 78,000,000 foot- 
 pounds per 100 pounds of coal. 
 
 The 6,ooo,ooo-gallon Reynolds pump of 1881 (Figs. 85 and 
 86) for Milwaukee shows a type similar to many of the European 
 pumps of this period in its general arrangement of cylinders, 
 except that there was an additional low-pressure cylinder 
 
 FIG. 86. Reynolds Milwaukee Engine. 
 
 added above the beam. The duty of this engine was 104,000,000 
 foot-pounds per 100 pounds coal. Following this came the 
 two 12,000,000 gallon Allegheny pumps of 1883 (Fig. 87), 
 in which there were three cylinders to each pump one high 
 pressure and two low pressure, The engine was an entirely 
 
94 
 
 PUMPING MACHINERY 
 
 new design for large engines, although small pumps had been 
 arranged in this general manner much earlier. The cranks 
 were placed at 120 and the number of working steam cylin- 
 
 FIG. 87. Allegheny Pump. 
 
 ders was reduced to three in place of four, so common in the 
 standard type of pump of the day. The arrangement of the 
 pump and steam cylinders is clear from the figure. The valve 
 

FIG. 89. Richardson' 
 
pie-expansion Engine. 
 
 (To face page 94) 
 

RECENT HISTORY 95 
 
 gear was of the Corliss type with fly ball and hand govern- 
 ing and receivers were used between the cylinders. The 
 
 FIG. 88. First Triple Expansion Pump. 
 
 cylinders were jacketed, tmt there was no heating coil in the 
 receiver. 
 
 The valves of the pumps were originally of the Cornish 
 
96 PUMPING MACHINERY 
 
 double-beat design, but these were later changed to small 
 valves supported on a cage or basket placed over the opening 
 for the Cornish valves. 
 
 The test of this engine gave 107,000,000 foot-pounds per 
 loop pounds of steam when tested in 1884 by Mr. C. A. Hague 
 and Professor David M. Green. It represented a type of 
 engine of great value for small floor space, as was the case of 
 the Eastbourne engine of 1880. 
 
 After the Allegheny engine of 1883 Allis built several 
 compound-beam fly-wheel engines of about 2,500,000 gallons, 
 and in one of them for Hannibal, Mo., the clearance was 
 reduced by putting the Corliss valves in the head of the 
 cylinder. This engine on test in November, 1885, by 
 Mr. C. A. Hague, gave a duty of 118,327,025 foot- 
 pounds per 100 pounds coal at a gauge pressure of 79 
 pounds. This engine then led to the design of the triple- 
 expansion engine. 
 
 The triple-expansion engine (Fig. 88) is the first of the most 
 popular type of large pumping engine, although many other 
 forms have been suggested and used. This pump was built 
 for the city of Milwaukee. The figure shows clearly the 
 arrangement of the pump and engine. This resembles, to some 
 extent, the design of frame used by Moreland in England. 
 It is in reality the same arrangement as that used for 'marine 
 engines with pumps added below the bed plate of the 
 engine. 
 
 The engine gave the high duty of 122,483,204 foot-pounds 
 per 100 pounds of coal with 80 pounds steam pressure. The 
 dimensions of the engine were as follows: 
 
 High-pressure cylinder diameter 21 ins. 
 
 Intermediate- pressure cylinder diameter 36 
 
 Low-pressure cylinder diameter 51 
 
 Pump plunger 23 \ 
 
 Stroke 36 
 
 Capacity per twenty-four hours 6,000,000 gals. 
 
 The first triple-expansion pump' in England (Fig. 89) was 
 introduced in 1891 by S. Richardson & Sons, and gave a steam 
 consumption of 13.53 pounds per I.H.P. hour. The use of 
 
RECENT HISTORY 
 
 97 
 
 triple-expansion for direct-acting pumps was also introduced 
 at this time by Davison in his water- works pumps. These 
 gave good results and increased the duty of direct-acting 
 pumps. 
 
 In 1885 Worthington brought out his high-duty engine, 
 
 built exactly as the -duplex engines, but with the addition of 
 compensating cylinders so as to use steam expansively. It 
 was not possible to obtain the high duties of the expansive 
 fly-wheel pumps with the ordinary duplex pump, and although 
 these could be sold much cheaper, their duty was so low as to 
 
98 
 
 PUMPING MACHINERY 
 
 prevent their use in many cases. The use of heavy recipro- 
 cating weights was not feasible, and Worthington improved 
 the invention of Mr. J. D. Davies of 1879 by which steam could 
 be used expansively. The arrangement as applied by Mr. 
 C. C. Worthington is shown in Fig. 90, where, as the piston 
 moves to the right, the tail rod of the pump forces the plungers 
 into the oscillating cylinders A A against water pressure from 
 the cylinder Z?, to which they are connected by pipes. The 
 pipes are connected through stuffing boxes to the trunnions 
 
 O O 0^0 O 
 
 FIG. 91. Independent Air Tank Compensator. 
 
 of the cylinders. It will be seen that the side thrust from these 
 is balanced on account of their symmetrical position. The 
 cylinder B is connected to the top of the air chamber in the 
 force main of the pump, and thus the pressure of the air cushion 
 above the fluid in B is always equal to that in the force main. 
 By the arrangement of multiplying cylinders at the base of 
 B the unit pressure on the plungers of A A is much greater 
 than that of the force main, as the area of the plunger is much 
 smaller than the piston of B. 
 
 The pistons are opposed by the pressure in the compensators 
 
RECENT HISTORY 
 
 99 
 
 until the engine reaches its mid-position; the plungers are here 
 in a vertical position. From this point to the end of the stroke 
 the pressure on the plungers is aiding the motion, the exact 
 amount being proportional to the cosine of the angle of incli- 
 nation to the horizontal. This then means that at the beginning 
 of the stroke the steam pressure will have to exceed the pressure 
 on the main pump plunger, diminishing until the mid-stroke, 
 and from there on it may be less than the pressure on the main 
 plunger. This permits the use of steam expansively. The 
 application of an independent air tank (Fig. 91) is used at 
 
 FIG. 92. Action of Compensator. 
 
 times with this system. The figure gives a clear idea of the 
 arrangement of the rams and plungers. 
 
 To better illustrate the action of this, Fig. 92 is given; in 
 it the indicator cards, the pump cards, the various positions of 
 compensating cylinders and the combined cards are shown. 
 At the lower right-hand side of the figure the total steam- 
 pressure card has been placed over the total water-pressure 
 card .and a dotted line has been drawn showing the variation 
 of total pressure exerted horizontally by the compensators. 
 It is evident that during the first half of the stroke the excess 
 
100 
 
 PUMPING MACHINERY 
 
 of total steam pressure just equals the positive resistance from 
 
 (i. 
 
 I 
 
 the compensators, while during the latter part, the excess of 
 
RECENT HISTORY- 101 
 
 water pressure is made up of the negative resistance or returned 
 energy from the compensators. This enables the pump to 
 obtain all of the advantages of the expansive use of steam and 
 tests of these engines have shown this. 
 
 Fig. 90 shows the application of this in 1885, wherein the 
 forms of water end and steam end are obvious. The steam 
 end is jacketed throughout and has separate steam and exhaust 
 passages with rotary cut-off valves in addition to the balanced 
 slide valves. This type was soon changed so far as the valve 
 was concerned, but the compensating principle was practically 
 the same. 
 
 A test of the first engine of this type built in 1885 for the 
 water works of New Bedford, Mass., gave a test duty of 
 79,238,160 foot-pounds on 100 pounds of coal, and this test 
 was followed by another by Mr. John G. Mair, a representative 
 of Simpson & Co., the English builders of pumping machinery. 
 Mr. Mair found that the engine used from 14.53 pounds to 15.05 
 pounds of steam per I.H.P. hour. 
 
 This engine was of the compound type, as shown in the 
 figure, and was developed extensively, being used in a vertical 
 position in places where ground area was limited, as in an 
 installation made at Memphis, Tenn. This same engine was 
 tested in 1891 and gave a duty of 117,325,000 foot-pounds per 
 1000 pounds of steam. 
 
 This compensated form of pump was also developed into 
 a triple-expansion steam end. One of the first built by 
 Worthington in 1891, for Turtle Creek, Pa., was of the vertical 
 form with the high-duty attachment. The triple-expansion 
 forms are shown in Figs. 93 and 94. 
 
 Fig. 93 shows the pump installed at Fall River, Mass., for 
 the city pumping station in 1909. It has three steam cylinders, 
 21, 33, and 60 inches in diameter, respectively, arranged in 
 tandem with the 24^-inch water cylinder. The common 
 stroke is 36 inches. The high-duty attachment is placed 
 between the high-pressure steam cylinder and the water cylinder. 
 
 The pump takes its suction from an open conduit and 
 pumps directly into the city mains at about 100 pounds per 
 
102 ,/VJ/P/;;^ MACHINERY 
 
 square inch. The capacity was 10,000,000 gallons per twenty- 
 
 
 four hours. The pump is of the outside -packed plunger type 
 
RECENT HISTORY 103 
 
 with special glands. The cylinders are jacketed and reheat er 
 coils are used between the cylinders. The auxiliary pumps for 
 charging the air chamber of the compensator, the air pump, 
 and jacket pump are driven from the main pump. The stroke 
 governor regulates the stroke, keeping it at its full value by 
 means of a change of the pressure in the compensator. 
 
 The official test of this pump gave a duty of 136,500,000 
 foot-pounds per 1000 pounds of dry steam. A pump similar 
 in principle to the above was installed at Montreal in 1908 and 
 gave a duty of 177,538,000 foot-pounds per 1000 pounds of steam. 
 
 Fig. 94 gives a good idea of the vertical form as installed 
 at the Central Park pumping station in Chicago, although 
 the figure is not of this particular pump. These pumps are 
 duplex with steam cylinders, 21, 33, and 60 inches in diameter, 
 and of 5o-inch stroke, while the double-acting water cylinders 
 are 34J-inch diameter and 5o-inch stroke. The steam cylinders 
 are jacketed on heads and barrels and the valve gearing is of 
 the four- valve rotative type. At 18 revolutions per minute 
 this pump lifted 20,000,000 gallons per twenty-four hours. 
 The duty on test was 174,735,801 foot-pounds per 1000 pounds 
 of superheated steam used. 
 
 By means of the balancing plungers shown in Fig. 94, the 
 weight of the moving parts is supported by plungers BB, which 
 force water into a tank C against air pressure. The pressure 
 in this tank is regulated by the pressure in the discharge main. 
 When this pressure falls, due to a break in the line, the weight 
 of the moving parts would not be supported by the water 
 pressure and the pump would soon come to rest. The steam 
 pressure is used to force the water on each stroke. At points 
 A A on the piston rods between the low-pressure cylinder and 
 the water cylinder the digh-huty attachment is applied which 
 permits of the expansive use of steam, as shown previously. 
 
 The suction pipe, which is connected to a large cistern in 
 the pump room, is joined to the surface condenser so that all 
 of the water passes through the tubes. In this way the con- 
 densation is cared for in a station situated in the center of the 
 city where condensing water is not at hand. 
 
104 
 
 PUMPING MACHINERY 
 
 One of the latest types of compensating, direct-acting 
 pump, permitting the expansive use of steam without the use 
 of a fly wheel, is the invention of Mr. Luigi d'Auria. The 
 pump (Fig. 95) has a cylinder A between the steam and water 
 cylinders, the ends of which form the extremities of a pipe 
 circuit which in some cases is used as the base of the pump. 
 The cylinder A contains a piston, and as this moves back and 
 forth the water with which the system is rilled is compelled 
 to travel first in one direction and then in the other. 
 
 At the beginning of the stroke the inertia of the water 
 
 FIG. 95. d'Auria Pump. 
 
 utilizes the excess steam pressure, while in bringing the water 
 to rest it will exert a force to supply the deficiency. In this 
 way steam may be used expansively in the steam cylinder to 
 drive the pump against a uniform water pressure. The steam 
 may be used in one or more cylinders, depending on the purpose 
 of the engine. When multiple expansion is wanted it can be 
 applied to this engine. The compensation is so perfect with 
 this machine that high speed may be employed with no danger 
 of pounding. 
 
 This method has also been applied to triple-expansion 
 pumps where high duties have been sought. 
 
RECENT HISTORY 
 
 105 
 
 About the end of the last century Mr. Charles L. Heisler 
 of Erie, Pa., introduced a new type of compensated engine 
 which he described in 1900. His design employs a duplex 
 arrangement of cylinders, as shown in Fig. 96, which is built for 
 triple expansion. The line drawings show that each piston rod 
 is connected to a vibrating radius rod A by a short connecting 
 rod B and that the radius rods are joined by the link C. Just 
 before the left-hand engine reaches the end of its upper stroke, 
 as in position i, the movement of the link C in compression 
 lifts the left radius rod and aids in the movement of that pump. 
 Position 2 shows that the left pump is being aided by com- 
 pression in the rod C, while the other positions show the pumps 
 
 Position 
 
 Position 2 Position 3 Position 4 
 
 FIG. 96. Heisler Pump. 
 
 at the two remaining starting points. In each of these posi- 
 tions the excess of pressure at the beginning of the stroke is 
 carried over to the other side and aids the steam when that 
 pressure is less than the resistance of the water. This method 
 gave fair results and the builders claim duties of 130,000,000. to 
 160,000,000 foot-pounds per 1000 pounds of steam, although 
 the author has no records of any tests giving such high duties. 
 The method of having one side of the pump aid the other 
 by means of a linkage was not new at this time, -for in 1874 
 Fielding of England proposed such a scheme and in 1887 Henry 
 Davey patented a less complex form. The Davey method of 
 compensation is given in Fig. 97. In this system single-acting 
 water plungers are connected with double-acting steam cylinders. 
 
106 
 
 PUMPING MACHINERY 
 
 When the plunger in A is moved to the right there is little 
 resistance to its motion, so that the excess of steam pressure 
 is carried through the rod D to the plate C and from it to the 
 other side of the pump by the rod E. At the beginning of the 
 stroke, when little of the pressure from the side A is needed, 
 it is seen that the rod D has little leverage about the center 
 of C while E has a long leverage. When, however, the plunger 
 of B has been driven to the left and the steam on that side has 
 expanded so that more pressure is needed, the leverage of the 
 rod D becomes greater, giving more pull on the rod E. At the 
 end of the stroke the low pressure on the side A is pulling with 
 full leverage and exerting a great force on E, which pulls on 
 
 FIG. 97. Davey Compensator. 
 
 the plunger of B. B requires this additional pressure, as the 
 steam on that side has expanded to a low pressure. The 
 valve gear is such that these events can take place as described, 
 the two pumps reversing at the same time. The pumps des- 
 cribed by Davey in London Engineering in 1877 and the Watt 
 engine described in the same magazine in 1885, are mentioned 
 as similar to the above in principle. 
 
 From the introduction of the triple-expansion fly-wheel 
 pump of Allis in 1886 to the present there have been a number 
 of improvements in the arrangement of steam valves and 
 reheaters, pump cylinders and valves, condensers and other 
 details. With the gradual increase in the steam pressure, the 
 
RECENT HISTORY 107 
 
 duty, when measured on the basis of 1000 pounds of steam, 
 has increased together with that of the duty on the basis of 
 1,000,000 B.T.U. The type first used by Allis has been adopted 
 by other large engine builders and this design in the hands of 
 the Worthington, Blake, Holly, Snow, and Southwark companies 
 has proven it to be a good one. Mr. E. D. Leavitt, Jr., has also 
 designed special types during these years which have given 
 excellent results, as will be seen later. The duties on the basis 
 of 1000 pounds of steam have gone from 122,452,729 foot-pounds 
 in 1886 to 154,048,704 foot-pounds in 1892, to 179,454,250 in 
 1900, and finally to 181,048,605 foot-pounds in 1906. This 
 method of stating the duty has the objection that the efficiency 
 is dependent on the amount of heat in each pound of steam 
 supplied. The amount of heat in 1000 pounds of dry steam 
 depends on the pressure of the steam and consequently it is 
 perfectly possible to have an engine with a smaller duty per 
 1000 pounds of steam more efficient actually than one with a 
 a higher looo-pound duty. The method of expressing the duty 
 as the number of foot-pounds per million British thermal units 
 supplied is far better, as in this case the real efficiency is being 
 measured. 
 
 The engine giving the highest actual thermal efficiency, 
 according to C. A. Hague, is the Norberg quadruple-expansion 
 engine of 1899, which gives 22.80 per cent; however, the thermal 
 efficiency based on useful work delivered by the pumps, obtained 
 by Professor Thurston in 1899, was 20.9 per cent. This result 
 was within 84 per cent of the theoretically perfect engine. 
 This pump was built for the Pennsylvania Water Company, 
 near Pittsburg. There were four steam cylinders as well as the 
 pump barrels and fly wheel. The pump was built on peculiar 
 lines because the water level of Allegheny River is subject to 
 much variation, and moreover the pump was to be placed with 
 two other similar 6,ooo,ooo-gallon pumps within a well of 38 
 feet diameter belonging to the company. The dimensions 
 of this engine are as follows: 
 
108 PUMPING MACHINERY 
 
 Diameter high-pressure cylinder 19 J ins. 
 
 first intermediate cylinder 29^ ' ' 
 
 second intermediate cylinder 49 " 
 
 low-pressure cylinder 5 7 ' 
 
 plunger (double-acting) cylinder 14! " 
 
 Stroke 42 
 
 Fly-wheel diameter 13 ft. 
 
 Capacity c 6,000,000 gals. 
 
 This engine gave the highest duty on the absolute basis, 
 although it does not give so high a value per 1000 pounds of 
 dry steam as those which have been built later; the best 
 accepted duty on this basis being about 180,000,000 foot- 
 pounds. 
 
 Figs. 98 and 99 illustrate two modern forms. The first 
 is a pump installed in 1909 at Brockton, Mass. The steam 
 cylinders are 19, 36, and 54 inches in diameter, and the stroke 
 is 36 inches. At 40 R.P.M. the 36-inch single-acting plungers 
 will lift 6,000,000 gallons per twenty-four hours against 300 
 feet head. The duty of this pump was 169,982,000 foot-pounds 
 per 1000 pounds of dry steam at 150 pounds pressure. 
 
 The second pump illustrates a horizontal cross-compound 
 condensing crank and fly-wheel Snow engine for Elyria, Ohio. 
 The capacity of the double-acting plungers, 14^- inches in diam : 
 eter, is 5,000,000 gallons per twenty-four hours at 36 R.P.M. 
 The steam cylinders are 22 and 48 inches in diameter with a 
 36-inch stroke. The steam pressure was 120 pounds per square 
 inch and the total water pressure happened to be the same. The 
 duty was 137,000,000 foot-pounds per 1000 pounds of dry steam. 
 
 These two pumps illustrate clearly the form of modern 
 high-duty pump for water works. Fig. 98 illustrates the 
 vertical self-supporting pump in which the valve boxes and 
 air chambers form the base for the pump, and the pump barrel is 
 a separate casting. The steam valve gearing is of the Corliss type. 
 The valves are placed in the sides of the high-pressure cylinder, 
 in the sides and heads of the intermediate, and in the heads 
 of the low. The side straddling rods leading from the cross- 
 head to the end of the plunger so as to clear the shaft and 
 crank are shown. The governors, valve controls, and oiling 
 devices are also apparent. Such pumps are usually placed in 
 basements so that the level of the main floor of the station is 
 just over the air chamber. 
 
RECENT. HISTORY 
 
 109 
 
 Fig. 99 is the Snow type of Horizontal pump designed to 
 take up little floor space and yet have the cylinders for the 
 
 
 FIG. 98. Holly Triple Expansion Pump. 
 
 water and steam ends in such a position that they may be 
 easily examined and repaired. The method of clearing the 
 
110 PUMPING MACHINERY 
 
 crank and shaft by two rods joining the cross-head to the 
 plunger rod is clearly shown a? well as the arrangement of 
 air chambers, pump cylinders, steam receiver, and valves. 
 
 In this period of the important developrr^ent of water-works 
 pumping engines strides were being made with the centrifugal 
 pump. As will be remembered, it was about the year 1846 
 that Andrews showed that curved vanes were better than the 
 straight vanes in the old Massachusetts pump (Fig. 39). This 
 development was carried out by Gwynne of England. At this 
 time Andrews also suggested the value of having the vanes of 
 
 FIG. 99. Snow Compound Pump. 
 
 the runner inclosed between two discs as shown in Fig. 40. 
 In 1851 Appold showed the advantage of the curved vane of 
 Lloyd's fans experimentally. 
 
 From now on the centrifugal pump was used extensively 
 and the patent gazettes are filled with the record of improve- 
 ments. The great field for this pump was the lifting of large 
 quantities of water through small distances, where there was 
 not much floor space available, and also where the original 
 cost of the apparatus was of some importance. The firm of 
 John & Henry Gwynne in England was the best known of 
 the early builders, and their pumps for the drainage of the low- 
 
RECENT HISTORY 111 
 
 lands of Holland and Denmark and the emptying of large 
 dry docks were very successful. 
 
 Not only did the Gwynnes try low lifts of 15 and 30 feet, 
 but in 1868 they built a pump with a runner 2 feet in diameter 
 to lift water against a total of 18 feet .suction and 114 feet 
 discharge. The pump was driven at a speed of 910 revolutions 
 per minute. 
 
 In 1869 these pumps were applied to condensers of steam 
 engines. For this service they are admirably adapted, as a 
 large body of water has to be lifted a small distance in the case 
 of stationary engines, while for marine practice the only work 
 the pump does is to overcome the resistance due to the moving 
 of the water and the friction of the pipes and tubes. 
 
 The early study of these pumps was quite meager and 
 their design was more or less empirical, but as their use extended 
 a better understanding of them was had. Although inexpensive 
 for the quantity of water handled and apparently of a form to 
 give good results the efficiency of this machine was low. The 
 earlier efficiencies of 50 per cent were increased to 65 and 70 
 per cent in 1885. Some records give as high as 75 per cent for 
 this type of pump. High efficiencies can be had only by careful 
 design and construction. 
 
 The application of this pump to other services increased. 
 In 1869 it was employed for dredging, although a patent of 
 Louis Schwartzkopff , granted in England in 1856, foreshadowed 
 this application of this simple machine, while for draining 
 wells vertical shafts were used in 1884. Before this time, 
 however, the manner of balancing end thrust by bringing in 
 water from each side was employed. The possibility of handling 
 large quantities of water is illustrated by a pump built by 
 R. Moreland & Sons for a dock at Malta, where it was required 
 to lift 30,000 gallons of water per minute against a compara- 
 tively low head. This was built in 1887. Another pump was 
 built a little earlier by the Southwark Foundry and Machine 
 Co., designed to raise 40,000 gallons per minute. 
 
 W. O. Weber showed by experiment that the efficiency of a 
 pump decreased after a certain head was reached, and so when 
 
112 
 
 PUMPING MACHINERY 
 
 greater heads were to be overcome by centrifugal pumps, 
 it was suggested by some (W. A. Booth claims the honor of 
 this) that two pumps be placed in a series (Fig. 100). This 
 then led to the multi-stage pump, where several pumps were 
 brought together in one casting. The water discharged from 
 one stage passes over to another stage, where its pressure is 
 increased. Continuing in this manner the pressure against 
 
 FIG. 100. Two-Stage Pump. 
 
 which the pump will act effectively can be made as great 
 as desired. The Allis-Chalmers Company have recently 
 installed a five-stage pump which lifts 3000 gallons of 
 water per minute against a total head of about 700 feet. 
 Further details of the centrifugal pump will be considered 
 in a later chapter. 
 
 The rotary pump has been the subject of many patents 
 during this later period. The principles involved, however, 
 
RECENT HISTORY 
 
 113 
 
 were all embodied in the earlier designs. A few will be described 
 to give some idea of the progress made in this form of pump. 
 
 The Behrens pump of 1867 (Fig. 101) represents a form of 
 rotary pump using rotating lobes with no sliding parts. The 
 shafts A A are geared to move in opposite directions with the 
 same velocity. The pistons B are so formed that they slide 
 over the bore of the main cylinder C and the inner abutment 
 liners Z), which are stationary. These liners are carried on the 
 heads on one side of the cylinder, while the" pistons B are carried 
 from the shaft A by means of a flange cast with the pistons. 
 This is seen better in the small perspective view. 
 
 On turning the pistons in the directions shown in the figure, 
 water is drawn into the space E while water in the space F is 
 
 FIG. 10 1. Behren's Rotary Pump. 
 
 being forced 'out of the pump. Water under pressure in F 
 cannot pass the surfaces of the pistons, which are always in 
 contact with the fixed surfaces of C or D. It is seen in the 
 figure that after a little motion, the right-hand piston touches 
 the surface on the left-hand liner D, and then it acts as an 
 abutment while the water in the space G is discharged. One 
 piston must come in contact with the liner of the opposite side 
 before the other piston breaks contact with its opposite liner. 
 In this manner there is always a barrier for a free passage from 
 the discharge to the suction. This was improved in the Port- 
 land rotary pump of 1882, when the contact with the central abut- 
 
114 
 
 PUMPING MACHINERY 
 
 ment made a line contact on the center line, making it unnecess 
 ary to have stationary sleeves, hence that part was eliminated. 
 
 t i i 
 
 FIG. 102. MacFarland's Rotary Pump. 
 
 The rotary pump of MacFarland, 1875 (Fig. 102), represents 
 a development of the Trotter type of rotary pump. The con- 
 struction of this pump as well as its action can be seen from the 
 
 FIG. 103. Phillip's Rotary Pump. 
 
 figure. The ring A, carrying the piston blades >, is supported 
 on a stationary ring B projecting from the back head of the 
 pump. The driving ring C is carried from the driving shaft, 
 which is not placed at the center of the cylindrical barrel of the 
 pump. The action of the blades on the water is clearly seen. 
 
RECENT HISTORY 
 
 115 
 
 A pump somewhat similar to this was brought out 
 by Phillips about ten years later (Fig. 103). In this 
 pump the vanes attached to a sliding rotating cylinder 
 were replaced by a cylindrical roller. This arrangement 
 would eliminate much of the friction, although, at the 
 center and periphery, there is still considerable friction 
 produced by the slipping of the rollers against the sides of 
 the driving slot. 
 
 FIG. 104. Wilkin's Two-Lobed Rotary Pump. 
 
 FIG. 105. Silsby Rotary Pump. 
 
 The pump (Fig. 104) of John T. Wilkin, designed about 
 the year 1892, consists of two similar rotating lobes each having 
 two hypocycloids and two epicycloids. These two curves will 
 work together with uniform velocity ratio. There is sliding 
 contact with the cylinder walls and at this point there may be 
 considerable leakage. 
 
116 
 
 PUMPING MACHINERY 
 
 The use of the curves employed for gear-tooth profiles is 
 not uncommon, as may be seen in Fig. 105, which represents 
 the form of pump used on the Silsby fire engine, and in Fig. 106, 
 
 FIG. 106. Root Rotary Pump. 
 
 FIG. 107. Allis Screw Pump. 
 
 which represents the form of propeller used by the Root Com- 
 pany. These last three pumps are modifications of the 
 early form of Serviere, illustrated in Chapter I. In the 
 Root form the number of teeth is reduced to two for each wheel. 
 A development of the early form of screw pump, Chapter I, 
 is shown in Fig. 107. This pump is for low lifts, being used in 
 
RtiCENT HISTORY 
 
 117 
 
 this country when water is to be raised a few feet into a stream, 
 there to mix with the sewage of a town. The screw acts on 
 the water, forcing against a static head. For low lifts these 
 
 FIG. 108. Wood Propeller Pump. 
 
 wheels, as well as the scoop wheels, are quite effective. A devel- 
 opment of the same idea is shown in the Wood propeller pump 
 
118 
 
 PUMPING MACHINERY 
 
 (Fig. 108). In this pump a vertical shaft guided by bearings 
 00 carries a series of helical surfaces or screws NN, which fit 
 close to the side of the pump pipe. Rotating this by means 
 of a belt, as shown in the picture, the water is forced upward. 
 An electric motor may be used for driving. 
 
 _FiG. 109. Helicoidal Pump. 
 
 FIG. no. Bolton and Imray Helical Pump. 
 
 Wade & Cherry's helicoidal pump (Fig. 109) is a combi- 
 nation of the screw pump and the centrifugal pump. This 
 was brought out in 1886. It is so arranged that the water is 
 first forced toward the centrifugal impellers by the screws, 
 after which it is discharged into the central wheel, 
 
RECENT HISTORY 
 
 119 
 
 from which place it passes into the volute chamber around 
 the wheel. 
 
 The Boulton & Imray helical pump (Fig. no) was 
 described in 1872. It is an impeller pump. The casing con- 
 tains a helical passage through which the water goes from one 
 level to another. The height of the passage is about twice its 
 width and two-thirds the pitch of the helix. In this manner 
 it is possible to place a series of square paddles A on a wheel 
 rim of such a width that they cut off one-half the height of the 
 passage, as seen in the figure. The action of the paddles is to 
 
 FIG. in. Gould Fire Engine. 
 
 impel the water through the casing. The pump shown in the 
 figure had a wheel 3 feet 6 inches in diameter outside of the 
 blades, the blades themselves being 6 inches square. The 
 cylinder of the driving engine was 10X8 inches. Originally 
 of the form shown ft was changed in a later design by the use 
 of the three-cylinder Brotherhood engine. 
 
 The steam fire pump brought out in the last decade of the first 
 period considered in this work was improved during the following 
 years, and by the year 1876 a number of them were in use. 
 
120 
 
 PUMPING MACHINERY 
 
 The Gould steam fire pump (Fig. in) exhibited at the 
 Centennial Exposition in 1876 illustrates to what degree the 
 steam fire engine had developed from the early form of Braith- 
 waite and Ericsson described in the last chapter. The boiler 
 has now been placed in a vertical position and the exhaust 
 steam has been used to produce a draft in place of the early 
 bellows. The engine has been made vertical also and a fly 
 wheel has been added to make the operation of the pump 
 more regular. This type of engine was likewise manufactured 
 by Ahrens and the firm of Clapp & Jones in America. Fig. 
 
 FIG. 112. Silsby Fire Engine. 
 
 112 illustrates the application of the rotary pump and engine 
 for fire service. The Silsby engine is one found in many of 
 our large cities, and although there is considerable slip in the 
 pump, its light weight is an important factor. These pumps 
 were introduced about the middle of the last century. The 
 rotary steam engine is seen near the boiler while the rotary 
 pump appears beneath the driver's seat. The pump was of 
 the type shown in Fig. 105, the steam engine being quite similar. 
 The use of horizontal engines is also found in the types of 
 American fire engines, as may be seen in Fig. 113. The Button 
 
RECENT HISTORY 
 
 121 
 
 engine is the product of one of the oldest builders of fire engines 
 and the type is of value in that the parts are less complex than 
 those in which the fly wheel is used. The early hand pumps 
 
 FIG. 113, Button Fire Engine. 
 
 FIG. 114. Hand Fire Pump of 1873. 
 
 were continued in many places even at this late date. Fig. 
 114 illustrates a pump exhibited at the Vienna Exposition of 
 1873. This machine was hauled by horses, but it was operated 
 
122 
 
 PUMPING MACHINERY 
 
 by men. The handles A and B are attached to the pump 
 plungers. The suction is connected to the pipe or hose C on 
 the rear of the pump, not clearly shown in the figure, while 
 the discharge is delivered through D. 
 
 Fig. 115 illustrates a modern type of fire. engine of recent 
 design. The pump is vertical with the steam cylinder near the 
 boiler; the pump end near the point of attachment to the fire 
 hydrant, air chambers on the suction and discharge sides of 
 the pumps, valve boxes in accessible places and all parts open. 
 The object sought in the modern fire engine is a machine always 
 
 FIG. 115. Modern Fire Engine, American-La France Co. 
 
 ready for a definite, positive service in which all parts are 
 accessible for operation, maintenance, and repair. 
 
 The air lift pump -was patented by James B. Frizell in 1880, 
 and Julius I. Pohle made a number of improvements covered by 
 patents granted in 1886. The pump (Fig. 116) consists of an 
 air pipe A which dips into a well B and passes below the lower 
 end of a larger pipe C. Compressed air is driven through the 
 pipe A and discharges against the water pressure due to the 
 depth of immersion of the discharge nozzle. The air escapes 
 
RECENT HISTORY 
 
 123 
 
 from the nozzle and acts on the water above considered as a 
 piston, or the air passes up through the water, which is thus 
 made lighter and is forced upward by the static pressure of 
 the water in the well. The figure shows beads of water acting 
 as pistons with air between, expanding as the head is reduced 
 by the upward motion of the water. This is probably the 
 
 FIG. 1 1 6. Air-Lift Pump. 
 
 manner of action of the pump after is it in operation, although 
 in starting it is likely that the reduction of density by aeration 
 causes the first flow. Although this form of pumping apparatus 
 was not extensively used or seriously thought of before the 
 work of Pohle in 1886, the method had been proposed earlier. 
 Collom's Lectures on Mining for 1876, delivered in Paris, 
 described this, and Gerlach in a paper points out that Loescher 
 
124 
 
 PUMPING MACHINERY 
 
 of Freiberg described a similar apparatus in a pamphlet printed 
 in 1797. 
 
 Another form of pumping apparatus using air is that in- 
 vented by Professor Elmo G. Harris about 1900. In this pump 
 (Fig. 117) air is compressed by the compressor A into a pipe 
 
 FIG. 117. Harris Air-Lift Pump. 
 
 leading to a tank B. This compressed air acts on top of the 
 water contained in B and drives it out of the discharge pipe D. 
 The suction side of the compressor is connected with a pipe 
 leading to the tank C and the reduction of pressure within the 
 tank draws water into C. At the proper time the valve E 
 is shifted, changing the suction and pressure sides so that water 
 
RECENT HISTORY 
 
 125 
 
 is sucked into B and driven from C. There is an equalization 
 of pressure through the valves of the compressor as soon as 
 the shifting valve E is turned and before the compressor starts 
 to drive out the water from C. This utilizes part of the energy 
 used in compressing the air into cylinder B. 
 
 These two air lift pumps of Pohle and Harris have the 
 advantage that a number of them may be driven from a central 
 plant and there is little loss from such an arrangement. The 
 air lift pump raises a larger quantity in a given time than 
 would be possible with a deep well pump. The Harris pump 
 may be installed in places where other pumps are impossible, 
 
 FIG. 118. Giffard Injector. 
 
 and forms a positive type of pump. These pumps have specific 
 advantages which will be pointed out when their design is 
 considered. 
 
 In 1858 Henri Jacques Giffard patented his invention ,of 
 the steam injector one of the simplest devices for pumping 
 water into boilers. In the earliest form (Fig. 118) steam 
 entered through the cock A and passed through a series of 
 small holes into the interior of a tube having a nozzle B at 
 its extremity. The core or rod C controlled by D closes the 
 nozzle or regulates the amount of opening. The steam acquires 
 a high velocity in the nozzle, causing a vacuum in the space 
 E, which raises water to that point, where it mixes with the 
 steam. As the steam condenses and imparts a high velocity to 
 the water, the cross-section of the mixture grows smaller and 
 
126 PUMPING MACHINERY 
 
 consequently the combining tube F is made convergent. To 
 change this high velocity into pressure, Giffard followed the 
 converging combining tube by a diverging delivery tube G. In 
 this manner the velocity was reduced and the pressure was 
 so increased that the water could enter against the pressure of 
 the boiler. 
 
 The idea of using the power of a moving jet was not new. 
 According to Kneass a crude injecting apparatus was used as 
 early as 1570 by Vitrio and Philibert de Lorme. In 1818 
 Mannoury d'Ectot patented a device which was quite similar 
 to the injector, and Bourdon in 1857 patented a combination 
 of convergent and divergent tubes for transforming the energy 
 of a moving jet. 
 
 The latter invention resembled Giffard' s, but the writings 
 of the inventor of the injector, made seven years before Bour- 
 don's invention, in which he explained the theory of the injector, 
 proved that he should receive the credit for this invention. 
 
 The manufacture of these machines was at once undertaken 
 in various countries by licensees. In America William Sellers 
 & Co. began their manufacture, while Sharp, Stewart & Co., 
 made them in England. 
 
 The Giffard design was an excellent one. Many ideas 
 were suggested in his patent specifications and pamphlets, 
 so that although changes were made in the injector as manu- 
 facturing processes were developed, and as more knowledge 
 was had of the properties of jets and steam, the original ideas 
 are still to be seen in the present-day injector. 
 
 The operation and design of injectors will be considered 
 later with the modern form of injector, yet it is well to men- 
 tion the names of some of those who added to the development 
 of the injector. They are: Millholland, Rue, Sellers, Han- 
 cock, Williams, Loftus, Bancroft, and Kneass in America; 
 Robinson & Gresham, Turck, Hamer, Metcalf, and Davies in 
 England; Haswell, Pradel & Krauss and Korting in Germany; 
 Cuau, Bouvret, Polonceau, and Delpeche in France. 
 
 A pump in which the driving steam touches the water is 
 known as the pulsometer (Fig. 119). This is really a develop- 
 
RECENT HISTORY 
 
 127 
 
 ment of the old Savery engine. Steam is admitted at A and 
 
 if the ball is resting over the left-hand opening steam will 
 
 enter C and drive out the 
 
 water. A sudden change in the 
 
 steam pressure when the water 
 
 surface reaches the level of the 
 
 discharge valv$ causes the ball 
 
 to close the right-hand side, 
 
 and then the condensation of 
 
 the steam on that side draws 
 
 water through the suction D 
 
 and suction valve E, as shown 
 
 in the figure. 
 
 When the water is driven 
 from B the ball is forced over 
 to the left and the operation is 
 repeated. 
 
 This type of pump is easily 
 applied, and in contracting work 
 
 it is USed for this reason, FIG. ng.-Pulsometer. 
 
 although it is most expensive 
 
 when its thermal efficiency is considered. For this kind of 
 rough and rapid work it may not cost more in money to operate 
 this than other forms of pumps. 
 
 FIG. 1 20. Quimby Screw Pump. 
 
128 
 
 PUMPING MACHINERY 
 
 The Quimby pump (Fig. 120) is a development of the pump 
 shown in Chapter I, as the invention of Revillion. Its oper- 
 ation is evident from the figure, which shows right- and left- 
 handed screws meshing^ together so as to form abutments for 
 each other, catching the liquid in the threads, thus forcing it 
 onward. 
 
 The latest form of pumping apparatus designed is that 
 of Herbert A. Humphrey, described by him in Engineering, 
 November 26, 1909. It is a combined gas engine and pump. 
 The pump (Fig. 121) consists of an explosion head A, a suction 
 
 FIG. 121. Explosion Pump of Humphrey. 
 
 portion B, a discharge pipe C. Assume that a compressed 
 charge of gas and air is exploded in the head A. The force 
 of this combustion drives the water from C into the reservoir, 
 and the energy given to the water by the excess pressure in A 
 is dissipated only after the pressure in A has been reduced to a 
 vacuum. When the pressure at B within the pipe system is 
 reduced sufficiently, water is drawn in through the suction 
 valves. After the water is brought to rest in C, the gases in A 
 are compressed by the static pressure from the reservoir, which, 
 forces the water backward, developing a certain amount of 
 velocity. During this operation, however, the valve D is 
 opened automatically, and the burned gases are exhausted. 
 
RECENT HISTORY 
 
 129 
 
 When the water reaches a point near the valve D the valve 
 is closed, and before the velocity which has been set up in C 
 can be dissipated the water and gas in A are compressed beyond 
 the pressure corresponding to the static head. This means 
 that there is another surge toward the reservoir, followed by a 
 vacuum in A, which is destroyed by opening the air and gas 
 valve E, permitting a fresh charge to enter. This suction 
 stroke is followed by another backward surge, during which 
 the explosive mixture in A is compressed. At the proper 
 time this is exploded by an electric spark, and the operation 
 is repeated. The valves E and D and the ignition apparatus 
 are operated by the surging water in the system. The device 
 is so arranged that E opens on an expansion stroke after D 
 closes at the end of a compression stroke. 
 
 Professor W. C. Unwin made a test on this pump in connec- 
 tion with a gas producer, developing 16 pump horse power on 
 1.063 pounds coal per delivered horse power hour or 12,243 
 B.T.U. per P.H.P. hour. The data for this test are given in 
 the table below: 
 
 Lift. 
 
 P.H.P. 
 
 Cu. ft. at 
 14.7-32 P. 
 
 Calorific 
 Value. 
 
 B.T.U. 
 per P.H.P. hr. 
 
 Lbs. Anth. 
 per P.H.P. hr. 
 
 32.87 
 
 16.15 
 
 83.12 
 
 !47-3 
 
 12,243 
 
 I .063 
 
 2 S-95 
 20.73 
 
 12.32 
 10.99 
 
 90.93 
 93.61 
 
 *43-5 
 I 4S-3 
 
 !3.37 
 *3,596 
 
 1.132 
 1.180 
 
 Mr. Humphrey does not claim to be the originator of a gas- 
 driven pump, as he says this matter dates back to 1868, but he 
 deserves credit for having built so simple a device which yet 
 gives efficiencies higher than those of the most improved steam 
 pumps. 
 
CHAPTER III 
 MODERN RECIPROCATING PUMPS 
 
 IN considering the actual pumps in use it is advisable to 
 classify them in several ways: (a) In regard to the form of 
 water displacer; (b) in regard to the number of displacements; 
 (c) in regard to the method of operation; (d) in regard to the 
 manner of packing; (e) in regard to the direction of the axis 
 of the pump cylinder; (/) in regard to the arrangement of cyl- 
 inders; (g) in regard to the use of the pump. 
 
 The general forms of that part of a pump used to displace 
 water are the plunger, the piston, and the bucket, although 
 such means as air, gas, and steam are used, as was shown in 
 Chapters I and II. The plunger is usually cylindrical in section 
 and forces the water of a pump by entering the space occupied 
 by the water. Fig. 122 shows the construction of a simple 
 plunger pump. This plunger enters the cavity filled with 
 water and displaces an amount equal to the increase of volume 
 of the plunger protruding into the cylinder. 
 
 The piston (Fig. 123) consists of a movable diaphragm 
 tightly fitting against the sides of a cylinder, and forcing the 
 water before it. This piston forces an amount of water equal 
 to the area of the piston multiplied by the stroke. 
 
 The bucket is a piston containing a number of valve-closed 
 passages through which the fluid may pass at the proper time. 
 Fig. 124 shows the construction of a bucket. 
 
 The use of any one of these devices on a pump gives its 
 name to the pump. Thus the names plunger pump, piston 
 pump, or bucket pump designate a pump in which one of these 
 displacers is used. 
 
 Single plungers and buckets, as ordinarily constructed, can 
 lift water on only one stroke, while the piston displaces water 
 
 130 
 
MODERN RECIPROCATING PUMPS 
 
 131 
 
 on both strokes. This gives rise to the classes, single- and 
 double-acting pumps. Of course two plungers may be united 
 
 O 
 
 LJ 
 
 PQ 
 
 IT 
 
 o 
 
 by an outside or an inside connection and form a double-acting 
 plunger pump as shown in Fig. 125, while a bucket or plunger 
 pump may be so arranged that although it lifts only on one 
 
132 
 
 PUMPING MACHINERY- 
 
 stroke of every two strokes, it discharges on each stroke. Such 
 pumps are known as differential pumps. Fig. 126 shows the 
 arrangement for a plunger, and the same could be used with a 
 bucket (Fig. 127). The piston rod is these cases is so enlarged 
 that it is practically another plunger of about one-half the area 
 
 f^^s^f v ^"^ 
 
 D 
 
 D 
 
 ^"'"^\^S?1^ 
 
 [] 
 
 n 
 u 
 
 
 
 FIG. 125. Double-Acting Plunger Pump. 
 
 of the main plunger or bucket. In this manner by connecting 
 the lifting side of the pump to the discharge main through the 
 other end of the pump, one-half of the water will be retained 
 on this side during 'the discharge stroke of the main plunger, 
 while one-half of the water is discharged. The portion of the 
 
 FIG. 126. Differential Plunger Pump. 
 
 FIG. 127. Differential 
 Bucket Pump. 
 
 water retained is delivered on the suction stroke of the main 
 plunger or bucket. In this manner the discharge is made more 
 regular, although the suction occurs on every other stroke. 
 
 The method of operation gives another classification. 
 Steam pumps are those driven by steam pistons, while com- 
 pressed-air pumps are those in which air is used in place of 
 
MODERN RECIPROCATING PUMPS 
 
 133 
 
 steam. Where the steam and water pistons are directly con- 
 nected these are known as direct acting, although if a fly wheel 
 is connected to the system they are known as fly-wheel pumps. 
 
 FIG. 128. Plunger and Ring Packing. 
 
 Power pumps are those driven by belts or gearing, and electric 
 pumps are those directly driven by electric motors. The use 
 of air or gas in direct contact with the water has given the 
 class of air-lift and gas-driven pumps described in Chapter II, 
 
 FIG. 1 29. Central Outside Packing. 
 
 while in the injector and pulsometer the water is moved by 
 the direct action of steam. 
 
 The classification due to the method of packing the water 
 piston or plunger divides pumps into outside- packed pumps 
 (Fig. 125), and inside-packed pumps (Fig. 128), as well as central- 
 
134 PUMPING MACHINERY 
 
 and end-packed machines. A special type of inside packing 
 is known as the plunger and ring type. Fig. 128 shows the 
 arrangement of this type of packing. The plunger in this case 
 passes through a long sleeve or ring in which the resistance 
 against the flow of water is so great that there is little or no 
 leakage. Fig. 125 shows an end packing while Fig. 129 shows 
 central packing. 
 
 Vertical and horizontal pumps are distinguished by the 
 direction of the axis of the pump. 
 
 A pump with a single water cylinder is called a simplex 
 pump, while two-cylinder pumps are called duplex, and three- 
 cylinder, triplex. The first two names, however, are usually 
 associated with direct-acting pumps of small or medium size, 
 while the last term is usually applied to a type of single-acting 
 power pump in which there are three cylinders. 
 
 The last classification of pumps is that due to the use of the 
 pump. It is one of the largest classifications and will demand 
 a more detailed consideration. In examining these different 
 classes the peculiar features of e.ach will be pointed out. For 
 those types which will not be considered later in this work, 
 a full description will be given here. 
 
 Boiler-Feed Pumps. This type of pump is, in all probability, 
 the most common. The conditions under which such pumps 
 are used determine many of their details. In general the pump 
 has a water piston or plunger which is about one-third the size 
 of the steam cylinder. The object of such a difference is to 
 have ample force with any amount of steam pressure to drive 
 water into the boiler against the steam pressure which operates 
 the pump. The pump must be as simple as possible, since it is 
 to be handled by unskilled men. It must have few parts 
 which are liable to breakage or disarrangement. As the varia- 
 tion in pressure in the discharge line is not important and as the 
 pumps are rarely run at full speed, it is quite customary to 
 install these pumps without air chambers on the discharga 
 or suction. 
 
 These pumps are either simplex or duplex, as may be seen 
 from Fig. 130 and Fig. 131. Fig. 130 shows two small Worth- 
 
 
MODERN RECIPROCATING PUMPS 
 
 135 
 
 ington pumps, the upper one being a piston pump and the 
 lower an end outside-packed plunger pump. These are both of 
 the same size, 6 inches diameter of steam cylinder, 4 inches 
 diameter of water cylinder, and a common stroke of 6 inches. 
 This is usually written as 6x4X6 inches. Fig. 131 shows a 
 
 FIG. 130. Duplex Boiler Feed Pumps. 
 (Sizes 6X4X6.) , 
 
 10x6x12 inch Knowles simplex boiler-feed pump. The water 
 end of this pump is of the piston type. 
 
 For heavier water pressures the boiler-feed pump is made 
 with steam cylinders, large when compared with the water 
 end, and the valves are placed above the cylinder casting in 
 separate valve pots as shown in Fig. 132. This pump, 12 inches 
 
136 
 
 PUMPING MACHINERY 
 
 and 17X10X15 inches, is a compound pump in which two 
 steam cylinders are used on each side of the duplex pump for 
 
 FIG. 131. Simplex Boiler Feed Pump. 
 (Size 10X6X12.) 
 
 the purpose of getting greater economy. Fig. 133 shows a 
 different design for the same type of pump. The outside 
 
 FIG. 132. Compound Outside Center Packed Boiler Feed Pump. 
 (Size 12 and 17X10X15.) 
 
 packing in this case is of the end type instead of the center 
 type shown in Fig. 132. 
 
 For use in marine installations where floor space is valuable 
 
FIG. 133. Outside Packed Plunger Boiler Feed Pump. 
 
 
 FIG. 134. Marine Boiler Feed Pump. 
 (Size 10X7X12.) 
 
 137 
 
138 
 
 PUMPING MACHINERY 
 
 vertical pumps are designed. These are often known as 
 Admiralty Pumps. Fig. 134 shows a Davidson duplex 
 piston pump bolted, to a bulkhead of a vessel. It is to be 
 noted that this pump has its valves so placed that they may 
 be examined readily. The suction and discharge pipes may 
 be attached to either side, and caps on the cylinder heads 
 allow an examination of the water piston. 
 
 To give some idea of sizes of boiler-feed pumps the following 
 tables have been taken from catalogues of pump makers: 
 
 THE WORTHINGTON BOILER-FEED PUMP 
 Pressure Pattern For 250 Pounds Pressure. 
 
 These pumps have four single-acting, outside-packed water 
 plungers, working through adjustable stuffing boxes in the 
 ends of the water cylinders. The valves are of brass, guided 
 from below by wings and controlled by composition springs, 
 and are located in separate valve chambers or pots designed 
 to withstand the heavy pressures to which this pump may be 
 subjected. 
 
 Diam- 
 eter of 
 
 Diam- 
 eter of 
 
 Length 
 
 e> 
 
 H.P. Boiler, 
 Based on 45 
 Ibs. Water pet 
 
 Sizes of Pipes for Short Lengths 
 to be Increased as Length 
 Increases. 
 
 Approximate 
 Space Occupied, 
 Feet and Inches. 
 
 Steam 
 
 Water 
 
 of 
 
 hour, which 
 
 
 
 
 
 
 
 Cylin- 
 ders. 
 
 Plun- 
 gers. 
 
 Stroke. 
 
 Pump will 
 Supply at 
 Slow Speed. 
 
 Steam 
 Pipe. 
 
 Ex- 
 haust 
 Pipe. 
 
 Suc- 
 tion 
 Pipe. 
 
 De- 
 livery 
 Pipe. 
 
 Length. 
 
 Width. 
 
 4* 
 
 2 
 
 4 
 
 70 
 
 * 
 
 ! 
 
 Il 
 
 I 
 
 3 91 
 
 i 3 
 
 si 
 
 3 
 
 5 
 
 190 
 
 i 
 
 i* 
 
 al 
 
 2 
 
 4 ii 
 
 ! 6 
 
 6 
 
 Si 
 
 6 
 
 290 
 
 I 
 
 it 
 
 2* 
 
 2 
 
 5 7 
 
 i 6* 
 
 71 
 
 4* 
 
 6 
 
 470 
 
 Ii 
 
 2 
 
 ,4 
 
 3 
 
 5 10 
 
 2 I 
 
 71 
 
 4i 
 
 10 
 
 670 
 
 ii 
 
 2 
 
 4 
 
 3 
 
 8 3 
 
 2 I 
 
 9 
 
 5 
 
 10 
 
 800 
 
 2 
 
 i 
 
 4 
 
 3 
 
 8 3 
 
 2 I 
 
 10 
 
 6 
 
 10 
 
 I2OO 
 
 2 
 
 *i 
 
 6 
 
 5 
 
 8 10 
 
 4 2 
 
 12 
 
 7i 
 
 IO 
 
 I4OO 
 
 1 
 
 3 
 
 6 
 
 5 
 
 9 ii 
 
 4 3 
 
 '14 
 
 *| 
 
 10 
 
 I800 
 
 *i 
 
 3 
 
 8- 
 
 7 
 
 IO I 
 
 3 3 
 
 12 
 
 71 
 
 *5 
 
 20OO 
 
 *i 
 
 3 
 
 6- 
 
 5 
 
 ii 9 
 
 4 3 
 
 1 14 
 
 81 
 
 IS 
 
 2700 
 
 *i 
 
 3 
 
 8 
 
 7 
 
 II 2 
 
 4 o 
 
 '17 
 
 10 
 
 IS 
 
 3700 
 
 2* 
 
 3i 
 
 8 
 
 7 
 
 12 2 
 
 4 3 
 
 1 20 
 
 12 
 
 15 
 
 5200 
 
 4 
 
 5 
 
 IO 
 
 8 
 
 M 5 
 
 4 5 
 
 These sizes have four center-packed plungers. 
 
MODERN RECIPROCATING PUMPS 
 
 139 
 
 PRESCOTT DUPLEX OUTSIDE-PACKED PLUNGER POT-FORM 
 BOILER-FEED PUMPS 
 
 For 300 Pounds Working Pressure. 
 
 The water ends are of the " pot form," and have four 
 single-acting outside-packed water plungers. The water valves 
 are designed for either hot or cold water and arranged so as to 
 be readily accessible. All water passages are large and direct. 
 
 The plunger stuffing boxes are very deep and fitted at the 
 bottom with a removable brass ring which can be replaced 
 when worn. 
 
 These pumps are especially adapted for boiler-feeding service 
 in electric lighting and railway stations or in other plants 
 where high pressures are carried. 
 
 Size. 
 
 (U"- 1 O 
 
 Diameter of Pipe Openings. 
 
 Space Occupied, 
 Feet and Inches. 
 
 ^ 
 
 
 
 S^e* 
 
 
 
 
 , 
 
 
 
 1 
 
 fe 
 
 jj 
 
 |"0ffi 
 
 
 *j 
 
 c 
 
 I 
 
 
 
 
 SJ 
 
 1 
 
 
 a 
 
 rt 
 
 
 
 1 
 
 Length. 
 
 Width. 
 
 z 
 
 * 
 
 
 
 
 GQ 
 
 w 
 
 
 s 
 
 
 
 8 
 
 4 
 
 12 
 
 500 
 
 Ii 
 
 2 
 
 3i 
 
 3 
 
 9 9i 
 
 4 6J 
 
 10 
 
 5 
 
 12 
 
 800 
 
 2 
 
 a* 
 
 4 
 
 2} 
 
 IO 2^ 
 
 4 6J 
 
 12 
 
 6 
 
 12 
 
 1,200 
 
 2 
 
 3 
 
 6 
 
 5 
 
 10 34 
 
 5 9 
 
 12 
 
 7 
 
 12 
 
 2,000 
 
 2 
 
 3 
 
 6 
 
 , 5 
 
 ii 4 
 
 5 5 
 
 14 
 
 8 
 
 18 
 
 4,3 
 
 a| 
 
 3} 
 
 7 
 
 6 
 
 J 4 3 
 
 4 3 
 
 16 
 
 10 
 
 18 
 
 6,850 
 
 2| 
 
 si 
 
 8 
 
 6 
 
 14 si 
 
 4 ii 
 
 18 
 
 10 
 
 18 
 
 6,850 
 
 3 
 
 4 
 
 8 
 
 6 
 
 M 7* 
 
 4 ii 
 
 24 
 
 14 
 
 24 
 
 25,000 
 
 5 
 
 6 
 
 12 
 
 10 
 
 19 oj 
 
 5 "4 
 
 THE DAVIDSON VERTICAL DUPLEX PUMP 
 For a Pressure of 250 Pounds. 
 
 In the Davidson vertical duplex pump there are but few 
 joints, all of which are visible and easily renewed, the water 
 valves are easily accessible for examination or renewal, the water 
 pistons can be packed from the upper end of cylinder and are 
 fitted with fibrous or metallic packing. Water ends of cast 
 iron are composition lined and fitted, or entirely of composi- 
 tion when specially ordered. 
 
140 
 
 PUMPING MACHINERY 
 
 
 
 
 
 H. P. Boiler, 
 
 
 
 
 
 Steam 
 .Cyl- 
 inder. 
 
 Water 
 Cyl- 
 inder. 
 
 Stroke, 
 Inches. 
 
 Gallons 
 per Single 
 Stroke of 
 each 
 
 based on 30 Ibs. 
 of Water oer 
 H. P. per- Hour, 
 which the 
 
 Steam 
 Pipe. 
 
 Exhaust 
 Pipe. 
 
 Suction 
 Pipe. 
 
 Dis- 
 charge 
 Pipe. 
 
 
 
 
 Piston. 
 
 Pump will sup- 
 
 
 
 
 
 
 
 
 
 ply with Ease. 
 
 
 
 
 
 4 
 
 *1 
 
 4 
 
 .084 
 
 165 
 
 | 
 
 1 
 
 2 
 
 Ii 
 
 4* 
 
 2| 
 
 6 
 
 154 
 
 300 
 
 \ 
 
 
 2 2 
 
 2 
 
 5i 
 
 3i 
 
 6 
 
 2 5 
 
 500 
 
 i 
 
 J i 
 
 3 
 
 ai 
 
 6 
 
 4 
 
 8 
 
 435 
 
 870 
 
 i 
 
 ii 
 
 3 
 
 M 
 
 7 
 
 4 
 
 8 
 
 435 
 
 870 
 
 i \ 
 
 i\ 
 
 3 
 
 2 4 
 
 7 
 
 4i 
 
 8 
 
 55 
 
 IIOO 
 
 ^ 
 
 Ii 
 
 4 
 
 3 
 
 8 
 
 5 
 
 10 
 
 -85 
 
 1700 
 
 ii 
 
 2 
 
 4 
 
 si 
 
 8 
 
 5 
 
 12 
 
 i .02 
 
 200O 
 
 ii[ 
 
 2 
 
 4 
 
 
 9 
 
 5i 
 
 IO 
 
 1.03 
 
 2OOO 
 
 if! 
 
 2 
 
 4^ 
 
 4 
 
 10 
 
 6 
 
 10 
 
 1.225 
 
 2450 
 
 2 
 
 2 1 
 
 *5 
 
 41 
 
 10 
 
 6 
 
 12 
 
 i .469 
 
 2900 
 
 2 
 
 2* 
 
 5 
 
 4i 
 
 12 
 
 7 
 
 12 
 
 2 .OO 
 
 4OOO 
 
 2 
 
 a| 
 
 6 
 
 5 
 
 14 
 
 8 
 
 12 
 
 2.6l 
 
 500O 
 
 2| 
 
 3 
 
 7 
 
 6 
 
 14 
 
 8i 
 
 12 
 
 2.94 
 
 6OOO 
 
 24 ' 
 
 3 
 
 7 
 
 6 
 
 Suction and discharge openings on both sides. 
 
 Capacities for boiler feeding are based on a speed of 60 single 
 
 FIG. 135. Worthington Packed-Plunger Pump. 
 (Size 7*X5X~6.) 
 
 strokes a side per minute; for other services, pumps should 
 be run at a piston speed of 30 to 80 feet a side per minute, and 
 
MODERN RECIPROCATING PUMPS 141 
 
 in cases of emergency can be speeded up greatly in excess 
 of this. 
 
 General Service Pumps are those which are intended for 
 the pumping of water for various purposes; drainage, elevator 
 work, small water supply or any other work of a general nature. 
 These pumps are usually designed for specific pressures. For 
 water pressures of 200, 250, and 300 pounds per square inch the 
 outside packed plunger water ends (Fig. 135 ) are used, while 
 
 FIG. 136. Tank Pump. 
 (Size 12X15X15.) 
 
 with pressures of about 150 pounds piston pumps are used. 
 When the water pressure is from 35 to 50 pound the pump is 
 known as a tank piimp, and although the general form is the 
 same as that of the other pumps, the parts are made lighter 
 and the steam cylinder is made smaller than the water end. 
 In all cases of these pumps the variation from one class to 
 another depends on the pressure to be carried. Fig. 136 shows 
 the type of tank pump, while the following table gives the 
 sizes in use by one manufacturer. 
 
142 
 
 PUMPING MACHINERY 
 
 THE WORTHINGTON PISTON PUMP 
 For Tank or Light Service. 
 
 Diameter of 
 Steam Cylinders. 
 
 Diameter of 
 Water Pistons. 
 
 Length of Stroke. 
 
 Gallons per 
 Revolution. 
 
 Maximum Revo- 
 lutions per 
 Minute. 
 
 Maximum Gallons 
 per Minute. 
 
 Sizes of Pipes for Short 
 Lengths to be increased 
 as Length increases. 
 
 Approximate 
 Space Occupied, 
 Feet and Inches. 
 
 | 
 
 JE 
 
 09 
 
 Exhaust 
 Pipe. 
 
 IJ 
 
 "u 
 3 
 O! 
 
 U 
 
 r 
 
 Length. 
 
 Width. 
 
 3 
 
 2| 
 
 3 
 
 3 
 
 80 
 
 24 
 
 i 
 
 i 
 
 It 
 
 ! 
 
 2 io 
 
 o 9} 
 
 4i 
 
 3i 
 
 4 
 
 75 
 
 75 
 
 56 
 
 i 
 
 1 
 
 2i 
 
 xi 
 
 2 II 
 
 i i 
 
 Si 
 
 4l 
 
 5 
 
 i-.S 1 
 
 70 
 
 106 
 
 i 
 
 1} 
 
 - 3 
 
 2 
 
 3 3 
 
 i 4 
 
 6 
 
 si 
 
 6 
 
 2.65 
 
 65 
 
 172 
 
 i 
 
 1} 
 
 4 
 
 3 
 
 3 I0 
 
 i 5 
 
 7i 
 
 5l 
 
 6 
 
 2.65 
 
 65 
 
 172 
 
 i| 
 
 2 
 
 4 
 
 3 
 
 3 JI 
 
 I 10 
 
 6 
 
 7i 
 
 6 
 
 4-54 
 
 65 
 
 295 
 
 i 
 
 1} 
 
 6 
 
 5 
 
 3 9 
 
 i 9 
 
 7i 
 
 7| 
 
 6 
 
 4-54 
 
 65 
 
 295 
 
 i* 
 
 2 
 
 6 
 
 5 
 
 3 9 
 
 i 9 
 
 6 
 
 *f 
 
 6 
 
 5-84 
 
 65 
 
 380 
 
 i 
 
 li 
 
 6 
 
 5 
 
 3 9 
 
 I IO 
 
 ri 
 
 8} 
 
 6 
 
 5-84 
 
 65 
 
 380 
 
 a 
 
 2 
 
 6 
 
 5 
 
 3 9 
 
 I IO 
 
 7i 
 
 6 
 
 10 
 
 4-75 
 
 54 
 
 256 
 
 ii 
 
 2 
 
 5 
 
 4 
 
 5 8 
 
 2 4 
 
 7i 
 
 7 
 
 IO 
 
 6.52- 
 
 54 
 
 ' 352 
 
 i 
 
 2 
 
 6 
 
 5 
 
 5 8* 
 
 2 4 
 
 IO 
 
 7 
 
 IO 
 
 6.52 
 
 54 
 
 352 
 
 2 
 
 2i 
 
 6 
 
 5 
 
 5 9 
 
 2 5 
 
 7i 
 
 8i 
 
 IO 
 
 9.68 
 
 54 
 
 522 
 
 1| 
 
 2 
 
 6 
 
 5 
 
 6 o 
 
 2 4 
 
 9 
 
 8J 
 
 IO 
 
 9.68 
 
 54 
 
 522 
 
 2 
 
 2* 
 
 6 
 
 5 
 
 6 o 
 
 2 4 
 
 71 
 
 roi 
 
 10 
 
 14.08 
 
 54 
 
 760 
 
 I* 
 
 2 
 
 IO 
 
 8 
 
 6 4 
 
 2 8 
 
 9 
 
 10} 
 
 IO 
 
 14.08 
 
 54 
 
 760 
 
 2 
 
 1 
 
 IO 
 
 8 
 
 6 5 
 
 2 8 
 
 12 
 
 ioi 
 
 IO 
 
 14.08 
 
 54 
 
 760 
 
 2i 
 
 3 
 
 10 
 
 8 
 
 6 9 
 
 2 8 
 
 12 
 
 12 
 
 IO 
 
 J 9-37 
 
 54 
 
 1045 
 
 1 
 
 3 
 
 IO 
 
 8 
 
 6 5 
 
 3 o 
 
 12 
 
 14 
 
 10 
 
 26.44 
 
 54 
 
 1427 
 
 2* 
 
 3 
 
 12 
 
 10 
 
 7 o 
 
 3 4 
 
 12 
 
 IS 
 
 IO 
 
 30.38 
 
 54 
 
 1640 
 
 2* 
 
 3 
 
 12 
 
 10 
 
 7 o 
 
 3 4 
 
 14 
 
 15 
 
 10 
 
 30.38 
 
 54 
 
 1640 
 
 2\ 
 
 3 
 
 12 
 
 IO 
 
 7 o 
 
 3 4 
 
 12 
 
 14 
 
 15 
 
 39-70 
 
 40 
 
 1588 
 
 ^ 
 
 3 
 
 12 
 
 10 
 
 7 10 
 
 3 10 
 
 12 
 
 ^5 
 
 15 
 
 45-6i 
 
 40 
 
 1824 
 
 M 
 
 3 
 
 12 
 
 IO 
 
 7 10 
 
 3 10 
 
 14 
 
 IS 
 
 15 
 
 45-6i 
 
 40 
 
 1824 
 
 ai 
 
 3 
 
 12 
 
 IO 
 
 7 10 
 
 3 10 
 
 17 
 
 IS 
 
 15 
 
 45- 61 
 
 40 
 
 1824 
 
 2i 
 
 3i 
 
 12 
 
 IO 
 
 8 o 
 
 3 10 
 
 12 
 
 J 7 
 
 15 
 
 58.66 
 
 40 
 
 2346 
 
 ai 
 
 3 
 
 14 
 
 12 
 
 8 5 
 
 4 i 
 
 14 
 
 i7 
 
 15 
 
 58.66 
 
 40 
 
 2346 
 
 ti 
 
 3 
 
 14 
 
 12 
 
 8 5 
 
 4 i 
 
 14 
 
 19 
 
 15 
 
 73-27 
 
 40 
 
 2931 
 
 2* 
 
 3 
 
 16 
 
 14 
 
 8 o 
 
 4 i 
 
 17 
 
 J 9 
 
 15 
 
 73-27 
 
 40 
 
 2931 
 
 2i 
 
 3i 
 
 16 
 
 14 
 
 8 ii 
 
 4 i 
 
 14 
 
 22 
 
 15 
 
 92.26 
 
 40 
 
 3940 
 
 2i 
 
 3 
 
 16 
 
 14 
 
 8 6 
 
 4 7 
 
 An additional . charge is made for Tobin-bronze piston rods, brass 
 water pistons, bed plates, or for any extras. 
 
 The water end is made of light construction for use on 
 
MODERN RECIPROCATING PUMPS 143 
 
 services where the total water pressure to be pumped against 
 is not over from 35 to 50 pounds per square inch. The ratios 
 of steam cylinders and water pistons are suitable for raising 
 liquids to moderate heights with ordinary steam pressures. 
 This design is intended for use at railway water stations, 
 breweries, distilleries, gas and oil works, tanneries, bleacheries, 
 refineries, etc. The valves in the liquid end are furnished of 
 material suitable for the liquid to be pumped. The liquid 
 ends up to the size 14 X 10 are designed for pressures up to 50 
 pounds per square inch and the larger ends for a pressure of 35 
 pounds per square inch. 
 
 A tank pump used for pumping water from the bottom of 
 a vessel is known as a bilge pump or ballast pump. These 
 pumps are usually placed in a vertical position owing to the 
 lack of room. On account of the small head against which 
 they work the steam cylinder is much smaller than the water 
 cylinder. Fig. 137 shows the form of ballast pump built by 
 Worthington and the covers over the valve chambers show 
 how simple it is to care for the valves. This is an important 
 matter in pumps for marine service where working space is 
 limited. 
 
 Fire Pumps are those installed for fire protection. When 
 these are built according to certain specifications adopted by 
 the Underwriters Associations they are known as under- 
 writers' pumps. Fig. 138 shows such a pump with ample 
 dimensions on the steam end. As will be seen later these 
 pumps have gauges on the steam and water ends, air chambers 
 on the suction and discharge pipes, direct hose connections on 
 the pumps, a relief valve on the discharge main, and a name 
 plate with certain data in regard to the pump on the air 
 chamber. The water parts are of bronze and there are other 
 peculiar features which will be considered. 
 
 THE KNOWLES UNDERWRITER FIRE PUMPS 
 1904 Pattern 
 
 In these underwriter fire pumps the water passages, valve 
 areas, and suction and discharge nozzles are all much larger 
 
144 
 
 PUMPING MACHINERY 
 
 than in any ordinary pump, so a greater amount of water can 
 be discharged without water hammer. Also, the steam and 
 exhaust ports and nozzles are designed so as to give unre- 
 stricted passage to the steam. The pump is " rust proof," 
 
 FIG. 137. Worthington Ballast Pump. 
 (Size 7iXioiXio.) 
 
 and will start instantly, after standing unused for a long time. 
 The piston rods and valve rods are made from Tobin bronze. 
 All stuffing boxes and glands are brass lined. Plungers and 
 plunger sleeves are of composition, but the metals are differ- 
 ently mixed to prevent cutting. This mixture of the metals 
 has been a subject of much experiment and study; it has now 
 
MODERN RECIPROCATING PUMPS 
 
 145 
 
 FIG. 138. Knowles Underwriters' Pump. 
 (Size 18X10X12; 1000 gallons per minute.) 
 
 FIG. 139. Compound Outside End-Packed Pressure Pump. 
 (Size 10X16X4X21.) 
 
146 
 
 PUMPING MACHINERY 
 
 brought great success, as the parts work perfectly upon each 
 other, without friction or impairment. 
 
 Each pump has the following fittings: 
 
 Capacity plate on discharge air chamber; stroke gauge, 
 graduated on each end; vacuum, or suction air chamber; 
 steam gauge, 5 inches diameter; water gauge, 5 inches diameter; 
 relief valve of large capacity; relief valve discharge cone; set 
 of brass priming pipes and special priming valves; required 
 number of 2 j~ inch Ludlow hose valves; one pint sight-feed 
 cylinder lubricator; one pint hand oil pump, all according to 
 the specifications. 
 
 On account of the larger passageways, brass parts and 
 attachments mentioned, the pump necessarily costs more than 
 an ordinary fire pump; but the cost by the gallon discharged 
 is less, since the underwriter pump can deliver a greater quantity 
 of water in the same length of time. It is also much heavier 
 and stronger, with superior workmanship, and better protected 
 from rust and accident. 
 
 Hose valves are threaded only when specially ordered and 
 at extra cost. If these valves are to be threaded a sample 
 thread must be supplied by the purchaser, as there is no estab- 
 lished standard, hose threads varying widely in different local- 
 ities. 
 
 Diameter 
 of Steam 
 Cylinder, 
 Inches. 
 
 Diameter 
 of Water 
 Plunger, 
 Inches. 
 
 Length of 
 Stroke, 
 Inches. 
 
 Normal 
 Capacities 
 Gals, per 
 Minute. 
 
 Number 
 of Fire 
 Streams. 
 
 Steam 
 Pipe, 
 Inches. 
 
 Exhaust 
 Pipe, 
 Inches, 
 
 Suction 
 Pipe, 
 Inches. 
 
 Dis- 
 charge 
 Pipe, 
 Inches. 
 
 14 
 
 7l 
 
 12 
 
 500 
 
 2 
 
 3 
 
 4 
 
 8 
 
 6 
 
 16 
 
 9 
 
 12 
 
 75 
 
 3 
 
 3* 
 
 4 
 
 IO 
 
 8 
 
 18 
 
 10 
 
 12 
 
 IOOO 
 
 4 
 
 4 
 
 5 
 
 12 
 
 8 
 
 2O 
 
 12 
 
 16 
 
 1500 
 
 6 
 
 5 
 
 6 
 
 14 
 
 IO 
 
 Underwriter fire pumps are built with compound steam ends, 
 and power underwriter fire pumps are arranged to run with 
 electric motors. 
 
 Pressure Pumps (Fig. 139) are those which are intended 
 for hydraulic pressure installations in which pressures of 500 
 pounds and over are used to operate presses. Although these 
 
MODERN RECIPROCATING PUMPS 147 
 
 pumps are sometimes of the fly-wheel tpye, the intermittent 
 operation of the pump in many plants makes this type unsuit- 
 able, hence the direct-acting type is used very extensively. 
 This pump is marked by small water cylinders, usually of the 
 outside-packed plunger type. The valves are placed in valve 
 boxes bolted to the water end on account of the complex 
 casting which would result if these parts were integral portions 
 of the water cylinders. 
 
 The Mine Pump (Fig. 140) is a form of pressure pump as 
 it is used to pump water from great depths. The water end is 
 
 FIG. 140. Compound Outside-Packed Mine Pump. 
 
 therefore quite similar in detail to the pressure pump. There 
 are other reasons for the construction of the water ends in 
 small parts. The action of the acid of mine waters is such 
 that parts of the pump are rapidly corroded, hence the use of a 
 number of small parts of similar form makes it possible to 
 renew the part affected and not replace a whole end, as would 
 be necessary if the water end was one complex casting. More- 
 over the number of different parts is reduced to a minimum by 
 having most of the small parts similar for both sides of the 
 pump. This cuts down the number of spare parts to be carried 
 in the storeroom. 
 
 The table below gives the sizes used by the Worthington 
 
148 
 
 PUMPING MACHINERY 
 
 Co. for their Lehigh pattern of pump used with pressures of 
 300 pounds per square inch and over. 
 
 THE WORTHINGTON MINE PUMP 
 Lehigh Pattern For 300 Pounds. 
 
 The Worthington mine pump (Lehigh pattern) is specially 
 designed to withstand the heavy pressures encountered in 
 deep workings. These pumps can be arranged to operate 
 either non-condensing or condensing, and are also made with 
 compound and triple-expansion steam cylinders, where a 
 saving of fuel is desirable. Pumps of this design of larger 
 size or for heavier service can be furnished. 
 
 1 
 
 I 
 
 1 
 
 
 \ 
 
 Minute. 
 
 Size of Pipes for Short 
 Lengths to be 
 Increased as Length 
 Increases. 
 
 Approximate 
 Space Occupied. 
 Feet and Inches. 
 
 *0 ! 
 
 *0 ,fl 
 
 fc& 
 
 CO 
 
 * 
 
 is 
 
 l 
 
 1 
 
 
 
 
 ' 
 
 
 ^ 
 
 
 
 gja 
 c >> 
 .2o 
 
 SB 
 
 II 
 
 Q 
 
 M 
 
 I 
 
 J 1 - 
 
 33 
 
 "o.S 
 
 I s 
 
 Gallons 
 
 ai 
 s 
 
 C/2 
 
 3oi 
 cd 5. 
 
 as 
 
 W 
 
 |4 
 
 f 
 
 H 
 
 | 
 
 jj 
 
 Width. 
 
 16 
 
 6 
 
 IO 
 
 4.89 
 
 43 
 
 210 
 
 i 
 
 3 
 
 6 
 
 5 
 
 9 ii 
 
 5 2 
 
 i8i 
 
 6 
 
 10 
 
 4-89 
 
 43 
 
 2IO 
 
 3 
 
 3i 
 
 6 
 
 5 
 
 IO O 
 
 5 2 
 
 16 
 
 7i 
 
 IO 
 
 7.64 
 
 43 
 
 328 
 
 2i 
 
 3 
 
 6 
 
 5 
 
 9 ii 
 
 5 2 
 
 i8i 
 
 7i 
 
 10 
 
 7.64 
 
 43 
 
 328 
 
 3 
 
 3i 
 
 6 
 
 1 5 
 
 IO O 
 
 5 2 
 
 *7 ' 
 
 6 
 
 IS 
 
 7-35 
 
 32 
 
 235 
 
 ai 
 
 3i 
 
 6 
 
 5 
 
 II IO 
 
 5 2 
 
 20 
 
 6 
 
 IS 
 
 7-35 
 
 32 
 
 235 
 
 4 
 
 5 
 
 6 
 
 5 
 
 II II 
 
 5 2 
 
 17 
 
 7i 
 
 IS 
 
 11.48 
 
 32 
 
 367 
 
 M 
 
 3i 
 
 6 
 
 5 
 
 II IO 
 
 5 2 
 
 20 
 
 7i 
 
 IS 
 
 ii .48 
 
 32 
 
 367 
 
 4 
 
 5 
 
 6 
 
 5 
 
 II II 
 
 5 2 
 
 18 
 
 7 
 
 18 
 
 12 .00 
 
 28 
 
 336 
 
 3 
 
 4 
 
 6 
 
 5 
 
 
 
 22 
 
 7 
 
 18 
 
 12.00 
 
 28 
 
 336 
 
 4 
 
 6 
 
 6 
 
 5 
 
 
 
 18 
 
 8 
 
 18 
 
 IS- 6 7 
 
 28 
 
 438 
 
 3 
 
 4 
 
 10 
 
 8 
 
 
 
 22 
 
 8 
 
 18 
 
 IS-67 
 
 28 
 
 438 
 
 4 
 
 6 
 
 IO 
 
 8 
 
 
 
 An additional charge is made when pumps are fitted with brass plungers 
 and brass-bushed glands. 
 
 To designate the sizes, give the diameters of the steam cylinders and 
 water plungers, and the length of stroke. 
 
 NOTE. One revolution means four strokes, counting both sides. 
 
 When lower pressures are present and the water is not 
 corrosive such a type shown in Fig. 141 may be used. The 
 
MODERN RECIPROCATING PUMPS 
 
 149 
 
 pump is of simple construction, with outside-packed plungers. 
 The following table gives sizes used for this type of pump. 
 
 THE WORTHINGTON PACKED-PLUNGER PUMP 
 Scranton Pattern For 250 Pounds Water Pressure. 
 
 6 
 
 1 
 
 g 
 
 | 
 
 c 
 
 & 
 
 Minute. 
 
 Size of Pipes for Short 
 Lengths to be 
 Increased as Length 
 Increases. 
 
 Approximate 
 Space Occupied, 
 Feet and Inches. 
 
 11 
 
 Is 
 
 C/3 
 
 "o 
 
 4 
 
 la 
 
 1 
 
 
 
 
 ^ 
 
 
 . 
 
 
 
 oj C 
 
 6^ 
 
 D C 
 
 6.3 
 
 3. 
 
 1! 
 
 3 3 
 
 "o.S 
 
 1 
 
 II 
 
 is. 
 
 ^4 PM 
 
 C OJ 
 
 > O. 
 
 xj 
 
 M 
 
 .d 
 
 
 
 
 
 9* 
 
 | S 
 
 cs 
 
 IS 
 
 gs 
 
 y'd, 
 
 " 
 
 C 
 flj 
 
 12 
 
 Q 
 
 H 
 
 _1 
 
 o 
 
 M 
 
 
 (73 
 
 
 
 V) 
 
 Q 
 
 J 
 
 ^ 
 
 14 . 
 
 84 
 
 10 
 
 9-56 
 
 54 
 
 Sl6 
 
 a | 
 
 3 
 
 8 
 
 6 
 
 9 8 
 
 3 2 
 
 16 
 
 84 
 
 10 
 
 9-56 
 
 54 
 
 5l6- 
 
 *4 
 
 3 
 
 8 
 
 6 
 
 9 9 
 
 3 10 
 
 18} 
 
 8| 
 
 10 
 
 9-56 
 
 54 
 
 516 
 
 3 
 
 3^ 
 
 8 
 
 6 
 
 9 10 
 
 4 o 
 
 16 
 
 10} 
 
 10 
 
 J 3-95 
 
 54 
 
 753 
 
 *i 
 
 3 
 
 10 
 
 8 
 
 10 9 
 
 3 10 
 
 184 
 
 lo i 
 
 10 
 
 J3 -95 
 
 54 
 
 753 
 
 3 
 
 34 
 
 10 
 
 8 
 
 10 9 
 
 4 o 
 
 18$. 
 
 12 
 
 10 
 
 19. 16 
 
 54 
 
 1035 
 
 4 
 
 34 
 
 12 
 
 IO 
 
 ii i 
 
 4 o 
 
 20 
 
 I 2 
 
 IO 
 
 19.16 
 
 54 
 
 1515 
 
 H 
 
 5 
 
 12 
 
 TO 
 
 11 2 
 
 4 2 
 
 T 7 
 
 8* 
 
 '5 
 
 14.14 
 
 40 
 
 565 
 
 4 
 
 3^ 
 
 8 
 
 6 
 
 10 5 
 
 3 ii 
 
 20 
 
 84 
 
 '5 
 
 14.14 
 
 40 
 
 5*5 
 
 a| 
 
 5 
 
 8 
 
 6 
 
 10 6 
 
 4 2 
 
 17 
 
 10} 
 
 15 
 
 20.83 
 
 40 
 
 833 
 
 4 
 
 34 
 
 10 
 
 8 
 
 TI 6 
 
 3 ii 
 
 20 
 
 10} 
 
 15 
 
 20.83 
 
 40 
 
 833 
 
 4 
 
 5 
 
 10 
 
 8 
 
 ii 8 
 
 4 i ^ 
 
 20 
 
 12 
 
 15 
 
 28.78 
 
 40 
 
 1151 
 
 5 
 
 5 
 
 12 
 
 10 
 
 ii 9 
 
 4 3 
 
 A sinking pump is one used in mines for dropping into a 
 shaft which has to be freed of water collected during a time 
 of disuse or which has come in when a subterranean stream 
 has been cut into. The pump is subject to rough usage and 
 must be applied quickly and easily to the timbers of the shaft. 
 Fig. 142 shows one form of sinker as built by the Prescott 
 Company. This pump is mounted on a frame which can be 
 hung from the mine timbers. The working parts are protected 
 from injury from falling objects by the projections of the cyl- 
 inder and steam chest. The chains or rods connecting the 
 link serve to support the pump from the derrick. The water- 
 valve boxes are accessible for repairs and the outside-packed 
 plungers make it possible to be sure that there is no leakage 
 from side to side. When electricity is applicable the electric- 
 
150 
 
 PUMPING MACHINERY 
 
 sinking pump of the form shown in Fig. 143 is used. This 
 pump is driven by the motor through a double reduction gear 
 and the plungers are of the form mentioned earlier. The 
 method of lowering, the protection of the parts from mechanical 
 injury and the motor from water, as well as the arrangement of 
 the valves for quick inspection are clearly seen. In each of 
 these sinking pumps the suction is connected to the bottom 
 of the pump and the discharge may be taken from the right 
 
 FIG. 141. Mine Pump Scranton Pattern. 
 (Size 16 and 25X14X15.) 
 
 or the left. These pumps are so built that they may run if 
 flooded by water, as such a condition may arise at any time. 
 The table on page 151 gives the sizes used with the Knowles 
 electric duplex sinking pump. 
 
 THE KNOWLES VERTICAL DUPLEX ELECTRIC SINKING 
 
 PUMPS 
 
 Double- Acting Outside Center- Packed Plunger Pattern. 
 
 This pump is light, compact, efficient, and of good capacity, 
 and not liable to damage from moisture or hard usage. The 
 entire motor mechanism is inclosed in a tight casing, but every 
 
MODERN RECIPROCATING PUMPS 
 
 151 
 
 1 
 
 
 c 
 
 U 
 
 
 
 
 
 
 
 
 Dimensions over all 
 
 o 
 
 c 
 
 <8 
 
 "3 
 
 u 
 
 
 
 
 . 
 
 _a 
 
 g 
 
 o 
 
 
 in Feet and Inches. 
 
 <j B 
 
 *l 
 
 1 
 
 a 
 33 
 
 C/3 
 
 is 
 
 -2 
 .2 
 
 L 
 
 a 
 cS! 
 
 L 
 
 
 o 
 
 s 
 
 1.1 
 
 
 
 
 
 V C 
 
 o 
 
 CJ ** 
 
 ^ ^ 
 
 c*S 
 
 c 3 
 
 OJ3 
 
 rtX 
 
 ^ I., 
 
 O t^ 
 
 ^ ^ 
 
 pi 
 
 
 +J 
 
 Ej3 
 
 T* 
 
 ._< 0) 
 
 "o.S 
 
 J 
 
 JJ 
 
 g o 
 
 Xo 
 
 .2 g 
 
 2 $ 
 
 D 1 
 
 $1 
 
 *c 
 
 
 .<s 
 
 2 
 
 fefc 
 
 
 
 
 
 SHH 
 * 
 
 1 
 
 * rl ^ 
 
 rtO 
 
 PHp/ 
 
 
 
 a 
 
 <n 
 
 
 Q 
 
 00 
 
 > 
 
 M 
 
 O 
 
 O 
 
 09 
 
 
 
 ^ 
 
 ffi 
 
 ^ 
 
 Q 
 
 W 
 
 3 
 
 6 
 
 400 
 
 55 
 
 73 
 
 40 
 
 3 
 
 2 
 
 3 5 to i 
 
 
 IO 
 
 3 o 
 
 2 9 
 
 8 8 
 
 4 
 
 6 
 
 20O 
 
 80 
 
 1.30 
 
 104 
 
 4 
 
 3 
 
 2 . 7 to i 
 
 
 7i 
 
 3 o 
 
 2 9 
 
 8 8 
 
 4 
 
 6 
 
 500 
 
 80 
 
 1.30 
 
 104 
 
 4 
 
 3 
 
 3 . 7 to i 
 
 5 
 
 20 
 
 3 8 
 
 3 6 
 
 9 4 
 
 
 
 
 
 
 
 
 
 
 o 
 
 
 
 
 
 5 
 
 6 
 
 300 
 
 75 
 
 2 .04 
 
 153 
 
 5 
 
 4 
 
 3 . 7 to i 
 
 a 
 
 20 
 
 3 8 
 
 4 2 
 
 10 2 
 
 5 
 
 6 
 
 4OO 
 
 80 
 
 2 .04 
 
 163 
 
 5 
 
 4 
 
 3 . 7 to i 
 
 *-* i 
 
 
 
 3 
 
 3 8 
 
 4 2 
 
 10 4 
 
 6 
 
 6 
 
 200 
 
 86 
 
 2-93 
 
 252 
 
 6 
 
 5 
 
 3 . 7 to i 
 
 d 
 
 0) 
 0) 
 
 20 
 
 3 9 
 
 4 2 
 
 10 4 
 
 6 
 
 6 
 
 2 5 
 
 86 
 
 2.93 
 
 252 
 
 6 
 
 5 
 
 3 7 to i 
 
 g. 
 
 2 5 
 
 3 9 
 
 4 2 
 
 10 4 
 
 6 
 
 8 
 
 400 
 
 62 
 
 3-91 
 
 242 
 
 6 
 
 5 
 
 3 5 to i 
 
 g 
 
 50 
 
 4 6 
 
 4 10 
 
 II 2 
 
 64 
 
 8 
 
 3OO 
 
 70 
 
 4-59 
 
 321 
 
 6 
 
 5 
 
 3 5 to i 
 
 w 
 O 
 
 5 
 
 4 6 
 
 4 10 
 
 112 
 
 
 
 
 
 
 
 
 
 
 H 
 
 
 
 
 
 7 
 
 8 
 
 300 
 
 68 
 
 5-32 
 
 362 
 
 8 
 
 6 
 
 3 5 to i 
 
 
 5 
 
 4 6 
 
 4 10 
 
 112 
 
 8 
 
 8 
 
 300 
 
 58 
 
 6.96 
 
 403 
 
 8 
 
 6 
 
 3 5 to i 
 
 
 50 
 
 4 6 
 
 4 10 
 
 112 
 
 part of motor and pump is readily accessible for examination 
 or repairs. This apparatus will stand hard usage without injury. 
 The main frame is designed especially to receive the specified 
 type and size of motor. As the plungers and piston rods are 
 packed from the outside, any leakage is readily noticeable. 
 Each pump is furnished with a discharge air chamber. In 
 sending inquiries or orders, always state as fully as possible the 
 intended service and requirements. 
 
 Centrifugal pumps are now being used for this purpose 
 at times. 
 
 The Water-Works Pump will be considered in a later 
 chapter in more^ detail, but at this point two simple forms will 
 be described. It is this kind of pump which has reached the 
 highest form of development, because such machines are used 
 for the sole purpose of raising water in large quantities and 
 continuously. They may be of the direct-acting type (Fig. 144) 
 or the fly-wheel type (Fig. 145). . They are usually compound or 
 triple expansion on the steam end, and precautions are taken 
 to cut down all losses. The pumps are made with great refine- 
 ment of parts, as they are usually handled by skilled engineers. 
 
152 PUMPING MACHINERY 
 
 In most cases there is a possibility of adjusting the steam- 
 
 FIG. 142. Prescott Steam Sinking 
 Pump. 
 
 FIG. 143. Electric Sinking Pump. 
 (Size 6X6.) 
 
 valve gearing, and the water valves may be examined with 
 ease. These pumps vary in size from those handling 50,000 
 
MODERN RECIPROCATING PUMPS 153 
 
 gallons per twenty-four hours to those handling 30,000,000 
 
 FIG. 144. Direct-Acting Water- Works Pump. 
 (Size 1 6 and 25X14^X18.) 
 
 FIG. 145. Water- Works Pump. 
 
 gallons in the same time. Such pumps are usually rated in 
 this manner gallons per twenty- four hours. 
 
154 
 
 PUMPING MACHINERY 
 
 In connection with 
 water works it becomes 
 necessary to lift water from 
 artesian or deep wells, and 
 for this work special deep- 
 well pumps have been 
 designed. These consist 
 usually of a steam cylinder 
 mounted at the top of the 
 deep well (Fig. 146). The 
 piston rod of such a pump 
 is usually made into a 
 plunger passing through a 
 stuffing box at the base of 
 the pump frame. This 
 plunger is then connected 
 to a bucket piston by 
 wooden rods with iron 
 joints. The bucket may 
 be several hundred feet 
 below the surface. This 
 distance is fixed by the 
 height at which the water 
 stands in the well when the 
 pump is in operation. The 
 foot valve is placed at the 
 end of the pipe line which 
 forms the pump barrel. 
 This foot valve is placed 
 in position by the pump 
 rods, although in some 
 cases it is lowered into 
 position by other means, 
 its weight holding it after 
 placing. The pump bucket 
 as shown by the figure is 
 
 FIG. 146. Fairbanks-Morse Deep- Well Pump, packed by means of CUp 
 
MODERN RECIPROCATING PUMPS 155 
 
 leathers and the foot valve and head valve shown are of the 
 ball type. This last section of pipe casing seen in the figure 
 may be of heavy drawn brass tubing to make a proper pump 
 barrel. In arranging the plunger at the top of the wooden 
 sucker rods, it should be made of net area equal to one-half 
 the net area of the bucket so as to discharge water on both 
 strokes of the pump. 
 
 The pump frame is so constructed that it may be moved 
 over from the top of the well in order that the rods or piping 
 can be taken up. This is done by hoisting the rods or casings 
 until one section is above ground, then clamping the line by 
 the bottom section until the upper section is unscrewed and 
 removed, when the operation is repeated. For this reason it 
 is necessary to construct the pump house over a deep well 
 with sufficient head room to permit the removal of casings 
 or rods; this is usually about 20 or 25 feet. The tables below 
 give data from catalogue of the Fairbanks-Morse Co. in regard 
 to their pumping engines. 
 
 THE FAIRBANKS-MORSE ARTESIAN- WELL ENGINE 
 
 This engine is placed directly over the well, and the piston 
 rod is continued to the required depth and connected to the 
 pump piston. The steam valve is perfectly controlled, and 
 the speed of the engine on both up and down strokes is uniform. 
 The apparatus may be run at a high speed without excessive 
 shock or jar. 
 
 These engines will pump from the deepest wells, forcing 
 the water in a steady stream into an elevated tank or 
 other reservoir. To remove the pump rods and pistons, 
 the bolts which connect base to the frame are loosened and 
 the steam cylinder and uprights are drawn back on the 
 base by a screw. The upper displacing cylinder discharges 
 one-half the volume pumped on the down stroke, thus tend- 
 ing to balance the machine and insure a smooth and easy 
 action. 
 
156 
 
 PUMPING MACHIXERY 
 
 SIZE OF ENGINE. 
 
 SIZE OF PIPE. 
 
 
 Diameter of 
 Steam Cylinder. 
 
 Length of 
 Stroke. 
 
 Steam. 
 
 Exhaust 
 
 Inches. 
 
 6 
 
 1 8 
 
 I 
 
 li 
 
 48 X20 
 
 8 
 
 24 
 
 *l 
 
 I* 
 
 5lX23 
 
 10 
 
 36 
 
 *! 
 
 2 
 
 62 X2 5 
 
 12 
 
 36 
 
 i* 
 
 2 
 
 62 X25 
 
 14 
 
 36 
 
 2 
 
 3 
 
 73 X 3 4i 
 
 16 
 
 36 
 
 2 
 
 3 
 
 73 X 3 4i 
 
 BRASS ARTESIAN-WELL CYLINDER 
 
 Inside 
 Diameter. 
 
 Length of 
 Stroke. 
 
 Capacity per 
 Stroke in 
 Gallons. 
 
 Outside 
 Diameter of 
 Caps. 
 
 Top and 
 Bottom Con- 
 necting Pipe in 
 Inches. 
 
 2i 
 
 18 
 
 3 1 
 
 3& 
 
 2^ 
 
 2f 
 
 18 
 
 .46 
 
 3i 
 
 3 
 
 3i 
 
 18 
 
 .64 
 
 4ik 
 
 3i 
 
 3i 
 
 18 
 
 .86 
 
 5* 
 
 4 
 
 
 1 
 
 24 
 
 .61 
 
 3i 
 
 3 
 
 3i 
 
 24 
 
 .86 
 
 43^ 
 
 3* 
 
 3i 
 
 24 
 
 .84 
 
 5* 
 
 4 
 
 4\ 
 
 \ 
 
 24 
 
 47 
 
 si 
 
 4) 
 
 4! 
 
 24 
 
 .84 
 
 6J 
 
 5 
 
 si 
 
 24 
 
 .68 
 
 7i 
 
 6 
 
 3; 
 
 
 36 
 
 .29 
 
 4i$ 
 
 3i 
 
 
 
 36 
 
 7 1 
 
 si 
 
 4 
 
 4i 
 
 
 36 
 
 2.21 
 
 5l 
 
 4i 
 
 4^ 
 
 
 36 
 
 2. 7 6 
 
 61 
 
 5 
 
 si 
 
 
 36 
 
 4-O2 
 
 7i 
 
 6 
 
 TABLE OF WOOD SUCKER RODS 
 
 Diameter of Rod in Inches, 
 it 
 
 Adapted for Working Barrels of a Diameter 
 2! to 4 1 
 4l to sf 
 5t to 7! 
 
 A gas engine geared to a deep-well pump is shown in Fig. 
 147, in which the same points are to be noted as in Fig. 146, 
 while Fig. 148 illustrates the method of using a horizontal 
 cylinder for this purpose, as proposed by Davidson. 
 
 A form of deep-well power pump used in the West is shown 
 by Fig. 149. In this, the Luitwieler pump, motion is given 
 
MODERN RECIPROCATING PUMPS 157 
 
 FlG. 147. Gas Engine Drive for Deep-Well Pump. 
 
 FIG. 148. Davidson Horizontal Steam Cylinder for Deep- Well Pump. 
 
158 
 
 PUMPING MACHINERY 
 
 to the spur wheel A by a belt, steam 
 engine, gas engine, or electric motor, 
 and this drives the gear B. On the 
 gear shaft are two cams arranged 
 opposite each other and so made that 
 they have a lifting motion during 53 
 per cent of the rotation, while the 
 \ descending motion occupies 47 per 
 J cent of a revolution. These cams 
 / drive cross-heads by means of fric- 
 J tionless rollers. The left-hand cross- 
 head is attached at its lower end to 
 a hollow pump rod and the right-hand 
 one at its upper end to a solid rod 
 passing through the hollow one. The 
 cross-heads are guided to prevent 
 turning. The hollow rod passes 
 through a stuffing box and has at its 
 upper end a stuffing box for the solid 
 rod. These rods are each attached 
 to a bucket piston as shown by the 
 figure. The outer rod is made of 
 larger bore than the diameter of the 
 solid rod so that there is no friction, 
 provided the rod does not whip and 
 is vertical. When the lower piston 
 is descending the upper one is ascend- 
 ing and forcing the water out, while 
 water will flow in between the two 
 pistons to fill the space formed there. 
 Before the upper piston reaches the 
 top of its stroke the lower piston 
 starts 'to move upward and when the 
 upper piston begins to descend the 
 lower piston has sufficient speed up- 
 
 F.C. i 49 .-Luitwieler Deep-Well Ward tO ke6 P the COlumn f Water in 
 
 Pump. the pipe in motion at the same ppeed. 
 
MODERN RECIPROCATING PUMPS 159 
 
 The water lifted by the lower bucket is forced through the 
 upper bucket. Before the lower bucket ceases to move upward 
 the upper one has completed its down stroke and has started 
 
 FIG. 150. Luitwieler Quadruplex Pump. 
 
 back again ready to continue the upward motion of the water 
 in the pipe as soon as the lower bucket begins to diminish its 
 upward rate and reverse its action. 
 
 The buckets, rods, and cross-heads balance each other 
 quite well, so that there is a fairly uniform torque on the driving 
 
ICO PUMPING MACHINERY 
 
 mechanism. The action is different from the ordinary deep- 
 well pump and its differential plunger in that not only the 
 stream above ground is moving but also that in the well barrel. 
 The pump lifts water on the suction side during one-half revolu- 
 tion, as in the case of the ordinary pump, but the discharge of 
 this through the well barrel is distributed over the whole revo- 
 lution. If it were possible to put the upper plunger of the deep- 
 
 FIG. 151. Triplex Pump. 
 (Size 5X8.) 
 
 well pump down near the working barrel the same action would 
 occur in that pump. The arrangement of cams may be such 
 that the motion of the water is uniform, the accelerating period 
 of the motion of one occurring during the time in which the 
 other is changing its motion. 
 
 The same arrangement of cams may be applied to duplex 
 or quadruplex pumps, where the motion given by the ordinary 
 crank is not sufficiently uniform. Such an arrangement is 
 
MODERN RECIPROCATING PUMPS 
 
 161 
 
 shown in Fig. 150. In this case the effort is made to sub- 
 merge the pump barrels to cut down all suction lift. 
 
 This leads to a large class of pumps known as power pumps. 
 Power pumps are those in which the pump proper is driven 
 through gears or belts. These pumps are of various forms, 
 some horizontal, some vertical, some with plungers, some with 
 
 FIG. 152. Horizontal Duplex Power Pump. 
 (Size 6X12.) 
 
 pistons, some with single cylinders, some with double cylinders, 
 and some triplex. The name triplex usually designates a 
 vertical, three-cylinder pump. In many cases it is of the 
 plunger type. Fig. 151 illustrates this form. The large valve 
 case with a simple arrangement for examining the valves, the 
 inlet and outlet openings, the short connections to the cyl- 
 inder, the method of supporting the side thrust, the method of 
 
162 
 
 PUMPING MACHINERY 
 
 packing, and the arrangement of cranks and gears are clearly 
 shown. The sizes of such pumps are given in the table below. 
 
 DEANE TRIPLEX VERTICAL SINGLE-ACTING POWER PUMP 
 
 Size. 
 
 I 
 
 4 
 
 Capacity. 
 
 Pipe 
 bizes. 
 
 Tight and 
 Loose Pulleys. 
 
 j 
 
 Approximate 
 Dimen. in Feet 
 and Inches. 
 
 
 
 
 c 
 
 <r. U 
 
 
 
 
 
 
 
 ' 
 
 0) 
 
 
 
 
 
 VM 
 
 eg 
 
 II 
 
 S 
 
 . 
 
 
 &> 
 
 . 
 
 
 
 O 
 
 
 
 
 il 
 
 5 
 
 Length c 
 
 Stroke 
 
 
 Gallons i 
 Revolt 
 
 Revoluti 
 per Mir 
 
 Il 
 
 O 
 
 Suction. 
 
 Dischar^ 
 
 Diamete 
 
 d 
 
 1 
 
 "3 
 PQ 
 
 Ratio of 
 
 ! 
 
 Width. 
 
 Height. 
 
 3 
 
 8 
 
 290 
 
 -73 
 
 45 
 
 33 
 
 3 
 
 24 
 
 3 
 
 54 
 
 D 
 
 4.7-1 
 
 3 10 
 
 3 9 
 
 5 7 
 
 4 
 
 8 
 
 290 
 
 1.30 
 
 45 
 
 58 
 
 3 
 
 24 
 
 3 
 
 54 
 
 D 
 
 5-o-i 
 
 4 8 
 
 4 ii 
 
 6 5 
 
 4 
 
 8 
 
 270 
 
 1.30 
 
 45 
 
 58 
 
 3 
 
 24 
 
 30 
 
 54 
 
 D 
 
 4.7-1 
 
 3 10 
 
 3 9 
 
 5 7 
 
 44 
 
 8 
 
 290 
 
 1-65 
 
 45 
 
 74 
 
 4 
 
 3 
 
 3 
 
 
 D 
 
 5-o-i 
 
 4 8 
 
 4 ii 
 
 6 5 
 
 5 
 
 8 
 
 254 
 
 2.04 
 
 45 
 
 91 
 
 4 
 
 3 
 
 30 
 
 74 
 
 D 
 
 5-0-1 
 
 4 8 
 
 4 ii 
 
 6 5 
 
 5 
 
 8 
 
 180 
 
 2 .04 
 
 45 
 
 91 
 
 4 
 
 3 
 
 3 
 
 6* 
 
 D 
 
 4-7-1 
 
 3 10 
 
 3 9 
 
 5 7 
 
 54 
 
 8 
 
 209 
 
 2. 4 6 
 
 45 
 
 no 
 
 4 
 
 3 
 
 3 
 
 74 
 
 D 
 
 5-o-i 
 
 4 8 
 
 4 ii 
 
 6 5 
 
 54 
 
 8 
 
 150 
 
 2 .46 
 
 45 
 
 I 10 
 
 4 
 
 3 
 
 3 
 
 64- 
 
 D 
 
 4-7-1 
 
 3 10 
 
 3 9 
 
 5 7 
 
 6 
 
 8 
 
 200 
 
 2-93 
 
 4b 
 
 132 
 
 4 
 
 3 
 
 30 
 
 74 
 
 D 
 
 5-o-i 
 
 4 8 
 
 4 ii 
 
 6 5 
 
 6 
 
 8 
 
 125 
 
 2-93 
 
 45 
 
 132 
 
 4 
 
 3 
 
 3 
 
 
 D 
 
 4-7-i 
 
 3 10 
 
 3 9 
 
 5 7 
 
 64 
 
 8 
 
 150 
 
 3-44 
 
 45 
 
 154 
 
 I 5 
 
 2 4 
 
 36 
 
 74 
 
 D 
 
 5-o-i 
 
 4 8 
 
 4 ii 
 
 6 5 
 
 64 
 
 8 
 
 107 
 
 3-44 
 
 45 
 
 
 
 2 4 
 
 3 
 
 64 
 
 D 
 
 4-7-1 
 
 3 10 
 
 4 i 
 
 5 7 
 
 7 
 
 8 
 
 150 
 
 4.00 
 
 45 
 
 180 
 
 2 5 
 
 2 4 
 
 36 
 
 84 
 
 D 
 
 5-o-i 
 
 4 84 ii 
 
 6 5 
 
 7 
 
 8 
 
 9 2 
 
 4.00 
 
 45 
 
 180 
 
 2 5 
 
 2 4 
 
 30 
 
 74 
 
 D 
 
 4-7-1 
 
 3 10 
 
 4 i 
 
 5 7 
 
 8 
 
 8 
 
 IOO 
 
 5-22 
 
 45 
 
 235 
 
 2 6 
 
 * 5 
 
 36 
 
 84 
 
 D 
 
 5-o-i 
 
 4 8 
 
 4 ii 
 
 6 5 
 
 1 D indicates double belt. 2 Flanged connections. 
 
 Above pumps may be safely run at somewhat greater speed than listed. 
 
 The pumps listed above have outside-packed plungers which 
 work through deep-stud gland stuffing boxes. All sizes have 
 cross-heads for guiding the travel of the plungers, and approved 
 means provided for taking up wear. The bearings are of 
 ample diameter and length. The gearing is a special mixture 
 of steel and iron, and the teeth are machine cut. The con- 
 necting rods are provided with adjustable boxes at both ends, 
 the adjustment being made by means of a wedge and screw. 
 
 Pumps of this type are used for general service and tank 
 service; as boiler feeders, elevator pumps, and water works. 
 
 This type of pump is built with 3-inch, 6-inch, 8-inch, and 
 
MODERN RECIPROCATING PUMPS 
 
 163 
 
 12-inch stroke, each class having the same general character- 
 istics as described above. 
 
 Fig. 152 shows a horizontal duplex power pump with 
 pistons. There are many different forms of this type of pump, 
 but these two will illustrate 'the ordinary styles. 
 
 Fig. 153 illustrates the power head for a deep-well pump, 
 
 FIG. 153. Power Head for Deep- Well Pump. 
 
 giving the method of operating such a pump with belting. 
 An electric motor could be attached by a coupling or belt 
 to the shaft carrying the belt pulley. The figure shows the 
 method of removing the top gearing from the well when it is 
 necessary to remove rods or casing. 
 
 A new form of power pump usually driven direct from 
 an electric motor without the use of gears to cut down the 
 
164 
 
 PUMPING MACHINERY 
 
 speed is the express pump. Such a pump, Fig. 154, is arranged 
 to run at high rotative speeds of about 200 to 300 R.P.M. The 
 Riedler pump of the Allis-Chalmers Company is arranged to 
 permit this high speed by having a clear direct passage for the 
 water through large valves and by making the discharge valve 
 positive in its closing. The water enters through the annular 
 space AB from the vacuum air chamber C. It is then dis- 
 charged into the air chamber K through the valve E. This 
 
 FIG. 154. Allis-Chalmers' Duplex Riedler Express Pump 
 
 valve is closed positively by means of an eccentric on the 
 shaft. The plunger GI is of one-half the area of G so that 
 water is passed back and forth through F, giving a discharge 
 on each stroke, although there is only one suction stroke to 
 each revolution. 
 
 The ring D, carried from the back of the plunger, closes 
 the annular valve when the plunger reaches the end of the 
 suction stroke, thus cutting down the chance of slip from the 
 slow action of the valve. 
 
MODERN RECIPROCATING PUMPS 165 
 
 The suction air chamber C insures a supply of water at all 
 times. 
 
 The express pumps of Riedler were among the earliest of 
 this type, but in later years there have been many new forms 
 introduced. In some cases the valves have been spring con- 
 trolled, eliminating the necessity of the valve gearing. 
 
 The simplex pump, the water- works pump, the centrifugal 
 and air-lift pumps, air pumps for condensers, and other special 
 pumps will be considered in later chapters. 
 
CHAPTER IV 
 SIMPLEX PUMPS 
 
 THE pump containing one steam cylinder and one water 
 cylinder is known as a simplex pump to distinguish it from 
 the duplex pump, so largely described in the previous chapter. 
 This type was the original form used by Worthington, and it 
 has been worked on by many inventors. It is the purpose 
 of this chapter to describe a number of the well-known types 
 which will show the general manner of solving the problem 
 of getting a definite reversal of the pump at the end of the 
 stroke no matter what the speed of the pump is. This has 
 been the aim of all simplex-pump designers. It is a simple 
 matter to reverse the action when the pump is moving rapidly, 
 but for slow action there must be some auxiliary apparatus, 
 as the spring of the first Worthington pump, Fig. 55. 
 
 The simplex pump is undoubtedly more complex than the 
 duplex pump, but with it there is little danger of short stroking. 
 In the duplex pump the valves must be carefully adjusted, as 
 the motion of one piston controls the action of the valve of 
 the other side. The simplex pump in many cases may be run 
 at higher speeds than are common with duplex pumps. The 
 loss due to the' clearance space is a total loss in direct-acting 
 pumps, as there is no expansion, and for this reason the clear- 
 ance should be made as small as possible and no short stroking 
 should be allowed. 
 
 The Cameron Pump (Fig. 155) is one of the simpler forms 
 of these pumps. As shown in the figure the B- valve G is ad- 
 mitting steam to the right-hand side of cylinder A and the 
 piston is being driven to the left. When the piston C over- 
 runs the steam passage at its left end the exhaust steam is 
 cushioned partially as it can only pass through a groove at the 
 
 166 
 
SIMPLEX PUMPS 
 
 167 
 
 top of the cylinder to the passage. The piston C strikes the pin 
 on the cylindrical valve 7, allowing the steam to the left of 
 the piston F to escape through E into a passage leading to the 
 exhaust. The steam which has leaked- into the space at 
 the right of the auxiliary piston F, through the small hole in the 
 center of the head, drives this to the left, moving with it 
 the valve G, and reverses the pump. When the piston C moves 
 to the right, live steam entering behind / through the passage K 
 
 FIG. 155. Cameron Pump. 
 
 forces the valve / over, closing the passage to the left of the 
 cylinder in which F travels while the steam which leaks through 
 the small hole in the end of F builds up the pressure for the 
 next reversal, when the right-hand valve / is opened. The 
 auxiliary piston F controls the valve G. It is made of two 
 hollow pistons with holes in the ends. The steam which leaks 
 through these openings represents the cost of operating the 
 valve. This may be an appreciable amount, as the steam 
 may fill the space at the left-hand end in the figure before 
 
168 PUMPING MACHINERY 
 
 the exhaust occurs. With a small hole, however, there need 
 not be much steam used, as it is only the steam on the smaller 
 end which is needed to reverse the valve. The handle H pro- 
 jecting from the side of the steam chest serves to move the 
 valve back and forth in starting. 
 
 The figure shows the arrangement of the suction and dis- 
 charge valves. These valves are controlled by springs, and 
 by taking out a common spindle the valves on one side may be 
 removed. The clear passage for the suction is seen and the 
 short length of this pump in comparison with its stroke is 
 important where space is necessary. 
 
 The table on p. 169 gives the sizes of this pump used for 
 boiler feeding as arranged by the makers. 
 
 
 FIG. 156 Cameron Pump. 
 
 CAMERON PISTON PUMP REGULAR BOILER-FEED 
 PATTERN 
 
 The main difficulty met with in fixing on the proper size 
 of pump to recommend is that the horse power of the boiler 
 for which the pump is required is about all the information 
 furnished. The expression " horse power," as applied to 
 boilers, is a very indefinite term; what should be given, if 
 
SIMPLEX PUMPS 
 
 169 
 
 -a 
 
 vO O rj-t^vocsj O vovo 
 CO w LO rj" vo CN CN CM CM 
 H CN CN co co "^t" vo vo t^* 
 
 N O O vo O r^ O 
 
 O t^. M CN O <N "<$ 
 
 CM t^ O M CO 00 M 
 
 H M CN CO CO ^ VO 
 
 ON O M <N 
 
 00 O M "fr 
 
 O O r- M 
 CO CO fO *t 
 
 XXXXXXXXXXXXXXXXX 
 
 01 O O IN t^ r* oo OC co 1*0 ~^t" ^ <^5 M M O ^ 
 ro rj" ^t" ^t" "^" ^" IO IO IO O O 'O t^* OO 00 s * O 
 
 10 10 10 o oo 
 
 OQ 
 
 10100000 O 
 
 i 
 
 XDn 
 
 w 
 
 BO) 
 
 1.S 
 
 000000*0000000000 
 
 \O O^ ^t" O^ 'O 'to (N O O 1 O O O O O O O 
 M M c^ Cv] co ^ 1 O l>* O ro ^O O fO O 
 
 rt fito 
 
 CJ O 
 
 I 00 
 
 COOOOOOOvoOOOOi-iOOOt^ 
 H CN tovoiOvOOO O voOO coO VOOO 
 
 sl-s 
 
 ^" ^O ^O t^ t^* t^ CN Cvi iN ro co ^O fO OO OO OO O 
 
 W W CS CO CO ^ CO ^J" ^" iO VO !> CO ON O O C4 
 
 M M M 
 
 Mb* 
 
 to "^ 1 O IO 'LO vO vO O t^- t"^- OO O CN Tj~ ^" vQ ^O 
 10 10 O *o O l o 10 10 O O O O O 
 
 CO vN 1 O O ^" O CS Tf" M LO O ^ CO 
 
 OOOOOOOOO^oioioO 
 
 oo w TJ-IO\O M M cooo CNr^.^o 
 
 QMC^C S 1CSCOCO<O 1 OLO- 
 
 
 
 ill 
 
 0*0 M 
 
 
170 
 
 PUMPING MACHINERY 
 
 possible, is the quantity of feed water required, and a pump 
 which will supply this quantity at about one-half its rated 
 capacity at ordinary speed will be right for cold water, and say 
 one- third speed for hot water. 
 
 In feeding hot water the pump should be placed below the 
 source of supply. The first six sizes are furnisheu wiin hand- 
 lever attachment when so desired. 
 
 An outside-packed type of the Cameron pump is shown 
 
 FIG. 157. Marsh Steam Pump. 
 (8X5X10, 8X5X12, 10X6X12.) 
 
 in Fig. 156. The view not only shows the type of water end, 
 but also shows the reversing valve cylinders on the steam 
 cylinder heads and the reversing handle. 
 
 The Marsh Pump (Fig. 157) is another one using a steam- 
 thrown valve. In the picture shown the piston travels to- 
 the right while the steam enters the cylinder through the annu- 
 lar space around the spindle valve and passes into the steam 
 
SIMPLEX PUMPS 
 
 171 
 
 passage. Should the water resistance on the water piston 
 decrease and the steam piston respond suddenly to this, the 
 pressure in the cylinder would fall and with it the pressure 
 on the left of the disc on the left end of the spindle-slide valve. 
 The pressure on the right-hand side of this disc would then 
 move the spindle, cutting down the steam supply and throttling 
 the exhaust, thus dampening the action of the steam piston. 
 
 FIG. 158. Marsh Pump. 
 
 When the piston reaches the end of its stroke it overrides 
 the small port leading to the right-hand end of the valve piston, 
 and the live steam which is introduced into the central space 
 between the two parts of the steam piston, then passes behind 
 the right-hand side of the right disc on the valve spindle, 
 and forces it to the left. This admits steam to the right-hand 
 end of the steam cylinder and opens the left-hand end to the 
 
172 
 
 PUMPING MACHINERY 
 
 exhaust. It is clear from the figure that the two faces of the 
 right-hand disc are exposed to exhaust pressure and the left- 
 hand disc has live steam on each side; the valve is thus balanced 
 until the central part of the main piston covers the auxiliary 
 passage, admitting live steam on one side and thus destroying 
 the balance in such a manner that the valve shifts readily. 
 The dotted lines show the passages carrying steam to the cen- 
 tral part of the piston and to the valve case. 
 
 The balanced feature of the valve means that the valve is 
 adjusted for speed so that the pump will not race. 
 
 This company sometimes builds the steam piston with the 
 central space as wide as the stroke So that steam may be intro- 
 duced into this space by a connecting passage in the cylinder 
 casting, saving 'the packed rod. This method means a much 
 longer steam cylinder and also considerable loss from steam 
 condensation. 
 
 The table below, gives the sizes of these pumps: 
 
 MARvSH STANDARD BOILER FEED-PUMPS 
 FOR HIGH PRESSURE. REGULAR LINE 
 
 1 
 
 1 
 
 
 <L> 
 
 0) 
 
 o 
 
 
 
 
 
 
 
 1 
 
 If 
 
 So 
 
 C| 
 
 2^ 
 fc? 
 
 u 
 
 1 
 
 2 ^ 
 o 
 
 W 
 
 15 a 
 
 ii 
 
 11 
 
 || 
 
 O 0) 
 
 
 
 U Cj 
 
 a 
 _t/j 
 
 A 
 
 I 
 
 o 
 
 
 
 ^ 
 
 
 
 O 
 
 
 
 O 
 
 09 
 
 w 
 
 CO 
 
 q 
 
 fe 
 
 K 
 
 ^ 
 
 7 
 
 4l 
 
 8 
 
 55 
 
 66 
 
 4,000 
 
 I 
 
 it 
 
 3i 
 
 l 
 
 15X40 
 
 400 
 
 650 
 
 7i 
 
 4i 
 
 IO 
 
 .68 
 
 75 
 
 4,5 
 
 I 
 
 J l 
 
 3i 
 
 2 2 
 
 18X48 450 
 
 900 
 
 8 
 
 5 
 
 IO 
 
 85 
 
 85 
 
 5,100 
 
 I 
 
 ij 
 
 
 2 i 
 
 18X48 500 
 
 1000 
 
 8 
 
 5 
 
 12 
 
 i .02 
 
 IOO 
 
 6,000 
 
 I 
 
 i 
 
 4 
 
 3 
 
 18X52 
 
 600 
 
 1150 
 
 10 
 
 6 
 
 12 
 
 i .46 
 
 146 
 
 8,75 
 
 1^ 
 
 ii 
 
 4 
 
 3 
 
 18X52 
 
 1000 
 
 1300 
 
 12 
 
 7l 
 
 12 
 
 2.14 
 
 216 
 
 13,000 
 
 !| 
 
 2 
 
 5 
 
 4 
 
 19X56 1500 
 
 2500 
 
 12 
 
 8 
 
 12 
 
 2.61 
 
 260 
 
 15,600 
 
 l 
 
 2 
 
 5 
 
 4 
 
 19X56 
 
 1650 
 
 1500 
 
 14 
 
 8J 
 
 12 
 
 2-94 
 
 295 
 
 17,700 
 
 2 
 
 2^ 
 
 6 
 
 5 
 
 22X62 
 
 2OOO 
 
 2400 
 
 16 
 
 10 
 
 16 
 
 5-44 
 
 408 
 
 25,000 
 
 3 
 
 3i 
 
 8 
 
 7 
 
 24X76 
 
 25OO 
 
 4800 
 
 16 
 
 9* 
 
 16 
 
 4-65 
 
 350 
 
 21,000 
 
 3 
 
 3* 
 
 8 
 
 7 
 
 24X76 
 
 2200 
 
 4800 
 
 16 
 
 10 
 
 20 
 
 6.80 
 
 450 
 
 27,000 
 
 3 
 
 3l 
 
 8 
 
 7 
 
 24X76 
 
 270O 
 
 5000 
 
 16 
 
 9} 
 
 20 
 
 5.81 
 
 400 
 
 24,000 
 
 3 
 
 3i 
 
 8 
 
 7 
 
 24X76 
 
 24OO 
 
 5000 
 
 20 
 
 12 
 
 2O 
 
 9.78 
 
 588 
 
 35,000 
 
 3i 
 
 4 
 
 8 
 
 7 
 
 26X93 
 
 35 
 
 6500 
 
 Capacity, Capacity rating above is intended to represent maximum 
 service recommended. It is always advisable to run pumps at moderate 
 speed, consequently select sizes large enough to meet ordinary require- 
 ments easily. Reserve power in a pump is quite as important as in 9. 
 boiler or engine. 
 
SIMPLEX PUMPS 
 
 173 
 
 The small tappets at each end of the valve casing are used 
 to start the pump in case it' does not operate. Fig. 158 gives 
 an outside view of this pump while the method of attaching the 
 suction is seen in the two figures. Steam may be exhausted 
 into the base where the suction enters by turning the handle 
 shown in Fig. 158. The exhaust steam heats the feed water 
 and is thus saved for useful application. 
 
 One of the oldest of the simplex pumps is the Knowles 
 pump. The Warren steam pump is quite similar to it, and 
 its action will be described. The operation of this pump is 
 
 FIG. 159. Knowles Pump. 
 
 shown in Fig. 159. When the steam piston reaches the left 
 end of its stroke the roller K strikes the lever R, forcing up the 
 rod L, which is attached to an arm D, projecting from the rod 
 of the auxiliary piston A. This rotates A through a small 
 angle and brings a groove in A extending to the right end over 
 a small steam passage while a groove on the left end of A is 
 brought over an exhaust passage. This forces the auxiliary 
 piston to the left, moving the B- valve into the position shown, 
 reversing the pump. At the right-hand end of the stroke the 
 roller K strikes the other end of R, pulling the rod L downward. 
 This rotation brings the groove on the right end of A over an 
 exhaust passage, while that on the left comes opposite a steam 
 
174 
 
 PUMPING MACHINERY 
 
 passage and the auxiliary piston is driven to the right. Should 
 the rotation fail to reverse the pump, the rod I would hit the 
 
 FIG. 160. Knowles Boiler Feed Pump. 
 (Size 10X6X12.) 
 
 FIG. 161. Warren Pump. 
 (Size 7^X7^X10.) 
 
 tappets on the rod from the auxiliary piston and drive the 
 valve to its opposite position. 
 
SIMPLEX PUMPS 
 
 175 
 
 The roller may be raised and lowered. This with the adjust- 
 ment in the connecting rod L makes it possible to adjust the 
 point of reversal. 
 
 The Knowles heavy pattern boiler-feed pump is shown in 
 Fig. 160, while Fig. 161 shows a Warren light-service pump. 
 
 The table below gives sizes of the Knowles pump. 
 
 THE KNOWLES HORIZONTAL BOILER-FEED OR PRESSURE 
 
 PUMPS 
 
 Piston Pattern For Hot or Cold Water or Other Liquids 
 
 All parts of this pump are interchangeable, and can therefore 
 be readily duplicated in case of accidental breakage or unusual 
 
 wear. 
 
 REGULAR PATTERN. FOR 125 POUNDS WATER PRESSURE. 
 
 
 
 
 
 
 Capacity per 
 
 
 
 
 
 
 
 <3 
 
 <5 
 
 v> 
 
 
 Minute at Max- 
 
 
 
 
 
 Jv " 
 
 
 B 
 
 
 a 
 
 B 
 
 o 
 
 L 
 
 <u 
 
 imum Speed. 
 
 
 
 i 
 
 Pn 
 
 <u 
 
 ex 
 
 .^H 
 
 I 
 
 
 
 sJ 
 
 1 
 B 
 
 o 
 || 
 
 ater Cy 
 Inches. 
 
 fl 
 
 i 
 
 a <u 
 
 3-g 
 ij 
 
 J 
 
 a 
 
 n 
 
 
 
 ^ ^ 
 
 B-g 
 
 
 PH . 
 
 C/3 
 
 ^ 
 
 fej 
 
 Ij* 
 
 
 fj 
 
 
 11 
 
 z 
 
 09 
 
 ^ 
 
 co 
 
 o 
 
 co 
 
 
 
 
 w 
 
 w 
 
 Q 
 
 --H W 
 
 000 
 
 i 
 
 i* 
 
 3 
 
 .023 
 
 15 
 
 3i 
 
 i 
 
 I 
 
 i 
 
 f 
 
 i8X 4 
 
 00 
 
 3 
 
 if 
 
 3 
 
 .031 
 
 
 
 i 
 
 1 
 
 | 
 
 
 i8X 5 
 
 
 
 3i 
 
 2 
 
 4 
 
 .05 
 
 150 
 
 72 
 
 i 
 
 1 
 
 J i 
 
 I 
 
 31X6 
 
 I 
 
 3i 
 
 2 i 
 
 4 
 
 .07 
 
 150 
 
 10^ 
 
 i 
 
 1 
 
 ij 
 
 I 
 
 2 9 X 7 
 
 2 
 
 4 
 
 | 
 
 5 
 
 . II 
 
 150 
 
 i6i 
 
 2 
 
 1 
 
 x 
 
 I 
 
 3 2X 7 
 
 3 
 
 5 
 
 3* 
 
 7 
 
 25 
 
 I2 5 
 
 31 
 
 I 
 
 I 
 
 2 
 
 J 2 
 
 44X10 
 
 4 
 
 5i 
 
 3l 
 
 7 
 
 33 
 
 I2 5 
 
 41 
 
 i 
 
 I 
 
 2 
 
 1^ 
 
 44X10 
 
 4i 
 
 7 
 
 4 
 
 7 
 
 38 
 
 125 
 
 47 
 
 i 
 
 tj 
 
 2 2 
 
 2 
 
 45Xio 
 
 6 
 
 7i 
 
 5 
 
 10 
 
 85 
 
 IOO 
 
 85 
 
 i- 
 
 ^i 
 
 3 
 
 2 i 
 
 56X16 
 
 6i 
 
 8 
 
 5 
 
 12 
 
 I .02 
 
 IOO 
 
 102 
 
 i 
 
 r i 
 
 4 
 
 4 
 
 64X16 
 
 7 
 
 10 
 
 6 
 
 12 
 
 1.47 
 
 IOO 
 
 147 
 
 *i 
 
 !i 
 
 4 
 
 4 
 
 68X19 
 
 8 
 
 12 
 
 7 
 
 12 
 
 2 . OO 
 
 IOO 
 
 2OO 
 
 2 
 
 2 2 
 
 5 
 
 5 
 
 68X19 
 
 9 
 
 14 
 
 8 
 
 12 
 
 2.6l 
 
 IOO 
 
 26l 
 
 2 
 
 2 i 
 
 5 
 
 5 
 
 68X20 
 
 10 
 
 16 
 
 10 
 
 16 
 
 5-44 
 
 75 
 
 408 
 
 2 i 
 
 3 
 
 6 
 
 6 
 
 81X21 
 
 
 
 16 
 
 10 
 
 18 
 
 6.12 
 
 70 
 
 428 
 
 2 i 
 
 3 
 
 8 
 
 6 
 
 96X22 
 
 
 18 
 
 IO 
 
 18 
 
 6.12 
 
 70 
 
 428 
 
 22 
 
 3 
 
 8 
 
 6 
 
 96X24 
 
 The Nos. o, i, 2, 3, 4, and 4$ pumps are provided with hand-power 
 attachments. By this means the pump can be used, when steam s 
 down, for filling .boilers after "blowing off/' washing decks, fire purposes, 
 etc. The hand lever can be easily removed by simply lifting it from 
 the pump. 
 
 Nos. ooo, oo, and o pumps may be built as "reinforced pattern," 
 designed for 400 pounds pressure. 
 
176 
 
 PUMPING MACHINERY 
 
 These pumps have large direct water passages and full 
 valve areas, which not only reduce water friction to a minimum, 
 but enable them to be run at a speed that makes them efficient 
 fire pumps. 
 
 HEAVY PATTERN. FOR 250 POUNDS WATER PRESSURE 
 
 I* 
 
 | 
 
 e/3 
 
 
 Capacity per 
 Minuteat Max- 
 
 
 
 
 
 ti 
 
 13 
 
 c 
 
 ^5 
 
 
 imum Speed. 
 
 
 a 
 
 9 
 
 a 
 
 '& 
 
 ^ 
 
 *>. . 
 
 CJ 
 
 C 
 
 s. . 
 
 
 0. 
 
 
 
 
 a. 
 
 o 
 
 
 
 Steam C 
 Inches 
 
 Water C 
 Inches 
 
 CO 
 
 Gallons 
 Stroke 
 
 Strokes. 
 
 Gallons. 
 
 11 
 
 CO 
 
 Exhaust 
 Inches 
 
 Suction 
 Inches 
 
 Delivery 
 Inches 
 
 co'S 
 
 I" 8 
 
 7 
 
 4 1 
 
 10 
 
 .69 
 
 IOO 
 
 69 
 
 I 
 
 z i 
 
 3 
 
 2.1 
 
 60X14 
 
 7$ 
 
 5 
 
 IO 
 
 8.S. 
 
 IOO 
 
 85 
 
 I 
 
 ij 
 
 3 
 
 2* 
 
 60X14 
 
 8 
 
 5 
 
 12 
 
 I .02 
 
 IOO 
 
 102 
 
 I 
 
 tl 
 
 4 
 
 4 
 
 66X14 
 
 10 
 
 6 
 
 12 
 
 I .47 
 
 IOO 
 
 *47 
 
 1^ 
 
 J i 
 
 4 
 
 4 
 
 69X15 
 
 12 
 
 7 
 
 12 
 
 2 .OO 
 
 IOO 
 
 200 
 
 2 
 
 2 i 
 
 5 
 
 5 
 
 72X20 
 
 14 
 
 8 
 
 12 
 
 2.6l 
 
 IOO 
 
 26l 
 
 2 
 
 2 
 
 5 
 
 5 
 
 72X20 
 
 Twice the above capacities can be had in emergencies ; but for contin- 
 uous work, such as boiler feeding, about half the speed stated is advised. 
 
 The Blake pump (Fig. 162) is operated in a different manner. 
 When the tappet A is struck as the piston moves to the left 
 the casting B containing the cavities C, Z), and E is moved 
 to the left. This motion causes the projection F on the casting 
 B to cover the passage carrying steam to the right of the aux- 
 iliary piston G while the projection H uncovers the passage 
 leading to the left end of G. The projection 1 on the other 
 side of B contains two cavities so that this motion to the left 
 covers the passage leading to the left end of the auxiliary 
 piston, while that leading to the right-hand end is connected 
 through one of the cavities under / to the exhaust. The steam 
 then entering the left of G when the cavity at the right is con- 
 nected to the exhaust causes the piston to travel to the right 
 and with it the main valve K, thus reversing the pump. The 
 action is then repeated in the reverse direction when the arm 
 strikes the tappet M. 
 
 Should the auxiliary piston fail to operate, the movement 
 of the casting B is such that the left-hand passage would be 
 moved to the left so far that steam would be admitted to the 
 
SIMPLEX PUMPS 177 
 
 left-hand end of the main cylinder, while exhaust would occur 
 
 (~ 
 
 p 
 
 
 _____ > 
 
 3 1 
 
 't%%%%%%^^ ^ 
 
 
 r? 
 
 tn 
 
 rtrib 
 
 L 
 
 FIG. 162. Blake Steam Pump. 
 
 FIG. 163. Deane Pump. 
 
 from the right. Such action might cease if the pump were 
 running slowly. 
 
178 
 
 PUMPING MACHINERY 
 
 The passages operating the auxiliary valve are made in 
 the cylinder casting, and although small, they carry sufficient 
 steam to operate the auxiliary piston. The casting B is really 
 the valve operating the auxiliary piston. It contains passages 
 leading to the main passage because it is necessary to connect 
 these with the main valve. The same action could be obtained 
 if the valve portion of this was made like a frame surrounding 
 
 FIG. 164. Blake Pump. 
 
 seat. 
 
 the main valve while the main valve moved on the lowe 1 
 Such an arrangement is used on the Deane pump. 
 
 The Deane pump contains a frame B surrounding the main 
 valve /. The tappet A (Fig. 163) is struck by the sleeve 
 and moves a frame B so that the right-hand end of the auxiliary 
 cylinder is connected to the exhaust, while the left end is con- 
 nected to the steam supply. In this way the auxiliary piston 
 is driven to the right and with it the main valve is carried over, 
 reversing the pump. Should the auxiliary piston cease to act, 
 
SIMPLEX PUMPS 179 
 
 the continued motion of the pump would pull the main valve 
 to the right by means of the reverse lever M and thus reverse 
 the pump. The frame B is really a valve operating the aux- 
 iliary piston. One end of this controls the admission and 
 exhaust of steam, from one end corresponding to one cylinder 
 end, while the other of the frame controls the other end. 
 These pumps are accurate in their action. Fig. 164 shows 
 
 FIG. 165. Deane Pressure Pump. 
 (Size i8X2jXi8.) 
 
 the exterior view of the Blake pump and Fig. 165 shows a 
 long-stroke pressure type of the Deane pump, to which the 
 table on p. 180 refers. 
 
 THE DEANE SINGLE OUTSIDE-PACKED DOUBLE-PLUNGER 
 PUMP POT-VALVE PATTERN 
 
 For pressures between 300 and 3000 Ibs. the style of 
 pump shown in Fig. 165 is recommended. As will be seen, the 
 arrangement of the plungers and cylinders is similar to the 
 regular outside plunger type, with the difference, however, 
 that the valves are placed in pots above the cylinder. These 
 valves are specially designed for the service; are each in separate 
 compartments, and may be readily inspected by the removal 
 of covers. The cylinders are made of steeline, a special hydraulic 
 cylinder mixture, for all pressures up to and including 1250 
 Ibs. For higher pressures water ends are made of open-hearth 
 steel castings. These pumps are adapted for hydraulic cranes, 
 presses, punches, shears, riveting machines, etc. All parts are 
 
180 
 
 PUMPING MACHINERY 
 
 quickly accessible, devoid of complications, and heavy in con- 
 struction. These pumps are also built with duplex compound 
 and triple-expansion steam ends. Water ends of pumps listed 
 below are of steeline good for a working pressure of 1250 Ibs. 
 
 SIZE. 
 
 CAPACITY. 
 
 PIPE SIZES. 
 
 Diam- 
 eter of 
 Steam 
 Cyl. 
 
 Diam- 
 eter of 
 Plun- 
 ger. 
 
 Length 
 of 
 Stroke. 
 
 Gallons 
 per 
 Stroke. 
 
 Strokes 
 per 
 Minute of 
 One 
 Plunger. 
 
 Gallons per 
 Minute of Both 
 Plungers. 
 
 Diam- 
 eter of 
 Steam 
 Pipe. 
 
 Diam- 
 eter of 
 Exh'st 
 Pipe. 
 
 Diam- 
 eter of 
 Suc- 
 tion 
 Pipe. 
 
 Diam- 
 eter of 
 Dis- 
 charge 
 Pipe. 
 
 IO 
 
 a* 
 
 12 
 
 . 21 
 
 40 to 80 
 
 8. 5 to 17 
 
 I* 
 
 2 
 
 2 
 
 li 
 
 12 
 
 a* 
 
 12 
 
 . 21 
 
 40 to 80 
 
 8 . 5 to 17 
 
 2 
 
 ai 
 
 2 
 
 li 
 
 14 
 
 H 
 
 12 
 
 . 21 
 
 40 to 80 
 
 8 . 5 to 17 
 
 2 
 
 a| 
 
 2 
 
 li 
 
 16 
 
 a* 
 
 18 
 
 3 1 
 
 40 to 80 
 
 1 2 . 5 to 25 
 
 2 
 
 ai 
 
 2 
 
 I* 
 
 18 
 
 ai 
 
 18 
 
 31 
 
 40 to 80 
 
 1 2 . 5 to 25 
 
 3 
 
 3i 
 
 2 
 
 ll 
 
 20 
 
 ai 
 
 24 
 
 .41 
 
 40 to 60 
 
 1 6 . 5 to 25 
 
 3 
 
 3* 
 
 2 
 
 li 
 
 2 4 
 
 *! 
 
 24 
 
 .41 
 
 40 to 60 
 
 1 6 . 5 to 25 
 
 4 
 
 4i 
 
 2 
 
 li 
 
 12 
 
 ai 
 
 12 
 
 25 
 
 40 to 80 
 
 10 to 20 
 
 2 
 
 a| 
 
 2 
 
 *i 
 
 14 
 
 2* 
 
 12 
 
 2 5 
 
 40 to 80 
 
 10 to 20 
 
 2 
 
 ai 
 
 2 
 
 li 
 
 16 
 
 ai 
 
 18 
 
 38 
 
 40 to 80 
 
 15 to 30 
 
 2 
 
 ai 
 
 2 
 
 li 
 
 18 
 
 ai 
 
 18 
 
 -38 
 
 40 to 80 
 
 15 to 30 
 
 3 
 
 si 
 
 2 
 
 li 
 
 20 
 
 a| 
 
 24 
 
 5 1 
 
 40 to 60 
 
 20 tO 30 
 
 3 
 
 si 
 
 2 
 
 li 
 
 24 
 
 ai 
 
 24 
 
 5 1 
 
 40 to 60 
 
 20 to 30 
 
 4 
 
 4i 
 
 2 
 
 ll 
 
 12 
 
 2| 
 
 12 
 
 3 1 
 
 40' to 80 
 
 1 2 . 5 to 25 
 
 2 
 
 ti 
 
 2 
 
 I-^ 
 
 14 
 
 a| 
 
 12 
 
 3i 
 
 40 to 80 
 
 1 2 . 5 to 25 
 
 2 
 
 ai 
 
 2 
 
 I^ 
 
 16 
 
 ** 
 
 18 
 
 .46 
 
 40 to 80 
 
 18.5 to 37 
 
 2 
 
 2i 
 
 2 
 
 li 
 
 18 
 
 at 
 
 18 
 
 .46 
 
 40 to 80 
 
 18.5 to 37 
 
 3 
 
 3i 
 
 2 
 
 li 
 
 20 
 
 at 
 
 24 
 
 .62 
 
 40 to 60 
 
 25 to 37 
 
 3 
 
 3l 
 
 2 
 
 ^ 
 
 24 
 
 a* 
 
 24 
 
 .62 
 
 40 to 60 
 
 25 to 37 
 
 4 
 
 4i 
 
 2 
 
 li 
 
 14 
 
 3 
 
 12 
 
 37 
 
 40 to 80 
 
 15 to 30 
 
 2 
 
 i 
 
 2 
 
 li 
 
 16 
 
 3 
 
 18 
 
 55 
 
 40 to 80 
 
 22 to 44 
 
 2 
 
 aj 
 
 2 
 
 li 
 
 18 
 
 3 
 
 18 
 
 55 
 
 40 to 80 
 
 22 to 44 
 
 3 
 
 $i 
 
 2 
 
 li 
 
 20 
 
 3 
 
 24 
 
 73 
 
 40 to 60 
 
 29 to 48 
 
 3 
 
 3i 
 
 2 
 
 li 
 
 24 
 
 3 
 
 24 
 
 73 
 
 40 to 60 
 
 29 to 48 
 
 4 
 
 4i 
 
 2 
 
 li 
 
 16 
 
 3i 
 
 18 
 
 75 
 
 40 to 80 
 
 30 to 60 
 
 2 
 
 ai 
 
 3 
 
 2 
 
 18 
 
 3l 
 
 18 
 
 75 
 
 40 to 80 
 
 30 to 60 
 
 3 
 
 3i 
 
 3 
 
 2 
 
 20 
 
 3i 
 
 24 
 
 i .00 
 
 40 to 60 
 
 40 to 60 
 
 3 
 
 3* 
 
 3 
 
 2 
 
 24 
 
 3i 
 
 24 
 
 i .00 
 
 40 to 60 
 
 40 to 60 
 
 4 
 
 4i 
 
 3 
 
 2 
 
 . 18 
 
 -4 
 
 18 
 
 .98 
 
 40 to 80 
 
 39 to 78 
 
 3 
 
 3i 
 
 3i 
 
 2i 
 
 20 
 
 4 
 
 2 4 
 
 1.30 
 
 40 to 60 
 
 52 to 104 
 
 3 
 
 3* 
 
 3i 
 
 2i 
 
 24 
 
 4 
 
 24 
 
 1.30 
 
 40 to 60 
 
 52 to 104 
 
 4 
 
 4i 
 
 3i 
 
 2i 
 
 Pumps of this type are made for heavier pressures. 
 
SIMPLEX PUMPS 
 
 181 
 
 The Davidson pump (Fig. 166) is operated by means of 
 the oscillation of the auxiliary valve. The arm A is attached 
 to the piston rod on the outside of the pump. The turning 
 
182 PUMPING MACHINERY 
 
 of the shaft B turns the cam C which causes the auxiliary 
 valve D to turn. When the piston gets to the end of the 
 stroke the pin puts the valve into the position shown so that 
 the passage E is connected with the steam and the passage F 
 with the exhaust. This causes the auxiliary pistons GG to be 
 driven to the left so that steam will enter the right end of the 
 main cylinder. The auxiliary valve is oscillated when it acts 
 as an auxiliary valve, but the same casting D acts as the main 
 valve by its longitudinal motion. The casting D fits into a 
 
 FIG. 167. Davidson Pump. 
 
 cavity in the frame of the auxiliary pistons GG, and although 
 it is carried to the top of the cylinder in which GG moves, it 
 does not completely fill the cylinder, so that the steam which 
 enters at H can pass around the valve D. The exhaust passes 
 through the cavity containing the pin of D and the cam C. 
 
 The valve D is twisted into the position shown in the small 
 section whenever it is preparing to drive the main valve to one 
 end of the cylinder. When it is in such a position that E and 
 F are covered, the openings to the main cylinder are open to 
 either steam or exhaust. In this way, when steam is turned 
 
SIMPLEX PUMPS 
 
 183 
 
 THE DAVIDSON PRESSURE PUMP, PISTON PATTERN 
 FOR BOILER FEEDING, OPERATING HYDRAULIC ELEVATORS, ETC. 
 
 For a pressure of 150 pounds. (If specially ordered these pumps 
 can be furnished for 250 pounds working pressure). 
 
 (Composition or bronze cylinders to order.) 
 
 SIZES AND DETAILS 
 
 
 
 
 
 
 i.P Boiler, 
 
 
 
 
 
 
 
 
 
 
 msed on 30 
 
 
 
 
 
 
 
 
 
 
 Ibs. of 
 
 
 
 
 
 Size. 
 No. 
 
 Steam 
 Cylin- 
 der. 
 
 Water 
 Cylin- 
 der. 
 
 Stroke 
 Inches 
 
 Gallons 
 per 
 Stroke. 
 
 Water per 
 H.P.perHr. 
 which 
 the Pump 
 
 Steam 
 Pipe. 
 
 Ex- 
 haust 
 Pipe. 
 
 Suc- 
 tion 
 Pipe. 
 
 Dis- 
 charge 
 Pipe. 
 
 
 
 
 
 
 will supply 
 
 
 
 
 
 
 
 
 
 
 with ease. 1 
 
 
 
 
 
 
 
 2* 
 
 ii 
 
 3 
 
 .022 
 
 20 
 
 1 
 
 1 
 
 1 
 
 | 
 
 i 
 
 3 
 
 if 
 
 4 
 
 .041 
 
 35 
 
 i 
 
 1 
 
 I 
 
 I 
 
 i 
 
 35 
 
 2 
 
 4 
 
 05 
 
 5 
 
 I 
 
 \ 
 
 il 
 
 i 
 
 ij 
 
 4 
 
 2 i 
 
 4 
 
 .069 
 
 65 
 
 i 
 
 I 
 
 il 
 
 i 
 
 2 
 
 4i 
 
 2 \ 
 
 6 
 
 13 
 
 I2 5 
 
 i 
 
 3 
 4 
 
 l| 
 
 il 
 
 2* 
 
 5 
 
 3 
 
 6 
 
 183 
 
 i75 
 
 8 
 
 i 
 
 2 
 
 ii 
 
 3 
 
 Si 
 
 3l 
 
 8 
 
 .28 
 
 275 
 
 i 
 
 I 
 
 2 
 
 ri 
 
 3^ 
 
 6 
 
 4 
 
 8 
 
 435 
 
 425 
 
 1 
 
 I 
 
 2$ 
 
 2 
 
 4 
 
 7 
 
 4 
 
 10 
 
 54 
 
 525 
 
 1 
 
 I 
 
 2^ 
 
 2 
 
 42 
 
 8 
 
 5 
 
 IO 
 
 85 
 
 850 
 
 i 
 
 J l 
 
 3 
 
 *i 
 
 5 
 
 9 
 
 5l 
 
 12 
 
 I . 12 
 
 I IOO 
 
 i 
 
 il 
 
 3 
 
 2 i 
 
 6 
 
 10 
 
 6 
 
 12 
 
 1-47 
 
 1400 
 
 1 
 
 ii 
 
 4 
 
 3 
 
 7 
 
 12 
 
 7 
 
 12 
 
 2 .OO 
 
 2OOO 
 
 ii 
 
 2 
 
 5 
 
 4 
 
 7^ 
 
 14 
 
 81 
 
 12 
 
 2.77 
 
 2750 
 
 i^ 
 
 2 
 
 6 
 
 5 
 
 8^ 
 
 14 
 
 81 
 
 14 
 
 3-23 
 
 3200 
 
 if 
 
 2 
 
 6 
 
 5 
 
 8 i 
 
 14 
 
 IO 
 
 14 
 
 4.76 
 
 475 
 
 ii 
 
 2 
 
 8 
 
 7 
 
 Q 
 
 16 
 
 9l 
 
 16 
 
 4.66 
 
 
 2 
 
 2* 
 
 7 
 
 6 
 
 9* 
 
 16 
 
 10 
 
 16 
 
 5-44 
 
 
 
 2 
 
 2 
 
 / 
 8 
 
 7 
 
 IO 
 
 18 
 
 ioi 
 
 18 
 
 6.74 
 
 
 2 l 
 
 J 
 
 8 
 
 7 
 
 10* 
 
 18 
 
 II* 
 
 18 
 
 *"* * / T" 
 
 8.09 
 
 
 2 1 
 
 O 
 J 
 
 o 
 
 / 
 8 
 
 j. w 2 
 I I 
 
 20 
 
 2 
 
 III 
 
 20 
 
 9 . oo 
 
 
 2 1 
 
 O 
 
 y 
 
 Q 
 
 8 
 
 II* 
 
 20 
 
 I -2 
 
 20 
 
 1 1 . 49 
 
 
 2* 
 
 -J 
 
 y 
 IO 
 
 
 A * 2 
 12 
 
 22 
 
 %J 
 I T. 
 
 22 
 
 12 . 64 
 
 
 2 
 
 
 "2 
 
 IO 
 
 o 
 
 13 
 
 24 
 
 O 
 
 14 
 
 24 
 
 16.00 
 
 
 
 3! 
 
 O 
 
 4 
 
 10 
 
 y 
 IO 
 
 1 Capacities for boiler feeding based on a speed of 60 single strokes per minute with 
 feed water at ordinary temperature. With high temperature feed water reduce H.P. 
 rating one-third. These pumps can be run as slowly as may be desired, and in cases of 
 emergency, speeded up to about twice ordinary capacities. 
 
 Suction and discharge openings on both sides. Hand levers furnished 
 with sizes No. o to No. 3^, and with No. 4 when ordered. Water-piston 
 packing and valves for hot or cold water as ordered. Every machine 
 thoroughly tested before leaving works. 
 
184 
 
 PUMPING MACHINERY 
 
 on either steam will enter E or F and drive the main valve to a 
 position to admit steam to one end and start the pump, or the 
 main valve itself will be in a position to start the pump. From 
 this it may be seen that there is no dead center possible with 
 this pump. 
 
 It is claimed that the action of the cam prevents the piston 
 
 FIG. 168. Burnham Steam Pump. 
 
 from striking the cylinder head. Moreover, the pump should 
 start from any position and make its full stroke. 
 
 Fig. 167 gives an outside view of this pump, showing the 
 method of driving the cam lever. 
 
 The table on the preceding page gives sizes used by the M. 
 T. Davidson Co. for one type of their pump. 
 
 In the Burnham steam pump (Fig. 1.68) the auxiliary valve 
 is placed on the side of the pump and is driven by the cam rod 
 
SIMPLEX PUMPS 
 
 185 
 
 A. When the pump reaches the position shown in the figure, 
 the rod B is moved to the right and this moves the aux- 
 iliary valve H to the right. 
 This valve is mounted on the 
 side of the steam chest and 
 its motion to the right will 
 admit steam to the left-hand 
 end of the auxiliary piston / 
 while it connects the right- 
 hand end to the exhaust. 
 This drives the auxiliary pis- 
 ton to the position shown in 
 the figure, and with it the 
 main valve. The pump is 
 then reversed. 
 
 The double ports for the 
 main piston and auxiliary 
 piston are arranged to prevent 
 pounding and give smooth 
 action . The small ports lead - 
 ing to the ends of the cylin- 
 ders are for the sole purpose 
 of admitting steam. The 
 exhaust takes place through 
 the passages shown full in the 
 section and when either the 
 main piston or the auxiliary 
 piston travel over these ports 
 the steam behind them is 
 trapped and forms a cushion. 
 In the positions shown in the 
 figure each of the pistons is 
 started on its return stroke 
 
 FIG. 169. Burnham Pump. 
 
 by the small quantities of 
 
 steam which are admitted through the small passages shown 
 dotted and extending to the ends of the cylinders. Fig. 169 
 illustrates the exterior appearance of the pump. It will be 
 
186 
 
 PUMPING MACHINERY 
 
 noted that the valve cam moves the valve rod at the end of 
 the stroke only, and moreover this motion is gradual and slow. 
 The table gives the sizes of these pumps used for tank 
 service. 
 
 DETAILS OF BURNHAM TANK OR LIGHT-SERVICE PUMPS 
 
 
 
 
 
 
 a; t-i 
 
 
 
 0-0 , 
 
 
 
 
 
 As length of connections 
 
 
 -M U 
 
 <A 
 
 s.S 
 
 
 1 
 
 I 
 
 1 
 
 increases, larger size 
 pipes should be used. 
 
 1 
 
 Z 
 
 |l 
 
 a s 
 
 
 D 
 
 sa 
 
 J5 
 
 CO 
 
 
 CO 
 
 IH ^ 
 
 VnCO 
 
 fc' 
 
 
 o 
 g 
 
 >> 
 
 
 8 
 
 *o 
 
 1 
 
 j 
 
 -M 
 * 
 
 
 
 is 
 
 1 
 
 joj 
 
 |I 
 
 S *" 
 
 bo 
 . 
 
 CO 
 
 i 
 
 J 
 
 co 
 
 X 
 
 w 
 
 I 
 
 1 
 
 "c3 
 O 
 
 rt rt O 
 
 o 
 
 3M 
 
 O 
 
 CO 
 
 C 
 
 8 
 
 10 
 
 12 
 
 
 
 
 8 
 
 6 
 
 4.08 
 
 204 to 408 
 
 24,400 
 
 50 to ioo 
 
 R 
 
 8 
 
 II 
 
 12 
 
 
 
 
 8 
 
 6 
 
 4-93 
 
 247 to 494 
 
 29,600 
 
 50 to ioo 
 
 R 
 
 8 
 
 12 
 
 12 
 
 
 
 
 8 
 
 6 
 
 5-87 
 
 293 to 587 
 
 35,200 
 
 50 to ioo 
 
 R 
 
 8j 
 
 6 
 
 10 
 
 
 
 
 4 
 
 3i 
 
 I . 22 
 
 73 to 147 
 
 8,800 
 
 60 tO 120 
 
 R 
 
 n 
 
 7 
 
 10 
 
 
 
 
 5 
 
 4 
 
 1.66 
 
 IOO tO 200 
 
 12,000 
 
 60 to 1 20 
 
 R 
 
 8f 
 
 8 
 
 10 
 
 
 
 
 5 
 
 4 
 
 2.17 
 
 130 to 261 
 
 15,600 
 
 60 tO 120 
 
 R 
 
 $ 
 
 9 
 
 10 
 
 
 
 
 6 
 
 5 
 
 2-75 
 
 165 to 330 
 
 19,800 
 
 60 to i 20 
 
 R 
 
 8i 
 
 10 
 
 10 
 
 
 
 
 6 
 
 5 
 
 3-40 
 
 204 to 408 
 
 24,400 
 
 60 tO 120 
 
 R 
 
 10 
 
 8 
 
 12 
 
 
 
 2 
 
 
 5 
 
 4 
 
 2.61 
 
 130 to 261 
 
 15,600 
 
 60 to 1 20 
 
 R 
 
 10 
 
 9 
 
 12 
 
 
 
 2 
 
 
 6 
 
 5 
 
 3-30 
 
 165 to 330 
 
 19,800 
 
 50 to ioo 
 
 R 
 
 10 
 
 10 
 
 12 
 
 
 
 2 
 
 
 8 
 
 6 
 
 4.08 
 
 204 to 408 
 
 24,400 
 
 50 to ioo 
 
 R 
 
 10 
 
 ii 
 
 12 
 
 
 
 2 
 
 
 8 
 
 6 
 
 4-93 
 
 247 to 494 
 
 29,600 
 
 50 to ioo 
 
 R 
 
 IO 
 
 12 
 
 12 
 
 
 
 2 
 
 
 8 
 
 6 
 
 5.87 
 
 293 to 587 
 
 35,200 
 
 50 to ioo 
 
 R 
 
 10 
 
 14 
 
 16 
 
 
 
 2 
 
 
 10 
 
 8 
 
 10.65 
 
 404 to 808 
 
 48,400 
 
 38 to 75 
 
 R 
 
 12 
 
 8 
 
 12 
 
 
 
 2- 
 
 
 5 
 
 4 
 
 2.61 
 
 130 to 261 
 
 15,600 
 
 50 to ioo 
 
 R 
 
 12 
 
 9 
 
 12 
 
 
 
 2 
 
 
 6 
 
 5 
 
 3-3C 
 
 165 to 330 
 
 19,800 
 
 50 to ioo 
 
 R 
 
 12 
 
 10 
 
 12 
 
 
 
 1 
 
 8 
 
 6 
 
 4.08 
 
 204 to 408 
 
 24,400 
 
 50 to ioo 
 
 R 
 
 12 
 
 ii 
 
 12 
 
 
 
 * 
 
 8 
 
 6 
 
 4-93 
 
 247 to 494 
 
 29,600 
 
 50 to ioo 
 
 R 
 
 12 
 
 12 
 
 12 
 
 
 
 I 
 
 8 
 
 6 
 
 5.8 7 
 
 293 to . 587 
 
 35.200 
 
 50 to ioo 
 
 R 
 
 12 
 
 14 
 
 12 
 
 
 
 I 
 
 10 
 
 8 
 
 8.00 
 
 400 to 800 
 
 48,000 
 
 50 to ioo 
 
 R 
 
 12 
 
 8 
 
 16 
 
 
 
 I 
 
 6 
 
 4 
 
 3-48 
 
 132 to 264 
 
 15,800 
 
 38 to 75 
 
 R 
 
 12 
 
 9 
 
 16 
 
 
 
 
 
 
 6 
 
 5 
 
 4-40 
 
 167 to 334 
 
 20,040 
 
 38 to 75 
 
 R 
 
 12 
 
 10 
 
 16 
 
 
 
 1 
 
 
 8 
 
 6 
 
 5-44 
 
 206 to 412 
 
 24,700 
 
 38 to 75 
 
 R 
 
 12 
 
 ii 
 
 16 
 
 
 
 
 
 8 
 
 6 
 
 6.58 
 
 250 to 500 
 
 30,000 
 
 38 to 75 
 
 R 
 
 12 
 
 12 
 
 16 
 
 
 
 j 
 
 
 8 
 
 6 
 
 7.82 
 
 297 to 594 
 
 35,6oo 
 
 38 to 75 
 
 R 
 
 12 
 
 14 
 
 16 
 
 
 
 l 
 
 
 10 
 
 8 
 
 10.65 
 
 404 to 808 
 
 48,400 
 
 38 to 75 
 
 R 
 
 14 
 
 10 
 
 12 
 
 
 i 
 
 
 8 
 
 6 
 
 4.08 
 
 204 to 408 
 
 24,400 
 
 50 tO TOO 
 
 R 
 
 14 
 
 12 
 
 12 
 
 
 ] 
 
 
 8 
 
 6 
 
 5-87 
 
 293 to 587 
 
 35,200 
 
 50 to ioo 
 
 R 
 
 14 
 
 14 
 
 12 
 
 
 
 
 10 
 
 8 
 
 8.00 
 
 400 to 800 
 
 48,000 
 
 50 to ioo 
 
 R 
 
 14 
 
 10 
 
 16 
 
 
 
 
 8 
 
 6 
 
 5-44 
 
 206 to 412 
 
 24,700 
 
 38 to 75 
 
 R 
 
 14 
 
 II 
 
 16 
 
 
 j 
 
 
 8 
 
 6 
 
 6.58 
 
 250 to 500 
 
 30,000 
 
 38 to. 75 
 
 R 
 
 14 
 
 12 
 
 16 
 
 
 ': 
 
 
 8 
 
 6 
 
 7.82 
 
 297 to 594 
 
 35.6oo 
 
 38 to 75 
 
 R 
 
 14 
 
 14 
 
 16 
 
 
 f 
 
 10 
 
 8 
 
 10.65 
 
 404 to 808 
 
 48,400 
 
 38 to 75 
 
 R 
 
 14 
 
 16 
 
 16 
 
 
 f 
 
 10 
 
 8 
 
 14.00 
 
 532 to 1064 
 
 63,800 
 
 38 to 75 
 
 R 
 
 14 
 
 16 
 18 
 
 20 
 
 
 
 
 12 
 
 10 
 
 17.40 
 
 522 to 1044 
 
 62,600 
 
 30 to 60 
 
 R 
 
 '4 
 16 
 
 I O 
 
 II 
 
 16 
 
 
 
 
 8 
 
 6 
 
 6.58 
 
 660 to 1321 
 250 to 494 
 
 79,200 
 29,600 
 
 30 to 60 
 38 to 75 
 
 R 
 
 16 
 
 12 
 
 16 
 
 
 j 
 
 
 8 
 
 6 
 
 7.82 
 
 297 to 594 
 
 35,6oo 
 
 38 to 75 
 
 R 
 
 16 
 
 14 
 
 16 
 
 
 
 
 10 
 
 8 
 
 10.65 
 
 404 to 808 
 
 48,400 
 
 38 to 75 
 
 R 
 
 16 
 
 16 
 
 16 
 
 
 
 
 10 
 
 8 
 
 14.00 
 
 532 to 1064 
 
 63,800 
 
 38 to 75 
 
 R 
 
 16 
 
 16 
 
 20 
 
 
 j 
 
 
 12 
 
 IO 
 
 17.40 
 
 522 to 1044 
 
 62,600 
 
 30 to 60 
 
 R 
 
 16 
 
 18 
 
 20 
 
 
 ; 
 
 
 12 
 
 10 
 
 22.03 
 
 660 to 1321 
 
 79,200 
 
 30 to 60 
 
 R 
 
 16 
 
 20 
 
 24 
 
 
 j 
 
 
 14 
 
 12 
 
 32.6- 
 
 816 to 1632 
 
 97,920 
 
 25 to 50 
 
 R 
 
 16 
 
 22 
 
 24 
 
 
 
 
 16 
 
 12 
 
 39-50 
 
 987 to 1975 
 
 
 25 to 50 
 
 R 
 
 18 
 
 14 
 
 16 
 
 \ 
 
 31 
 
 
 IO 
 
 8 
 
 10.65 
 
 405 to 810 
 
 48,600 
 
 38 to 75 
 
 R 
 
 18 
 
 16 
 
 20 
 
 1 
 
 3 ? 
 
 
 12 
 
 10 
 
 17.40 
 
 52? to 1044 
 
 62,600 
 
 30 to 60 
 
 R 
 
 18 
 
 18 
 
 20 
 
 1 
 
 3* 
 
 12 
 
 10 
 
 22.03 
 
 660 to 1321 
 
 79,200 
 
 30 to 60 
 
 R 
 
 20 
 
 16 
 
 20 
 
 L 
 
 3* 
 
 12 
 
 10 
 
 17.40 
 
 522 to 1044 
 
 62,600 
 
 30 to 60 
 
 R 
 
 20 
 
 18 
 
 20 
 
 ? \ 
 
 3* 
 
 12 
 
 10 
 
 22.03 
 
 660 to 1321 
 
 79,200 
 
 30 to 60 
 
 R 
 
 20 
 
 20 
 
 24 
 
 2 \ 
 
 3* 
 
 14 
 
 12 
 
 32.63 
 
 816 to 1632 
 
 97,920 
 
 25 to 50 
 
 R 
 
 2O 
 
 22 
 
 24 
 
 * 
 
 31 
 
 16 
 
 12 
 
 39-50 
 
 987 to 1975 
 
 118,500 
 
 25 to 50 
 
 R 
 
 "R" signifies removable bronze lining in water cylinder. 
 pressed in. 
 
 "P" signifies bronze lining 
 
SIMPLEX PUMPS 
 
 187 
 
 The Dean Brothers pump (Fig. 170) illustrates another 
 method of obtaining these results. As in the Burnham pump 
 the auxiliary valve is placed on the side of the main steam 
 chest. This pump differs from the others considered in that the 
 motion of the auxiliary valve is continuous and not intermittent. 
 This valve is driven from a reverse lever attached by a link 
 to the piston rod, and the valve is moved by a link at the upper 
 end of this. When the main piston reaches the extreme right 
 of its stroke the auxiliary valve is at its extreme left, and the 
 edge of the valve uncovers a small circular port of the passage 
 
 FIG. 170. Dean Brothers Pump. 
 
 A leading to the right end of the auxiliary cylinder. At this 
 time one of the two diagonal grooves on the under side of the 
 valve connects the other passage B to the exhaust port C. 
 This then forces the auxiliary piston to its extreme left, 
 and with it the main D slide valve, admitting steam to the 
 right-hand end of the cylinder, reversing the motion. The 
 auxiliary valve begins to move to the right, cutting off the 
 steam and exhaust from the auxiliary valve. This condition 
 does not change until the valve almost reaches the end of 
 its stroke, when the "other diagonal groove connects A to C, 
 while B is uncovered to the action of the live steam. The 
 auxiliary valve then goes over to the other position and the 
 
188 PUMPING MACHINERY 
 
 pump reverses. The link Z), which moves the auxiliary valve 
 rod, may be moved in a slot closer to or farther from the pivot 
 of the reverse lever, making its stroke shorter or longer; this 
 makes the stroke of the main pump longer or shorter, as it 
 means more or less motion of the main piston to move the 
 auxiliary valve a sufficient distance to bring the ports into the 
 position for shifting the auxiliary piston. 
 
 The small spindle projecting from the top of the main valve 
 casing is used to shift the auxiliary piston and with it the main 
 valve which rests in a slot in the auxiliary piston. Such a 
 contrivance is necessary, as the main valve may be left in a 
 
 FIG. 171. Dean Brothers Pump. 
 
 position so that steam could not enter either side. This con- 
 dition rarely occurs, but in most of these pumps such a contin- 
 gency is guarded against by some device like this or the pro- 
 jecting rods of the Marsh pump. 
 
 Fig. 171 illustrates one of these pumps designed for boiler 
 feeding, in which several special features can be seen. The 
 steam and water valves are easily examined without the neces- 
 sity of interfering with any pipes; the frame is of steel; the 
 cylinders are removable and may be replaced without disturbing 
 any other parts; the piston rod has a cross-head fitting over 
 the lower rod frame, making it impossible to twist the cross- 
 head so as to jam the valve gear. 
 
SIMPLEX PUMPS 
 
 1S9 
 
 The following table gives the sizes used by these makers 
 for one line of their pumps. 
 
 DEAN BROTHERS SIMPLE PUMPS WITH PACKED PISTONS 
 
 (For feeding boilers or pumping against pressure. They will elevate water 200 feet with 
 
 60 pounds steam.) 
 
 TABLE OF DIMENSIONS 
 
 
 fed 
 
 0) 
 
 'O 
 
 1 
 
 B 
 
 
 1 Capacity per 
 Minute. 
 
 
 
 
 
 $ 
 
 i. M 
 O C 
 
 SSQJ 
 
 
 S9 
 
 .^H 
 
 o 
 
 B 
 
 
 & 
 
 o"! 
 
 & 1 
 
 CL 
 
 
 i 
 
 $ 
 a| 
 jl 
 
 >> 
 
 ?j 
 
 %* 
 
 C 
 | 
 
 ii 
 
 Tito 
 
 1 
 
 t 
 
 
 13 
 
 E o 
 
 xhaust ] 
 Inches. 
 
 PL, ^ 
 
 II 
 
 fl 
 
 u-i (^ T 
 ^ S 
 
 w 
 
 w 
 
 
 
 03 
 
 o 
 
 & 
 
 O 
 
 & 
 
 W 
 
 oS 
 
 Q 
 
 3 
 
 C 
 
 3 2 
 
 4 
 
 0.054 
 
 2OO 
 
 IO 
 
 1 
 
 | 
 
 if 
 
 i 
 
 5 
 
 D 
 
 4 
 
 tf 
 
 6 
 
 0.14 
 
 I4O 
 
 20 
 
 i 
 
 4 
 
 J i 
 
 if 
 
 IOO 
 
 E 
 
 5i 
 
 3f 
 
 7 
 
 33 
 
 125 
 
 42 
 
 1 
 
 
 2 
 
 
 220 
 
 F 
 
 6 
 
 4 
 
 10 
 
 55 
 
 125 
 
 68 
 
 i 
 
 J i 
 
 3 
 
 2i 
 
 400 
 
 G 
 
 8 5 
 
 12 
 
 i .02 
 
 IOO 
 
 102 
 
 i 
 
 r i 
 
 4 
 
 3 
 
 700 
 
 H 
 
 9 6 
 
 I 2 
 
 1.47 
 
 IOO 
 
 147 
 
 ij 
 
 2 
 
 4 
 
 3 
 
 IOOO 
 
 I 
 
 10 7 
 
 12 
 
 2 .00 
 
 IOO 
 
 200 
 
 ii 
 
 2 
 
 5 
 
 4 
 
 1300 
 
 K 
 
 12 8 
 
 12 
 
 2.6l 
 
 IOO 
 
 26l 
 
 2 
 
 2 i 
 
 6 
 
 5 
 
 1600 
 
 L 
 
 14 9 
 
 16 
 
 4.40 
 
 80 
 
 352 
 
 2 
 
 3 
 
 7 
 
 6 
 
 2500 
 
 M 
 
 14 10 
 
 20 
 
 6.80 
 
 70 
 
 476 
 
 2 
 
 3 
 
 8 
 
 7 
 
 35 ' 
 
 1 In an emergency more than the above capacity can be had, but for continuous work, 
 such as feeding boilers, not more than half the strokes given above are advised. The valve 
 motion secures a smooth action and admits of regulation, so as to deliver a steady supply 
 of water, exactly equal to the amount evaporated. 
 
 The pumps described in this chapter are taken to repre- 
 sent this class, and although there are many other different 
 forms, the principles shown in. the these will aid in under- 
 standing the action of any other simplex pump. All of the 
 valve gears aim to move the valve positively whatever be its 
 speed. From the spring- thrown valve of Worthington to the 
 latest form, this is the governing idea. 
 
CHAPTER V 
 DYNAMICS OF WATER END 
 
 Velocity and Acceleration. To make the velocity of dis- 
 charge from a pump more definite a pump with a fly wheel 
 and connecting rod will be considered. The figure below (Fig. 
 172) represents a plunger pump in which the rate of discharge 
 or suction at any instant is proportional to the velocity cf the 
 plunger. From the diagram the movement from the end cf 
 the stroke is * 
 
 x=AB+BC=r(i-cosd)+nr(i-cosa), . . (i) 
 
 FIG. 172. Diagram of Pump. 
 
 where = the angle the crank has moved from its head dead 
 
 center; 
 a=the inclination of the connecting rod with the center 
 
 line; 
 
 r = the crank radius in feet; 
 
 w=the ratio of length of connecting rod to crank radius. 
 Now 
 
 . 
 
 sin a = sin 0. 
 nr 
 
 Hence 
 
 
 cos a 
 
 ~\ 
 
 rip* 
 
 * V=I \ 
 
 ? w 2 
 
 8 0, approx. 
 
 Therefore 
 
 
 
 
 
 
 
 
 % 
 
 -r[i 
 
 cos - 
 
 I 1 
 
 1 
 
 v 1 
 
 (2) 
 
 
 
 L 
 
 
 F ' w sm2 
 
 J 
 
 
 * See page 257 for table of symbols. 
 
 190 
 
DYNAMICS OF WATER END 
 
 dx sin cos 6\ 
 Velocity =-jr= S= r sin 0+- 
 
 = r sin 0+ sin 26 \co. 
 2n 
 
 d 2 
 
 Acceleration = -r 
 at* 
 
 2* r 
 p-=r[. 
 
 H cos 0-\ 
 
 cos 201 
 
 n 
 
 t>. 
 
 -8 
 
 30 
 
 -Stroke 
 
 30 0- 
 
 90 x e. 
 
 20 
 
 \ 
 
 -Stroke 
 
 Head End Q 6=30 0=60 
 
 191 
 (3) 
 
 (4) 
 
 Craak End 
 
 FIG. 173. Velocity and Acceleration. 
 
 !i . 
 
 These give diagrams such as shown in Fig. 173, wjien S 
 and a are plotted against x or piston position as asbcissaeJ 
 
 This approximate solution is sufficiently close for. general 
 work, as may be seen from the table below, which gives the 
 
192 
 
 PUMPING MACHINERY 
 
 FIG. 174. Curves of x, S, a for Different Crank Angles. 
 
DYNAMICS OF WATER END 
 
 193 
 
 values of a by the approximate formula and the exact formula 
 obtained without expanding the equation for cos . The 
 exact equations are 
 
 r . sm e cos e l 
 
 S=cor sm#+ 
 
 Vtt 2 -sin 2 /9J 
 
 COS0 + 
 
 n 2 cos 2 
 
 sin 2 # 
 
 (5) 
 
 (6) 
 
 For a discussion of the use of the exact and approximate 
 values of acceleration the reader is referred to the papers of 
 D. S. Jacobus, A. S. M. E. Transactions, Vol. XI, p. 492 et seq., 
 and p. 1116 et seq. 
 
 In the exact formula the angular velocity is considered 
 as constant, and since this is not the case there is really an error 
 in the exact formula. 
 
 The curves, Fig. 174, give the values of x, 5, and a as func- 
 tions of crank movement. These curves are symmetrical, and 
 in computing the values of the points, as in the table below, 
 for a unit value of a> and r, this symmetry is seen. 
 
 
 
 cos 
 
 sin d 
 
 cos 26 
 
 sin 20 
 
 X 
 
 S 
 
 a 
 
 a exact 
 
 
 
 I 
 
 
 
 I 
 
 
 
 
 
 
 
 I . 167 
 
 I . 167 
 
 I 5 
 
 96593 
 
 .25882 
 
 .86603 
 
 5 
 
 .040 
 
 .300 
 
 I . I 10 
 
 I . ill 
 
 3 
 
 .86603 
 
 5 
 
 5 
 
 .86603 
 
 J55 
 
 572 
 
 949 
 
 95 
 
 45 
 
 .70711 
 
 .70711 
 
 o 
 
 i 
 
 335 
 
 .790 
 
 .707 
 
 .706 
 
 60 
 
 5 
 
 .86603 
 
 - -5 
 
 .86603 
 
 563 
 
 938 
 
 .417 
 
 .417 
 
 75 
 
 .25882 
 
 96593 
 
 - .86603 
 
 5 
 
 .819 
 
 1 .007 
 
 .114 
 
 **J 
 
 90 
 
 
 
 I 
 
 i 
 
 o 
 
 1.083 
 
 1 .000 
 
 -.167 
 
 . 169 
 
 i5 
 
 - .25882 
 
 96593 
 
 - .86603 
 
 - -5 
 
 i-346 
 
 .924 
 
 - .403 
 
 - -405 
 
 I2O 
 
 - -5 
 
 .86603 
 
 -5 
 
 .86603 
 
 1.562 
 
 794 
 
 -.583 
 
 -.583 
 
 135 
 
 - .70711 
 
 .70711 
 
 o 
 
 i 
 
 1-749 
 
 .624 
 
 -.707 
 
 - .709 
 
 J 5 
 
 .86603 
 
 5 
 
 5 
 
 .86603 
 
 i .908 
 
 .428 
 
 -783 
 
 -.782 
 
 165 
 
 - -96593 
 
 .25882 
 
 .86603 
 
 - -5 
 
 1.972 
 
 .217 
 
 - .822 
 
 - .821 
 
 180 
 
 i 
 
 
 
 i 
 
 
 
 2 
 
 o 
 
 -833 
 
 -833 
 
 To get actual S the values are multiplied by ra) lt a by ra> 2 , and x by r. 
 
 Discharge and Discharge Curves. The momentary rate of 
 discharge from any pump will be A X velocity, where A rep- 
 resents the effective area of the piston or plunger. A has 
 
194 PUMPING MACHINERY 
 
 various values for the different types of pumps. If D = diam- 
 eter of cylinder and d = diameter of rod, the following results: 
 
 TtD 2 
 
 CASE i. For the plunger pump, A = . 
 
 4 
 
 For the piston pump, on end without rod, . 
 
 CASE 2. , 4 , 
 
 For the piston pump, on end with rod, ( D 2 d 2 } . 
 
 4 \ / 
 
 CASE 3. 
 
 TO 
 
 For the bucket pump down stroke, . 
 
 4 
 
 For the bucket pump up stroke, -f/) 2 d 2 }. 
 
 Differential pump, down stroke, - . 
 
 r i '" .'; 4 
 
 Differential pump, up stroke,- -( D 2 d 2 j . (See Fig. 228.) 
 The discharge then becomes 
 
 ^'; K Total Q-fdQ-A%dt, - : -_. V . (7) 
 
 Q =A I dx=A X distance passed over, 
 
 = ALN'. . . . . . , . , . . . ' . (8) 
 
 L = length of stroke, 
 N'= effective pumping strokes in given time. 
 
 The action of the discharge or suction may be shown by a 
 diagram, Fig. 175, in which abscissae represent time and ordi- 
 
 nates, the quantity A-rr. The area of this curve is Q and the 
 mean height of the curve would be 
 
 **$Q-4!f-ALS (9) 
 
 N" = effective strokes per second. 
 To show the variation of the momentary values from the 
 
DYNAMICS OF WATER END 
 
 195 
 
 30 GO 90 120 150 180 210 240 270 300 330 
 
 30 60 90 120 150 180 210 240 270 300 330 360 
 
 30 60 90 120 150 180 210 240 270 300 330 360 
 
 CaseS 
 
 120 150 180 210 240 270 300 330 360 
 
 30 60 90 120 150 180 210 240 270 300 330 860 
 
 Case 5 
 
 FIG. 175. Pump Discharges. 
 
 mean, the mean value may be plotted (Fig. 175 ) as a dot and 
 dash line on the diagram. Two additional cases have been 
 
196 PUMPING MACHINERY 
 
 added: One, Case 4, in which there are two double-acting pumps 
 and another, Case 5, where three single-acting pumps are used. 
 The series of diagrams are drawn for pumps in which 
 the diameter of the pump barrel is 20 inches; the diameter 
 of the piston rod, where used, 2 inches; the diameter of the 
 equalizing plunger rod, where used, 14 inches, and the stroke 
 30 inches. The speed of the pumps will be taken at sixty 
 revolutions per minute and n=6. The quantities discharged 
 per second for the various cases are as follows: 
 
 .16 
 
 CASE I. q= xLxN"=-~ X-Xi =5-4475- 
 4 
 
 CASE 2. q=\ +-(D*-d*)\xLxN 
 L 4 4 J 
 
 [314.16 311. 
 + 
 
 CASE 3. 2= 
 4 
 
 xd 2 
 
 
 
 4 
 
 144 
 
 *) =1^ =2.160. 
 144 
 
 .16 
 
 CASE 4. q = zq of Case 2. 
 
 A h and A c same as Case 2. 
 CASE 5. q=$q of Case i. 
 
 A =same as Case i. 
 
 The curves show clearly the way the quantity varies as 
 the pump moves. The variation due to the piston rod is 
 
DYNAMICS OF WATER END 
 
 197 
 
 shown in the figure by the different heights of the curve on 
 the forward and back stroke of the double-acting pump. The 
 effect of a number of cylinders on the uniformity of flow in 
 the discharge pipe is seen by comparing the curve from the 
 
 Li and L- 2 represent length 
 of pipes; 
 
 h 2 and h' 2 represent vertical 
 heights of lift. 
 
 FIG. 176. Pump Arrangement. 
 
 single-cylinder pump with the resultant curve from the multi- 
 cylinder pump. 
 
 Forces on Piston. The velocities in the discharge pipe arid 
 suction pipe are dependent on the velocity of the piston, as 
 
198 PUMPING MACHINERY 
 
 water may be considered as incompressible. Since this will 
 mean a variation in pressure due to the change of momentum, 
 it is necessary to consider the forces acting in the suction pipe 
 and delivery pipe. 
 
 Consider one side of the pump, Fig. 176, having a piston 
 area A and a stroke L. To make this a general case the cylinder 
 will be inclined at an angle < to the horizontal. The dimen- 
 sions L represents lengths, while A\, A 2 are the areas of section 
 of the various channels through which the water passes, and h, 
 the pressures measured in feet of water at the various points. 
 
 The first pressure necessary to find is that acting on the 
 lower side of the piston during an upward stroke. This may 
 be written as 
 
 Sm 0), . . .. (lo) 
 
 h 8 = pressure on suction side of piston expressed in feet; 
 /* a =head corresponding to the pressure of the atmosphere; 
 h 2 = height to end of piston stroke from water level; 
 A 3 = friction and other losses except that of inertia; 
 A 4 = force due to inertia expressed in feet head; 
 /r s .v. = friction in suction valve; 
 
 z'sin < = the vertical movement of the piston. 
 
 The pressure h s is the mean pressure on the piston and the 
 actual pressure at the center of the piston; there is a slight 
 variation over the area. 
 
 Losses in Pipes. The friction losses in the suction side 
 are made up of three principal parts: That in the suction 
 pipe, that in the valve boxes, and that in the passages of 
 the pump. The general formula for -this loss at any instant, 
 as given in works on theoretical hydraulics, is 
 
 or using later experimental work, 
 
 h = klv n . .... (12) 
 
DYNAMICS OF WATER END 
 
 199 
 
 The velocity of the water in any member of the pump 
 varies in the same manner as that of the pistcn provided the 
 water stream does not separate but follows the piston. Hence 
 a mean value of the loss per cubic foot must be obtained. 
 Since v 2 varies, the mean loss per cu.ft. is 
 
 2g 
 
 i 
 
 -- dq 
 o di 2g a 
 
 q q 
 
 since dq=AS dt, where 5 = the speed of the piston. 
 Now v\ = . This gives 
 Al, 
 
 To integrate this for any given pipe the values of the coeffi- 
 
 FIG. 177. Curve of fS 3 . 
 
 cient / for different velocities v\ in the pipe are found, and the 
 product fS 3 is found at different intervals of time, or what is 
 the same thing, for different values of 6 and are plotted on a 
 base. This gives a diagram shown in Fig. 177. 
 
 The area of this figure, found by a planimeter or calculated 
 by Simpson's Rules, is the value of the integral and from this 
 the mean h loss may be found. 
 
200 PUMPING MACHINERY 
 
 The instantaneous value of the loss at any point per pound 
 reduces to 
 
 This form of an expression holds for the loss in the pas- 
 sages leading from the valve to the cylinder and through the 
 valve, although there is also a direct loss due to the pressure 
 required to hold up the suction valves. The form will also 
 represent the losses at entrance to the pipe, and other obstruc- 
 tion. 
 
 The values for these various terms will now be discussed. 
 
 For many years it has been customary to consider that the 
 loss of head in a pipe line was given by Eq. (n), 
 
 where h p =\oss of pressure expressed in feet; 
 
 /=a coefficient which varies with the kind of pipe, diam- 
 
 eter, and velocity; 
 /= length of pipe in feet; 
 v= velocity in feet per second; 
 d= diameter of pipe in feet; 
 g = acceleration of gravity in feet per sec. per sec. 
 
 Since / varies with the velocity an attempt was made to 
 eliminate from this term the velocity, and for this purpose the 
 equation was put in the form Eq. (12): 
 
 h=kv n l ...... . (12') 
 
 By plotting the results of many experimenters on log- 
 arithmic paper and by the methods of least squares Reynolds 
 and others have found that n has a value of i for low velocities. 
 The value of n changes to about 1.75 when the velocity exceeds 
 a certain amount. The point at which this change occurs is 
 called the critical velocity. Reynolds showed, by injecting a 
 colored liquid in a fine stream into the water in a glass tube, 
 that so long as the velocity was less than the critical velocity 
 
DYNAMICS OF WATER END 201 
 
 the colored stream remained as a thread, showing that the 
 v ater was traveling in stream lines; but as the critical velocity 
 vras reached the color became mixed with the other water, 
 showing that there were eddies throughout the whole stream. 
 Some have found this same effort by other means. 
 
 If the results of a great number of experimenters are exam- 
 ined, n will be found to have various values, varying from 1.7 
 to 2, rough pipe seeming to give the higher value. The value 
 of k of the formula depends on d, and F. C. Lea in his " Hy- 
 draulics " shows that 
 
 k=kd- ij *. 
 
 This gives, then, 
 
 , k'iv 1 - 75 k'lv 
 
 to 
 
 There are thus two formulae giving the loss in a pipe line, 
 one in which the loss is expressed in terms of the square of the 
 velocity and a factor which varies with the diameter and velocity, 
 and another in which the loss is expressed in terms of powers 
 of the velocity and diameter and a factor which is a constant 
 for one kind of pipe. 
 
 For general purposes in cast-iron pipes the loss given by the 
 two formulae is by the approximate forms : 
 
 For ordinary cast- "-^ U T ( J 5') 
 
 iron pipe, not 
 clean 
 
 V 2 
 
 h s = 0.00055^ 
 
 The value of / for clean pipe is given in tables in books such 
 as Merriman's " Hydraulics " and Lea's " Hydraulics." The 
 tabular value for / is about 0.02 for clean pipe when an average 
 value only is needed. For an exact value when the table is 
 net at hand the following may be used for clean pipe: 
 
 , 0.026 
 
202 PUMPING MACHINERY 
 
 -,2 
 
 The two expressions for Joss may be put in the form h s = 
 
 where has the value/- or 125 25 . In each case has a value 
 
 7 
 - 125 
 
 which varies with the velocity to a small degree, and as the other 
 losses are expressed in terms of fl 2 , it is well to express this loss 
 in the same manner. 
 
 The less at entrance into a pipe line depends on the arrange- 
 ment of the end. If a plain end flush with the wall of the fore 
 bay is used the loss at entrance has been shown (Merriman's 
 
 V 2 
 
 " Hydraulics/' page 207) to be 0.49-^, while if the pipe pro- 
 
 2 
 
 jects through the wall the coefficient becomes 0.93, and if a 
 mouthpiece is used on the end of the pipe the coefficient is 
 very small. In general it may be taken that the loss at en- 
 trance is given by the equation 
 
 *.=- 
 
 The loss due to bends and obstructions, such as valves, is given 
 in the form 
 
 7 z_r I" / \ 
 
 h = kf, ....... (19) 
 
 d2g 
 
 for bends, and 
 
 h = m , . . . -v . v . (20) 
 
 2 
 
 for valves and cocks, where k is a multiplier of the / of straight 
 pipe and has values depending on the ratio of the radius of 
 the bend to the radius of the pipe while / is the length of the center 
 line of the bend, m depends on the amount the valve or cock 
 is closed. 
 
 The values of k and m for the different cases are given in 
 the forms of curves (Figs. 178, 179, 180). These have been 
 constructed from the reported results of Williams, Hubbel 
 and Fenkell, Weisbach and Grashof. In these R = radius of 
 pipe bend of 90, d = diameter of pipe, the quantity d f = amount 
 
DYNAMICS OF WATER END 
 
 203 
 
 
 
 
 
 
 
 
 
 k - 
 
 
 2.0 4 
 
 
 t 
 
 
 
 
 
 
 
 
 
 
 
 
 1.8 ^- 
 
 
 
 
 
 
 
 
 \, 
 
 
 
 
 
 
 
 
 1.6 - ^ 
 
 
 
 
 __s v _ 
 
 
 
 V^ 
 
 
 
 1.4 ---------- -----I~I-~ ----!-- --L. 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 1.2 
 
 
 5 10 1 
 
 5 , FU 20 25 30 
 
 V rl ' 
 
 FIG. 178. Values of k for Loss in Bends. . 
 
 10 
 100 
 
 FIG. 179. Values of m for Gate Valves. 
 (For Upper Curve use Lower Values.) 
 
 12 
 120 
 
 14, 
 140 
 
204 
 
 PUMPING MACHINERY 
 
 80 120 160 300 240 
 
 FIG. 180. Values of m for Cocks and Butterfly Valves. 
 
 Upper Curve Cocks. 
 
 Lower Curve Butterfly Valves. 
 
 the valve has been closed, and 6 the angle the cock or butter- 
 fly valve has been turned. Fig. 181 shows the form of valves 
 used. 
 
 The loss due to sudden enlargement which occurs when the 
 
 FIG. 181. Gate Valve, Cock and Butterfly Valve. 
 
 water enters the cylinder from the valves or from the suction 
 pipe is equal to 
 
 h fa-") 2 
 
 in which v s is the velocity before enlargement and v the velocity 
 after this occurs. 
 
DYNAMICS OF WATER END 205 
 
 The loss due to sudden contraction is much less for the 
 same change of section, and is equal to 
 
 2,, 2 
 
 where 
 
 , - . . , x 
 
 =0.582+-^-, ...... (23) 
 
 ^ = ratio of the diameter of the small pipe to that of the large pipe. 
 
 The losses of pressure include those of velocity heads, and 
 
 since the water has finally a velocity 5 in the pump, the velocity 
 
 S 2 
 head, , must be one of the terms of the reduction in pressure, 
 
 2g 
 
 as there was originally no velocity in the fore bay or place 
 
 AS 
 from which the water is drawn. The initial velocity, , 
 
 AI 
 
 in the suction pipe is equivalent to, and requires a pressure 
 
 i S 2 
 
 , which is greater than . This greater velocity 
 
 2 2g 
 
 head, however, is changed into pressure head as the velocity 
 decreases, so that when the piston is reached the only part 
 of velocity head left which requires a pressure from the piston 
 
 . S 2 
 is . 
 
 2g 
 
 Lach less may be reduced by means of the equation 
 
 so that each term is expressed as a function of S 2 . 
 Each equation reduce's then to the form 
 
 A 
 
 These various equations give the values for each one of 
 the losses, and hence the expression 
 
206 PUMPING MACHINERY 
 
 may be found. This represents the variation in head due to 
 the pipe line. 
 
 Water Inertia Force. The loss h is computed by considering 
 the inertia of the water. 
 
 The masses of water in the various parts of the pump are 
 as follows: 
 
 In the pump Aocw in pounds. 
 
 In the pump spaces A 2 L 2 w in pounds. 
 
 In suction pipe A\L\w in pounds. 
 
 It is to be remembered that L designates the length of the 
 pipe and not the lift necessarily. 
 
 These masses when multiplied by the acceleration will give 
 the force required to accelerate them in poundals, while by 
 dividing by g the result is in engineering units of pounds. It 
 is to be noted that the acceleration of the piston is felt in all 
 of the water in the system, and is inversely proportional to 
 the areas, as in the case with the velocities. From these the 
 following formulae may be derived, remembering that the force 
 is exerted over the areas, A, AI, A 2 , etc.: 
 
 Axwa , , 
 
 - ; . . . ... .... (25) 
 
 6 
 
 A 2 L 2 wa 2 A 2 L 2 wAa , ., 
 
 =^--j - ; .... (26) 
 
 g gA 2 
 
 A\L\wa\ A\L\wAa 
 
 , N 
 
 - .(27) 
 
 These are typical terms, and if there is a series of changes 
 in the lines these reduce to 
 
 (28) 
 
 g\ ^2 -AI/ 
 
 where 
 
 x = r\ i cos 0+ sin 2 ; 
 L 2n J 
 
 5 = roj sin 0+ sin 20 ; 
 L 2n J 
 
 a = ru> 2 \ cos 0H cos 20 . 
 
 
 
DYNAMICS OF WATER END 207 
 
 This total term, h\, may be divided into two parts, one - 
 
 and the other z + 
 
 \ Ai A 2 /g 
 
 The value of xa is found by multiplying the expressions for 
 x and a in Eqs. (2) and (4) or by multiplying the values of x 
 and a if these have been computed as on page 193, where the 
 value of n is 6. 
 
 By the first method: 
 
 xa 
 
 = im i -cos 0+ sin 2 6 cos 0+- cos 26 
 I 2n J I. n J 
 
 = ^ 2 r 2 ] cos 6 cos 2 0H sin 2 6 cos 0H cos 20 cos cos 20 
 L 2^ n n 
 
 +-i- sin 2 cos 20 . 
 
 2W 2 J 
 
 The third and last terms of this could be omitted, but the 
 other terms are of importance; thus the approximate value 
 would be 
 
 xa = & 2 r 2 [cos 0(i cos 0) + - cos 20(i cos 0)] 
 
 n 
 
 = 0)V[cOS0-f- COS20][l-COS0]. ...... (29) 
 
 By the second method for n = 6, the values are: 
 
 e= 
 
 15 
 
 30 
 
 45 
 
 60 
 
 75 
 
 90 
 
 xa 
 
 
 
 
 
 
 
 tfV 2 ~ 
 
 0.044 
 
 0.147 
 
 0.237 
 
 0-235 
 
 0.093 
 
 -0.181 
 
 6= 
 
 105 
 
 1 2O 
 
 135 
 
 150 
 
 165 
 
 180 
 
 xa 
 
 
 
 
 
 
 
 
 -0.592 
 
 O.9II 
 
 -1.236 
 
 -1-494 
 
 1.621 
 
 -1.667 
 
 W 2 r 2 
 
 
 
 
 
 
 
 These values can be used to find the inertia force at any 
 position of crank and piston for a given pump, determining 
 the quantity h of equations (10) and (28). 
 
 Valve Losses. One term to require explanation is the 
 resistance h sv of the suction valve. The discussion following 
 is due to Hartmann and Knoke and is based on the work of 
 Bach. 
 
208 
 
 PUMPING MACHINERY 
 
 There are three types of valves: lift or beat valves, clack 
 valves, and slide valves, as shown in Figs. 182, 183, and 184. 
 
 FIG. 182. Lift Valve. FIG. 183. Clack Valve. FIG. 184. Slide Valve. 
 
 The valves serve the purpose of alternately connecting the 
 pump cylinder to the suction or discharge pipe and with the 
 exception of the last form of valve they are operated by 
 the water of the pump. The slide valve is operated by some 
 external means as an eccentric. At times, however, the clack 
 or lift valve may be operated in part by some positive gear. 
 
 Now the resistance of the passage of water through valves 
 as investigated by Bach is divided into two parts: (a) that 
 due to the opening of the valve, and (b) that due to the passage 
 of the water through the valves. 
 
 Pressure Drop at Opening of Valves. The valve has the 
 velocity and acceleration of the water under it, the latter being 
 
 A 
 
 a-;-, where A p is the area of the passage beneath the valve. 
 A p 
 
 W a A 
 If the valve weight is W v its inertia will be - - -j- . If the 
 
 6 ^-P 
 
 total spring pressure is taken as F s and the area of the upper 
 side of the valve is A v , the following equation exists between 
 pressures above and below the valve, when valve is just opening: 
 
 6 
 
 pip u = loss in pressure in opening valve. 
 A,-A\ .W.F. W,aA 
 
DYNAMICS OF WATER END 209 
 
 On the upper side of the discharge valve the following 
 relation holds: 
 
 where h a represents the atmospheric pressure, H d the static 
 
 pressure in the pipe line, and - - -j-a the force required to 
 
 **d 
 
 accelerate the water in the discharge pipe. 
 
 This value may be determined in any given case before 
 finding the quantity hdv given below for loss in delivery valve. 
 
 This may be expressed in head as 
 
 For the suction valve this equation becomes: 
 
 TT/" A 
 
 pi A p = puA v + Wv + F 8 + ~<TT-. 
 
 6 ^-P 
 
 Wv F. Wv A 
 A- + T p + -i a A^\' (33) 
 
 an expression exactly similar in form. In this^case, however, 
 it is better to find pi originally by the formula: 
 
 . y <. . (34) 
 
 to suction valve, i.e., lift, losses and velocity 
 head, and L = length to this point) and then p u by the 
 formula: 
 
 W A 
 
 A p p l ~A v pu + W.+F, + -^- r a. . . . (35) 
 
 6 ^P 
 
 Then 
 
 ++ - w 
 
 Before numerical values can be found for these expressions 
 the value of F s must be determined. The Eqs. (32) and (36) 
 show the dependence of the pressure required to open the 
 
210 PUMPING MACHINERY 
 
 valve on the pressure above the valve, the weight of the valve, 
 the spring pressure, and the acceleration which depends on the 
 speed of the pump in revolutions per minute. The resistance 
 or pressure increases with all of these. The effect of area in 
 
 -^^3 KvW AkAui I! ...ju^ A<A___xl>.V%- . ' ~3w A"l "Ivvv 
 
 MSSSSl {^^^^^m_^ ^^^^^x^^^ imm> 
 
 i ii in iv y 
 
 FIG. 185. Valve Forms. 
 
 decreasing the value of this resistance shows what an impor- 
 tant part the valve area plays in the design of pumps. The 
 pressure drops considered above occur as valve is just opening. 
 Valve Friction. The resistance of valves to the flow of 
 water beneath them during the open period of the valve has 
 been found to be given by the expression 
 
 (37) 
 
 where v\= velocity of water in passage leading to valves; h v = 
 the loss in feet head while valve is open; ^ = a coefficient, the 
 value of which is given by formulae below, depends en the 
 form of valve and has been investigated for the forms shown 
 in Fig. 185 in which Form I is a plain disc valve with no 
 guides, Form II is a disc with guide vanes attached to the 
 lower face, Form III the conical valve with no lower guides, 
 Form IV a complete conical valve, and Form V a spherical valve. 
 
 , for valves of Forms I and IV. . . (38) 
 for valves of Form IL (39) 
 
 \ 2 
 
 ; = a +ft- + r (-) , for valves of Forms III and V. (40) 
 
 Bach gives the various values for the constants and pro- 
 portions for the parts which reduce to the forms below: 
 
DYNAMICS OF WATER END 211 
 
 For Form I: 
 
 i, dl + dl 
 h = to ; 
 
 10 4 
 
 =0.55 +4 
 
 /? = o.i5 to 0.16. 
 
 h = rise of valve in feet; 
 di = diameter of passage leading to valve; 
 
 bi = breadth of seat in feet = to ; 
 
 10 4' 
 
 i= number of ribs; 
 
 s = thickness of ribs at end. 
 
 For Form II: 
 
 d, d, 
 h = ^ to -; 
 4 
 
 10 4 
 
 a = a for Form I multiplied by 0.8 to 1.6; 
 /? = i.7o to 1.75. 
 
 For Form III: 
 
 di di 
 
 h= to , &i = 
 10 4 
 
 a =2.6; 
 /?=-o.3; 
 ^=0.14. 
 For Form IV: 
 
 L dl fn ^1. 
 k= ~8 13 4' 
 
 a =0.6; 
 /? = o.i5; 
 For Form V: 
 
 d l d l 
 h= to ; 
 TO 4 
 
 a =2.7; 
 /?=-o.8; 
 
212 PUMPING MACHINERY 
 
 The values of these coefficients were determined when the 
 clear space between the valves was 1.8, the area of the passages 
 between the valves, or 
 
 d 2 = diameter of space to point midway between valves, 
 d = diameter of valve, and 
 di = diameter of passage. 
 
 With these coefficients the loss through the valve may be 
 determined if the value vi is known. This will vary with the 
 speed of the plunger. 
 
 The discharge area beneath a valve changes as the valve 
 lifts, and it may change so that the velocity under the valve, 
 designated by c, may remain constant. 
 
 (41) 
 
 where A =area of piston or plunger; 
 
 5= speed of plunger at any instant; 
 ^4 p =area of passage beneath valve; 
 vi = velocity of water in passage beneath valve; 
 T = coefficient of discharge; 
 v = velocity of discharge radially; 
 d= outside diameter of valve; 
 h = lift of valve. 
 
 In this case, however, the movement of the valve itseif 
 will cause a change in the velocity v or in the height h, because 
 
 r.d 2 
 a quantity of water 02 , where V2 is the velocity of the valve, 
 
 is held beneath the valve during the lifting of it or is dis- 
 charged on the dropping of the valve. This then gives 
 
 2 , .... (42) 
 calling xd = l, 
 
 .AS = A P vi = rvlhA v V2 ..... (43) 
 
 (+ during rise of valve; during fall.) 
 
DYNAMICS OF WATER END 
 
 213 
 
 This means that the motion of the water in the passage 
 below the valve relative to the valve is viV2, and therefore 
 
 b _ 
 c ~ 
 
 / N 
 (44) 
 
 Hartmann and Knoke investigated this loss during the 
 stroke of a pump, finding the values 
 of h at different parts of a stroke 
 experimentally by an indicator whose 
 pencil motion was attached to the 
 valve. From this they computed 
 by Eq. (38) at various points. They 
 then computed the velocity of the 
 water c and from the curve of move- 
 ment of the valve determined v%. The 
 computation of h showed that the 
 head over a large part of the stroke 
 
 FIG. 186. Valve Friction. 
 
 was practically constant, the variations. 
 
 being at the end. The curve found was similar to Fig. 186. 
 This showed that the loss could be computed as if the water 
 were moving at the speed determined by the crank position 
 of 90 with a zero velocity of the valve or at a point where the 
 valve has reached the top of its stroke. 
 Here 
 
 yioQo Atoff . ^ 
 
 -J-. (45) 
 
 A 
 
 and 
 
 (46) 
 
 Size of Valves. To determine the vajues of and h v and 
 also to find h sv and h dv the size of the valve or the number of 
 valves of a certain size for a given pump must be known. The 
 following discussion may be used to determine the size or num- 
 ber of the valves used on a pump. 
 
 Now near the center of the stroke 
 
 A p w 
 
 (47) 
 
214 PUMPING MACHINERY 
 
 where F s is the spring tension, W v is the weight of the valve, 
 and h c . is the resistance from the valve expressed in feet of 
 water. From this 
 
 (48) 
 
 ^^ C ^ ^~Lp(JU 
 
 Now, 
 
 where f is the coefficient of discharge beneath the valve and v 
 is the radial velocity through the valve opening of lift h. 
 
 flh 
 
 or 
 
 i W,+F,A 9 . . 
 
 7=\/ 2 A~ ~77 > (49) 
 
 or using the area A v of the valve and a new coefficient by 
 analogy this becomes 
 
 It is to be noted that this is expressed in terms of feet head 
 of water exerted over the complete valve area A c . 
 The equation then given for h is 
 
 h = 
 
 T^ / W,+F.' 
 
 V^\ 2g Aw 
 or 
 
 B.^ 
 
 ......... (51) 
 
 where b expresses the feet of water equivalent to the pressure 
 
 Y 
 
 on the whole area of the valve and n = ~y=-. 
 
 V tt 
 
 Since 
 
 or 
 
 n>=/A/2& ....... . (52) 
 
DYNAMICS OF WATER END 215 
 
 This formula is true for a definite velocity v and a definite 
 spring force F for the given compression. It is necessary to 
 investigate the value of /x from actual tests. Equation (43) 
 becomes, at the end of the stroke, 
 
 fv A ird 2 
 
 I ho 4irdho 4/^0* 
 
 In this ho is the height of the valve from the seat at the end 
 of the stroke. Showing that at the end of the stroke the valve 
 is not' on its seat. This might permit slip to occur From (52) 
 
 is value of b at this time 
 
 4- ...... ... (S3) 
 
 V2 40 
 
 If V2 is assumed to vary with bo it is evident that /* will vary 
 
 4-. 
 
 4^0 
 
 Hartmann and Knoke measured the movement of the valves 
 of a certain pump with given springs by means of an indicator 
 and in this they could measure the quantities of Eq. (43) to 
 find pi. By measurements they obtained b and using this and 
 the value of "p (52) the value of M could be found and plotted 
 
 against of different speeds. The results of this are seen in 
 4 
 
 Table I below and Fig. 187. 
 
 It is seen that /* has a maximum value at about -7=25. 
 
 4# 
 
 There is little variation from this when -7 = 50. This latter 
 
 value gives a va'ue of M slightly different from the previous 
 value, but does give a value of h one-half as large as before for a 
 given value of d. This is of importance in fixing the sizes of 
 values for initial discharge. In general ho is made -^d and 
 for this n = 0.845 or '85 
 
216 
 
 PUMPING MACHINERY 
 
 TABLE I 
 
 VALUES OF 
 
 
 h 
 
 d 
 
 
 
 k 
 
 d 
 
 
 mm. 
 
 ins. 
 
 4* 
 
 ti 
 
 mm. 
 
 ins. 
 
 . 4 h 
 
 in 
 
 o.o 
 
 0.000 
 
 
 o . 650 
 
 6.0 
 
 0.236 
 
 2.50 
 
 -53 2 
 
 . i 
 
 .004 
 
 150 .OO 
 
 .710 
 
 6-5 
 
 .256 
 
 2.31 
 
 523 
 
 . 2 
 
 .008 
 
 75.00 
 
 .780 
 
 7.0 
 
 .2 7 6 
 
 2.14 
 
 3*3 
 
 3 
 
 .012 
 
 5 
 
 845 
 
 7-5 
 
 295 
 
 .00 
 
 57 
 
 4 
 
 .016 
 
 37-5 
 
 .890 
 
 8.0 
 
 315 
 
 .87 
 
 5 
 
 5 
 
 .020 
 
 30.00 
 
 .911 
 
 8-5 
 
 335 
 
 .76 
 
 493 
 
 .6 
 
 .024 
 
 25.00 
 
 913 
 
 9.0 
 
 354 
 
 .67 
 
 485 
 
 .8 
 
 .031 
 
 i8-75 
 
 .902 
 
 9-5 
 
 374 
 
 58 
 
 477 
 
 I .0 
 
 039 
 
 15 .00 
 
 .870 
 
 IO.O 
 
 394 
 
 5 
 
 .472 
 
 J -5 
 
 059 
 
 10.00 
 
 .788 
 
 ii .0 
 
 433 
 
 36 
 
 459 
 
 2 .O 
 
 .079 
 
 7-50 
 
 73 2 
 
 12 .O 
 
 '472 
 
 2 5 
 
 445 
 
 2-5 
 
 .098 
 
 6.00 
 
 .690 
 
 I 3 .0 
 
 512 
 
 i5 
 
 431 
 
 3-o 
 
 .118 
 
 5 - 
 
 .650 
 
 14 . o 
 
 55 1 
 
 .07 
 
 .420 
 
 3-5 
 
 138 
 
 4.28 
 
 .622 
 
 I 5- 
 
 591 
 
 .00 
 
 .407 
 
 4.0 
 
 J 57 
 
 3-75 
 
 599 
 
 16.0 
 
 .630 
 
 94 
 
 395 
 
 4-5 
 
 .177 
 
 3-33 
 
 578 
 
 17.0 
 
 .669 
 
 .88 
 
 -381 
 
 5-o 
 
 .197 
 
 3 - 
 
 .560 
 
 18.0 
 
 .709 
 
 33 
 
 37 
 
 5-5 
 
 .217 
 
 2-73 
 
 545 
 
 
 
 
 
 (From Hartmann-Knoke, "Die Pumpen.") 
 
 At any point of the stroke A=I o 
 
 0.9 
 fvlh=ApViA c V2, From (52) 
 
 dh d /ASA,V2\ 
 
 (54) 
 
 At the end of the stroke v% 
 may be assumed constant and 
 small of value, while at a crank 
 
 0.8 
 0.7 
 0.6 
 0.5 
 
 0.4 
 0.3 
 0.2 
 0.1 
 
 01234 6 8 10 12 14 16 18=7i 
 
 position of 90 the valves may FlG " ' 8 7-V*tion of jM*h Vaccord- 
 
 ing to Hartman and Knoke. 
 
 be assumed to be wide open 
 
 and V2 may be taken as zero. At each of these points b may be 
 
 considered constant. These assumptions then give the following: 
 
DYNAMICS OF WATER END 217 
 
 * 
 
 Aru> 2 ( cos 8+- cos 26 ) 
 
 dt \ n ] 
 
 Mo 
 at 90 
 
 If now the value of bo be assumed and ho be equated to 
 the size of the valve or valves may be found from (55). Eq. 
 (56) may be used to find A max if & max is assumed or by assuming 
 /jtnax the value of b max may be found. Of course the value of 
 
 ju depends on the value -7 and must be assumed and then the 
 
 correct value used for the second approximation. The equa- 
 tions above are reduced into other forms. 
 
 N 
 Air- =q= A2T , 
 
 60 " 27r' 
 
 , A pTT^CO 4 60 ___ I 
 
 = = (0.85 X 2 ~ 20 ' 
 
 ^ T q , . 
 
 bod== ~ =^> =o ' ii2 w N or aii2 4-/^ (57) 
 
 2 
 
 or 
 
 N 
 
 27T 
 
 - (58) 
 
 In these two formulas bo may be assumed (usually as i to 3 
 feet) and / or d found. 
 
218 PUMPING MACHINERY 
 
 If one valve is used q cared for by this valve is found, but 
 if n f valves are used each valve is designed for t cubic feet of 
 
 water per second. 
 
 The Eq. (56) becomes , 
 
 7 QT Q / \ 
 
 /*max= == ^== (59) 
 
 In this /^max is assumed and then from the value of 
 
 n is found in Fig. 187 and then Z> max is found. 
 
 The force exerted by a spring is proportional to the amount 
 of compression and in general if yo is the amount of compression 
 the force F 3 is given by 
 
 when C is the force required to compress the spring i foot. 
 
 If yo is the amount of initial compression in the spring in 
 feet the values of b and 6 max are used in the following equations: 
 
 (6o) 
 
 _ 
 - ;;.v. . (60 
 
 Solving for yo and C there results 
 
 \ . ...... (62) 
 
 Z, f*\ 
 
 - 4 no ..... (63) 
 
 These are the constants for the spring to give the desired 
 pressures, and are used in fixing the size of the spring wire, 
 number of turns and other dimensions. 
 
 The formulae just used were for definite velocities, but actual 
 velocities are variable. This will change certain of the forms, 
 as observed before in the discussion of the resistance to the 
 passage of the water, 
 
DYNAMICS OF WATER END 219 
 
 The movement of the valve and its change in velocity intro- 
 duces a change in the term b, which has been used to denote 
 
 W,+F, 
 
 A v w 
 
 On account of friction of the valve the force to lift is given by 
 P = W,+F S R, (64) 
 
 where R = resistance due to friction. 
 
 By properly designing the valve and its connections the 
 term R may be made small. Hence, although not strictly 
 correct, the resistance to valve motion may be considered as 
 Wv+Fs or when measured in feet of water on area of valve, 
 W,+F. 
 A v w 
 
 Valve Pressure for Stroke. The pressure loss for the valve 
 varied from k n or h^ Eqs. (36) and (33), at the beginning of 
 the stroke to fe, Eq. (46), during the major part of the stroke 
 as shown in Fig. 186. 
 
 Simple Methods of Design of Size and Number of Valves. 
 The method used above for the design of valves is based on 
 theory. Two practical rules for the same purpose used in 
 the United States are those due to Hague and to Reynolds. 
 Mr. C. A. Hague recommends using 50 per cent of the piston 
 area for the valve area on discharge or suction side when the 
 piston speed is 100 feet per minute and 150 per cent when the 
 piston speed is 300 feet per minute. For intermediate piston 
 speeds a proportional value of the percentage is used. Mr. I. 
 H. Reynolds uses a very simple rule: i square foot of valve 
 area is required on suction and on discharge for each million 
 gallons capacity. In this way the piston speed need not be 
 considered, as it is the quantity of water in a given time which 
 fixes the area of the valves. 
 
 The springs in this case are not designed but sizes and diam- 
 eter are assumed as shown on page 317. 
 
 All points have now been covered in the design of valves 
 and the value of for the valves may be found and the results 
 computed. The general equation for h s may be formed. 
 
220 PVMPING MACHINERY 
 
 The variation in pressure of the water on the bottom of 
 the piston on the up stroke may now be written: 
 
 The quantities h no or h dvo on discharge are the resistances at 
 the time of opening the valves. Their values may be quite 
 large, and in order to reduce them and the acceleration term 
 air chambers are used, as will be seen later. It is well to 
 remember that h s , or the pressure under the piston during the 
 suction stroke, can never be less than zero. This means that 
 the sum 1 of the resistances and the lift on the suction side must 
 be less than 34 feet. In many cases pumps are run so fast that 
 this quantity is exceeded causing the water column to break, 
 and pounding results. 
 
 In designing pumps with clack valves the same methods 
 as used with the lift valves may be employed, remembering 
 that moments of forces about the pivot must be considered 
 and also that the perimeter of discharge is not the complete 
 perimeter of the valve. The lift h may be taken as the lift 
 at the outer end. 
 
 Problem in Valve Size and Pressures. A problem will be 
 computed to illustrate the use of the above formulae: 
 
 Suppose a 20 by 30-inch duplex single-acting pump is driven 
 at 60 R.P.M. and that there is 9 feet of 1 8-inch suction pipe 
 with one bend followed by i foot lift at pump to deck of valves, 
 and 300 feet of 1 8-inch discharge pipe, rising 200 feet and con- 
 taining three right-angle bends of 6-foot radius. The number 
 of valves and the variation of suction pressure will be found. 
 
 Losses in Pipe Lines. 
 
 Area of suction and discharge pipes: 
 
 ^(i. S )*-i.77. 
 
 4 
 
DYNAMICS OF WATER END 221 
 
 Area of water piston 
 
 4 \ 12 / 
 Loss at entrance 
 
 2g \i-777 
 In suction pipe 
 
 d 2g 
 For dirty pipes 
 
 /-0.03; 
 
 ._ Q 
 
 \I-77/ 2g 2g 
 
 One bend 
 
 = 6 feet; 
 
 k}~ d = 2. 15X0.03X^=0.41; 
 
 / 2 .i8\ 2 5 2 , 5 2 
 ^ = 0.41 (- - = 0.01 . 
 
 \1.77/ 2g 2g 
 
 Sudden enlargement into suction chamber, 
 
 / AA 2 v s 2 [ i.77\ 2 /2.i8\ 2 5 2 S 2 
 
 h= i = i -- " ' = 0-0 
 
 -05 
 
 A / 2g \ 2.l6/ \1.77/ 2g 2g 
 
 The total term will then be 
 
 (A \S 2 S 2 S 2 
 
 2w- hi = [l. 5+0.27-4-0.61 +0.05 + 1] =3.40 . 
 Al )2g 2g 2g 
 
 The terms show the importance of the various friction losses. 
 In the present instance it is seen that the loss at entrance is 
 the most important item. 
 
 On the discharge side the equivalents of these terms are 
 given below: 
 
222 PUMPING MACHINERY 
 
 Loss in friction of pipe 
 
 = 0.03X300/2. i8\ 2 5 2 
 i-5 \-77/ 2g 
 
 S 2 
 = 9. 
 
 Loss in bends 
 
 Loss due to sudden contraction 
 
 Final velocity head 
 
 S 2 
 
 0.04 . 
 
 i.77 
 S 2 
 
 TzrVsT - 3 
 
 The next term of the expression for h, is that involving 
 the acceleration. Let 
 
 Ai 1.77 
 and 
 
 ^^ j 
 
 This total term may be divided into two parts, one and 
 
 o 
 
 the other + ^- The latter equals ^a = . 37 a. 
 
 AI A 2 >g 3 2 - 2 
 
DYNAMICS OF WATER END 223 
 
 The value of xa is found by multiplying the expression for 
 X and a in Eqs. (2) and (4) or by multiplying the values of x, 
 and a if the numerical values of x and a at different angles are 
 found as shown on page 207 in terms of r 2 u 2 . 
 
 These values may be plotted for different piston positions. 
 
 Size and Number of Valves. The American method of 
 designing pumps is to use a large number of small valves in 
 place of a small number of large valves. 
 
 The reason for this is the fact that since the area of a valve 
 varies as the square of its diameter, and its discharge area 
 varies with the first power only, smaller valves have a larger 
 discharging area for the same area of valve deck. To util'ze 
 the central part of large valves, they are made multiported, 
 as shown in Fig. 188. It is this type of valve which will be 
 compared with the small American valves For one pump 
 cylinder: 
 
 A N 30X314. i6 v 60 , . ,. 
 
 q = 2rA = - - ~ X = 5.5 cubic feet per sec. 
 
 DO 1728 OO 
 
 0.1124 ,N 
 6 =-~-; 
 
 assuming 
 
 2 ft. 
 
 0.1124X5.5X60 
 - - 
 
 = 57 valves; 
 
 fo = -!-</ = =0.00167 ft. 
 
 200 200 
 
 To compare the results for the 4-inch valves with another 
 method used to determine the number of valves, consider the 
 rule that the velocity through the area occupied by the valves 
 should be between 100 and 125 feet per minute. Since the 
 water is delivered through the valves in one stroke or a half 
 revolution, this gives 
 
 29X60 ^2X5.5X60 
 ~ 100 X (area)" '_4^ '^' 
 
 100 X 144 
 
224 
 
 PUMPING MACHINERY 
 
 FIG. 188. Multiported Valve. 
 
 American practice is to use 4-inch valve, and this diameter 
 will be assumed. 
 
DYNAMICS OF WATER END 225 
 
 This means about 200 to 250 feet per minute through the 
 opening in the valve deck. If this is figured on the maximum 
 velocity through the opening in place of the mean this number 
 
 75 will have to be increased by the ratio ofir:2or^ = -X75 = n8. 
 
 This result is different from the previous result. Using the 
 Hague and Reynolds methods the following results: 
 
 Piston speed 2LN = 2 X if X 60 = 300. 
 
 Hague's method: 
 
 Area piston =2.18 sq. ft.; 
 
 Valve area =1.50X2.18 -=-3. 25 sq. ft.; 
 
 Number of valves = =37 valves. 
 
 49 
 Reynolds' method: 
 
 5.5X60X60X24X1728 
 Capacity of f pump - =3,570,000 
 
 gallons per 24 hours. 
 Area = 3. 5 7 sq. ft. 
 
 Number of valves -^ =42 valves. 
 
 7T I 
 
 49 
 
 These two results agree fairly well, but they differ from the 
 value 57 first obtained. The reason for this difference is the 
 fact that bo = 2 feet is smaller than the amount used in either of 
 these latter methods. 
 
 For the multiported valve several formulae will be derived 
 referring to Fig. 188: 
 
 for each 
 
 (66) 
 (67) 
 
226 PUMPING MACHINERY 
 
 Hence 
 
 A, ?A irbiZd bi 
 
 ~~- ; ..... (68) 
 
 i A, bi 
 
 h Q = -- - = . 
 50 / 100 
 
 From Fig. 184 
 
 Now W = - 
 
 Assuming, as before, bo = 2 ft, 
 
 .. 
 
 27T 
 
 From the figure 
 
 -(d 1 -x) 2 +ird 1 X+Trd 2 X+ . . . =-[ 
 4 4 
 
 or 
 
 irnx 2 + wn 
 
 n 2 -n 
 Assuming d\ = 6 in. and n = 5, 
 
 
 
 Let ^1=0. 19 ft; 
 
 62=0.15 ft; 
 d ut=di+(2n- 1)^ = ^+9X0.34 = 3. 56; 
 
 , I , O.IO 
 
 ho = bi= = 0.0019. 
 100 100 
 
DYNAMICS OF WATER END 227 
 
 Comparing results, assuming 50 per cent as the excess area 
 required for valve deck, it is seen that the 57 valves require 
 745 square feet of area while the multiported valve requires 
 12.7 square feet. 
 
 This calculation shows the necessity of a large number of 
 valves when a high-speed pump is used. If the speed of the 
 pump is decreased the number of valves will be diminished in 
 proportion to the square of the speed, as the number varies 
 as the product qN, in which q depends on N. Thus for a 20- 
 revolution pump the number of valves would be decreased 
 to iXii2 = i2, and the diameter of the multiported valve 
 would be changed to 7X3.35, or 1.12 feet. The area required 
 would be 0.83 square foot and 1.94 square feet, respectively, 
 for the small valves and the multiported valve. By changing 
 bo the number and size could be diminished. This, with a 
 3-foot pressure for b Q in the original calculation, results in the 
 following: 
 
 n = $S. 
 
 Jo = 2. 5 approx. 
 
 Area valve deck =4.96 sq. ft. 
 
 Area multiported valve = 3 : 52 sq. ft. 
 
 To find the spring pressure at the position ho use the follow- 
 ing formula: 
 
 For small valve 
 
 6i.5X2X^ 
 144 
 
 assuming W v = %lbj 
 
 10.90=1+^0; 
 
 .'. FQ = 10.40. 
 For large valves 
 
 62.5X2X^X0.19X9.3 = 
 
 6Q3= 
 Let 
 
 PF P = 9 ol 
 
 /. ^0 = 603. 
 
228 PUMPING MACHINERY 
 
 Now 
 
 Since h msix will be greater than ho, consider it as 0.035 f 
 the first approximation for small valves. 
 
 - -S" - = ^ = 2.4; 
 
 M=-53; 
 
 6m= WV^J = \57Xo.53X8.o2X.33X.o35J = 3 -87 ft. 
 For the large valve let 
 
 h m = .02 
 so as to give less distance for this large valve to seat, 
 
 2km 2X.02 
 
 Hence 
 
 ^1 = 0.64; 
 
 _( 5-5 r V n, ft 
 
 \.64X8.02X 2 Xg.3X.02/ 
 
 Since 
 
 W,+F = 62. 5 X8.4X7TX. 19X9.3 = 2920. 
 
 Spring Constants. In the case of the small valve, the spring 
 tension which was 10.40 Ibs. at .0017 movement will be 20.6 Ibs. 
 
 20.6 10.40 
 at .035 movement, or / - - x =26 Ib. per inch of com- 
 
 pression (approx.). 
 
 For the large valve this becomes 
 
 (28^0) 603 2227 
 
 T^- ON = ' =9660 Ib. per inch, approx. 
 
 (.02 .0008)12 .0192X12 
 
DYNAMICS OF WATER END 229 
 
 In this case there would be several springs to hold such a 
 large valve to its seat, say ten. This would give 966 Ib. per 
 inch for the modulus of these springs. The large valve as 
 designed would be very cumbersome and would undoubtedly 
 be divided among a number of smaller valves, changing the 
 results considerably. It is suggested that the student work 
 out the data, using six valves of the same style in place of one 
 large valve. 
 
 The method of designing the springs will be considered 
 in the next chapter. 
 
 Valve Resistance. The resistance of the 4-inch valves during 
 the operation of the pump is 
 
 For disc valves the diameter of the passage is 4 2X1^ = 3.2 
 inches or 0.27 foot and from this the following results: 
 
 ;-wr 
 
 = .55+0.00+8.9 = 9.4 
 
 1.18 
 
 64.32X3.28 I2 
 
 =3.8 ft. 
 
 
 The value for hv is the resistance during the stroke of the 
 pump. It should be slightly larger than the valve b meai because 
 it includes the losses in the passages. max was taken as 3 feet 
 in working out the spring constants. 
 
 The resistance to water passage through both tH suction 
 and discharge valves is the same. The resistance, however, at 
 each end of the stroke is different for these, as at that time 
 inertia plays an important part, and this effect depends on the 
 
230 PUMPING MACHINERY 
 
 length of water column, which is greater on the discharge side. 
 The equations 
 
 i\A.-A. W.F. W,a A 
 
 and 
 
 give the loss of pressure at opening of the valves, and to get 
 numerical values for these the pumps with the 4-inch valves 
 will be considered, assuming that the suction lift is equivalent 
 to 10 feet, including friction; that the discharge head is equiva- 
 lent to 232 feet with friction; that there are 10 feet of i8-'nch 
 suction pipe and 300 feet of 1 8-inch discharge pipe, and that 
 the 4-inch valves have 3.2-inch openings beneath them. 
 At the end of the stroke 
 
 a = 
 
 = 58 f *' per sec ' per sec ' 
 
 57^.272 = 3. 28 sq. ft; 
 
 A = 2.18 sq.'ft.; 
 
 == ^ 7 =i-77 sq.ft. 
 4 J 44 
 
 From Eq. (31 a) 
 
 f 
 -- 
 
 = 934X62.5 = 58200 Ib. per sq. ft. 
 
 The effect of inertia in the pipe line is seen to be much 
 greater than the static head. Of course this water probably 
 does not start from rest each time and hence such a high value 
 
DYNAMICS OF WATER END 231 
 
 of pud may not occur. If this does not occur the water column 
 must separate and pounding will be observed. 
 
 7-3.28 57X0.5 10.40X57 
 _ 
 
 +57X^5 ^ 8oft 
 32.16 3 3-28 2 J 
 
 This large resistance at opening of the valve is due to the 
 fact that a large discharge pipe is used with no air chamber; 
 the effect of such a chamber will be seen later. 
 
 p ls = 2410 58 62.5 = 2X62.5 = 125.0 Ib. per sq. ft. 
 
 The result shows that the net pressure causing flow in the 
 suction pipe is only 2 feet, and hence a slight increase in the 
 speed of the pump would cause the suction column to break. 
 
 , i [ 4-97-3-28 ,57X0.5 10.40X57 
 
 Pn y I I 2 S 
 
 62.51 4-97 4-97 4-97 
 
 57X0.5 r 2.18 1 
 
 + 7^587 TT ^ =2.7 ft., approximately. 
 
 32.16 3 (4.97X3.28)] 
 
 Plotting Pressures for Points of Stroke. The expressions for 
 all of the terms of Eq. (65) are known at different positions of the 
 pistons. The terms are the following: 
 
 1 . The curve of friction varying as S 2 known as the KS 2 curve. 
 
 2. The piston movement curve or the x sin $ curve. (This 
 is zero for horizontal pumps.) 
 
 3. The inertia curves; one for th? a term or Ka curve and the 
 other for the xa term or the Kxa curve. 
 
 4. The valve curve. 
 
 These values of the different terms of Eq. (65) are now 
 plotted in Fig. 189, in which the resultant pressures above 
 the absolute zero are given when = 90. 
 
 It is to be remembered in this work that dimensions are 
 in feet. 
 
 In Fig. 189 the various curves are plotted for different 
 piston positions. The positions of the piston for the crank posi- 
 tions of o, 30, 60, 90, 120, 150, and 180 have been 
 
232 
 
 PUMPING MACHINERY 
 
 shown by vertical lines. The important lines of pressure are 
 marked in different ways. The atmospheric pressure of zero 
 gauge pressure, or 34 feet absolute is shown by a dot-and-dash 
 line while the lift h 2 of 10 feet is represented by a dash-and- 
 three-dot line. 
 
 The friction loss varies with S 2 and is given by a dash-and- 
 two-dot line. This curve starts from zero at the piston posi- 
 tion for o and reaches a maximum just before 90, while the 
 value at 180 is again zero. 
 
 The valve friction is represented by a dash and two-dot 
 line and remains practically constant during the larger por- 
 
 -34G 
 
 Atmospheric Pressuro 34 'Absolute 
 
 FIG. 189. Pressure on Suction without Air Chamber. 
 
 tion of the stroke according to the experiments of Hartmann- 
 Knoke. 
 
 The portion of the inertia effect due to the water in the 
 cylinder has been shown to vary with the product xa. This 
 is represented by a dash line which starts at zero, increases to 
 a maximum about 45, passes through zero just before '90 
 and increases negatively to the end of the stroke. This 
 quantity is the same for each end of the cylinder and for 
 each stroke. 
 
 The chief disturbing element in the suction pressure is the 
 
DYNAMICS OF WATER END 233 
 
 inertia of the water in the pipe line. This is the quantity 
 given by Y = ka and is shown by the dotted line. The water 
 in the suction pipe has to have an acceleration proportional to 
 that of the piston if the column remains intact, and conse- 
 quently to produce this a great head results. 
 
 The curve representing the head lost due to piston travel, 
 x, is shown by a slanting dot-and-dash line. 
 
 The combined curve is drawn as a solid line. At the begin- 
 ning of the stroke the gauge pressure is 34 feet, or absolute 
 zero. This means that the column would probably break and 
 produce knocking. The pressure increases rapidly, however, 
 reaching a value above the atmosphere at the end of the stroke. 
 This shows that although the water is lifted 10 feet + 2|- feet, 
 the inertia of the water in the pipe is so great that on trying to 
 stop it the pressure increased to the value shown in the figure. 
 
 The pressures marked on this and the other pressure dia- 
 grams, Figs. 191, 197 and 198, are measured in feet, starting 
 from atmospheric pressure. The absolute zero of pressure is 
 34 feet below this. 
 
 The absolute pressure of this figure at its lowest point can 
 never be a negative quantity, as there is never a negative abso- 
 lute pressure. Should the diagram give this in the combina- 
 tion of the various lines this would indicate that the water 
 column would separate and shock would result. This occurs 
 when the pressure assumes an apparent negative value. h s is 
 therefore limited to zero. This value cannot be obtained, as 
 the pressure of boiling for the water temperature would fix the 
 lowest value. When this pressure is reached the column will 
 separate, since steam will form. This gives then as a limit 
 
 
 (I A \ 2 \S 2 
 1 '\~A~J + /2 
 
 At the end of the upper stroke when S = o 
 
 i "~-iou -A -#180 / \ 
 
 -/^ -2-JpL , (71) 
 
 1 o 
 
234 
 
 PUMPING MACHINERY 
 
 or 
 
 L< 
 
 a so 
 
 g 
 
 . . (72) 
 
 at 90' 
 
 L< 
 
 - (73) 
 
 Ai g 
 LIMITS OF HEAD FOR PRIMING 
 
 These give the limiting values of L with which the pump 
 
 will operate without breaking 
 the column of water or pound- 
 ing. Should L be made 
 greater the pump will pound. 
 At times small pumps are 
 built with foot valves at the 
 base of the pump and a 
 bucket plunger in the same 
 barrel at some height above. 
 Such pumps have a limiting 
 distance between the piston 
 and foot valve. Consider the 
 pump (Fig. 190) to have the 
 same bore as the pipe below, 
 and that the water has been 
 sucked in until it reaches the 
 height h, at which time the 
 action stops because when 
 the piston is at its lower end 
 of stroke the pressure of the 
 air is just atmospheric, while 
 
 FIG. I90.-Suction Lift. 
 
 That is, the pressure will not be sufficiently low to allow 
 any more water to enter when at the top of its stroke. The 
 following equation holds then from Boyle's law: 
 
DYNAMICS OF WATER END 235 
 
 (h a -h)A(s+d-h)=h a A(d-h); 
 
 (h a -h)(s+d-h)=h a (d-h); 
 
 h a s+h a (d-h)-(s+d)h+h*=h a (d-h)i 
 
 (74) 
 
 Now when ^(s-}-d) 2 =h a s there are two equal roots for h 
 or the water will rise to a certain point. 
 
 If ^(s+d) 2 >h a s there are two real roots, which means that 
 the water will rise to a certain point, but if it could be moved 
 beyond this point it would go still further to another point h. 
 
 If i(s+d) 2 <h a s there will be no real roots or the water 
 will be sucked up and the pump will operate. 
 
 This gives 
 
 or 
 
 d<2\/h~J-s ......... (75) 
 
 This then gives the limiting lift for such a pump when the 
 suction valve is placed in the bottom of the pipe. 
 
 The same principle is applicable to pumps in which the 
 volume of the cylinder above valves is large compared with 
 the displacement of the pump. Let V c be the volume of the 
 cylinder above valves when the piston is at the end of the 
 stroke (called clearance on steam cylinders) and V r the volume 
 displaced by the piston, then if the pump is to operate under 
 the suction lift h 8 to the suction valves without filling the 
 pump barrel with water to prime the pump for starting, the 
 following equation must hold: 
 
 Under this condition the air in the cylinder on the stroke at 
 which water just reaches the suction valve will be compressed 
 to a little over atmospheric pressure and be driven out. This 
 gives 
 
236 PUMPING MACHINERY 
 
 V c is really what might be called the clearance volume, and 
 it is seen that when this is zero the limiting suction head is 7/ a ; 
 when it is equal to V r , h a becomes \h a at the limit. 
 
 The limitations on this deduction must be kept in mind. 
 The value of h s in formula (76) is the suction head against 
 which a pump with air-tight valves will operate without priming. 
 If the pump is primed, however, this water fills the clearance 
 space and V c becomes equal to zero, and then the pump will 
 operate against the limiting head h a if no friction is considered. 
 If friction [is considered the values given in Eqs. (72) and (73) 
 are the limiting suction lengths used under these conditions. 
 
 Discharge Pressure Curves. Passing now to the discharge 
 side it is seen that the expression for the pressure below the 
 piston on the down stroke will be 
 
 As before the various losses will be plotted in Fig. 191. 
 To plot these losses for the problem stated throughout this 
 chapter, the following numerical values hold: 
 
 ^ = 345 
 3 = 12.6-; 
 h f 2 = 200; 
 
 7^0 = 480 ft; 
 7^ = 3.8. 
 
 Assume that all valves are open wide. 
 xa 
 
 L+2Ll=H , = ? 
 
 Ai A 2 )g Vi-77 
 
 These are now plotted in Fig. 191. 
 
DYNAMICS OF WATER END 
 
 237 
 
 This figure shows the great effect of the inertia of the water, 
 and in working it out the effect of the compression of the 
 water has been neglected, as the use of the air chamber makes 
 
 FIG. 191. Pressure on Discharge without Air Chamber. 
 
 such a condition impossible. However, the problem has been 
 computed to show what is eliminated by the use of the air 
 chamber. 
 
 The scale of head has been increased so that the heads 
 could be plotted and so great is the scale that the loss, due to 
 
 Xd 
 
 friction, the larger part of the valve friction, and the =Z 
 
 o 
 
 terms do not show. The important curves are the static 
 head, 234 ieet above the absolute zero, the curve due to acceler- 
 ation of the water in the pipe, Y=Ka, and the atmospheric 
 line at 34 feet. These combine with the initial valve friction to 
 give a line which starts with a value of over 800 feet gauge 
 pressure and ends with a pressure below absolute zero. This 
 last pressure is impossible, and as a result water will separate 
 from the piston or water may be sucked through from the 
 suction side. 
 
 The great variation should be eliminated, and one method 
 is to run the pump more slowly or cut down the mass of water 
 which has to be accelerated. This latter is done by the use 
 of the air chamber, as will be shown. 
 
 The diagrams seen here and in Figs. 174, 175, 189, 197, 
 
238 
 
 PUMPING MACHINERY 
 
 and 198 have been on time and space bases, and it is well to 
 see how a curve on one base may be changed to one on another 
 base. 
 
 Fig. 192 gives a velocity diagram ABC on a time base. 
 To change this to a space base the curve DEF has been drawn 
 with an integraph such that the height i" i'" directly over 
 the point i is proportional to the area An', 2" 2'" = area 
 .422', etc. The planimeter may be used for the determination 
 of these areas and from them the ordinates i" i'", 2" 2'", etc. 
 
 Now area An' = I vdt= space. Hence the ordinates of 
 JQ 
 
 the curve DEF are the spaces passed over in the times repre- 
 
 Velocity 
 
 Time 
 FIG. 192. Velocity-space from Velocity-time Diagram. 
 
 sented by the points i" 2", etc., or the curve is a time space 
 curve. 
 
 If now the horizontal lines i" i iv , 2" 2 iv be drawn and the 
 distances i iv i v , 2 iv 2 V , be laid off equal to i' i, 2' 2 respec- 
 tively, the curve FGH is obtained which is the velocity curve 
 on a space base. 
 
 The curve in Fig; 193 is a velocity curve on a space base. 
 
DYNAMICS OF WATER END 
 
 Velocity 
 
 239 
 
 N i \\ \ \! 
 1 C 62 
 
 Space 
 FIG. 193. Velocity-space Diagram to Velocity Time. 
 
 In this the method of getting to a time base is not direct. The 
 following method is used: 
 
 ds' = vdt; 
 
 &=* 
 
 (itf. 
 
 J v 
 
 Hence the time taken to go a distance s may be expressed 
 as the area of a curve if the quantity is an ordinate and the 
 
 space passed over is the abscissa. 
 
 The method of procedure is to construct a reciprocal curve 
 
 to the one given in which the ordinates are quantities and to 
 
 draw the integral curve of this. 
 
 To get the reciprocal curve of BCD construct the line EF 
 at any distance A from the base and on the ordinate i, i v 
 desired swing la down to b, join i v and b and draw ac parallel 
 to i v b, and the distance ic will be proportional to the reciprocal 
 
240 PUMPING MACHINERY 
 
 of v. This distance is then swung up to i 1 giving i, i 1 propor- 
 
 lal to - 
 
 v 
 
 This is seen as follows: 
 
 ic ib 
 
 tional to - 
 v 
 
 la 1 
 
 hence x=A 2 -. 
 
 . v , 
 
 The ordinate of the curve HI] when divided by A 2 gives 
 the reciprocal of v. If A be made unity 
 
 The integration of the areas under HIJ gives the integral 
 curve KLM as before, and the velocity curve is constructed 
 by laying off the distances i, i v , 2, 2 V on the proper horizontal 
 lines. 
 
 For purposes of studying forces acceleration diagrams are 
 desired. To construct such from various points of the crank 
 or piston movement of a pump when the velocity is known, the 
 following methods are used: 
 
 Let Fig. 194 represent the variation of velocity with the 
 space passed over by the body having that velocity. 
 
 In this -3- = tangent of the angle of inclination of the tangent 
 
 to the curve with the horizontal. 
 Now 
 
 \c= 
 ds r dt ~~ dt~ 
 
 Hence, if V is multiplied by , the acceleration is obtained. 
 
DYNAMICS OF WATER END 
 
 241 
 
 FIG. 194. Acceleration from Velocity-space Diagram. 
 
 V-j-=ab if ac = V. The subnormal of the sv curve on the s 
 ds 
 
 axis represents the acceleration. 
 
 If V is on a time base (Fig. 195) the operation is different. 
 
 dv 
 
 In this case -^-=a=the tangent of the angle of inclination of 
 (M> 
 
 the tangent of the curve with the axis of t. 
 
 Tfme 
 
 FIG. 195. Acceleration from Velocity-time Diagram. 
 
 If then ad is laid off from the point of tangency equal to 
 unity, ed will equal the acceleration. 
 Since 
 
 , , , ,dv dv 
 
 ed = adta.n ead = ad-jr = iX-Tr=a. 
 at at 
 
 In the case of the figures in Fig. 175, showing the discharge 
 of the water from pumps of various kinds, these are really on a 
 time base as they represent velocities of the pistons, or quanti- 
 ties of water passing through the pumps per second at any 
 instant, and therefore velocity in the main, for different posi- 
 tions of the crank which is moving at a uniform rate. 
 
 The acceleration diagrams could therefore be computed by 
 the second method if desired. 
 
 In the figures and in the computation of the h 8Va and h dvo 
 the effect of inertia in the suction pipe and discharge pipe can 
 
242 
 
 PUMPING MACHINERY 
 
 readily be seen. These masses of water have to change their 
 velocity as the piston changes its velocity or if the pump is a 
 multicy Under pump the curves which have been computed for 
 the resultant discharge (Fig. 175) give the velocity variations 
 which exist. 
 
 FIG. 196. Pump with Air Chambers. 
 
 To cut down the variation of the velocity change in the 
 suction and delivery mains and to make the inertia forces smaller, 
 air chambers are introduced on the suction and discharge sides 
 of the pump. The object of the air chamber is attained where 
 so designed that the variation of pressure in it will be slight 
 and hence keep the water in the main under a practically 
 
DYNAMICS OF WATER END 243 
 
 constant velocity while the water which is accelerated is only 
 that which lies between the pump and the air chamber. 
 
 The air chambers on the two sides of the pump (Fig. 196) 
 are subject to a certain amount of variation of pressure and in 
 making the preliminary investigation it is well to consider 
 the mean pressure in the air chamber as constant. 
 
 This pressure is that within the suction air chamber which 
 causes a flow from the forebay overcoming the resistance of 
 entrance, friction, and lift. There is no velocity considered 
 since the head used in causing the velocity in the pipe is re- 
 turned when the water is I rought to rest in the suction air 
 chamber. In the discharge air chamber the pressure is used 
 to give the velocity, lift the water, and overcome friction of all 
 kinds. The pressure in the suction chamber is less than at- 
 mospheric pressure in general, although water might be sup- 
 plied to the pump under pressure from a higher source. The 
 pressure in the discharge chamber is greater than that of the 
 atmosphere. In any case pressures are always measured from 
 absolute zero. The pressure in the air chamber is measured 
 when there is no velocity head. 
 
 h sc = h a -(^ + ^~-y^ . .-. . . . (80) 
 
 A 
 
 +^4, . . . . (81) 
 
 where the coefficients , refer to the losses in the pipes due 
 to friction, and the quantities y are the lifts. 
 
 These pressures in the air chambers are now the datum 
 planes from which to compute the pressure in the cylinder 
 beneath the plunger on the two strokes; the water in the con- 
 necting pipes and cylinder between the piston and the air 
 vessels being the only water which has to be considered in the 
 expression for inertia and friction. 
 
 The expressions for pressure now become 
 
 
244 PUMPING MACHINERY 
 
 The term L is much less than L, and consequently that 
 term is smaller. h s v does not have such a large value as h no , 
 although h sv is the same over a large part of its range. 
 
 h s must be greater than h,-, and this inequality will give 
 the limiting suction height or stroke for given conditions, as in 
 Eqs. (72) and (73). 
 
 The computations below are made for suction chambers 
 of the original pump with the chamber placed two feet back 
 from the entrance to pump. 
 
 From Eq. (80) 
 
 / 2.l8\ 2 
 
 8 \(5- x - 
 
 +0.41) I>77 
 I -5 / 4-3 
 
 = 34 .92 8 = 25.08 ft. 
 From Eq. (21) the various parts making up h s are 
 
 fee = 25.08 ft; 
 hi = 2 it.; 
 x sin (j> = X', 
 A,, = 3. 8. 
 
 The value h svo will have to be computed, as the pressure 
 in the air chamber will give a new value of pi s . There will be 
 2 feet of water to be accelerated between the chamber and the 
 pump. 
 
 4-97 4-97 4-97 
 
 57X05 
 
 32.16 3.28X4.97 
 
 ^JA \ 2 A 2 \ ; ^ /2.i6\ 2 s 2 ^ 
 
 2U-- - =o.03X - - = 0.06 ; 
 \Ai/ \2g/ i.5\ I -77/ 2g 2g 
 
 
 xa , . 
 = Z as before ; 
 
 (A T \a , ,a 2.2 
 
 2- L* -=(1.2 + 1)-= - a = .o6^a. 
 \ A 2 /g g 32.2 
 
DYNAMICS OF \\ATKR END 
 
 245 
 
 -84'G 
 
 Atmospheric I ressure 84/Absolute 
 
 Air Chamber Pressure 25.08 Ft.Abs< 
 
 6P 
 
 Lift 
 
 Line of Valve Friction 
 Line of X 1 2 - 
 Line of -*| 
 Line of uv Z 
 Line of Y 
 
 90 
 
 FIG. 197. Pressure on Suction with Air Chamber. 
 
 These values are plotted in Fig. 197. 
 
 The pressure on the discharge stroke is given by the equa- 
 
 tion 
 
 A \ 2 9 2 
 
 The computations and curves for the previous problem on 
 the assumption that the air chamber is 2 feet from the pump 
 are given below: 
 
 From Eq. (81) 
 
 -+20O 
 
 = 34+4.9+200 = 238.9. 
 The parts of Eq. (83) now become 
 ^ = 238.9; 
 
 x sn 
 
246 PUMPING MACHINERY 
 
 ,JA \ 2 /S 2 \ 2 /2.I6 
 
 SU- ( =0.03 X - 
 \Ai/ \2g/ i.\i-77/ 
 
 8 s 2 
 
 = .12 ; 
 
 #- = Z as before; 
 * 
 
 _<4 r a 2.18 # 
 2 Z, 5 - = -- X 2 -- = 0.0760. 
 ^2 g i-77 3 2 - 2 
 
 The value ^ has also to be found when the air chamber 
 is used. In this case 
 
 = I 239+-^ ^-58 162.5 = 244.4X62.5; 
 
 In the two quantities h svo and /// the effect of the air 
 chamber is seen in the smaller value of the last term due to a 
 smaller value of L of the water to be accelerated. The effect 
 of the first term is also to be noted. This has increased the 
 value of h no , while the last term has been changed so much 
 in the expression for hd, that the resultant head here has 
 been made smaller. 
 
 These values afe plotted in Fig. 198. 
 
 In Fig. 198 the curves are all marked and from the descrip- 
 tion of Fig. 189 these may be followed. The term involving 
 S 2 in (83) has a negative coefficient when combined and con- 
 sequently this curve is below the line 34' gauge pressure. 
 The effect of the reduction of the mass of the water to be accel- 
 erated is seen by the diminished height of the curve Y = Ka. 
 The curve Z = Kax and the curve of x are the same as in Fig. 
 189. The value of the valve friction at the suction end cf the 
 stroke has been increased, although the major portion of the line 
 remains at the same height as before. 
 
 By comparing Fig. 197 with Fig. 189 the great advantage 
 of the air chamber in making the pressure more uniform and 
 
DYNAMICS OF WATER END 
 
 247 
 
 in making it possible to draw water at the high speed of the 
 pump is evident. The suction pressure in the pump cylinder 
 never falls to such a low value that the column would break 
 when a proper air chamber is used. 
 
 200G 
 
 look 
 
 Atm< 
 
 Air Chamber 
 
 m 
 
 phere 
 
 -M'GT- -JL^-rfz^: 
 
 re 239' Abs. 
 
 20' 
 
 Resultant 
 
 210 
 
 Term X 
 
 210 / 
 
 FIG. 198. Pressure on Discharge with Air Chamber. 
 
 Fig. 198 shows the action of the air chamber in reducing 
 the Y effect and the friction loss at the opening of the valve 
 at the beginning of the stroke. All of the curves of the various 
 terms are shown except that of the negative term involvingS 2 . 
 This is so small that it would not be appreciable on the scale 
 of this figure. By comparing this with Fig. 191 the great 
 value of the chamber may be seen. 
 
 In Eqs. (80) and (82) the following conditions must hold: 
 
 h sc > h sc t, 
 
 The first of these equations may be used to find the length 
 
248 PUMPING MACHINERY 
 
 3'3, and the second to get the limiting -length h+yz, or the 
 limiting speed 2LN of the pump piston. 
 
 The second equation will give the limiting length to the 
 cylinder if the expression for h sc from Eq. (80) is inserted in 
 the expression for h s in Eq. (82). 
 
 SIZE OF THE AIR CHAMBER 
 
 The suction air chamber may be considered to receive 
 water at a steady rate A s v s per second, while it gives to the 
 pump an amount which varies for different crank positions 
 as shown by the curves of Fig. 175. The areas of these figures 
 represent quantity handled by the pump so that the area 
 above the mean curve represents the amount supplied by the 
 suction air chamber during a certain part of the stroke while 
 the area below the mean curve represents the amount sent into 
 the air chamber when the pump does not require it. This 
 statement is reversed in the discharge or pressure air chamber 
 when uniform discharge is assumed A d V d in the discharge 
 pipe. In this case the area above the line represents the amount 
 given to the air chamber while that below the line represents 
 the amount given up by the chamber when the pump is giving 
 out less than the quantity discharged through the main. 
 
 Calling the amount of water in the air chamber at a, Q, the 
 
 excess areas abcve mean line ei, e 2 , etc., and the deficiencies d\. 
 
 dz, etc., then the quantities at the various points b, c, d, etc., are 
 
 At (a) Q, 
 
 " (c) 
 " (e) 
 
 ..... (84) 
 
 The upper sign would be used for the discharge air chamber 
 and the lower sign for the suction air chamber. 
 
 The difference between the greatest and least of these 
 quantities represents the greatest variation of volume of the 
 water in the chamber. Call the difference F. 
 
DYNAMICS OF WATER END 249 
 
 The volume of the air varies from a maximum to a minimum 
 as this water comes from the pump or suction, so that 
 
 ioo Fmax ~ Fmin = 100^ * the percentage 
 
 variation in volume. Since this compression of air in the 
 chamber is isothermal, pV =k. The pressure maybe expressed 
 in any units. So the formula will hold for the pressure ex- 
 pressed in feet head, hence 
 
 ^sc ' sc == max ' mm ~ "min ' max* 
 
 since the maximum pressure occurs with the minimum volume 
 and vice versa. Solving for F max and F min in te t rms of V 8C , 
 and substituting in the equation for the variation of volume 
 there results 
 
 * c V 
 
 SC - ' 8C 
 
 % variation = AV = 
 
 "max ~" 
 
 8C ^max ^min ' 
 
 If ^max^min is taken as h 2 sc which makes h 8c a geometric 
 mean or mean proportional between A max and A min , the value 
 of AV becomes 
 
 A yJ< m ax-Amin =J ^ 
 
 or the percentage variation of pressure is the same as the per- 
 centage variation in volume. Hence 
 
 or 
 
 F 8C =J. :.;:;.. (86) 
 
 If then Ah or the permissible variation in pressure is assumed, 
 the volume of the air chamber to give this result is known for a 
 given F. The equation shows clearly that the air chamber will 
 vary with F and that whenever the resultant piston discharge 
 approaches a mean line, the volume of the air chamber decreases. 
 
250 PUMPING MACHINERY 
 
 If the variation Ah is to be made small, the volume V 8C will 
 be made large. It is to be noted that Ah is a ratio of variation, 
 and hence if the same ratio of variation is assumed, the volume 
 of the air chamber for a pump will be the same for all pressures 
 or suction lifts. Moreover, according to this the air chamber 
 on the suction side is to be the same as that on the discharge 
 side if the same percentage variation is used, the actual 
 variation being much less on the suction side. This is as it should 
 be, for the pressure change on the suction side is most impor- 
 tant, as there is only the atmospheric pressure to do the driving. 
 The size of the complete air chamber should be such that 
 when the pump is shut down the increase of volume due to 
 the decrease of pressure from the elimination of friction, since 
 the water is at rest, will not be sufficient to cause the water 
 to be entirely, driven from the air chamber. The value V&, 
 computed previously, is the volume of the air in the chamber 
 when it is under the pressure of operation. Calling V' dc the 
 volume of the air .when the pump is at rest 
 
 Where h d is the head when the pump is stationary, and h^ 
 is that while the pump is operating. This volume F'^ is then 
 the minimum volume to be given to the air chamber so 
 that it will never lose the air when the pump is brought to 
 rest. 
 
 Although this simple method could be used, Hartmann and 
 Knoke consider the discharge air chamber from the stand- 
 point of supply from the pump when the pump is started 
 up. 
 
 Considering the pipe leading to the air chamber to be of 
 area A c , in which the velocity is assumed to be C c , it may be 
 said that the quantity entering the air chamber per second 
 is A C C C cubic feet. The amount of water leaving the air chamber 
 per second is AdCa cubic feet, where A d is the area of the dis- 
 charge pipe and C the mean velocity at any instant in the 
 cross-section. 
 
DYNAMICS OF WATER END 251 
 
 If now C c be taken as constant the net amount of water 
 entering the air chamber in any time t after the pump has 
 started is 
 
 q=A c Cj- CA d C d dt. . . . . . (88) 
 
 J o 
 
 It is assumed that it takes this time to get the pump into uni- 
 form motion and at the end of the time that the air chamber 
 is giving back as much water as it receives. 
 
 The air in the air chamber is originally under the pressure 
 hds = ha+hd in feet of water where h a is the atmospheric pressure, 
 and hd is the vertical distance from the end of the pipe to the 
 water level in the air chamber. As this quantity q enters, 
 the total pressure becomes greater, say h d2 due to compression, 
 so far as the chamber is concerned, but produced in reality 
 by the inertia and friction of the water in the discharge pipe. 
 Then if V ' ^ is the volume of the air chamber 
 
 = (V dc -q)h d2 , ..... (90) 
 because the temperature of the air may be assumed constant. 
 
 -, 
 
 * dc 
 
 if h dz is considered variable. 
 Now from (88), 
 
 dq=A c C c dt-A d C d dt. 
 Hence 
 
 (91) 
 
 The pressure in the air chamber at a certain instant is A, 
 while that at the entrance to the air chamber due to the static 
 head is h , and there is an unbalanced pressure of A d (h-h ds )w, 
 which may be utilized. 
 
252 fVMPlNG MACHINERY 
 
 This acts on the water in the pipe line Id, of weight 
 
 -rr 
 at 
 
 and produces the acceleration -rr. The force of inertia is 
 
 IdAjuvT- expressed in pounds. 
 Hence 
 
 or 
 
 _ Id dcd 
 ctt = "^ j , 
 gh-kj 
 
 dh 
 V dc">ds-', 
 
 , \ r A ^ ^J 
 --(A c C c -A d Cd)dc = 
 
 /, ~\hd 
 h+^l . . (92) 
 
 n Jhdo 
 
 where C = C d at the time h=h d . 
 
 A d C d =A c C c because at the instant when the greatest com- 
 pression occurs in the air chamber there is no net flow into 
 the chamber, 
 
 r d - AcCc 
 
 Ld ~ A, ' 
 Then 
 
 g 2A d 
 
 or 
 
 ^c = -P^ 
 
 r hd h do 
 
 To illustrate the method used in the design of air chambers, 
 
DYNAMICS OF WATER END 253 
 
 the size of chamber for the pump will be computed. The 
 pump will be assumed to be double acting, and of the dimen- 
 sions given before. The curves of Case 2, Fig. 175, represent 
 the action of this pump, and by using a planimeter on the 
 original drawing the right-hand excess area was found to be 
 0.42 square inch, the left 0.39, the right- and left-end deficiency 
 areas which go together 0.38 square inch, and the middle 
 deficiency 0.43. This gives the following variation in the 
 quantity of water in the air chamber: 
 
 At i, v, 
 
 " 2, 0+0.39; 
 
 " 3, v + 0.039 -0.43=0 -0.04; 
 " 4, 0-0.04 + 0.42=0 + 0.38; 
 " i, 0+0.38 -0.38=0. 
 
 The greatest variation is from 00.04 to ^ + -39 or 43 
 square inch of diagram area. 
 
 The figure was originally drawn 6 inches long and five-eighths 
 inch for the mean height. The length represented the time of 
 one revolution or the angle turned through, and the quantity 
 discharged bv the pump was 10.85 cubic feet per second. The 
 scales of the figure are therefore one-sixth second per inch of 
 length and 17.38 cubic feet per second per inch of height. The 
 area scale is therefore 2.89 cubic feet per square inch. 
 
 The change in the volume of water in the air chamber is 
 therefore 
 
 0.43 X 2.89 = 1.24 cu.ft. 
 
 The variation of pressure Ah is now assumed to be 5 per 
 cent and the volume of the air in the chamber is, by Eq. (86), 
 
 1.24 
 v*c = = 25 cu.ft. 
 
 To permit the pump to be shut down the air in the discharge 
 chamber on expanding from 256 feet pressure (the pressure 
 in the discharge chamber during action) to 222 feet pressure 
 (static head), the volume of the air vessel should be such that 
 this air is not driven out. By Eq. (87) 
 
254 PUMPING MACHINERY 
 
 The volume of the cylinder is 5.5 cubic feet, and the volume 
 of the air chamber above is about five times this. The vessel 
 would be 30 inches diameter by 48 inches length. The great 
 size of this is due to the kind of pump [Term F of Eq. (86)], 
 and the allowable variation in pressure. 
 
 The method of design sometimes employed is to assume 
 the ratio of chamber volume to cylinder volume, and use this 
 only. The ratios suggested are: 3 and 6 times the cylinder 
 volume for single-acting pumps, and one-half to two- thirds of 
 this for double-acting pumps. 
 
 From the above formulae the following steps may be taken 
 to determine the leading dimension of the water system when 
 a given quantity of water, Q per second, is to be pumped: 
 
 SIZE OF PIPES AND AIR CHAMBERS 
 
 The suction pipe should be large and as free from bends as 
 possible. In the first approximation a velocity of 3 feet per 
 second may be assumed for the velocity in the suction pipe. 
 
 Then A 8 = . 
 
 The size of the suction air chamber is given by Eqs. (86) 
 and (87). 
 
 The discharge pipe is in many cases so long that it will 
 pay to compute several sizes. Suppose a pipe is found in 
 which the lost head due to friction is h/ feet and by using a 
 larger pipe this may be reduced to h'f feet. The gain by this 
 enlargement is 
 
 (hf-h' f }wQ 
 
 550 X eft. 
 
 = H.P. 
 
 If now the power costs M dollars per horse power hour 
 the saving per year of T hours will be 
 
DYNAMICS OF WATER END 255 
 
 This same could be used to pay the difference in yearly 
 cost between the cost of the small and large pipes. The 
 increased cost of iron and installation could then be found, 
 and if the interest, depreciation, taxes, and insurance on this 
 cost is just equal to the amount saved per year, there would 
 be no economic advantage in putting in the larger pipe. If 
 the amount is less than the saving, a still larger pipe should 
 be tried, while, if the saving is less, it would be well to make 
 an investigation with a smaller pipe to see if the gain in interest, 
 depreciation, taxes, and insurance of a smaller pipe would not 
 be greater than the increased cost of power. This does not 
 consider the development for future service, which would alter 
 the problem. 
 
 Having the area Ad of the discharge pipe, its length, and 
 the bends in it, the resistance from such a line can be found, 
 and from it the size of the air chamber for the discharge. Eqs. 
 (86) and (87) are used for this. 
 
 NUMBER OF REVOLUTIONS 
 
 Ratio of L to D 
 
 These two quantities are mutually dependent. In many 
 cases the quantity zLN or piston speed is the quantity assumed, 
 and then D is known from the formula 
 
 D = 
 
 where n is the number of active strokes in one revolution. 
 With 2LN=S as soon as N is known L is given by 
 
 L _A 
 
 ~2N' 
 
 The speed of the piston, S, is not a constant, but may vary 
 from 50 to 700 feet per minute. The equation shows the 
 common tendency, however, to cut down the stroke as the 
 number of revolutions increases. The piston speeds used with 
 steam engines vary from about 350 to 800 feet per minute, 
 
256 PUMPING MACHINERY 
 
 and where the pump piston is mounted in tandem with the 
 steam piston, the piston speed for the pump would fix that 
 for the engine. 
 
 The length of the machine may determine the stroke to be 
 used. Where a short machine is desired for any given reason, 
 % whether the pump be horizontal or vertical, a short stroke is 
 chosen. 
 
 This choice of a short stroke, however, may mean a large 
 diameter and with it much heavier parts, cylinders, piston 
 rods, cross-heads, connecting rods, pins, and other parts. No 
 general rule can be given; each problem arising must have its 
 own solution. 
 
 The value of N is determined by many things. Until 
 within the last twenty years pumps were usually run at from 
 20 to 80 revolutions per minute, and at the higher speeds 
 trouble was experienced. In later years by increasing the valve 
 area or the number of valves, and in cases using positively 
 actuated valves, by giving large passages and cylinder space, 
 and by the use of large air chambers much higher speeds of 
 revolution were used. As was mentioned earlier, Riedler was 
 one of the first to design high rotative speed pumps. The 
 speeds have been carried up to 250 R.P.M. By using such 
 speeds it has become possible to operate pumps by direct 
 connection to electric motors and gas engines without the 
 use of intermediate gears. Such arrangements save room, 
 although on account of the special design for such pumps the 
 cost may not be reduced. 
 
 In America the practice is to operate large pumps with 
 piston speeds of about 500 feet per minute, and a rotative speed 
 of 25 R.P.M. This selection has its advantage, as the inertia 
 forces vary as the square of the number of revolutions, and as 
 the first power of the crank radius or stroke. With smaller 
 pumps 60 R.P.M. to 80 ^R.P.M. is often found, and with special 
 pumps the higher speeds of 150 to 200 R.P.M. are used. The 
 latter express pumps are of value where direct connection is 
 necessary, and space or weight limited. 
 
 Having then the size of the cylinders, revolutions, size of 
 
DYNAMICS OF WATER END 257 
 
 suction and discharge pipes, the number of valves should be 
 determined, and then the capacity of the air chamber. 
 
 TABLE OF SYMBOLS 
 
 A = area of piston in sq. ft. 
 yl p = area of passages beneath valves at one end of 
 
 cylinder in sq. ft. 
 
 'A\ = area of valve at one end of cylinder in sq. ft. 
 AI, A2, AS, etc. =area of pipes in sq. ft. 
 
 C = velocity at air chamber in ft. per sec. and con- 
 stant for spring. 
 D = diameter of cylinder in ft. 
 F = difference in discharge from air chamber. 
 F s = pressure from spring in pounds. 
 H = head in feet. 
 Li, 2 = lengths of pipes in feet. 
 L = length of stroke in feet. 
 N = number of revolutions per minute. 
 N' = number of effective strokes in / seconds. 
 N" = number of effective strokes in i second. 
 Q = quantity of water in cu. ft. discharged in t sec. 
 R = radius of pipe bends in feet. 
 5" = velocity of piston in feet per sec. 
 S' = space passed over. 
 V = volume in cu. ft. 
 W v = weight of valves in pounds. 
 a = acceleration in ft. per sec. per sec. 
 b = equivalent head for weight and spring pressure. 
 bi = breadth of valve seat in ft. 
 d = diameter of piston rod at stuffing-box in feet 
 
 also diameter of valve in ft. 
 
 di, d2, d% = diameters of pipe in ft. (see below also). 
 d' = amount face valve is closed. 
 d\ diameter of passage below valve in ft. 
 d 2 = pitch of valves in ft. 
 e = excess or deficiency area. 
 
258 PUMPING MACHINERY 
 
 f= friction factor for flow in pipes. 
 g = acceleration of gravity in feet per sec. per sec. 
 h = head in ft. of water and lift of valve in ft. 
 h a = head in ft. corresponding to atmospheric pres- 
 sure. 
 
 hdv = head loss at opening of discharge valve. 
 h s v = head loss at opening of suction valve. 
 h s = pressure head on suction side of pump piston. 
 hd = pressure head on discharge side of pump piston. 
 fi2 = height in ft. from fore bay to lower end of pis- 
 ton stroke. 
 // 2 ' = similar height on discharge side from end of 
 
 piston stroke, 
 fe = head lost in friction, velocity, enlargement, etc., 
 
 in ft. 
 
 h = head due to inertia in ft. 
 h$ = head from level in suction air chamber to end 
 
 of stroke. 
 hs = head from level in discharge air chamber to end 
 
 of stroke. 
 
 h p = hea.d loss in pipe measured in feet of water. 
 i = number of ribs in valve seat. 
 k = constant for pipe, pipe bends. 
 / = length of pipe in ft. and periphery of valve. 
 m = constant for valve friction. 
 n = ratio of length of connecting rod to crank 
 
 radius. 
 
 w c = number of effective strokes in i rev. 
 n' = number of valves at one end. 
 p c = pressure below valve in Ibs. per sq. ft. 
 p u = pressure above valve in Ibs. per sq. ft. 
 q = discharge in cu. ft. per sec. 
 r = crank radius in ft. 
 s = thickness of ribs at end or lap of multiported 
 
 valve. 
 
 s' = space passed over. 
 t = time in seconds. 
 
DYNAMICS OF WATER END 259 
 
 z> = velocity of water in ft. per sec., velocity under 
 
 valve edge. 
 
 D 1 = velocity of water in valve passage. 
 D2 = velocity of valve. 
 v s = velocity in smaller pipe where sections change 
 
 suddenly. 
 
 w = weight of i cu. ft. of water in Ibs. 
 x = movement of piston or cross head from head 
 
 end of stroke in feet and pitch of rings in 
 
 multiported valve. 
 ^2 = lifts in ft. 
 yo = compression of spring. 
 a = angular inclination of connecting rod. 
 a = coefficient in Bach's formula. 
 |8 = coefficient in Bach's formula. 
 7 = coefficient in Bach's formula. 
 7 = coefficient of discharge. 
 
 6 = angular motion of crank from head dead center. 
 = inclination of cylinder. 
 p = ratio of diameter of small pipe to that of large 
 
 pipe. 
 
 ju = coefficient for valve velocity. 
 = coefficient for lost head, coefficient for valve 
 
 loss, coefficient for valve discharge, 
 co = angular velocity. { 
 
CHAPTER VI 
 DESIGN OF PARTS 
 
 WATER CYLINDERS 
 
 AFTER the diameter and stroke have been determined, the 
 cylinder is designed. The arrangements -of the bore, the 
 passages leading to the valves, the valves, the valve decks, 
 and the valve chests are all of importance, and are varied 
 according to the peculiarities of the designer., The general 
 principle is to make the path of the water as direct as possible; 
 the construction of the casting simple, and the position of the 
 valves such that they may be easily examined and replaced. 
 To show the arrangements proposed and used by pump designers 
 a number of typical pump cylinders have been chosen, each 
 illustrating some special point. Although a few might have 
 answered the purpose, the number has been increased to famil- 
 iarize the student with the construction of modern pumps, 
 to show many ways of constructing machines, and to illustrate 
 how constructions may be simplified. 
 
 The water pistons take one of four principal forms: ist, 
 the piston (A, Fig. 199); 2d, the plunger and ring (B, Fig. 199); 
 3d, the packed plunger (C, Fig. 199), and 4th, the bucket 
 (D, Fig. 199). 
 
 These are packed in different ways. Fig. 199^ . shows a 
 canvas packing inserted on a ledge formed on the piston casting. 
 The packing is usually made of layers of canvas or square- 
 woven cotton indurated with rubber, or it may be a square 
 flax packing. The form of packing shown at B has been 
 used for many years; the long sleeve makes practically a 
 water-tight joint, and for clear water this lasts for some time. 
 The sleeve may be renewed when necessary. The plunger 
 (C, Fig. 199) is packed on the inside at the center. Such packing 
 
 260 
 
DESIGN OF PARTS 
 
 261 
 
 may be placed at the outside. The form is identically the same 
 in both cases. The kind of packing used is the same as that 
 employed on the piston. The cup-leather packing (D, Fig. 199) 
 is used with deep-well pumps and with high-pressure pumps. 
 The cup leather forms a good packing in such cases, as the 
 
 B D 
 
 FTG. 199. Pistons, Plungers and Bucket. 
 
 pressure exerted by the leather varies with the water pressure, 
 and so the friction on the suction stroke is small. Fig. 200 
 shows the form- for a double-acting piston with cup-leather 
 packing. When packings are used for the outside they are 
 practically the same in form. The piston rod is packed as 
 shown in Fig. 201 or Fig. 202; the plunger for ordinary pres- 
 
262 
 
 PUMPING MACHINERY 
 
 sures with a light packing (Fig. 203), while for heavy pressures in 
 
 pumps or hydraulic jacks U-leather or 
 hydraulic packing (Fig. 204) is used. 
 
 The design of these packings is 
 usually empirical, and the figures 
 shown are marked with numbers 
 which are given in terms of a unit. 
 
 In Figs. 201 and 202 the propor- 
 tional unit is \d,+y where d= diam- 
 eter of piston rod. 
 
 In Fig. 203 the size and number 
 of the bolts and the thickness of the 
 stuffing box will have to be designed. 
 The thickness A of the packing and 
 the depth B depend to a certain ex- 
 tent on the size of the plunger. B=o.2 to 0.4^ and A = o.id 
 
 FIG. 200. Double-leather 
 Packing. 
 
 " up to i" may be used as a guide. 
 
 FIG. 20 1. Stuffing Box. 
 
 The thickness C is designed for high pressures as a cylinder 
 wall under the pressure of the fluid. 
 
 The bolt area, a, or the number of bolts, n, may be found 
 by the following formula if one of these quantities is assumed: 
 
DESIGN OF PARTS 
 
 263 
 
 The cup leathers (Fig. 206) are usually made by soaking 
 the leather to make it pliable, and then forcing it into the 
 mold (Fig. 206) by a press or bolt. After it is forced down 
 
 FIG. 202. Stuffing Box. 
 
 it is allowed to dry and set, when it is trimmed on the edge by 
 turning. The wear on cup and other leather packing occurs 
 near the bend in the leather (Fig. 205) as it is this point which 
 
 FIG. 203. Plunger Packing. 
 
 FIG. 204. U-Leather Packing. 
 
 is driven against the plunger or cylinder wall. Since this is 
 the case there is no necessity for making the dimensions h 
 greater than the amount given below. 
 
 The U-leathers are proportioned, as shown in Fig. 207, for 
 
264 
 
 PUMPING MACHINERY 
 
 
 FIG. 205. Wear in Cup Leather. 
 
 FIG. 207. U Leather. 
 
 FIG. 206. Formation of Cup Leather. 
 
 FIG. 209. Hat-leather Packing. 
 
 FIG. 208. Formation of U Leather. FIG. 210. Formation of Hat Leather. 
 
DESIGN OF PARTS 265 
 
 which the various dimensions are given later. These are 
 made by formers or molds, after soaking, as shown in Fig. 
 208. This operation may best be performed on a hydraulic 
 press. The first fold is made as in the second part of the figure, 
 after which the last core is introduced. 
 
 The hat leather used as shown in Fig. 209 is formed in the 
 same manner as the cup and U leathers by formers (Fig. 210). 
 After drying, the center is cut out and the edge chamfered. 
 
 For cup leathers (Fig. 205), 
 
 A = i"to i"; 
 
 Friction =cdp\ 
 
 0=0.03 to 0.05; 
 
 p= hydrostatic pressure per sq. in. 
 
 For U leathers (Figs. 204, and 207), 
 
 Diameter box = diameter plunger + 2w J" ; 
 
 Outside diameter leather = diamet er plunger + 2 w^" 
 Inside diameter leather = diameter plunger 
 Height h = 1.6 width =" to 
 
 Width =w=yto%"; 
 
 -\(D 
 
 Bolt area =a= -^ 
 
 n^t 
 
 Height of flange A" = 1.5 diameter bolts; 
 
 h' = 3 diameter bolts ; 
 Width of cylinder flange =3 diameter bolts; 
 
 P 
 
 Friction =0.04-3 = o 
 
 The Marsh pump (Fig. 211) has one of the simplest forms 
 of water end. In this design there is no lining to the cylinder 
 bore. The suction chamber A is formed by coring, and the 
 discharge chamber B is formed by the valve deck plate C and 
 the cover D. The valve seats are forced into holes on the 
 valve decks. The piston is packed with cup leathers. This 
 
266 
 
 PUMPING MACHINERY 
 
 particular pump is built for pumping milk and for that reason 
 it is so constructed that it may be taken apart quickly and 
 easily for cleaning. All nuts are wing nuts, and the bolts are 
 hinged so that they may be swung out. The piston rod is 
 light and the stuffing box is simple. 
 
 Another form of Marsh pump (Fig. 212) is so constructed 
 that the cylinder bore is fitted with a brass liner to take the 
 wear of the piston. The liner is held in place by screws passing 
 through a flange. It may be renewed easily when necessary. 
 The piston body F is fastened to the piston rod by a nut, while 
 
 FIG. 211. Marsh Milk Pump. 
 
 the follower plate G is held on the same thread by another nut. 
 The removal of this plate permits one to examine or repair 
 the piston packing. The shoulder on the piston rod holds 
 the body F in place, preventing any play. 
 
 The stuffing box H is of the cap form, and is made separate 
 from the cylinder casting. This simplifies the foundry work 
 and lessens the difficulty of the machine work in the manu- 
 facture. 
 
 The method of dividing the valve chambers of the cylinder 
 ends by a partition carried to the discharge valve deck as well 
 as the method of attaching the air chamber, and the form of 
 
DESIGN OF PARTS 
 
 267 
 
 the air chamber are all to be noted. The cap on the end of 
 the air chamber has been employed so that a core print may 
 be used at each end of the pattern. 
 
 FIG. 212. Marsh Pump. 
 
 To simplify the water end, the Marsh Company build a 
 pump (Fig. 213) in which the water-cylinder bore is made by 
 
268 
 
 PUMPING MACHINERY 
 
 using a piece of solid-drawn brass pipe. This makes a very 
 simple casting, as the suction chamber A is carried up from 
 the base of the pump. The cap forming the discharge chamber 
 
 FIG. 213. Marsh Pump. 
 
 B is cast solid with the air chamber. The valves used in this 
 pump are peculiar in form and the guides are so arranged that 
 the valves are not subject to a lifting pressure after they rise 
 to a certain height. The valve seats are inserted in openings 
 
DESIGN OF PARTS 
 
 269 
 
 ^tfttf^^ 
 
 FIG. 214. Marsh Pump. 
 
 in the valve decks, simplifying the casting and making repairing 
 easy. The peculiar form of packed piston, the stuffing box, 
 
270 PUMPING MACHINERY 
 
 and the simple form of casting are to be observed in the 
 figure. 
 
 For larger sizes the Marsh pump water end is built as shown 
 in Fig. 214. In this case the suction chamber is a separate 
 casting, and forms the base of the water end. The suction 
 valve deck is reinforced against the water pressure from the 
 cylinder by a rib 5. Such a construction is necessary, as the 
 valve deck is a broad, flat plate, subject to a downward pres- 
 sure. A large suction pipe enters the side of the suction 
 -chamber. The discharge valve deck is part of the cylinder 
 proper, and is stiffened by the rib T which extends across the 
 casting, while the discharge chamber is bolted to the cylinder. 
 The air chamber is attached to this and is similar in form to 
 those used on the smaller pumps. Hand holes are arranged 
 in the side of the cylinder casting, and in the discharge chamber 
 at MM, so that the valves may be inspected or repaired without 
 the necessity of removing the large castings or opening large 
 covers. Such construction is always necessary in pumps of 
 any size. 
 
 The valves are of a different form from those shown before. 
 These are spring-controlled valves. Their action is similar 
 to that of the others illustrated. The seats are made separate, 
 a good construction, in order that seats may be renewed if 
 broken, and, moreover, the casting is thereby simplified. The 
 follower ring R on the piston replaces the plate used on the 
 other pistons, and the bolted gland in the stuffing box replaces 
 the cap gland. The liner of brass is used to simplify 
 renewal when ,the wear from the piston packing becomes 
 excessive. 
 
 One of the important features in pump design is to have an 
 ample and simple direct passage for the water through the 
 pump. The pumps shown in the preceding figures have been 
 good in this respect, but the Fairbanks-Morse pump in Fig. 
 215 shows in a somewhat better manner, the ample passages. 
 The suction enters the chamber A in an easy sweep from D, 
 and then goes through a large passage to the pump cylinder. 
 The discharge from the other side is forced into B and leaves 
 
DESIGN OF PARTS 
 
 271 
 
 through a passage to the discharge E. The general features 
 of this design are seen on inspection of Fig. 215. 
 
 Another pattern with large passages is shown in the Worth- 
 ington pump (Fig. 216). In both of these pumps the cylinder 
 bore is lined by using a piece of brass tubing. This tubing is 
 withdrawn when worn, and another piece inserted in its place. 
 In pumps intended to lift acids which may attack the metal, 
 
 FIG. 215. Fairbanks-Morse Pump. 
 
 . 
 
 a lining of wood has been used for the bore of the cylinder, 
 and the valve chambers as well as the piping. 
 
 . In Fig. 216 the peculiar form of follower ring or plate is to 
 be noted. This gives a very long wearing surface. 
 
 The Worthington plunger and ring form of water end (Fig. 
 217) is one which has many advantages. The water enters 
 the suction chamber A through C and passes up into the pump 
 cylinders, and from there it is forced into the discharge pipe 
 
272 
 
 PUMPING MACHINERY 
 
 D through the chamber B. The plunger is fastened to the 
 piston rod by two nuts. It operates through a sleeve E, which 
 is held in the partition between the sides by a ring F. The 
 sleeve is called a ring. There is no tight packing. The joint 
 
 FIG. 216. Worthington Pump. 
 
 is so long that there is practically no leakage. This type of 
 pump has been quite successful. 
 
 Outside-packed plunger pumps have the advantage that 
 the leakage past their displacing parts is visible. They are of 
 various forms. The piston rod is at times carried through the 
 two plungers (Fig. 218), the rod passing through a sleeve or 
 stuffing box. In this pump the suction enters at A and is 
 
DESIGN OF PARTS 
 
 273 
 
 carried to the valves, and finally discharges into the chamber 
 C and from it into B. This type of end outside-packed plunger 
 
 FIG. 217. Plunger and Ring 
 
 FIG. 218. Worthington Pump. 
 
 pump is simple, and the water passages are all ample. The 
 great disadvantage is in the possibility of leakage around the 
 piston rod. To obviate this leakage the plungers are -con- 
 nected on the outside by a trombone frame (Fig. 219). The 
 
274 
 
 PUMPING MACHINERY 
 
 rods E and F join the two cross heads G and H together. This 
 pump may be the same in form as Fig. 218, with the exception 
 of the connecting rod and sleeve. 
 
 A center outside-packed pump (Fig. 220) does away with 
 
 FIG. 219. Outside-packed Plungers. 
 
 FIG. 220. Center Outside Packing. 
 
 the necessity of outside rods and gives a pump in which all 
 leakage past the plungers or rods is visible; the stuffing boxes 
 are also clearly shown. 
 
 Water enters the suction chamber A from B, passes directly 
 into the pump, and is driven into chamber C, and from there 
 
DESIGN OF PARTS 
 
 275 
 
 into the discharge D. The space F is opened to the outside so 
 that the stuffing boxes are always visible. As constructed in 
 this figure the chambers C and A are cast solid with the other 
 parts. This is riot always the method of construction, as in 
 many cases the two parts of the water end are separate castings 
 connected by the chambers A and C. 
 
 The discharge valves are placed beneath hand plates G 
 and H, while the suction valves may be examined and repaired 
 through side hand holes in the pump barrel. The valves have 
 removable seats and these are so placed that there is a direct 
 
 SECTION THROUGH x-y 
 
 FIG. 221. Fire Pump. 
 
 path of sufficient dimensions through the pump. Such arrange- 
 ments cut down the friction and give a more smoothly acting 
 pump. 
 
 These same points of a direct path, simplicity and ease of 
 repair, are seen in the double-acting fire pump (Fig. 221), 
 where the water enters at A and leaves at B. The piston is 
 packed with cup leathers, and the passages from it to the 
 suction valves are very large. The reason for this is the high 
 rotative speed of these pumps and the necessity for having no 
 interruption in the water column. There is not the need for 
 such large passages on the discharge side as the water is being 
 forced out. By removal of certain parts, these valves may be 
 
276 
 
 PUMPING MACHINERY 
 
 A 
 
 FIG. 222. Metropolitan Fire Pump. 
 
DESIGN OP PARTS 
 
 277 
 
 examined, although this is not so easily done as in most other 
 pumps. These pumps are so closely built on account of the 
 lack of room that many desirable. features have to be sacrificed 
 in order that more urgent needs may be met. The Ahrens 
 
 FIG. 223. Railroad Pump. 
 
 pump has been improved by the American Fire Engine Co. 
 in their Metropolitan engine. The suction valves (Fig. 222) 
 have been replaced at each end of the cylinder through five 
 suction valves of large area. The two discharge valves have 
 not the same area as the suction, but -friction here, although 
 costly, will not interfere with the smooth running of the pump. 
 
278 PUMPING MACHINERY 
 
 The discharge occurs through B. C is the discharge air chamber 
 and D is the suction air chamber. 
 
 The valves of the pump are easily examined or replaced 
 by the removal of the cylinder heads EE or hand-hole plate 
 -F over the discharge valves. When necessary the cylinder 
 liner may be renewed. The piston is made deep and con- 
 tains a number of grooves. 
 
 Fig. 223 is a detail of the water end of a pump used for 
 railroad work. It is driven by a gas engine as shown in Fig. 
 147. There are four valves arranged around the pump barrel; 
 two suction and two discharge. The suction valve shown in 
 the figure is connected with the upper end of the cylinder, 
 while the discharge valve is connected with the lower end. 
 The valves not shown are connected with the opposite ends. 
 The piston rod A is guided by the arm B while the cross-head 
 C is connected with the pitman bar from a pin on a gear wheel. 
 The air chamber D serves to steady the discharge. 
 
 The valve boxes are provided with covers which are held 
 in place by yoke pieces so that the valves may be easily exam- 
 ined. The valve boxes are of ample size. 
 
 The barrels of deep-well pumps (Fig. 224) are usually made 
 of a piece of brass tubing lowered into the well. The valves 
 are of the disc-lift type or the ball type. The foot or suction 
 valve is either lowered into place and held to its seat by its 
 own weight or it may be put in place by attaching it to the 
 pump rods and forcing it into position. It is necessary to 
 have the foot valve water tight, as there will be lifting should 
 the water be forced around it. These foot valves and buckets 
 are usually packed with cup leathers. 
 
 The pump rods are often made of close-grain lumber with 
 iron armored ends. One end is made into a nut, the other 
 end is threaded as a bolt. In threading these ends considerable 
 taper is used, so that it is only necessary to turn the rod three 
 or four times to have ten or fifteen threads in contact when 
 joining the sections. Precautions must be taken to prevent 
 these screws from backing off, as the pump line depends on 
 their holding. 
 
DESIGN OF PARTS 279 
 
 Pressure pumps (Fig. 225 ) are made with heavy .cylinder 
 
 FIG. 224. Deep-Well Pump and Rods. 
 
 y 
 
 A 
 
 y 
 
 walls, and in most cases they are of the plunger type, as this 
 design is very suitable to high pressures. The packing may be 
 
280 
 
 PUMPING MACHINERY 
 
DESIGN OF PARTS 281 
 
 of the ordinary hemp form shown in the figure or leather packing 
 may be used. The valve chambers are usually single castings. 
 They vary in form; sometimes the casting contains both suc- 
 tion and discharge valves, and at other times only one valve 
 is in each casting. The valve castings A A are bolted to the 
 suction casting B, and the discharge casting C as well as to the 
 pump barrels DD. This arrangement gives a simple pump 
 cylinder, and although the valve chambers are not simple, 
 they are easily built, and machined. The valves are so placed 
 that they may be examined and repaired by the removal of a cap. 
 When these pressure pumps are made larger and more valves are 
 required, the valve-box castings are increased in number. 
 
 FIG. 226. Pressure Pump. 
 
 In smaller pumps under very high pressure the valve 
 chambers may be cast with the pump barrel as shown in Fig. 
 226, which represents a high-pressure Burnham pump. The 
 small valves are so arranged that by removing one plug both 
 suction and discharge valves may be ' examined. Attention 
 is called to the great difference between the plunger and the 
 steam piston areas and to the thickness of the cylinder wall. 
 The valve seats are renewable, so that if any of these should 
 wear it could easily be replaced without disconnecting the 
 pump. 
 
 The water end of large pumps has been greatly simplified. 
 One of the older pumps for the Boston sewage system, designed 
 by Mr. E. D. Leavitt, Jr., is shown in Fig. 227. In this case 
 the valves were rectangular clack valves, 3f Xi3i inches. Six 
 
282 
 
 PUMPING MACHINERY 
 
 of them were attached to a frame and placed on the openings of 
 the suction, while three discharge valves were attached to ea.ch 
 frame on the other side. The frame with the suction valves 
 
 open as would occur on the 
 suction stroke is shown on 
 the left, while on the right 
 the frame alone is to be 
 seen in section and at the 
 upper right hand the dis- 
 charge valves are shown 
 closed. 
 
 The suction valves and 
 discharge valves are in the 
 proportion of 36 to 27. The 
 reason for this was seen in 
 Chapter V. In all cases an 
 endeavor is made to cut 
 down the losses on the 
 suction side. The clear 
 passage to and from the 
 valves is necessary, and 
 valves of large areas with 
 no supports obstructing the 
 passage are required to 
 pass the solid matter which 
 is found in sewage. Mr. 
 Leavitt states that there 
 is a record of this pump 
 having passed through its 
 water cylinder a plank 
 2X12X36 inches. 
 
 The water end is 48 
 inches in diameter and of 
 
 form 
 
 FIG. 227 .-Sewage Pump. 
 
 of the plunger is dotted in the lowest position. The form of 
 stuffing box is illustrated together with the grooves on the 
 inside of the gland and bushing. This form of so-called 
 
DESIGN OF PARTS 
 
 283 
 
 labyrinth packing has been proven of little value. The 
 manhole in the center ring of the pump permits entrance 
 
 FIG. 228. Lawrence Pump. 
 
 for the examination of the suction valves, while a manhole 
 in the discharge chamber permits the examination of the other 
 valves. The cylindrical casting is made in three principal 
 
284 
 
 PUMPING MACHINERY 
 
 parts to simplify the foundry work and to make shipping arid 
 erection less difficult. 
 
 FIG. 229. Ontario Pump. 
 
 Fig. 228 illustrates the Lawrence Water Works pump of 
 Mr. E. D. Leavitt, Jr. It is the bucket form of pump with a 
 
DESIGN OF PARTS 
 
 285 
 
 supplementary discharge at A for the purpose of giving a 
 passage to the water should the valve in the bucket cease to 
 operate. When the pump operates as a bucket pump the 
 course of the water is direct through the pump barrel. The 
 use of a plunger enlargement on the rod of one-half the area 
 of the piston serves to produce a discharge on each stroke, 
 while suction occurs on every other stroke. 
 
 The valves for this pump were originally 1 6-inch brass 
 double-beat valves shown in Fig. 252, but as the friction of the 
 
 
 /; 
 
 
 s5< 
 
 ^ 
 
 1 
 
 FIG. 230. Milwaukee Pump. 
 
 pump was excessive the valves were changed to the annular 
 form as shown in Figs. 228 and 250. This reduced the friction 
 very materially. The castings making up the pump, the air 
 chamber B, the manholes, valve decks and other details are 
 seen by a study of the picture. 
 
 A single-suction double-discharge plunger pump is shown 
 in Fig. 229. In this type of pump for Ontario, Leavitt used a 
 large number of small valves and on the suction side small 
 draft tubes were placed below the valves. The plunger was 
 
286 
 
 PUMPING MACHINERY 
 
 shaped to fit into the bottom of the cylinder and cause all of 
 the water to be in circulation. Such a construction may be 
 questioned unless there is a lack of head room and it is thought 
 necessary to point the plunger. The plunger has an extension 
 sleeve on it of one-half the area of the plunger. The castings 
 are simple and well braced by brackets and webs, as well as 
 
 FIG. 231. Cincinnati Pump. 
 
 reinforced by rings around the barrel. The path of the water 
 is direct, and the head pressure on the discharge valve deck is 
 supported by the webs A A, which act as girders. 
 
 The valves are spring controlled, of small size, and shown in 
 greater detail in Fig. 252. 
 
 The Allis pump of Milwaukee (Fig. 230) is .another case in 
 which a series of lift valves were used to replace large double- 
 beat valves. In this case, each large valve was replaced by a 
 box or cage on the sides and top of which were openings for 
 
DESIGN OF PARTS 
 
 287 
 
 small valves as shown in detail in Fig. 254. The construction 
 shown in Fig. 230 illustrates how simple the castings of a large 
 purnp may be. The valve decks, which are reinforced by cross 
 ribs or girders to withstand the pressure, are the end parts of 
 individual castings; on account of this the faces may be easily 
 machined in the shop, and erected in the field. The path of 
 
 FIG. 232. Snow Pump End. 
 
 the water is more or less direct and the suction pipe is of proper 
 size for this slow-running pump. 
 
 Fig. 231 illustrates a 3o,ooo,ooo-gallon water end built by 
 the Holly Pump Co. The suction pipe enters the valve chamber 
 on each side of the pump at AA\ the valve decks BB are made 
 of similar castings of considerable depth to withstand the high 
 working pressure. The valves are mounted on cages as in the 
 previous figure. The discharge occurs at CC. The space D 
 in the top of the discharge acts as an air chamber. The pipes 
 EE are used to introduce compressed air into this space. The 
 
288 
 
 PUMPING MACHINERY 
 
 plunger F draws water from the suction valves on each side 
 of it on its up stroke and discharges through the two sets of 
 discharge valves on its down stroke. The connection H made 
 between the discharge side and pump space is used to prime 
 the suction valves when necessary, but may also be used in 
 starting the pump to equalize the pressure on the plunger 
 on each stroke. A large suction air chamber is shown at K. 
 Fig. 232 shows the water end of a large horizontal pump. 
 
 FIG. 234. Metal Hinge Clack Valve. 
 
 FIG. 233. Leather Clack Valve. 
 
 FIG. 235. Clack Valve. 
 
 The water end is made up of four principal castings: A suction 
 chamber A\ two pump ends BB; and a discharge chamber 
 C, containing the air chambers DD. The valve decks in the 
 pump ends BB contain a large number of small valves; they 
 are well braced by cross ribs as seen in the figure. The man- 
 hole cover at E, which is hung from an arm, serves for the exam- 
 ination of the suction valves, while that at F is used for the 
 discharge. The figure illustrates very clearly the method of 
 packing the plunger and piston rods and the simple manner in 
 
DESIGN OF PARTS 
 
 289 
 
 which the castings can be made. The trough G beneath the 
 intermediate stuffing boxes is intended to catch the drip. 
 
 One of the simplest valves found in small hand pumps is 
 the leather clack valve (Fig. 233). A circular piece of leather 
 shown in plan in the figure has a groove cut out of it and after 
 fastening two iron washers to this, the leather is held beneath 
 the pump barrel and the valve seat casting. The small part 
 of the leather left after cutting the groove forms a hinge. The 
 upper washer is not only used to weight the clack, but it also 
 supports the weight of water above the leather, making a water- 
 tight valve. 
 
 In Fig. 234 the leather hinge has been replaced by a pin, 
 and in Fig. 235 the leather hinge is retained for a valve which is 
 
 FIG. 236. Butterfly Valve. 
 
 rectangular. Such a valve is used for sewage pumping (Fig. 
 227). A double clack valve (Fig. 236) is sometimes called a 
 butterfly valve. In this form the valve is made rectangular. 
 Such a shape is often found when clack valves are used. The 
 valve seat is formed separate from the pump casting and is 
 bolted in position. Stops are usually employed to keep the 
 valve from opening too much. 
 
 A metal clack valve (Fig. 237) may be used at times where 
 the liquid handled would destroy the leather.. The pivot in 
 this case moves in a slot to allow for the wear of the valve. 
 A pin through a hole would not allow the valve to seat properly 
 after wear had occurred. 
 
 The conical valve of Fig. 238 is guided by wings on its lower 
 face. The seat of this valve is often made of bronze, and 
 
290 
 
 PUMPING MACHINERY 
 
 inserted in the valve deck. This renders it an easy matter 
 to renew the seat and permit one to use a better metal for it, 
 the valve deck being made of a soft iron. The valve is fitted 
 to the seat by grinding. The operation consists in turning the 
 valve against the seat by a screwdriver, after oil and emery 
 are introduced between them. To give the valve in seating 
 
 FIG. 237. Metal Clack Valve. 
 
 a turning motion, the wings below the valve disc are bent into 
 a helical form (Fig. 239). The action of the water on the 
 vanes is to rotate the disc. This allows the valve to seat at 
 different points each time, thus eliminating the excessive 
 grooving which occurs when a valve always seats at the same 
 point after the formation of an incipient groove. The rotary 
 motion also gives the valve a wiping action in seating. 
 
 n 
 
 ^^^ 
 
 s 
 
 
 
 / 
 
 \\\\\\^ 
 
 
 
 
 
 
 
 
 
 
 
 
 
 n 
 
 FIG. 238. Conical Valve. 
 
 FIG. 239. Helical Wings. 
 
 The ball valve (Fig. 240) is an effective type. The action 
 of the cage over the ball and the removable nature of all parts 
 are evident from the figure. 
 
 One of the common forms of valves used with all forms of 
 pumps is shown in Fig. 241. The valve seat is made of com- 
 position brass, and screws into the valve deck. The valve is 
 made of a composition of rubber and other substances. It is 
 backed by a piece of sheet brass against which the spring presses. 
 
DESIGN OF PARTS 
 
 201 
 
 The spring is held beneath a nut on the valve spindle. The 
 spindle is screwed down tight against a shoulder so that it 
 
 FIG. 240. Ball Valve. 
 
 will not back off. A split pin put in the hole in the top of the 
 spindle prevents the nut from unscrewing. All parts shown 
 
 FIG. 241. Disc Valve. 
 
 in the figure with the exception of the disc are made of brass. 
 The figure shows the best method of installing the valve spindle, 
 as the spindle remains in the seat when the valve is removed. 
 
292 
 
 PUMPING MACHINERY 
 
 At times, however, the spindle and nut are combined into one 
 piece, Fig. 242, and in placing this in position a plug wrench 
 is used. 
 
 The valve disc is sometimes backed up by a brass cup 
 
 FIG. 242. Valve Spindle. 
 
 FIG. 243. Valve Barking. 
 
 FIG. 244. Metal Disc Valve. 
 
 FIG. 245. Cameron Valves. 
 
 into which it fits (Fig. 243).- This gives more stiffness to the 
 valve. 
 
 When hot water is used brass valves replace the composition 
 valves which are intended for cold water, although there are 
 special compositions used for hot water. Fig. 244 illustrates 
 the form often used for metal valves. 
 
 A neat arrangement for the suction and discharge valves 
 is made by the Cameron Company. These valves (Fig. 245 ) 
 
DESIGN OF PARTS 
 
 293 
 
 are placed on a common spindle which is held in place by a set 
 screw in a cap, screwed into the wall of the valve chest. This 
 spindle also serves to hold the valve seats in place. These 
 valves are backed with metal cups. 
 
 The Marsh pump is equipped with metal valves the spindles 
 
 FIG. 246. Marsh Valves. 
 
 of- which are guided by holes in the metal at the center of the 
 valve seat. This center is so formed that the water is deflected 
 under the valve as shown in Fig. 246. This central part is 
 almost as large as the cavity of the valve and when the valves 
 are raised the distance shown in the figure, there is no tendency 
 
 FIG. 247. Rubber Valve with Guard. 
 
 for them to lift higher. This puts them into a position where 
 they may seat quickly, but without shock, as the deflector acts 
 as the piston of a dash pot. 
 
 When rubber valves are made of large diameter they are 
 designed as shown in Fig. 247. The guard on the back of the 
 valve keeps the valve from opening too far. The. holes in the 
 
294 
 
 PUMPING MACHINERY 
 
 center permit the water to act on the back of the disc and aid 
 in quick closing. The method of supporting the spindle, and 
 thus holding the seat in position, is clear from the figure. 
 
 FIG. 248. Double Ported Valve. 
 
 To increase the opening in the valve discs of large diameter 
 they are made multiported; in reality they become a series 
 of concentric rings, joined by radial arms, as was explained in 
 
 FIG. 249. Riedler Valve. 
 
 Chapter V. Fig. 248 illustrates such a valve with seat, spindle, 
 spring, and nut where one ring only is used. The form of this 
 type of valve, used by the Allis-Chalmers Co. in their Riedler 
 pumps, is best shown by Fig. 249. In this valve a leather 
 
DESIGN OF PARTS 
 
 295 
 
 washer is used in addition to the conical-faced ring and seat. 
 The leather is held between two iron rings. This leather really 
 makes the joint when the pump is in action. The valve may 
 be raised from its seat by the water pressure when the arms 
 CC are raised. The amount of motion is limited by the nut 
 D on the spindle. At the end of the stroke the arms CC are 
 driven down against the metal sleeve B by the action of an 
 eccentric. This presses against a rubber collar A and forces 
 the valve to its seat. The rubber collar gives a yielding 
 
 FIG. 250. Weighted Valve. 
 
 connection to care for the possibility of solid objects getting 
 beneath the valve. Before the pump reaches the other end 
 of its stroke the arms CC are raised so that the valve may 
 open as soon as the pressure is sufficient. The Riedler valve 
 was first employed with express pumps. 
 
 A weighted ring valve of American design is shown in 
 Fig. 250, although for smaller sizes springs could be used, as 
 shown in Fig. 251. This valve (Fig. 250) was employed on the 
 Lawrence pump to replace the double-beat valve of Fig. 252. 
 Where large area is required, double-beat valves (Figs. 252 
 
296 
 
 PUMPING MACHINERY 
 
 and 253) are used. Double-beat valves are valves containing 
 two seats, as at A and B. The upper seat B is made smaller 
 than that at A , and hence an upward pressure from below causes 
 the valve to lift from its seat and the water to escape by these 
 two openings. This gives a large area and by 
 properly arranging the size of each seat, a 
 proper amount of pressure increase can be had 
 to lift the valve weight. The motion of the 
 valve is limited by the nut on the spindle. In 
 Fig. 252 the valve is open to its full extent. 
 Although these valves have been used to some 
 FIG. 251. Double degree, the American practice is to use a num- 
 ber of small valves, and where sufficient valve 
 deck area is not available valve caps (Fig. 254) may be placed 
 over the openings used for double-beat valves. In this manner 
 a large discharge area may be obtained. The valve cap or 
 box is held to the main valve deck by a large through bolt. 
 A series of webs are employed within this to stiffen it. 
 
 FIG. 252. Double Beat Valve. 
 
 FIG. 253. Double Beat Valve. 
 
 The valves shown in Fig. 255 illustrate the method used for 
 high-speed pumps. These valves have been recently introduced 
 by Witting of England. In them, two brass rings are placed 
 around each opening. The rubber rings BB force the brass 
 rings together when the water pressure on each side of the 
 brass ring is the same. The rings touch each other, and are 
 
DESIGN OF PARTS 
 
 297 
 
 held together by the water pressure on the outside as well as 
 by the rubber. When the internal pressure is sufficient to force 
 
 FIG. 254. Valve Box. 
 
 these apart, the brass ring? are separated and this give? a 
 series of large openings to which the water is guided. These 
 valves may be made with leather facings, and helical springs 
 may replace the rubber. The lower figure illustrates a method 
 
298 
 
 MACHINERY 
 
 mm 
 
 FIG. 255. Witting's Metallic Valves. 
 
 FIG. 256. Pressure Valve. 
 
DESIGN OF PARTS 
 
 299 
 
 of increasing the valve area. These valves are so constructed 
 that they may replace other valves, the cages being bolted 
 to the valve decks. These valves are somewhat similar to the 
 
 rm 
 
 FIG. 257. Pressure Valves. 
 
 valves used on Merryweather's fire engine of 1870, in which 
 rubber replaced the metal rings. 
 
 In high-pressure pumps the valves are made deep for stiff- 
 ness; the springs are usually made heavier on the discharge 
 side to close the valve quickly, as considerable spring pressure 
 is not objectionable on this side. The cap over the valve 
 and the walls are made heavy and the valve should be access- 
 ible. Fig. 256 illustrates a simple form of valve, as used with 
 
300 
 
 PUMPING MACHINERY 
 
 the Burnham pump, while Fig. 257 shows a series of valves 
 on one deck. The springs are held in place by a ring. This 
 makes it possible to remove the cover of the valve chamber 
 for inspection without disturbing the springs. In these two 
 arrangements the springs are guided by spindles on the caps, 
 frame, or valve. When this cannot be done a frame is formed 
 on top of the valve and placed around the valve and its frame 
 (Fig. 258). Such a cage may be held down by a ring and bolts 
 (Fig. 259), as was also used in Fig. 257. The massive construc- 
 tion shown in the last four figures is due to the excessive 
 
 FIG. 258. Pressure Valve and Cage. 
 
 FIG. 259. Valve Pot. 
 
 pressure carried. In all of them, it is seen that the parts may be 
 easily removed and repaired, and quick seating of the valve 
 is possible so that the slip may be cut to a minimum. 
 
 A form of valve used by German pump builders is known as 
 the Gutermuth valve. 
 
 In this valve (Fig. 260) a sheet of steel or bronze is bent 
 into the form of a spiral which forms a spring for the end of the 
 strip, which is left flat to form the valve. A pressure from 
 beneath drives the plate into the dotted position, and the water 
 has a path through which the flow occurs almost straight from 
 the channel leading to the valve. This channel is not arranged 
 
DESIGN OF PARTS 
 
 301 
 
 normal to the valve disc, as it is desired to continue the flow 
 in as straight a line as possible. 
 
 FIG. 260. Gutermuth Valve. 
 
 JTL 
 
 
 FIG. 261. Borsig Valve. 
 
 The Borsig Co., of Germany, use a metal valve of very 
 simple form. It is made of thin spring metal of the form shown 
 in Fig. 261. The points A A are bolted to the valve deck, 
 
302 
 
 PUMPING MACHINERY 
 
 and the arms extending to the outer ring form the springs to 
 seat the valve. At times additional helical springs are placed 
 on the back of this. 
 
 The discharge air chambers for pumps are made in various 
 
 <Q [j@5 
 
 forms, as shown in Fig. 262. For small pumps the first form 
 is often used with the main body of sheet copper, shaped into a 
 conical vessel. The other forms are usually made of cast iron, 
 and are placed on top of the discharge valve chamber. The 
 second form is well arranged, as the main discharge is cared 
 for by the bend in the pipe. The sizes of these are fixed by 
 
DESIGN OF PARTS 
 
 303 
 
 considerations given in Chapter V, and the thickness is deter- 
 mined by the formulas given in this chapter. As was 
 pointed out earlier, the size of the air chamber will depend on 
 the details of an installation, so that a standard size cannot 
 be fixed for stock pumps which will be suitable for any 
 purpose. 
 
 The air from the chamber is gradually absorbed by the 
 water, and to keep the chambers charged air compressors may 
 be used. A simple device (Fig. 263) is sometimes applied. 
 This consists of piping attached to 
 the pump cylinder and air chamber 
 in which are placed two check valves 
 and one gate valve. The check 
 valves open toward the air chamber. 
 When the valve is opened air is 
 drawn into the vertical line on the 
 suction stroke of the pump, and on 
 the discharge stroke this is com- 
 pressed and sent over into the air 
 chamber, as the air cannot escape to 
 the atmosphere and the pressure in 
 the cylinder is greater than the pres- 
 sure in the air chamber. On each 
 stroke there is a slight discharge of 
 air into the chamber, and after suf- 
 ficient air is obtained the stop valve is shut off and the 
 operation of this air pump ceases. 
 
 The air chambers on the suction side of the pump are 
 important adjuncts, as was seen in Chapter V. They should 
 be placed so that they can receive and care for the water which 
 it is necessary to store for the variation in the suction of the 
 pump. The air chambers of Fig. 264 are all placed at the end 
 of a suction line, while the water is taken to the valve chest 
 from the side of this line. The various arrangements illustrate 
 how the air chamber can be placed when different conditions 
 of piping exist. In all cases the water is flowing directly into 
 the air chamber. 
 
 FIG. 263. Air Chamber 
 Pump. 
 
301 
 
 PUMPING MACHINERY 
 
 Fig. 265 shows two air chambers; the left-hand being a 
 simple one, made of pipe fittings, and of proper design, while 
 
 Ift 
 
 the right is improperly arranged, since the water, in passing 
 at right angles to the air chamber, must exert some force on the 
 valves beyond before the air in the chamber is compressed. 
 
DESIGN OF PARTS 
 
 305 
 
 The left-hand chamber discharges from the end B at a varying 
 rate, while the supply at A is uniform. The excess or deficiency 
 is made up from the water which passes into the space around 
 the upper pipe. This flow goes naturally into the space, as the 
 stream enlarges suddenly. To cut down the loss due to sudden 
 enlargement and yet get the advantage of it, the suction air 
 chamber could be built as shown in Fig. 266, and in this manner 
 
 FIG. 265. Suction Air Chambers. 
 
 FIG. 266. Suction Air Chamber. 
 
 the losses of sudden enlargement and contraction could be 
 reduced. 
 
 To protect the pump valves against leaves, twigs, and other 
 foreign particles which might lodge beneath the valves and 
 prevent their proper action, strainers (Figs. 267, 268, 269 and 
 270) are used. These are made of wires, grills, or drilled plates. 
 They should always be so constructed that they may be cleaned 
 easily. At times the strainer is combined with a foot valve, 
 as in Figs. 269 and 270. A foot valve is a large valve or series 
 of small valves attached to a valve deck as shown in the figures, 
 and placed on the end of the suction main. The purpose of 
 
PUMPING MACHINERY 
 
 the foot valve is to hold the water in the suction main when the 
 pump is shut down, and to keep the line filled in case it is very 
 long. The large foot valve may be of the lift form, as shown 
 
 FIG. 267. Suction Strainer. 
 
 
 IT 
 
 FIG. 268. Strainer. 
 
 (\ 
 
 
 
 
 s 
 
 ( "N 
 
 oooc 
 oooo 
 oooo 
 oooc 
 oooo 
 oooc 
 oooc 
 nooc 
 
 000000 OOO00 
 
 ooooooooooooo 
 ooooooooooooo 
 
 ooooooooooooo 
 
 ooooooooooooo 
 ooooooooooooo 
 
 0000000 OOOO^gjJ 
 
 o o o o o o o 
 o o o o o 
 
 o o o o o ' 
 o o o o o rt 
 
 O O O O 0^3 
 
 FIG. 269. Foot Valve and Strainer. 
 
 oo y oo 
 
 poo cc 
 oooo 
 oooo 
 
 FIG. 270. Foot Valve and Strainer. 
 
 The 
 
 in Fig. 270, or it may be of the form shown in Fig. 271. 
 type shown in Fig. 268 is arranged as shown in Fig. 271. 
 
 The suction pipe should be of large diameter, as straight 
 and as short as possible, and air tight. The smallest leak is 
 apt to destroy the proper action of the pump. It is well to 
 test these pipes under pressure before covering them. The 
 
DESIGN OF PARTS 
 
 307 
 
 pipe must be laid on a uniform grade to prevent air pockets. 
 Everything should be done to cut down the losses, as the total 
 loss with the lift must be kept less than 34 feet. 
 
 With these details in mind, the design will now be con- 
 tinued. The number of revolutions, the size of the pump, 
 the size and number of the valves, the sizes of the pipe lines, 
 
 FIG. 271. Foot Valves. 
 
 and the size of the air chambers are determined as in Chapter 
 V. From these, the pressure in the water cylinder can be 
 found. 
 
 DESIGN OF CYLINDER 
 
 In the water cylinder (Fig. 272) the first quantity to be 
 determined is the thickness of the walls. 
 
 When the cylinder walls are cylindrical and the surface 
 is not subject to wear, the formula: 
 
 pd = 2iS or its equivalent may be used; 
 p=the water pressure in Ibs. per sq.in.; 
 d = diameter to center of walls in inches; 
 t = thickness in inches; 
 5= allowable stress in Ibs. per sq.in. 
 
 This formula is for thin cylinders; for thick cylinders used 
 
308 
 
 PUMPING MACHINERY 
 
 with heavy pressures the formulae of Banow and Lam6 are 
 employed. 
 
 Before proceeding with these formulae it will be well to 
 note the table (page 309) of the strengths and other properties 
 of the materials to be used in design together with the safety 
 factors required. These values have been taken from standard 
 
 FIG. 272. Pump Section. 
 
 texts on machine design as well as from engineering handbooks 
 and current literature. 
 
 The formula above where cast iron is used as the cylinder 
 wall is put into the form 
 
 . 
 
 5000 
 
 When the cylinder is bored for the reception of the piston 
 and allowance is made for reboring, this equation becomes 
 
 .. 
 
 5000 
 
 In making the head as shown in the Fig. 272, the radius of 
 the spherical portion should be made equal to the diameter 
 
DESIGN OF PARTS 
 
 309 
 
 C <-. *- rj 
 
 || ^J 
 
 oo 
 
 <H 1 S ? i r?? i ? i i i 
 
 1 1 1 1 
 
 'So w.2*: 
 
 
 
 
 )> PH Q) r^ O 
 
 VO 00 
 T 1- 
 
 ? ? ^^isiia^i^^^^ 
 
 s 5 5 
 
 ^ 
 ' o 
 
 00 N 
 
 HMMM^HM! 1 1 HH) I ^^^?^M 
 
 1 
 
 JJ -S 
 
 ea w 
 
 
 
 
 o_. 0)13 
 
 > rt 
 
 <M \O X 
 
 ' MI o 'o-o'e e O.'Q e o 
 
 00^0 
 
 E J >3 
 
 
 
 
 y r s' r s 
 
 
 
 
 fe 3H 
 
 
 M 
 
 H 
 
 w > 
 
 H 
 
 ^ w 
 
 20,000,000 
 29,000,000 
 
 00 0000 0000 
 
 oo oooo oooo 
 o" o" 1 1 1 o" o" 1 o" o" I | 1 1 w" o" o" o" 
 
 Sol I loolool I I IMQOO 
 O OO"">O >Ot^<3 
 o" o" o" 10 ro 't H" H M 
 
 f> to ro w H H 
 
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310 PUMPING MACHINERY 
 
 of the cylinder, and then the thickness should just equal that of 
 the walls. 
 
 If the head has to be made flat the formula for the thickness 
 when well supported by bolts is 
 
 for a round plate (d = diameter), while 
 
 * =0.350 \l 
 
 'S 
 
 holds for a square plate (a= side of square), and 
 
 ab 
 
 for a rectangular plate of length a and width b. 
 
 For manhole or handhole covers which are of elliptical 
 form, the thickness is fixed by the formula 
 
 ab 
 
 In all cases the endeavor should be made to eliminate flat 
 places, and when such must exist they must be designed for 
 thickness as outlined above. At times ribs are used to stiffen 
 flat places, and at times stay bolts are used crossing from side 
 to side. 
 
 The thickness of heavy-pressure pump cylinders is deter- 
 mined by either one of two formulae. Barlow's: 
 
 f P r * 
 ~S-p' 
 or Lamp's: 
 
 = S' 
 
 t= thickness in inches; 
 p= hydrostatic pressure in Ibs. per sq.in.; 
 TI = the outside radius of cylinder, in inches; 
 y 2 =the inside radius of cylinder in inches; 
 S=maximum stress in Ibs. per sq.in. 
 
DESIGN OF PARTS 
 
 311 
 
 In designing the cylinders for hydraulic presses or pressure 
 pumps the metal thickness should be uniform, as thicker parts 
 
 of castings cool after the other parts 
 have solidified. There is danger, there- 
 fore, that cooling strains will be set up 
 and cracks may develop. 
 
 There should be no sudden changes 
 in the thickness of the section and all 
 corners must be well filleted. These 
 points should be observed in all cylinder 
 design or in any other castings. 
 
 When the core of the heavy cyl- 
 inder castings has to be supported at each end a hole must be 
 left at the bottom of the cylinder, and when such is the case 
 it is closed by driving in a plug or by using a packed joint 
 (Fig. 274) in which a hydraulic packing of some form is 
 used. 
 
 FIG. 273. Pressure 
 Cylinder. 
 
 FIG. 274. Core Hole Cap. 
 
 FIG. 275. Twin-Cylinder Casting. 
 
 When two hydraulic cylinders are to be used in one casting, 
 there must be sufficient metal where the two cylinders come 
 together to support the total load in each cylinder. The upper 
 part of Fig. 275 shows the incorrect method, while in the case 
 below there is sufficient metal to carry the load. 
 
 When there are to be openings made into cylinder walls 
 
312 
 
 PUMPING MACHINERY 
 
 for pipes, manholes, or other purposes, the edge of the hole 
 should be reinforced by a ring or boss, and the amount of metal 
 added should equal or should exceed the amount of metal 
 removed in the section, as shown in Fig. 276. In this figure 
 the amount of metal cut away is dt square inches, and this 
 
 * t 
 
 1 
 
 FIG. 276. Branch Connection Reinforcement. 
 
 should be equal to or less than (ad}h. These openings 
 should be avoided, but when necessary in every case the rein- 
 forcement should be made. 
 
 FLANGES 
 
 The flanges to be used on cylinders, heads, manhole covers, 
 and covers for other parts of the pumps are in general made 
 equal to 3 diameters of bolt in width; this gives i^d from 
 the center of bolt to edge, a proportion which is common in all 
 design work. Until the size of the bolts is determined this 
 cannot be fixed. It is best, therefore, to assume the diameter 
 of bolt, and design the number to be used. The bolts should 
 be taken as large as possible. Small bolts are apt to be twisted 
 off. It is well to use nothing less than f -inch bolts except when 
 very small parts are used. On cylinders from 8 to 15 inches, 
 J-mch bolts should be assumed; from 15 to 24 inches, i inch; 
 and for larger sizes, the diameter should be more than 
 i inch. 
 
 The flange width is now made equal to 3^. 
 
 The thickness of the flange, t" , is made 1.2 to 1.4^; where 
 t= thickness of cylinder walls. 
 
 The bolts support the pressure on the area A of the 
 
DESIGN OF PARTS 313- 
 
 cover exposed to the water pressure. The number n is given 
 by 
 
 a = area of bolts at root of thread found in table of 
 
 bolts in sq.in.; 
 
 5= allowable stress in Ibs. per sq.in.; 
 A =area supported in sq.in. (usually to the inner edge 
 of cover, although if packing is not tight this area 
 may be increased); 
 p = water pressure in Ibs. per sq.in. 
 
 After finding the number of bolts, the pitch of the bolts 
 should be found, and if this is greater than 
 
 where t" is the thickness of the flange, more bolts must be 
 used. The quantity above is the maximum allowable pitch for 
 a tight joint. 
 
 WATER PISTONS 
 
 The water piston is made up of various parts, but it may 
 be considered as a flat plate supported at the center and loaded 
 uniformly. In this case the thickness would be about two- 
 thirds the thickness of the cylinder head. If the piston is 
 cored, as in Fig. 297, Chapter VIII, this thickness is made less, 
 as the casting acts like a deep beam. The thickness of each 
 part could be made less than one-third the thickness of the 
 head but this amount may be used. If the two parts are 
 separate, as in Fig. 199, then each should be of the full 
 thickness. 
 
 Piston design is in most cases empirical and the proportional 
 
 number is - - for cast iron or bronze. This will be con- 
 100 
 
 sidered in Chapter VIII, and until that place the proportions 
 will not be given. 
 
314 PUMPING MACHINERY 
 
 The packing used on water cylinders is often made of 
 square-plaited flax, hemp or cotton, with a rubber core, indurated 
 with rubber or alternate layers of rubber and cotton. The 
 sizes are J-XJ inch, f Xf inch, JXJ inch, etc. In putting this 
 in, the packing should be cut so that the ends do not touch by 
 about the thickness of the packing. This packing swells on 
 being wet. The depth of the packing is 0.4^ on Fig. 199. 
 
 PISTON RODS 
 
 The piston rod is designed to take compression only in a 
 single-acting pump, and to take compression and tension in a 
 double-acting pump. In many cases these rods are of steel, 
 but in some small pumps where corrosion is to be avoided, 
 phospor bronze, Muntz metal, or some other form of non- 
 rusting metal is used. 
 
 Tobin bronze is an alloy of copper (59%), zinc (38.4%), 
 tin (2.16%). Iron (0.11%) and lead (0.31%) are sometimes 
 found in the mixtures. It has a tensile strength as high as 
 75,000 when rolled, and at times 100,000 pounds per square 
 inch, has been obtained after cold rolling. At a cherry red 
 this bronze can be stamped and forged. 
 
 Muntz metal is an alloy of 3 parts of copper with two of 
 zinc. It has a strength of about 50,000 pounds per square inch. 
 
 The piston rod is treated as a column in all cases, as it is 
 the compressive load which fixes the size. 
 
 The figures of this chapter illustrate the method of attach- 
 ment of the piston rod. It will be seen that the end of the 
 rod has a taper extending from a shoulder or collar to a cylin- 
 drical-threaded portion. The threaded portion takes plain ten- 
 sion; the shoulder and the tapered portion take the com- 
 pression. It is well in forming the shoulder or collar to fillet 
 the corner instead of leaving it absolutely square, as this fillet 
 increases the strength, although it is not so easy to get a fit 
 of the rod to a rounded corner. 
 
 The load. on the rod is 
 
DESIGN OF PARTS 315 
 
 The area at the root of the thread is 
 
 _P 
 
 tt c* 
 
 The diameter of this thread is found from a table of standard 
 threads for the area a. 
 
 The rod is designed by a column formula: 
 
 P S c 
 
 , Rankin's Formula; 
 
 p i 
 
 -jz=S c K', Straight-Line Formula; 
 
 TtCl Y 
 
 T 
 
 P = total load on one rod in Ibs; 
 d = diameter of rod in inches; 
 
 d 
 
 r = radius of gyration =-; 
 
 5 C = allowable working stress in Ibs. per sq.in.; 
 /= greatest length of rod between supports in inches; 
 
 K= - -- for steel; round ends; 
 6250 
 
 K'= 284 for steel; round ends. 
 
 These equations must be solved by trial by first getting the 
 
 p 
 value of d from ^ = 5 C , and then using this to find r\ solve 
 
 4 
 for d after the approximate value of r has been inserted. 
 
 The shoulder used on the rod should be j- to J inch on small 
 rod, while \ inch may be used if a collar is formed. The taper 
 part of the rod should reduce the section to the diameter at the 
 thread. A taper such that the diameter is reduced 3 inches in 
 a toot is a good one to use. 
 
 STUFFING BOXES 
 
 The stuffing-box design has been given earlier in this 
 chapter. 
 
316 PUMPING MACHINERY 
 
 VALVES AND SPRINGS 
 
 The proportion of valves will be seen from the various 
 figures. The diameter of the valves has been discussed in 
 Chapter V. The spring design will now be considered. The 
 formulae for springs are: 
 
 <* 
 
 where 5 =safe shearing strength in Ibs. per sq.in.; 
 P=loadinlbs.; 
 n = total number of coils; Jj 
 d = diameter of wire in inches; 
 =torsional modulus of elasticity; 
 X= compression in inches. 
 
 From formula (a) the radius of coil may be found to carry 
 the load for an assumed diameter and stress for the spring 
 wire. With this radius the number n may be found to give 
 the necessary compression with a given change of load. 
 
 The pitch of the helix is found by assuming the height 
 allowable, .and putting n turns in this height. If this result 
 is not thought proper a different diameter of wire may be taken; 
 a new r and n, found. This should be tried until one of the 
 proper form is found. Although r and n are definite and would 
 give the desired result, they do not always fit the other parts 
 of the design. 
 
 The following proportions for valves are used in practice, 
 according to Barr: 
 
DESIGN OF PARTS 317 
 
 VALVE Discs 
 
 Diameter. Thickness. Hole. 
 
 2 " f" 1" 
 
 21" A" r 
 
 3 " r ft" 
 3i" f" f" 
 
 4 " f " ft" 
 4i" I" ft" 
 
 r> 3" 13" 
 
 5 4 T5 
 
 Taper of thread on brass valve seat to fit valve deck i inch per 
 foot. Valve stem ^ inch less than diameter of hole. Plate on 
 top of valve ^ inch thick, and of diameter equal to three-fourths 
 diameter of disc. Plate ^ inch thick for valve? 4^ inches and 
 over. Springs are made of the following sizes: 
 
 Diameter of Valve. Size of Wire. 
 
 2 " NO. 12 
 2j" NO. 12 
 
 3 " No. 10 
 3i" No. 10 
 
 4 " No. 8 
 4i" No." 8 
 
 Springs are usually coiled with a diameter of one-half the valve 
 diameter, and with five turns they have sufficient elasticity. 
 
 Fig. 277 illustrates the method of installing a small modern 
 pump and applies the general principles outlined in these two 
 chapters. The foot valve A is applied at the lower end of the 
 suction pipe. The suction air chamber B is applied so as to 
 receive directly the impulse from the water. The strainer at 
 C is conveniently placed for cleaning. A check valve D per- 
 mits one to relieve the valve chamber of pressure when necessary 
 to examine the valves without draining the discharge main. 
 The priming pipe E is used to prime the pumps. By opening 
 the waste-drain pipe F and the primer E the pump is filled 
 with water, the air being driven out. The waste pipe may be 
 opened in starting the pump so that air may be removed before 
 the full pressure is exerted. Such an arrangement may be 
 
318 
 
 PUMPING MACHINERY 
 
 necessary in starting a compound pump where pressure is needed 
 on both pistons to start the pump against full head. In such 
 a case the pump could not start unless live steam were by-passed 
 into the low pressure cylinder. Another method sometimes 
 used, which is the equivalent of this, is to connect the two 
 ends of the water cylinder by a cross connection to eliminate 
 water pressure. After both cylinders have received the proper 
 
 FIG. 277. Arrangement of Pump. 
 
 steam on finishing a few strokes the waste or the connection 
 between the ends is closed, and the pump discharges against 
 the working head. 
 
 The figure illustrates large-size suction pipes and short, 
 direct connections, so necessary for proper action. Care must 
 be taken in making up these joints to have them air tight. 
 
 The following quotations from the catalogue of the Snow 
 Pump Works are valuable, although many of them have been 
 given previously in this chapter: 
 
DESIGN OF PARTS 319 
 
 " INFORMATION CONCERNING PIPE CONNECTIONS 
 
 " Fig. 277 shows in a general way the proper method of 
 piping a pump. Faulty connections are generally the cause 
 of the improper action of a pump, and great care should, there- 
 fore, be taken to have everything right before starting. To 
 accomplish this, note carefully and understand thoroughly 
 the following: 
 
 " Be sure that the quantity of water you desire to pump 
 is available and that your pump is within easy reach of it 
 when it is at its lowest level. 
 
 " Locate your pump as near the source of suction supply, 
 both vertically and horizontally, as is possible or convenient; 
 but never place it in such a location that the sum of the follow- 
 ing three items would exceed a total of 26 feet: 
 
 " i. Height in feet from the discharge valves of the pump 
 to the lowest level of the surface of the suction water. 
 
 " 2. Total frictional loss in suction pipe in feet head. 
 
 "3. Total frictional loss in feet head, due to elbows and 
 tees (assumed as being the equivalent of the frictional loss due 
 to 100 feet of same size of pipe, for each elbow or tee). 
 
 " EXAMPLE. Would a pump having an easy capacity of 
 750 U. S. gallons per minute operate satisfactorily at this 
 capacity if the height from the delivery valves to the surface 
 of the suction supply was 20 feet, the suction pipe 8 inches 
 diameter, and 800 feet long, and having two 8-inch standard 
 elbows? 
 
 "ANSWER. No. (Ascertained as follows): 
 
 " Height from delivery valves to surface of water. . . = 20. ft. 
 " Total friction in 800 ft.of 8-in.pipe ( = .53 x8 X2.3I) = 9.79 ft. 
 "Total friction in two 8-in.elbows ( = .53x2.31X2)= 2.45 ft. 
 
 Total 32.24 ft. 
 
 "The sum of these three items is in this case about 32 feet, 
 or 6 feet greater than the 26 feet above mentioned. Therefore, 
 the pump should be lowered 6 feet, or the frictional resistance 
 
320 PUMPING MACHINERY 
 
 in the suction pipe reduced by about 6 feet, by increasing the 
 size of the suction pipe. 
 
 " Lay your suction pipe so that it slopes away from the 
 pump gradually. A suction pipe should have no air pockets 
 in its entire length, but should be so laid that if air be admitted 
 to it, near the intake end, with the pump standing still, the 
 air would rise to the pump or suction air chamber, and not be 
 pocketed in some high part of the suction pipe. A slope of 
 i per cent (i foot drop in 100 feet of length) will be found 
 very satisfactory. 
 
 " Be sure that your suction piping is absolutely tight, for 
 a very small air leak will cause a pump to work improperly. 
 The suction pipe should be tested with about 20 pounds water 
 pressure after it has been laid and before it is covered. If the 
 test shows up a leak, fix it; it is not ' good enough.' 
 
 " Keep the end of your suction pipe well under water. 
 It should never have less than 3 feet above it and 6 or 8 feet 
 would be much batter. 
 
 " When you take suction from a tank into which the returns 
 from hydraulic elevators or presses empty, take care that the 
 returns enter the tank as far away from the suction-pipe opening 
 as possible, for the pump is liable to get air if the returns empty 
 near the suction opening. 
 
 " If two or more pumps draw from the same suction pipe, 
 or if water comes to the pump under a head, a gate valve should 
 be placed on the suction pipe of each pump, to enable you to 
 open up any one pump cylinder for repairs or examination 
 without interfering with the operation of the other pumps. 
 We recommend on larger sizes, when practicable and when 
 not too costly, that each pump have a separate individual 
 suction line entirely independent of the suction line of any 
 other pump. 
 
 " A suction air chamber will be found desirable in all cases, 
 and indispensable in cases where the sum of the three items 
 referred to in a previous paragraph exceeds 10 feet, or when 
 the suction pipe is long. 
 
 " A foot valve is desirable in all cases (except when suction 
 
DESIGN OF PARTS 321 
 
 water comes to the pump under a head) and indispensable 
 when the suction lift exceeds 10 feet. By its use the pump 
 and suction pipe are kept primed when the pump is shut down, 
 and permits of easily priming the pump and suction pipe if 
 purposely emptied, thus enabling the pump to be easily started 
 at any time. 
 
 " In all cases where the water contains sticks, weeds, rags, 
 or other rubbish a strainer should be used on the suction pipe, 
 to prevent them from getting into the pump and clogging 
 valves and passages. If a foot valve is used, a strainer placed 
 outside the foot valve is the best; but if no foot valve is used, 
 a box strainer, placed near the pump, and so designed. that by 
 removing the strainer cover all accumulations can be removed, 
 will be found the most desirable. Keep the strainer clear from 
 accumulation of rubbish. 
 
 " When a foot valve is used, a drain valve should be placed 
 near the surface of the water, to enable the suction pipe to be 
 drained when desired. 
 
 " A relief valve, set to blow at about 20 pounds pressure, 
 should also be placed on the suction pipe near the pump, to 
 prevent delivery pressure, if over 50 pounds, from accumu- 
 lating in the suction chamber of the pump or the suction 
 pipe. This does not cost much and may sometime save you 
 the cost of replacing a broken pump cylinder or foot valve, 
 due to carelessness. 
 
 " In cases where the pump gets its supply directly from 
 driven wells, it will be found most satisfactory to place a large 
 tank on end in some part of the pump house, and to run the 
 pipes from the wells into the side of the tank not far from the 
 bottom. Take the suction for your pump from the bottom 
 of the tank, and lead a small pipe from the top of the tank 
 to an air pump provided for the purpose, or to the condenser 
 of the main pump, if its air pump is large enough. A gauge 
 glass on the side of the tank will show the level of- the water 
 in same, and by opening the valve on the small air pipe a certain 
 amount of the air may be removed from the tank and the 
 pump will then get no air, but pump its full displacement of 
 
322 * PUMPING MACHINERY 
 
 water. The air may be extracted automatically from the 
 tank by means of a float arrangement. 
 
 " A check valve D on the discharge pipe will be found very 
 convenient. 
 
 " A gate valve should always be placed on the discharge 
 pipe outside of the check valve. 
 
 " A priming pipe should always be connected from the 
 discharge pipe, outside of the gate, to the suction pipe, if a 
 foot valve is used. This will enable the pump cylinders and 
 suction piping to be primed, if empty, before starting. If you 
 have no suitable relief valve on the suction pipe, be very careful, 
 in priming with this pipe, that you do not let delivery pressure 
 accumulate in the suction pipe. This will be prevented by 
 having the starting waste valve F open before you open the 
 priming pipe valve E. This should always be open before start- 
 ing your pump (whether you have a foot valve or not), as by this 
 means the pump is enabled to discharge the air from the pump 
 cylinders and suction pipe through this starting valve against 
 a light pressure. As soon as water is discharged through the 
 starting valve, shut it and open your steam throttle valve, 
 and the pump will then discharge through the discharge main, 
 opening the check valve automatically. If you have a foot 
 valve or a gate on your suction pipe, and no relief valve, be 
 careful to open the starting valve at the instant you shut the 
 pump down and leave it open until after you have started 
 again, as by so doing you prevent the possibility of pressure 
 accumulating in the suction. The pet cock in the force chamber 
 of small-size pumps is intended to be used in the same manner 
 as the starting valve above referred to. 
 
 " When shutting the pump down in late fall, winter, or 
 early spring, be sure and open all steam and pump-end 
 drains, and leave them open, otherwise your steam or pump 
 cylinders or other parts may be cracked, due to water freez- 
 ing inside of them, and you may be forced to purchase 
 practically a new pump. A little care will prevent this occur- 
 rence. 
 
 " Do not try to raise hot water, crude petroleum, or any 
 
DESIGN OF PARTS 323 
 
 thick liquids by suction, but wherever possible have the liquid 
 flow to the pump under a suction head. 
 
 " Keep the steam cylinders and valve motion of your pump 
 well lubricated. Oil is cheaper than repair parts. 
 
 " Do not pack the stuffing boxes too tightly, and do not 
 let the packing stay in until it gets hard and cuts the piston 
 rods or plungers. Renew it sufficiently often to keep it soft 
 and pliable. If the pump runs badly, make sure that the 
 pump valves, packed pistons, or plungers, and suction and 
 discharge connections are all right before examining the steam 
 end." 
 
CHAPTER VII 
 
 DYNAMICS OF STEAM END 
 
 IN the design of a pump the variation of pressure has been 
 found together with the size of the pump cylinders, and it is 
 now necessary to find the size of the steam cylinders to operate 
 the water end, the size of the electric motor to drive the pump, 
 or the size of the water wheel which is required. 
 
 The curves constructed in Figs. 197 and 198 can be com- 
 bined to give the indicator card of the head end of the pump. 
 The crank-end card of the water end, if the pump were double 
 
 n acting, would be found in a 
 
 Chan her I ressu ;e_222|ft 
 
 Atmosph 
 
 to eric Line 
 
 |. 
 
 FIG. 278. Pump Card. 
 
 similar manner. From this 
 combined card (Fig. 278), 
 which is the indicator card, 
 the power to be given to the 
 water may be computed. 
 
 Assume that there are 
 three single-acting pumps giv- 
 ing cards similar to Fig. 278 
 connected together, the dis- 
 charge pipes of each cylinder being 18 inches in diameter, and 
 continued separately to the height considered in Chapter V. 
 The area of the card then would give the work done on the 
 water. The mean height of -the card is 235.2 feet or. 102.1 
 pounds per square inch. This includes all friction of the water, 
 and would give the power required if an allowance is made for 
 the mechanical efficiency of the pump. 
 The water indicated horse power is 
 
 I.H.P. - 
 
 3PLAN 
 33,000 ! 
 
 324 
 
DYNAMICS OF STEAM END 325 
 
 where P=mean effective pressure in Ibs. per sq.in.; 
 L= stroke in feet; 
 A =area of piston in sq.in.; 
 N = number of revolutions per minute; 
 IHp= 3Xio2.iXffX 3 i4X6o 
 
 33,OOO 
 
 The mechanical efficiency of the water ends of pumps will 
 vary with the construction and condition of packing. With 
 a plunger pump an efficiency of 97 per cent is not uncommon; 
 and this will be taken as the mechanical efficiency in the problem 
 studied. With inside-packed pumps the friction may increase 
 so that for such pumps 92 per cent will be used. These values 
 will vary, and if the glands or packing rings on pistons are 
 tightened too much these values will be smaller. 
 
 The mechanical efficiency of the complete steam-driven 
 unit, including the power for the air pump of the condenser 
 on large pumps is about 95 per cent, while 90 per cent or even 
 80 per cent should be used for small pumps. With a gear- 
 driven pump the gears will absorb about 5 per cent of the 
 power per pair, so that with a back gear the loss of the gearing 
 amounts to 10 per cent. The efficiency of engines and electric 
 motors may be taken as 90 per cent to 95 per cent, and the 
 efficiency of water turbines as 80 per cent. Using the figures 
 above, the indicated horse power of the engine for driving, the 
 electrical power and the water horse power applied to the 
 wheel will be respectively: 
 
 W.H.P.= 437 ' 2 Q =626 W.H.P. 
 
 .97 x. 90 x. 80 
 
 To carry further the design of the steam end for the purpose 
 of finding the size of fly wheels or other parts it will be assumed 
 that this pump of three cylinders is to be driven by a triple- 
 
326 
 
 PUMPING MACHINERY 
 
 expansion engine, with three cylinders, one of them in tandem 
 with each of the water cylinders. The plungers are single 
 acting, and it will be assumed that they are built solid to give 
 a greater pressure on the downward stroke. 
 
 To find the size of the engine, assume that the steam expands 
 in one cylinder, and that the theoretical card is of the form 
 shown in Fig. 279. This card assumes no clearance and com- 
 pression, the effect of these together being equal to a decrease 
 in area of about 3 per cent. If the initial absolute pressure 
 
 FIG. 279. Indicator Cards with no Clearance. 
 
 be called pi, the pressure at the end of expansion p x , the back 
 pressure p^ and if the curve of expansion be taken as a rectang- 
 ular hyperbola, the mean height of the figure becomes 
 
 Mean height = 
 
 This quantity is multiplied by 0.97 to allow for the effect 
 of clearance and compression. There is another effect for 
 
DYNAMICS OF STEAM END 327 
 
 which allowance must be made. This is the effect of the valve 
 gearing on the indicator card. The corners of the card are 
 rounded, due to slow valve action, and there is some change 
 of pressure between the various cards as shown in Fig. 280. 
 This means that a diagram factor must be applied to give 
 the proportion of the theoretical card which may be actually 
 obtained. The value of this will be taken 95 per cent for D- 
 slide valve engines, and 97 per cent for Corliss engines. 
 
 FIG. 280. Combined Cards. 
 
 The probable mean effective pressure will then be given by 
 M.E.P. ^diagram factor X .97 X \px(i +log ? |- l j -p b \ . 
 
 The size of the low-pressure cylinder will be found by the 
 formula 
 
 33,000 
 
 In this formula the values of P, L and N for the tandem- 
 steam cylinder are known at this point, and A may be found. 
 If the engine is to drive by gears 2LN may be assumed, next 
 
328 PUMPING MACHINERY 
 
 N, after which L and A are computed. In some cases a ratio 
 of L to D is assumed. Having the size of the low-pressure 
 cylinder the size of the other cylinders may be assumed by 
 taking the ratio of cylinder volumes from practice. A better 
 way is to divide the area of Fig. 279 into three equal parts, 
 giving equal work to each cylinder. After this the sizes of the 
 cylinders are found from the figure. 
 
 If the receiver is assumed to be very large the pressure 
 between stages will be constant. 
 
 If p\ is the initial pressure in the intermediate cylinder, 
 and p"\ that in the low-pressure cylinder, the following formula 
 may be used to find these, assuming equal work: 
 
 for low pressure; 
 
 for low and intermediate together. 
 From these: 
 
 log, p"i =i[log^,/>* 2 + 2( -i)] 
 
 For 
 
 />i = 150.3 lb- gauge, 
 px= -6.7lb. gauge, 
 p b = -ii 7 lb. gauge, 
 
 the following pressures result: 
 
 p'i =50.4 absolute; 
 p"i = 14.47 at solute. 
 
 These values have been marked on Fig. 279. From these, 
 the relative dimension of the cylinders may be found, since the 
 lines ab, cd, and ef represent respectively the volumes of the 
 
DYNAMICS OF STEAM END 329 
 
 high-pressure cylinder, of the intermediate-pressure cylinder, 
 and of the low-pressure cylinder, hence 
 
 These ratios are the theoretical ratios of the cylinders if 
 the expansion is complete in each cylinder. If there is free 
 expansion in the different cylinders, the proportionate drops 
 in each are assumed, and then the points b and d are moved 
 
 to b' and d', and the ratios above are changed to 7 and 7. 
 
 With such ratios the volume of the receiver and piping is 
 expressed in terms of the cylinder volumes from practice as 
 are the clearance volumes. So soon as the low-pressure cylinder 
 volume is computed from the horse-power equation, the volumes 
 of the other parts are all known and the combined individual 
 cards may be drawn as in Fig. 280. The receiver and piping 
 between each cylinder is taken as about 250 per cent of the 
 cylinder from which the steam is discharging, and the clearance 
 on the three cylinders, to be 5 per cent for the high-pressure 
 cylinder, 2.8 per cent for the intermediate-pressure cylinder, 
 and 1.5 per cent for the low-pressure cylinder. These values 
 are given as a guide so that these quantities may be computed, 
 but a designer after making a number of designs would use 
 the tables of proportions which he had computed from previous 
 experience. 
 
 The value of a reheater is questioned, a matter which will 
 be taken up later. 
 
 With the values above, the data for the problem considered 
 give the following results: 
 
 M.E.P. = .96 X -9o[8(i +log e ip) -3]; 
 ,= 25.2 Ibs. per sq. in.; 
 
 I.H.P.=46o = 
 
 33,000 
 
330 PUMPING MACHINERY 
 
 A =2020 sq.in.; 
 Z)=5i in.; 
 2LN =2 Xfi X6o = 300 ft. per. min., which is not too high; 
 
 8 
 
 F,. p . 14.47 1.81' 
 
 F h . p i.8i i 
 K lp . 6.3 -3.48' 
 
 The strokes are all the same, hence 
 - 1 13.48:6.3, 
 
 Since 
 
 A =51 in.; 
 DA = 20.33 in- 
 A- =37.8 in. 
 
 The first approximation of cylinder sizes will therefore be 
 
 20X30 
 37X30 
 51X30 
 
 The clearance volumes are 
 
 470 cu.in. 
 
 880 " 
 890 " 
 
 The receiver volumes are 
 
 23,530 cu.in. 
 81,884 " 
 
 From these volumes and pressures the theoretical indicator 
 cards can be drawn as in Fig. 280. The complete expansion of 
 steam to the receiver pressure is found in most pumping engines, 
 and for this reason these cards have been constructed in this 
 manner. The cranks are assumed at 120, and the pressures 
 and volumes of admission to the various cylinders have been 
 
DYNAMICS OF STEAM END 331 
 
 constructed graphically. The receiver volumes and clearance 
 volumes have been used in constructing curves where possible; 
 where not possible, the pressures have been computed. As an 
 example, the point to which the letter a (Fig. 280) refers, has 
 been computed by assuming the product of a final p and v 
 as equal to the sum of the original products, pv, before mixing, 
 
 Pa(v ah +V aI +V R )=p b (v bh +V R )+p c V cI . 
 
 The pressures in the receivers have been made the terminal 
 pressure of expansion, and the exhaust has been arranged to 
 give this result. 
 
 After the cards have been drawn they have been used to 
 construct cards 4 inches long, and to various scales. The 
 cards drawn are seen to be of different size from the original 
 card with no clearance; the H.P. card has been reduced, the 
 L.P. card increased, and there is little change in the I.P. card. 
 To make certain that the steam cylinders are of sufficient size 
 the I. H.P. of the total engine should be found in terms of the 
 L.P. cylinder displacement, assuming that the volumes of the 
 cylinders are of the ratios given originally. Allowance is also 
 made for the piston rod by considering it to be 2j per cent 
 of the piston area. After finding the M.E.P. from each theoretic 
 card in Fig. 280 or Fig. 281, the following formula may be used: 
 
 J.H.P. = - (pt+pi M* +P,)(Disp) low N. 
 
 33,000 Xi2\ 6.3 6.3 V v 
 
 From this a new size of L.P. may be computed. 
 
 The M.E.P. from the indicator cards give 47 pounds for 
 the high-pressure cylinder, 15.9 pounds for the intermediate, 
 and 8.2 pounds for low-pressure cylinder. Substituting these, 
 the equation becomes 
 
 i.Q75X6o 747 3.48 \ 
 
 460= v/3 - ( |^- +~- 15.0 +8.2 )D; 
 33,000 X 12 V6.3 6.3 / 
 
 Z)= 56,220 cu.in.; 
 
 56,220 
 Area = -^-= 1874 sq.m.; 
 
 Diameter =49 in. 
 
332 PUMPING MACHINERY 
 
 Hence 
 
 Diameter intermediate =36.4 in.; 
 Diameter high = 19. 5 in. 
 
 Use 50, 37 and 20. It would be well at this time to rede- 
 sign the cylinder and cards of Fig. 280 so as to make the work 
 on each cylinder the same. 
 
 The areas of these cards together' will be greater than the 
 area of the water cards by the amount of friction. The fric- 
 tion of the steam and water pistons and the friction of the 
 stuffing box will be considered to be three-fourths of the total 
 friction, and the remaining part will be taken uniformly by the 
 .bearings of the shaft. 
 
 This gives 22.8 H.P. (460 437.2) as the total to be absorbed, 
 of which 17.1 H.P. is used in the packing and 5.7 in the journals. 
 
 17.1 H.P. means - = 9405 ft.lbs. per revolution. 
 
 * DO 
 
 The mean pressure from this over 2j-foot stroke would be 
 
 =627 pounds' for each unit above and the same amount 
 5X3 
 
 below the zero of pressure, as this force operates to require 
 work and always to hinder motion. This quantity is to be 
 divided by the areas of the pistons in order to reduce this to 
 the same basis as that used in the diagrams. 
 
 The weights of the various parts will be assumed to be 
 those given in the table below. The weights would ordinarily 
 be determined from the design of the piston, piston rod, cross- 
 head, and connecting rod, which would be. made before pro- 
 ceeding with this part of the work. For their, design, see 
 Chapter VIII. 
 
 Weight of H.P. piston 1 225 Ibs. 
 
 Weight of I. P. piston 900 " 
 
 Weight of L.P. piston 1400 " 
 
 Weight of each piston rod 125 " 
 
 Weight of each cross-head 800 " 
 
 Weight of each connecting rod 350 " 
 
 Weight of 4 rods to plunger from cross-head. 450 " 
 Weight of plunger 5000 " 
 
DYNAMICS OF STEAM END 333 
 
 This gives as the sum of the weights of the reciprocating 
 parts of each cylinder the following: 
 
 High^pressure unit 6900 Ibs. 
 
 Intermediate-pressure unit 7575 " 
 
 Low-pressure unit 8075 " 
 
 Reducing these to pressure per square inch of specific steam 
 cylinder these values become: 
 
 High-pressure piston 22 Ibs. per sq.in. 
 Intermediate-pressure piston, 7 Ibs. per sq.in.: 
 Low-pressure piston, 4.1 Ibs. per sq.in. 
 Friction H.P. piston = 2.0 Ibs. per sqjn. 
 Friction I.P. 'piston =0.575 Ib. per sq.in. 
 Friction L.P. piston =0.318 Ib. per sq.in. 
 
 Taking the indicator cards from the separate steam and 
 water cylinders, as shown in Fig. 281, it is necessary to add 
 22 pounds per square inch for the high-pressure unit on the 
 down stroke, to get the effective pressure, and subtract the 
 same on the up stroke. Moreover, the back steam pressure on 
 the crank end must be subtracted from the head end on the 
 down stroke in order to get the net pressure from the steam. 
 As these piston areas are not the same on each end of a piston, 
 to reduce them to the same scale the crank-end pressures 
 
 A c 
 
 should be multiplied by -p, where area A c is the area of the 
 
 A-h 
 
 crank end of the cylinder and the area Ah is the area of the 
 head end. In the case of a 2o-inch cylinder with a 3-inch rod, 
 this ratio is o . 98, and hence it may be considered as unity. 
 On the other pistons it is still closer to unity. 
 
 The reciprocating parts require accelerating and therefore 
 a force equal to 
 
 w w / I \ 
 
 -a=-a> 2 r(cos 6+- cos 20) 
 g g \ n 9 
 
 is required on each square inch of piston area, if w represents 
 the weight of the parts per square inch of piston area. The 
 values of a for the different positions may be taken from Chapter 
 
PVMPINC MACHINES? 
 
 _Atmpspheric Line 
 
 . , 
 
 (270 1240 J510 
 
 Atmospheric Line N 
 
 N Atmosphcric Line 
 
 B. Intermediate Unit. 
 
 FIG. 281. Combined Cards. 
 
DYNAMICS OF STEAM END 
 
 335 
 
 V, page 192, and plotted on stroke position after multiplying 
 by -aj 2 r for each cylinder. These curves are plotted in Fig. 
 
 o 
 
 281. At the beginning of each stroke the inertia force must be 
 
 ^Atmospheric Line 
 
 C. Low-pressure Unit. 
 FIG. 281. Combined Cards. 
 
 subtracted from the net steam force, while at the end of the 
 stroke it is added. The curve repeats itself with the same 
 numerical values, but of opposite signs for 210 as for 150, 
 
336 PUMPING MACHINERY 
 
 for 240 as for 120, etc. Hence the curve need not be redrawn, 
 but the same curve may represent each stroke, provided the 
 sign of the quantities be changed for the two strokes. 
 
 Having now these various forces which act on the system, 
 it is a simple matter to find out how much of it remains unbal- 
 anced and must therefore pass from this system into the others, 
 or into the fly wheel by way of the connecting rod and shaft. 
 
 Take any point A on ' the high-pressure set of cards and 
 consider this to be on the up stroke (head to crank), when the 
 fly wheel and crank are assumed to be above the steam and 
 water cylinders which are adjacent. The steam pressure on 
 the upper side is aA, on the lower side bA. The weight is cA^ 
 the inertia is dA, the friction of packing is Af and the water 
 pressure is hA below the atmosphere. The water pressure is 
 measured from the atmosphere because the air pressure is 
 pressing down on an area equal to the area of the plunger minus 
 the area of the steam piston rod (a small quantity compared 
 with the piston or plunger area) and this balances practically 
 14.7 pounds of the water pressure. The pressure on the water 
 cylinder has been reduced to a figure which represents pressure 
 per square inch of steam piston area, by redrawing the diagram 
 
 A 
 
 of Fig. 278, using for pressures P-j , and using the 
 
 ^ steam end 
 
 same scale as used for the other diagrams. The net pressure 
 remaining in the system per square inch of steam cylinder area 
 is 
 
 P=aA -bA -cA -dA -Af-hA = CB. 
 
 On the down stroke at the point A the net pressure becomes 
 P l = a'A -b'A + cA +dA -Af-h'A = CB'. 
 
 (Note that arrows showing directions -of piston movement 
 aid in this construction.) The figure is drawn by joining points 
 for different positions. In the problem assumed, the work 
 of each steam cylinder was taken about the same and that 
 of each water cylinder was the same, hence the area of this 
 resultant figure which represents the net work done by the 
 
DYNAMICS OF STEAM END 
 
 337 
 
 system outside of itself, must just equal its share of the work 
 done in journal friction. 
 
 Since the probable cards are not arranged to give absolutely 
 the same work on each cylinder, this is not quite true. The 
 diagrams for the different cylinders have been made of the 
 same length, but to make them of proper area, the high pressure 
 was drawn with a scale of 50 pounds per square inch to the inch 
 
 of height, the intermediate with 
 
 3.48 
 
 or 14.38 pounds per 
 50 
 
 square inch to the inch of height and the. low pressure with ^ 
 or 7 . 95 pounds per square inch to the inch of height. 
 
 FIG. 282. Tangential Effort Construction. 
 
 The cards, Fig. 281, have been constructed by laying off 
 positive pressures above the base line EF on the crank head 
 stroke, while for the head-crank stroke, positive pressures are 
 laid off below. This gives a figure, the area of which repiesents 
 the work or energy transmitted from this system to the others 
 through the connecting rod and shaft. 
 
 Having now the force per square inch of piston which is 
 transmitted to the crank by the connecting rod, the next oper- 
 ation is to find the value of the force produced by this in a 
 direction tangent to the crank arm. In Fig. 282 the crank 
 and connecting rod are shown by lines in a given position 
 
338 
 
 PUMPING MACHINERY 
 
 of the crank. The end A of the rod is moving horizontally and 
 therefore it is moving instantaneously about some point on 
 a perpendicular to the path, as on AC for the position shown. 
 The end B is moving in a "tangent to the crank circle and there- 
 fore about some point in line with the crank radius OBC. Now 
 
 30 60 QQ^ igft i*m^>-re/T '210 
 H.P. Effort. 
 
 240 \ 270 800 
 
 330 
 
 30 
 120 H. 
 
 60 
 180 H. 
 
 90 \120 
 
 180 210 240 
 300 H. 360 H. 
 
 I.P. Effort. 
 
 120 H. 
 
 FIG. 283. Tangential Efforts. 
 
 one end of the rod is moving about a point in one of the lines 
 and the other end about a point on another line and the 
 only point which satisfies both of these conditions is the 
 point C. That is, for the instant considered the connecting 
 rod is turning about the point C as much as if it were 
 rigidly connected to a pivot at this point. If then it turns 
 about this point, the force perpendicular to OB produced 
 
DYNAMICS OF STEAM END 339 
 
 by the piston force P, perpendicular to AC, is given by the 
 equality 
 
 TCB=PAC. 
 
 The effect of inertia of the rod is such that this equation 
 is not strictly true, but Jacobus has shown in a paper (A.S.M.E., 
 XI, pp. 492 and 1116), that, for the design of the fly wheel, 
 the approximate method of considering the connecting rod 
 as having the motion of the piston and including it with the 
 other reciprocating parts and neglecting the real inertia effect 
 at this point, is sufficiently accurate. 
 
 To find the value of T graphically, lay off op equal to the 
 pressure from the piston rod, and draw pt parallel to the connect- 
 ing rod. 
 
 The triangles otp and CAB are similar, hence 
 
 ' Ot ~0t' 
 = OtBC] 
 
 PAC=TBC~ 
 
 hence Ot = T. 
 
 For the piston pressure P at the cross-head the tangential 
 effort T has been found. This is done for several points in 
 Fig. 283 and plotted on a line representing the travel of the 
 crank. The other two cards are drawn as shown. 
 
 The area scale of these figures, Figs. 281 and 283, is equal 
 to the product of the two linear scales. The various scales are 
 given below: 
 
 Card. Length Scale. Height Scale. Area Scale. 
 
 H.P. f ft. per. in. 50 Ibs. per sq.in. H.P. 31.25 ft.-lbs. . per sq.in. 
 
 piston per in. H.P. area per sq.in. 
 
 I. P. f ft. per in. 14.38 Ibs. per sq.in. of I. P. 8.99 ft.-lbs. per sq.in of 
 
 piston per in. I. P. area per sq.in. 
 
 L.P. | ft. per in. 7.95 lbs.persq.in.of L.P. 4.96 ft.-lbs. per sq.in. of 
 
 piston per sq.in. L.P. area per sq.in. 
 
 If now total work is desired, the areas of the diagrams will 
 be multiplied by the area scale and the area of each piston. 
 
340 PUMPING MACHINERY 
 
 To refer these, however, to the L. P. area, the pressure scale 
 
 A 
 
 for the H.P. cylinder will be multiplied by ~j ? - and the 
 
 **Lp 
 
 A- 
 I. P. scale by --. When this is done it is found that all area 
 
 ^I.P. 
 
 scales are the same. 
 
 Since then the tangential diagrams are of the same scale, 
 they may be combined by addition. The customary arrange- 
 ment of cranks as seen in Fig. 312 is to have them at 120 apart. 
 Hence the three positions of 6=0 will be placed at 120 apart 
 as in Fig. 283 and the resultant line found. The area of this 
 resultant curve will be the energy put into the journal friction 
 and if the height .of the tangential effort required for this be 
 laid off above the zero, it will be found that this will be the 
 mean height of the curve arid the amount of area above the 
 line will be equal to the amount of area beneath the line. 
 
 These areas represent work or energy because the height 
 represents tangential effort or force, while the abscissae represent 
 crank movement or motion in the direction of the force. To 
 reduce these areas to actual total force they are multiplied 
 by the area scale and the area of the low-pressure piston. 
 
 Calling the areas above the mean line excess areas, lf 
 2, ^3, etc., and those below the line deficiency areas, D^ D 2 , 
 D 3 , etc., the variation of energy in the fly wheel may be found 
 as follows: 
 
 At i the energy in the fly wheel may be called F; then by 
 the time the crank has rotated to 2, the three systems have 
 developed an excess of E\ units of energy over the amount 
 required by the pump. This energy must be absorbed by the fly 
 wheel in an increase of speed. By the time 3 is reached, D\ units 
 have been abstracted, and so on for the various points; the 
 energy at these is therefore a^ follows; 
 
DYNAMICS OF STEAM END 341 
 
 Point. Energy. 
 
 1 F 
 
 2 F+P^i 
 
 3 F+E l -D l 
 
 i F+E 1 -D l +E 2 -D 2 =F 
 
 If now the difference between the greatest and the least 
 of these is found, that quantity, when multiplied by the area 
 scale and the area of the low-pressure cylinder will give the 
 amount of energy to be stored up in the fly wheel to change 
 it from its lowest rate of speed during one revolution to its 
 highest rate during that revolution. Calling tl>is area difference 
 AE and the moment of inertia of the fly wheel WR 2 , where R 
 is radius of gyration in feet and W the weight of the wheel and 
 N' and N" the highest and lowest rates of speed in R. P. M. in 
 one turn of the wheel, the following results: 
 
 AE Xarea scale Xarea L P piston 
 &* 
 
 now N'z - N" 2 = (N' + N" ) (N f - N" ) = 2NAN 
 
 where N = mean speed = R. P. M. 
 
 AN = variation in speed. . 
 
 The value of AN depends on the kind of machine considered. 
 
 N N 
 
 For slow pumps may be used while would be better for 
 
 high speed pumps. (In the case of spinning mill engines and 
 
 N 
 
 electric light engines the value is used.) 
 
 100 
 
 Assuming AN and the R.P.M. for a given pump, the 
 quantity WR 2 is given by the formula. If R is known for a 
 given shape, the weight W could be found. 
 
 In the case of the ordinary flywheel, the rim is the most 
 
342 PUMPING MACHINERY 
 
 important portion, and in many cases the rim is designed to 
 give the necessary WR 2 . In the case of the flywheel or rim, 
 of width b, outer radius r\\ inner rz\ with hub, of width b l t 
 outer radius r 3 , inner radius r and with n arms, the following 
 results: 
 
 WR 2 for rim = | * {w?xrbdr)r* 
 
 Jn 
 
 nbw 
 -- 
 
 2 
 
 WP? for spoke, uniform elliptical section, 
 
 r 2 dr 
 
 4 
 
 4 3 
 W- 
 
 For whole wheel, 
 
 In this equation the last two terms may be found by taking 
 data from practice, and then by assuming r\ of the first term, 
 the only unknown would be the term b of the weight W. If 
 desired to assume b this may be done and then the unknown 
 portion of the first term would be r^ r 2 4 , in which r 2 is assumed, 
 fi 4 may be found. 
 
DYNAMICS OF STEAM END 343 
 
 If two flywheels are to be used, \AE is used for each wheel. 
 
 The first tangential diagram of Fig. 283 is the amount of 
 twist which is transmitted from the first crank to the first fly- 
 wheel, while the third tangential diagram is the amount which 
 is transmitted from the third crank to the second flywheel. 
 The middle diagram is the amount which is transmitted from 
 the center crank at the middle to the two wheels. This tan- 
 gential effort, when multiplied by the crank radius, gives the 
 twisting moment for which the shaft must be designed. 
 
 The net piston force from Fig. 281 gives the force which is 
 applied to the crank shaft to produce bending. The amount 
 in the direction of the piston is increased by. the inclination of 
 the connecting rod. The maximum increase amounts to only 
 i|- per cent, and so this increase may be neglected in design, 
 and the height from Fig. 281, when multiplied by the pressure 
 scale and the area of the piston, will give the force which causes 
 bending on the shaft. However this may be, after the pump 
 is running, if anything should happen to the machine in one 
 of the inner pump barrels, the plunger might jam and then the 
 whole steam pressure from any one piston would reach the crank 
 pin and hence the pin should be designed to support this. 
 
CHAPTER VIII 
 STEAM END DETAILS 
 
 THE steam end of the pump has been illustrated in chapters 
 III and IV quite fully, and it remains to examine peculiarities 
 of design. The action of the steam end of simplex pumps has 
 been fully studied and an explanation was made of the action 
 of the valve of the duplex pump. 
 
 It was mentioned that in order to operate duplex pumps 
 properly the steam valve should have a certain amount of play on 
 the stem. This is accomplished in several ways. In Fig. 284 
 
 FIG. 284. Steam Valve of Duplex Pump. 
 
 the method of using a central nut is shown. The valve is made 
 without any lap, that is, the valve just reaches from the outer 
 edge of one steam port A to the outer edge of the other, which 
 would mean that as soon as steam was cut off from one end 
 the other would open for steam; reverse the pump and so start 
 the other pump, which in turn would reverse the front pump. 
 By having the valve rod slide through slots in the projections 
 on the back of the valve and by using the split nut of proper 
 length to drive the valve, the action will be as follows: 
 Suppose the valve is moved to the right by the motion of the 
 other pump, the piston controlled by this valve will move to 
 the right and it will start to move the valve rod on the other 
 
 344 
 
STEAM END DETAILS 
 
 345 
 
 pump to one side; however, the valve will not start until the 
 piston, which is moving, reaches a point near the end of its 
 stroke, whereupon the other piston starts and actuates the 
 valve rod of Fig. 284 to move to the left. The valve does not 
 operate until the rod has traveled the distance equal to the 
 play between the nuts and the projection, at which time the 
 valve will be moved to the left, closing the left-hand steam port 
 and opening the right, thus starting the piston on its return. 
 The time taken for one side to make the major part of its stroke 
 is the time taken for the other side to come to rest at the end 
 of its stroke and reverse. In this way there is a very steady 
 discharge, as was explained earlier, since one piston is moving 
 at full speed while the other is reversing. There is a small period 
 of rest at the end of the stroke when the cylinder is full of steam, 
 while the other piston is moving to throw over the valve. The 
 driving piece C should be made in two parts so that they may 
 be jammed to hold them at one place on the rod. 
 
 Another method of accomplishing this is by the use of out- 
 side striking pieces, as shown 
 in Fig. 285. This arrangement 
 is better, as the valve may be 
 adjusted at either end. When 
 this adjustment must be made 
 often, it is done outside of the 
 cylinder. One method is illus- 
 
 FIG. 285. Steam Valve. 
 
 trated in Fig. 286. Here the valve stem A is provided with 
 
 FIG. 286. Valve Rod Yoke. 
 
 a box or yoke F on the end. This is driven through the 
 cross head C by the pin B attached to the lever, connected to 
 the other side of the pump. The amount of play is regulated 
 
346 
 
 PUMPING MACHINERY 
 
 by the set screws DD which are fixed in place by the jam 
 nut. 
 
 In Fig. 287 another device is shown to accomplish the same 
 result. The reverse lever A is driven from the piston rod B by 
 a link. The lever is pivoted behind the valve rod. It moves 
 
 FIG. 287. Driving Mechanism for Valves. 
 
 the sleeve C by means of a link and this sleeve strikes the jam 
 nuts DD and moves the valve rod at the proper time. 
 
 These last two devices are used on large and small direct- 
 acting pumps. They are used at times on the rotating steam 
 valves which are the equivalent of D valves. 
 
 FIG. 288. Triple Expansion Cylinders. 
 
 The cylinders of the single expansion duplex pumps are 
 quite simple. The figures of chapters III and IV show their 
 construction. The action of the five ports, the form of valve 
 and piston, the arrangement of flanges for the cylinder heads, 
 the valve chest covers, and the method of carrying the cylinders 
 may be seen by reference to these figures. 
 
 The arrangement of the cylinders of a triple expansion direct- 
 acting pump, Fig. 288, will serve to illustrate the methods of 
 
STEAM END DETAILS 
 
 347 
 
 building these. The cylinders are cast usually with single walls, 
 and the cylinder ends are flanged to receive the cylinder head. 
 Heads may be cast solid with the walls. This latter method 
 serves to cut down the number of joints to be packed and does 
 not materially increase the cost of construction. The heads of 
 the cylinders are dished or cored to receive the nuts at the end 
 of the piston. 
 
 The arrangement of piston rods shown in the figure is designed 
 to reduce the number of inside packing boxes and make it 
 possible to examine each piston without disturbing any cylinder. 
 
 When the cylinders are jacketed, several methods are 
 
 "a 
 
 
 u 
 
 (i 
 
 
 8 
 
 FIG. 289. Leavitt Cylinder and Jacket. 
 
 available. The jacket may be made in the cylinder casting 
 or it may be made by the use of a liner. 
 
 When the jacket is made in one piece with the cylinder 
 barrel, a casting difficult to produce results, the outer shell 
 is apt to cool first and when the inner part cools, strains are 
 set up which crack the casting. To prevent this Mr. E. D. 
 Leavitt, Jr., devised the scheme of casting the outer shell of 
 the jacket with a division in the outer wall, Fig. 289. The 
 opening thus made was closed by a copper ring A, with one 
 corrugation. This ring was attached to the edges of the opening, 
 by tap bolts with a sharp ring on the lower face of the head, 
 which was driven into the copper and made a steam-tight joint. 
 The edge of the copper was caulked against the cast iron. 
 
 The Snow Steam Pump Company makes jackets at times 
 with a cast corrugation around the center of the shell to 
 allow for this expansion and contraction. This .cylinder, 
 
STEAM END DETAILS 
 
 349 
 
 Fig. 290, is jacketed on the heads as well as on the barrel. 
 The steam enters the barrel jacket at A and the condensation 
 is removed at B. The condensation from the heads is removed 
 by the drains C, which in some cases serve as admission pipes 
 also, the system being the equivalent to a single pipe system 
 of heating. The air valves at DD relieve the jacket of air. 
 
 FIG. 291. Valve Gear with Cut-off Controlled by Speed Governor. 
 
 The method of attaching the heads by stud bolts is clear. The 
 stuffing box is shown at K. 
 
 The half cross sections through the valves and center, Fig. 
 290, show how the steam entering through EE passes the 
 Corliss steam valves to the exhaust passages FF, which are 
 cross connected to the manifold G. 
 
 The ribs at H and / are to stiffen the casting where passages 
 are made. 
 
 The steam and exhaust valves and gear are shown in this 
 figure and in Figs. 291, 292. In these the knock-off cam L 
 and arm M are seen. In Fig. 291 the position of L is fixed by 
 
350 
 
 PUMPING MACHINERY 
 
 the position of the governor through the rod N from the governor, 
 and depends on the speed, while in Fig. 292 the position is 
 controlled by the pressure in the force main from the pump. 
 
 The steam pipe connects the inlets E leading to the valve 
 chests, while the exhaust pipe G connects the exhaust outlets. 
 The dash pots QQ are carried by the exhaust piping. 
 
 The jacket is sometimes constructed by introducing a 
 liner within the cylinder, Fig. 293. In such design the liner 
 
 FIG. 292. Valve Gear with Cut-off Controlled by Pressure Regulator. 
 
 may be made of a hard iron, while the main casting is of a 
 soft cast iron, which is less liable to crack. The liner is fastened 
 at its lower end by bolts, while at its upper end a copper ring 
 is used, fastened by tap bolts with sharp circular lips on the 
 lower faces. The jacket is supported by rings on its outer 
 surface in contact with rings projecting from the inner surface 
 of the shell. The surfaces of the projections are the only portions 
 of the inner bore of the cylinder or the outer surface of the 
 liner which require turning. Moreover, by making these of 
 
STEAM END DETAILS 
 
 351 
 
 decreasing diameter toward the closed end of the cylinder 
 and by making each surface slightly tapered the liner may be 
 introduced with little or no forcing and yet will be held 
 tight. 
 
 The cylinder shown in Fig. 291 is intended for a horizontal 
 
 FIG. 293. Independent Liner. 
 
 FIG. 294. Head Valves. 
 
 engine, but the same construction would be used if for a vertical 
 one. To cut down the clearance by eliminating the long passages 
 extending from the straight valve face to the circular bore of 
 the cylinder, engines have been built with the valves in the 
 heads. Fig. 294 shows such an arrangement for Corliss valves. 
 Many companies restrict this head valve to the low-pressure 
 cylinders only. The system has the great advantage of simplify- 
 
352 
 
 PUMPING MACHINERY 
 
 ing the cylinder casting, but for small high-pressure cylinders 
 this is not so important. 
 
 The simple form of cored piston, Fig. 295, with two piston 
 rings is a very common type. It is attached to the tapered 
 end of the piston rod by a nut. A collar is used to keep the 
 rod in position and support part of the thrust. 
 
 The half rings shown beside the piston illustrate the two 
 forms of rings used with small pistons, one of constant cross 
 section and the other with variable cross section. 
 
 Fig. 296 is the type used for large cylinders. The main 
 
 m 
 
 FIG. 295. Steam Piston. 
 
 portion A is attached to the piston rod as shown, and from 
 the ring B a number of bolts radiate, supporting a bull ring C. 
 This bull ring contains a groove in which is placed a ring made 
 of a series of overlapping sections. These are held out against 
 the cylinder bore by helical springs in the cavities D. The 
 follower plate E is held in position by the tap bolts FF. The 
 ring B is supported by a series of webs from the central hub of 
 the piston. 
 
 Fig. 297 illustrates another one of these pistons in which 
 the sections are held out by carriage springs and the bull ring 
 has two additional grooves which are filled with a bearing 
 compound, so that the weight of the piston may be taken on 
 
STEAM END DETAILS 
 
 353 
 
 a substance which will not wear away the bore, but all wear will 
 occur in the ring which may be turned as soon as the cast iron 
 of the bull ring comes in contact with the bore. In this way 
 
 these bearing rings will last for some time, as it will take eight 
 or ten movements to use up the complete circumference. 
 In Fig. 298 the bull ring is provided with two sectional 
 rings. 
 
354 
 
 PUMPING MACHINERY 
 
 These figures with those of the previous chapter illustrate 
 the method of attaching pistons. 
 
 The simple stuffing boxes have been discussed under the 
 
 FIG. 297. Piston with Sectional Ring. 
 
 head of water-end details. The metallic packing is quite com- 
 mon and as an example of one form Fig. 299 is given. In this 
 the rings A are forced into the conical cups BE by the springs 
 
 FIG. 298. Piston. 
 
 C through the ring D. The cup B has a ground steam-tight joint 
 against the ball-ended ring E. The ball E fits the socket F. 
 The socket F can not be moved, the ball and socket allow for 
 the alignment of the rod and the ground joint between B and 
 
STEAM END DETAILS 
 
 355 
 
 E allows the rod to move across the cylinder when wear of 
 the piston or cross head makes this necessary. This is possible 
 
 FIG. 299. Metallic Packing. 
 
 since the rings or cups E and F are all of a larger inside diameter 
 than that of the rod. 
 
 DESIGN OF CYLINDER 
 
 The thickness of the cylinder is given by the same formula 
 as suggested in chapter VI. (See Figs. 290,. 293 and 294.) 
 
 5000 
 
 t\ v =^i iv =o . 6 thickness of single plate cover. 
 
 Thickness of single plate cover = o. 55^ \ , 
 
 / vi =i" to i" for liner jackets. 
 = i" to 3" for cast jackets. 
 
 The valve chest walls and covers have their thicknesses 
 determined by the cylinder thickness formula or the flat sur- 
 face formula. 
 
 /v =0.35^ \- (or 0.55^ \- for round plates), 
 s \ s /. 
 
356 PUMPING MACHINERY 
 
 Flanges here are made i . 2$t vii in thickness. 
 
 The diameter of the bolt should vary with the size of cylinders. 
 |" bolts are the smallest to be used and these may be increased 
 to ij" on very large cylinders, f-" and i" bolts are common 
 on cylinders of 20". After the diameter is assumed the number 
 of bolts is found by the formula 
 
 # = area at root of thread of bolt in sq.in. from table of 
 
 standard threads. 
 S t = allowable stress in pounds per sq.in. 
 
 =4000 Ibs. per sq.in. to allow for straining. 
 n = number of bolts. 
 
 p = maximum steam pressure in Ibs. per sq.in. 
 d = diameter of cylinder in inches. 
 Flange width 3^ at least. 
 
 Pitch of bolts < 40 \- 
 
 STEAM PASSAGES 
 The area of steam passages is given by 
 
 2LNA 
 
 6000 ' 
 
 while the exhaust passages are given by 
 
 2LAN 
 
 a = -- . 
 4500 
 
 In these a is the area of cross section of tne passage in square 
 inches; L, the stroke in feet; N, the revolutions per minute, and 
 A y the area of the piston in square inches. 
 
 The length of the port is made about o . 8 diameter of cylinder, 
 
 then, width = 5-3. The value 6000 for the velocity of steam 
 o.oa 
 
 should only be used to find the steam passages. The port- 
 opening is found by using this value when cut-off is at 0.7 
 
STEAM END DETAILS 357 
 
 stroke. For 0.25 stroke, the port-opening is found by using 
 18,000 for the velocity. For cut-offs between these use a pro- 
 portional amount. 
 
 CLEARANCE 
 
 The clearance distance is used to allow for wear in the connect- 
 ing rod and bearings, unevenness in the castings of the piston 
 and head, and errors in alignment. The clearance distance 
 is made from ^" to ^" in different sizes of engines, the first 
 being used for strokes of 15" and under. The latter for strokes 
 to 6 feet. 
 
 The clearance volume is given by 
 
 LA 
 
 
 
 In these LI and LI are the length of the steam and exhaust 
 passages. 
 
 The clearance has an important effect, as has been shown 
 in many engine tests, so that it is desired to keep the value 
 of this as small as possible. The following table will give some 
 idea of the variation of this quantity. 
 
 QiYY-iTVIa TT-nrrlnck' Clearance Volume in Percent- 
 
 Simple Engine . age of Cylinder Displacement. 
 
 D-Slide valve 6 to 15 
 
 4 valve engines 2 to 4 
 
 Corliss valves 2 to 4 
 
 Piston valves 7 to 15 
 
 Compound: H.P. L.P. 
 
 Slide valve 7 to 15 5 to 10 
 
 Corliss valves in side 2 to 4 1.3 to 3 
 
 Triple: H.P. I.P. L.P. 
 
 Slide valves 10-15 5-10 5-7 
 
 Corliss valves 1.5-1.75 1.25-2.0 1.5-2.5 
 
 Corliss valves in head. . 4 i ^-i 
 
358 PUMPING MACHINERY 
 
 PISTON PROPORTIONS 
 
 (See Figs. 295 and 296.) 
 
 The design of the piston is empirical and the following 
 proportions are recommended by various authors: 
 
 Flat disc i= -- - for cast iron. 
 
 100 ... 
 
 f 
 
 for steel. 
 
 150 
 
 This value will be used as the unit for all pistons; and dimen- 
 sions on the figures are in terms of this: 
 
 B =0.40 and for large cylinders ^/LD (inches). 
 Thickness of faces ='=0.5^ 
 Thickness of face when follower ring is used =o.& 
 Thickness of center boss -/" = 0.5* +*'". 
 Thickness of ribs = *'" = 0.3* 
 Distance from edge to ring, b' =Q.$t. 
 Thickness of junk ring or follower plate / iv =o.5^. 
 Diameter of junk ring bolts ^' = 0.25 t. 
 Number of webs =o.iD + 2. 
 Thickness of bull ring at edge t v = o.6t. 
 Thickness of bull ring under packing ring f fi = .$t. 
 Thickness of spring packing rings t vii =o.o^D. 
 Width of spring packing rings &" = " to f". 
 Thickness of solid piston wall beneath ring / Vlii =t. 
 Diameter of piston ring = 1.0157). 
 
 The basis for the proportions used for the piston rings above 
 as given by Unwin in his "Machine Design," is as follows: 
 
 Assume the piston ring, Fig. 295, to be loaded with p' pounds 
 pressure per square inch and to be changed by this from a 
 radius R at any point to the radius r of the cylinder. If the 
 width of the ring is 6" the resultant pressure causing bending 
 at the angle of 6 from the end after fitting into the cylinder is 
 
 = 2r sn 
 
STEAM END DETAILS 359 
 
 and its moment is 
 
 = \2rsin ^ 
 
 SI 
 
 This bending moment is equal to -~ or 
 
 * 
 
 . 
 
 6 2 
 
 The change in curvature also depends on this moment or 
 
 I__M 
 
 r~R~EI' 
 
 This last expression is used to determine the variation in 
 curvature when I is constant, or the variation of / if the curvature 
 before springing into the cylinder is constant. The first equation 
 is used to determine the thickness at the point in the ring half 
 way from the ends. In this case # = 180 a ad 
 
 . 
 6 
 
 Using 6000 for 5 and 2 pounds for p' the following results: 
 
 6 
 
 . 
 oooo 
 
 If now the section is made uniform, as on the left of Fig. 
 295, the value of R will have to be different at various points. 
 
 
 
 2-b b r 2 sin 2 
 i 2 
 
 R~^~E Xo.ooooi8b"~r*' 
 
360 PUMPING MACHINERY 
 
 for />' = 2lbs. and = 15,000,000. 
 
 6 
 
 Ar 2 sin 2 
 I I 2 
 
 2 Ar 2 sin 2 270 4 sin 2 
 
 For different values of 6 the values of R are given below: 
 
 = o 30 60 90 120 150 180 
 
 # = 1.000?' 1.003^ 1.004;' 1.007^ i.our 1.014? i.oi5r. 
 
 The curve of the original curvatures is drawn in Fig. 295; 
 the radius for 180 being used for a length .of arc equal to 15 
 of the cylinder circumference on each side of the 180 position. 
 This gives the point a, and on the radius to this point a new 
 center is taken equal to that for # = 150. This arc is made 
 equal to 30 of the" cylinder circumference, giving the point b. 
 On the radius to -this point a new center is taken for the curve 
 to c, then in a similar manner d, e, f and g are found. When 
 this ring is sprung together until the points touch, it will just 
 fit in the cylinder and will exert a uniform pressure of 2 pounds 
 per square inch. 
 
 If R is constant the expression for curvature becomes 
 
 or 
 
STEAM END DETAILS 361 
 
 Hence the following results in terms of t]^' 
 
 6=0 
 
 5 
 
 10 
 
 15 
 
 30 
 
 45 
 
 60 
 
 t vii =o 
 
 .124 
 
 .197 
 
 257 
 
 .405 
 
 529 
 
 .630 
 
 
 90 
 
 120 
 
 
 150 
 
 180 
 
 
 
 794 
 
 .908' 
 
 
 .971 
 
 i 
 
 
 The thickness t] Q is made the same in this as in the former 
 method. 
 
 PISTON ROD 
 
 The piston rod is designed as described in chapter VI, 
 page 314. 
 
 For the remainder of this chapter the maximum pressure 
 on the piston will be called P, where 
 
 A- P - 
 
 **t c ' 
 
 A i = area at root of thread; 
 A =area of main rod. 
 
 Shoulder $" to J" with taper of 3" in 12". 
 
 STUFFING Box 
 
 The stuffing-box design has been given in chapter VI, 
 page 201. 
 
 CROSS HEADS AND CONNECTING RODS 
 
 The cross head is necessary to guide the piston rod and keep 
 it from deflecting under compression and also to take the side 
 thrust from the connecting rod. The amount of this side thrust 
 
362 
 
 PUMPING MACHINERY 
 
 is shown by a force diagram in Fig. 300. The three forces, P 
 from the piston rod, P' from the connecting rod and 7? from 
 the guides are in equilibrium. R and P' will be the largest 
 
 for a given value of P when a is a maximum. Sin a = . ac is 
 
 ac 
 
 constant. Hence sin a is a maximum when ab is a maximum. 
 This occurs when ab=*oa or is 90. In this case the length of 
 
 FIG. 300. Force Diagram. 
 
 P' in force polygon is n times the length of R and the following 
 relations hold: 
 
 P' = 
 
 =P, 
 
 (n= ratio of connecting rod length to crank arm.) 
 Although P rarely continues its full value until the crank 
 has moved 90, since cut off occurs earlier than this, it is well in 
 design to consider this to happen and design accordingly. 
 This load might be developed if the full pressure were con- 
 tinued. 
 
 The form of cross-head used in pumps when the water end is 
 in tandem with the head end of the steam cylinder or where the 
 pump is driven from another part of the shaft, as in a triplex 
 pump, is shown in Fig. 301. 
 
STEAM END DETAILS 
 
 363 
 
 i ^ 
 i 
 
 
 
 | 
 
 M 
 
 
364 
 
 PUMPING MACHINERY 
 
 This type is not the principal type used, but it is of good 
 form and is therefore given. 
 
 The main casting A is made like a rectangular box with one 
 end open and bosses BB placed on two sides, while a longer boss 
 or tube C is placed on the side opposite the open end. On the 
 upper and lower sides of the main casting A are placed slippers 
 or shoes EE which are supported by wedges D. The lower 
 face of the wedge D rests on the horizontal surface of A , but its 
 upper face is inclined and supports the inclined face of E. The 
 wedge may be moved by turning the bolt F, which is j ammed 
 after moving the wedge far enough to lift the slipper out from 
 
 FIG. 302. Cross-head. 
 
 the casting an amount equal to the wear. The slippers are 
 faced with a bearing metal such as babbit. 
 
 The pin G is slightly tapered where it passes through the 
 sides of A. This makes it possible to have the pin tight without 
 driving. The pin is fastened by a nut to the plate H which is 
 held in position by two tap bolts // while the set screws // 
 jam the plate so that there is no danger of the bolts // backing 
 off. 
 
 The piston rod is screwed into the boss C and a jam nut 
 keeps it from backing off. 
 
 The cross-head, Fig. 302, illustrates a form which is used 
 when four guide bars are employed. The main casting consists 
 of two blocks A A connected by the yoke B. The cross-head 
 
STEAM END DETAILS 
 
 365 
 
 pin C is separate from the casting and is held in position in the 
 grooves made for it by the pin D. The piston rod fits into the 
 tapered hole at E and is held in place by a cotter. 
 
 Plunges 
 
 FIG. 303. Crank and Reciprocating Parts. 
 
 FIG. 304. Two-rod Cross-head. 
 
 Fig. 303 shows the arrangement of the steam piston and 
 water plunger when these are in tandem with the crank-shaft 
 between them. When this is the case the cross-head has to be 
 so arranged that the rods which are used to cross over the 
 
 . : FIG. 305. Pump End Cross-head. 
 
 crank and shaft may be fastened to the cross-head. As shown, 
 this is accomplished by two rods, one above and behind the 
 center line of the piston rod and the other below and in front 
 so fixed that they are symmetrical about the center line. 
 
 The cross-head at the steam end is showm in Fig. 304. It is 
 
366 
 
 PUMPING MACHINERY 
 
STEAM END DETAILS 
 
 367 
 
 
368 
 
 PUMPING MACHINERY 
 
 of the same form as Fig. 301 with the addition of the arms A A. 
 The water end cross-head, Fig. 305, is only used to prevent 
 buckling of the rods when under compression and hence is not 
 so large as the other. 
 
 In some cases the designer prefers to use four rods "sym- 
 metrically placed, and in such a case the form used is that 
 shown in Fig. 306. This form is modified on account of the 
 great distance between the frames in which the guide surfaces 
 
 FIG. 309. Strap End. 
 
 must be placed. The sliding blocks, as seen in the figure, are 
 cast blocks fitted to the pin at the center, which passes through 
 the main portion and forms the wrist pin. These blocks are 
 used for guiding only, the largest part of the steam forces 
 going directly from the piston rod to the pump rods. 
 
 The types of connecting rods used on these machines are 
 shown in Figs. 307 and 308. In Fig. 307 the main forging is 
 made with boxes at each end and is circular in section at A, 
 being turned from A to B as a cone with its faces slabbed off. 
 The box ends C and D are fitted with brasses EE which are 
 
STEAM END DETAILS 
 
 369 
 
 tightened up for wear by the wedges FF. The wedges are 
 moved by the bolts G which are jammed by the nut on their 
 opposite faces. The wedges are arranged to shorten one end, 
 as wear occurs while the other end is lengthened. 
 
 In Fig. 308 the rod is made with a conical section in which 
 the slabbing of the sides occurs near the crank end. In this 
 
 FIG. 310. Marine End. 
 
 rod the box at the crank end is made by inserting a block G 
 between two fork ends HH and bolting it. This is necessary 
 when the rod is used with the center crank A of Fig. 312. The 
 rod 307 could only be used with the overhung crank B of Figs. 
 311 and 312. 
 
 The strap end 309 and the marine end 310 may be used with 
 center cranks. In the strap end, Fig. 309, the strap A is held 
 on the square end of the rod by the gib B and the key or cotter 
 
370 
 
 PUMPING MACHINERY 
 
 C inserted in a slot through the end of the rod. The key is 
 prevented from coming out of the slot by a set screw, and the 
 end of this set screw engages with the key in a groove so that 
 the burr formed does not interfere with the removal or adjust- 
 ment of the key. The brasses and other parts are the same as 
 in the previous rods. The marine end, Fig. 310, has two brasses 
 
 c 
 
 FIG. 311. Crank Shaft. 
 
 A A held to the club end B of the rod by a keeper C and two 
 bolts. The liners D allow the bolts to be made tight and jammed 
 without gripping the crank pin. When wear occurs these are 
 reduced in thickness. 
 
 The crank-shaft for a two cylinder engine is shown in Fig. 
 311, while in Fig. 312 a three-crank shaft is illustrated. In these 
 the arms BB may be driven on and keyed or in some cases they 
 
 FIG. 312. Three-crank Pin Shaft with Return Cranks. 
 
 are forged solid in one piece. The bearing journals are at DD 
 and at times the shaft is enlarged at the point of attachment of 
 the flywheel, as at C. A key-way is left for the flywheel. These 
 shafts are sometimes forged hollow to reduce weight and to 
 remove metal which is of little value. 
 
 DESIGN OF CROSS-HEAD, CONNECTING ROD AND SHAFT 
 
 The designs of the cross-head, connecting rod and shaft are 
 interdependent so that it is well to take up the designs together. 
 
STEAM END DETAILS 371 
 
 THE CROSS-HEAD 
 
 The pressure on the cross-head to be carried from the piston 
 is P pounds, P' pounds from the connecting rod and R from 
 the guides, where 
 
 If the connection with the piston rod is made by a thread, 
 the amount of thread to be covered by the boss of the cross- 
 head should be equal to the diameter of the rod. If a key is 
 used to key the rod to the cross-head the thickness of the key 
 (t, Fig. 302) should be o.zd and the width is given by 
 
 2Xo.2dS 8 ' 
 
 The pin should be investigated for crushing and the rod 
 for bearing by the formulae: 
 
 o.2dXdS r >P 
 
 The cross-head pin may be designed empirically and then 
 investigated for bearing, bending, and deflection. A good rule 
 is to make the length of the pin (/ Figs. 301 and 302) equal to 
 
 x/P 7 x/P 7 
 , and the diameter (d f Fig. 301 and 302) . 
 
 P' 
 
 The bearing pressure jj, then becomes 1200 pounds per 
 let 
 
 square inch, which is allowable for wrist pins. 
 
372 PUMPING MACHINERY 
 
 The investigations for stress and deflection are given by the 
 formulae: 
 
 s- / 
 
 *~>8 
 
 nd' 2 ' 
 2 
 
 4 
 
 100 _ ^TTtf 4 ' 
 
 384^ 
 
 These are usually fulfilled, but it is well for the student to 
 investigate expirical design. 
 
 The area of the shoe is designed so as to give a pressure of 40 
 pounds per square inch in fast-running engines with cast iron 
 slippers and 250 pounds in slow engines when the shoe is lined 
 with babbit metal and 80 pounds in case of cast-iron slippers 
 against cast-iron guides for slow running. 
 
 A'=l'w'=j,. 
 
 p' is the allowable pressure for any given problem. 
 
 The relative values of /' and w' are fixed by the design. 
 
 \l r =w f may be used. 
 
 The pin is held by the sides of the main casting. In Fig. 301 
 the supporting surface is zd't". 
 
 P' 
 
 The details of the main casting (Fig. 301 ) are given in terms 
 oid'. 
 
 Thickness side, t'" = 0.3^', but must be not less than J", 
 Thickness boss, *" = *'" + \", 
 Thickness bottom, ? v =0.25 d', 
 Thickness front, F = o.2$d'. 
 
STEAM END DETAILS 373 
 
 In the cross-head shown in Fig. 302 the design is somewhat 
 different. Here the thickness must be calculated. 
 
 
 
 ~ 
 
 <*"' = 
 
 The cross-head pin is cut off for 20 at top and bottom as 
 this is not of much value and it distributes the oil better 
 when so treated. 
 
 The thickness of the side blocks t vii is assumed and then 
 investigated for strength. The piece is under combined bending 
 and direct stress and is investigated as follows: 
 
 Fig. 302 shows the general arrangement of the forces and 
 their positions. The thickness t v{i may be assumed and then 
 the center of gravity c, moment of inertia /, and area of section 
 A', shown at side, are found. The point of application of the 
 
 P 
 
 load is at ty vl from the edge so that there is a lever arm of 
 
 c~^t v{ causing bending. On the crank-head stroke the maxi- 
 mum tension is 
 
 P Pc-F { c 
 
 *~ 
 
 2l 
 
 on the other stroke the maximum compression is the same. 
 Since the tensile strength is usually less than that of the com- 
 pressive strength for the material used for cross-heads, the 
 investigation for S t is the only one made. If S* is found to be 
 greater than the allowable S t , the section must be increased or 
 cast-steel used in place of cast-iron. 
 
 The yoke is designed as a cantilever beam. At the center 
 it is best to assume d viii and compute I"' by the formula 
 
 p 
 
 - 
 
374 PUMPING MACHINERY 
 
 For a section at x from the center, the thickness d iv is 
 assumed and the distance / iv is found by the formula: 
 
 In Fig. 305 the arms are designed as cantilever beams by 
 assuming t and finding w from the formula: 
 
 In Fig. 306 the bending moment is %Pb. Although the force 
 P is in general not the whole force on the steam piston, yet the 
 design should be made to carry this load since that force might 
 come upon it. 
 
 The length of the crank pin is determined by the heating 
 caused by friction. The total piston pressure P will produce 
 the amount of friction //P where /* is the coefficient of friction. 
 This force moves through the distance rdd. The work per 
 minute will be 
 
 W = 
 
 If this is divided by the product of the length and the diameter 
 of the pin I'd', the result will be the amount of work per square 
 inch of projected area of pin. This quantity then becomes 
 
 . 
 
 f 2 * 
 
 If for I PdO the approximate value, 2nP mean is used, the 
 J 
 
 above expression becomes 
 
 W 
 
 I' 
 
 By using the values of , N, P and / in a number of engines 
 which have run without heating, a mean value of w may be 
 found and from this value on substitution a formula for / may 
 be derived: 
 
since P mean = 
 
 STEAM END DETAILS 375 
 
 I.H.P.X33,ooo 
 
 2LN . ' 
 
 These two formulae may be used to determine /' by making 
 7^ = 0.00003 an d K' =o.j to i.o. 
 
 The value of u. is 0.04 for proper lubrication and o.io for 
 poor lubrication. 
 
 After the length of the crank pin is determined its diameter 
 is found to care for strength, bearing pressure and deflection. 
 These three are considered separately. If the crank pin is 
 overhung, the pin acts as a cantilever beam. Then 
 
 Pi' 
 
 32 
 
 Forbearing 
 
 
 In this p is the unit bearing pressure. For slow engines 
 Unwin allows 800 to 900 pounds per square inch. While on 
 fast engines the value is from 500 to 800, in marine engines 
 400 to ^oo. For deflection the formula is 
 
 i _ 
 * El ~ 
 
 The amount of deflection is limited to o.oi inch. If the crank 
 pin is a center pin, the design for strength is to be considered 
 with the design of the shaft, and for that reason the shaft design 
 will now be studied. 
 
 In beginning to design a shaft, assumptions must be made 
 
376 
 
 PUMPING MACHINERY 
 
 of the various lengths. The length of the pin /' has been found 
 and this will serve as a guide for certain other parts. It may 
 be assumed that crank arms have the width /' and that the 
 bearings have a length 2,1' . With these dimensions in view an 
 approximate sketch is made. Fig. 313 shows this for an 
 
 f*^f *-* 
 
 E- 
 
 X 
 
 FIG. 313. Shaft Sketch. 
 
 overhung crank and Fig. 314 for a triple crank-shaft, both of 
 approximate form. 
 
 The distances A are assumed of sufficient size for the fly- 
 wheel or to bring the cylinders of the multicylinder engine at 
 the proper distances apart. The quantities, a, b, c, d, e, etc., are 
 now found in actual lengths in inches. 
 
 From the diagrams of Fig. 281 the greatest steam pressure 
 
 ^-+ A ^ 
 
 
 e- 
 
 
 _^ A 1_,_ 
 
 xi bxc 
 
 XT IX 
 
 i d ,i. 
 
 ^u f ^ 
 
 FIG. 314. Shaft Sketch. 
 
 can be found and the crank position at which it acts can be 
 constructed, giving a diagram as Fig. 315 from which P' is 
 found. This force acts on the crank pin; but its equivalent 
 P f acts at the center of the shaft in connection with a couple. 
 This is seen in the figure by adding two equal and opposite 
 forces at the center of the shaft. One of these combines with 
 P' to form a couple which produces twisting only while the 
 
STEAM END DETAILS 
 
 377 
 
 other force left at the center of the shaft produces bending 
 only. The other piston forces on a triple crank-shaft may be 
 found by obtaining the positions of the other pistons from a 
 diagram similar to Fig. 282, and then from the piston position 
 obtaining the pressures from Fig. 281 by multiplying the heights 
 
 p' 
 
 FIG. 315. Forces on Shaft. 
 
 by the spring scale and the area of the piston. The direction 
 and magnitude of the rod pressure are found in a manner similar 
 to that shown in Fig. 315. 
 
 The weights of the flywheels are now taken from the deter- 
 mination of chapter VII, and after resolving the connecting 
 
 FIG. 316. Loading Diagram. 
 
 rod pressures into vertical and horizontal components, the 
 shaft is treated as a beam, first subject to vertical forces and 
 then to horizontal forces. After the bending moments or bend- 
 ing moment diagrams for these are computed, the resultant 
 moments may be found and the maximum determined. The 
 loading diagrams for a vertical, overhung engine is given in 
 Fig. 316. In these figures the forces are arranged as if they 
 
378 
 
 PUMPING MACHINERY 
 
 were acting in the same direction. They may in reality be 
 reversed and when this is the case the sign of the term involving 
 any such should be changed. 
 
 The critical points in Fig. 313 are at the left support and 
 under the load. 
 
 These may be computed 
 
 M v (at left support )=P v 'a, 
 M h =Ph'a. 
 
 At load: 
 
 The resultant of M v and M h = VM V 2 +Mh 2 . 
 
 FIG. 317. Loading Diagram. 
 
 The largest of these is the one which fixes the size of the 
 shaft. 
 
 In the case of the three-cylinder engine, the shaft is a contin- 
 uous beam of three spans loaded with concentrated loads and 
 hence the theorem of three moments must be applied to find 
 the moment at the supports. 
 
 The general form of the theorem for this case is 
 
STEAM END DETAILS 379 
 
 The equations for this -particular case for which the loading 
 diagram is Fig. 317, are as follows: 
 
 From these four equations the values of M over each support 
 may be found and then the shear to the right of each support 
 is given by 
 
 I/ ^2,-Mj W 
 d K 2 J 
 
 M 3v -M 2 P' v 
 
 V2= ~7~~ + T' 
 
 MA. Mo.. W 
 
 f ' 2- 
 
 (These are of this form since load is at center.*) 
 The moments at the load will then be 
 
 V 2 e 
 
 This same operation is performed for the horizontal forces, 
 remembering that the terms involving W are zero. From the 
 values of the positive and negative moments the resultants are 
 found and the maximum determined. In this manner deter- 
 mination should be made for various positions of the crank and 
 the maximum of the maxima, ascertained. 
 
 This may be the correct method, but a better method would 
 
 *The general form for Vi is Vi = Mz ~ Ml + +P(i-Ki). 
 
380 PUMPING MACHINERY 
 
 be to assume the engine fixed on the dead point on one cylinder 
 with full steam pressure, and have the full net steam pressure 
 on one of the others the same while the third has exhaust pressure. 
 With this, work out the equation above. A determination 
 should also be made of full boiler steam on the first crank alone 
 with no pressure on the other pins. In these problems the 
 pressures from the pistons might be considered as acting directly 
 on the pins with connecting rods of n =6, as the effect of angu- 
 larity is not great enough to affect the result. 
 
 After making a number of investigations the values of M 
 at the various pins are known. 
 
 From the tangential effort diagram, Fig. 283, the twisting 
 moment of any shaft can be found as well as the twist at the 
 various pins. Twisting moment is equal to the tangential 
 effort multiplied by crank radius. The diagram of H.P. cylinder 
 gives the twist from the first crank to the first flywheel, and 
 the third diagram from the L.P. cylinder gives the twist going 
 to the flywheel near it. The amount going to the flywheels 
 is given by the combined diagram, and if one half of this 
 is subtracted from the H.P. diagram the amount which 
 remains shows the quantity going to the second crank from 
 H.P. The twist from the L.P. diagram may be treated in the 
 same manner and the amount received by the second crank 
 from L.P. is given. The quantities from the two ends when 
 added should give the amount from the I. P. diagram. If the 
 amount from the H.P. diagram reaching the second crank is of 
 the same sign as that on the I. P. diagram, the whole twist from 
 H.P. must be transmitted through the second crank pin. If 
 the amount of twist from the I. P. diagram is less than that from 
 the H.P. but of different sign, then the arithmetic difference 
 is transmitted. In this way a diagram may be drawn for the 
 twist which is transmitted across the center crank. 
 
 If now from these diagrams the bending moments and twist- 
 ing moments are obtained at various points in the revolution 
 for the pins and shafts below the wheel, the values of T\ = 
 could be found. 
 
 It has been shown by Grashoff that when bending and 
 
STEAM END DETAILS 381 
 
 torsion occur at the same time in a shaft, the shaft should be 
 designed as subject to a bending moment of MI where 
 
 when M= bending moment, 
 
 T = twisting moment. 
 
 MI = combined equivalent bending moment 
 
 Rankin develops a slightly different formula by considering 
 maximum stress, in place of maximum deformation and arrives 
 at the formula 
 
 Mr. J. J. Guest, within the last few years, has examined bodies 
 subject to tension and torsion and finds that the controlling 
 force is not one of tension but of shear, and so he recommends 
 that the combined equivalent shear be used in designing when 
 these two stresses occur. This would give for the moment 
 formula: 
 
 The diameter of the shaft or pin may now be found by this 
 formula. 
 
 It is to be noted in passing that when a triple-expansion 
 three-cylinder engine drives a pump or other machine at one 
 end of the shaft, the crank pin nearest the point of driving 
 transmits the sum of the twists from the cylinders placed in 
 front of it. 
 
 The investigation for bearing should be made with center 
 cranks at this point, although d' is so large here on account of the 
 bending moment that it will rarely be found necessary to enlarge 
 d' to that given by the formula: 
 
 -4 
 
 Having now the diameter of the crank pin for the various 
 forms of cranks, with the length, the boxes for the ends may be 
 designed. 
 
382 PUMPING MACHINERY 
 
 The proportions recommended by Unwin are given in Fig. 
 318, where the unit is 
 
 when the pin is an overhung pin, but with center pins, the unit is 
 
 The connecting rod proper is subject to stress produced by 
 four causes: ist, Direct tension from the piston pull when the 
 piston moves toward the head; 2nd, Direct compression on the 
 return; 3d, The bending stresses produced by the column 
 flexure, and 4th, The bending stresses produced by the whipping 
 of the rod as it changes the direction of motion at the top and 
 
 FIG. 318. Box for Connecting Rod. 
 
 bottom positions of the crank in a horizontal engine. The stress 
 in the first case is equal to that in the second, and as the second 
 and third are included in the ordinary column formula, investi- 
 gation by that formula is sufficient. The fourth stresses are 
 produced by the inertia of the rod, and if m be the mass for any 
 unit length of the rod, and a the. maximum acceleration, ma 
 represents the inertia force over that length tending to bend 
 the rod. The acceleration in the rod is a maximum near the 
 crank position 90 from the dead point. The rod is pivoted at 
 the cross-head pin, and hence the acceleration varies as the 
 distance from that point. If the connecting rod is uniform in 
 section, the effect of inertia is the same as if there were a beam 
 loaded with a distributed load which increased from zero at one 
 
STEAM END DETAILS 
 
 383 
 
 end to a certain value at the other, as in the upper part of Fig. 
 319. The bending moment is a maximum at O.577/, and it is 
 at this point the dangerous section for bending occurs. 
 
 If the section of the rod is not uniform, but increases toward 
 the crank end, the loading does not vary along a straight line 
 but along some curve, as shown in the lower part of Fig. 319. 
 In this case the dangerous section for the bending is further out. 
 
 Now the dangerous section of the column is at J/, but owing 
 to the bending effect from inertia, consider o.6/ as the dangerous 
 section with a rod of varying cross-section for both bending, 
 
 FIG. 319. Inertia Load on Connecting Rod. 
 
 due to the column and to inertia. If the rod is uniform the 
 dangerous section would be between \l and 0.577^. 
 
 The stress produced at the dangerous section, treating the 
 rod as a column, is given by the formula: 
 
 while the stress by the inertia bending is given by 
 
 Me 
 
 Now this latter could only be determined after knowing 
 the shape of the rod, and hence the shape would have 1 to be 
 assumed for a first approximation. In place of this such a 
 value of S in the column formula may be taken that this inertia 
 effect would be covered. If such a value is used and the cross-sec- 
 tion deterrhined from formula, the shape is found at the danger- 
 ous section, which is considered to be at o.6l. 
 
384 PUMPING MACHINERY 
 
 The rod may now be designed. The cross-head end of most 
 rods is made smaller than the remainder of the rod and is designed 
 to take direct tension or compression. This is therefore designed 
 first. If the rod is to be uniform in section this end is not de- 
 signed, as the size is determined by the cross-section at the center. 
 
 The force which this rod may have to stand is, as was men- 
 tioned under crank pins, 
 
 - * 
 
 Vn 2 -i 
 
 In many engines the cut off occurs before the crank has 
 reached its mid position, but it is best to use this value as the 
 valve may be reset so that the pressure could be carried to the 
 90 crank position. 
 
 The area at this smallest section is 
 
 P f 
 4000' 
 
 The value 5=4000 being found, from practice, to be a fair 
 average. 
 
 If the rod is rectangular, 
 
 A" = b"d", 
 in which V or d" is assumed; 
 
 .. n =A", - ':. , ; :- ,-...;, V/ 
 
 gives the diameter if the neck is circular. The connecting rod, 
 considered as a column in the plane of oscilla- 
 tion, has round ends, while in a perpendicular 
 plane it has fixed ends. In- investigating rods 
 of standard engines, considering the dangerous 
 section of rods similar to Fig. 320 as a rectangle, 
 which would circumscribe the section and not 
 as the actual section, it was found that a 
 
 FIG. 320. Standard r , , . . 
 
 Rod Cross-section. stress of 4ooo pounds used in the column 
 formula would give sections found in modern 
 engines. Hence this includes the effect of whipping as well as 
 the column effect. 
 
STEAM END DETAILS 385 
 
 Taking now the plane of oscillation, for the section at o.6/, 
 the following results : 
 
 P' 4000 
 
 'A JTf= 4 &' 
 
 25000 r 2 
 
 or in case the width is the same as the diameter at the smallest 
 section, 
 
 P f 4000 
 
 ^ 7r/= ZTTJL' 
 
 + 25000 d'"* 
 
 12 
 
 From these d'" at the dangerous section is found. 
 
 Now in the plane perpendicular to the plane of oscillation, 
 consider the rod as a column with fixed ends. This is not strictly 
 true, as there is some play, but since in this plane there is no 
 whipping effect, the value, 5=4000, will equalize the decrease 
 of strength due to the rod not having strictly fixed ends in that 
 direction. 
 
 P' 4000 
 
 d"d'" 
 
 25000 ^2 
 
 12 
 
 Solving this for d'", another value is found and the larger of 
 these two is taken. 
 
 Barr treats rectangular sections another way. The danger- 
 ous section is at the center, and if d f " is made 2d", columns. of 
 equal strength in the two planes result, hence only solve for 
 d" f once. The formula used is that of Euler, which reduces to 
 
 d"=c\/Di 9 
 
 For steel: 0=0.057 mean, 
 
 =0.07 maximum, 
 =0.043 minimum, 
 
386 PUMPING MACHINERY 
 
 In order to overcome the whipping effect df" is made greater 
 than zd". The mean value is 
 
 4.^" maximum, 
 = 2.2d" minimum. 
 
 The factor of safety here is 13. In this case the taper is 
 obtained by giving the rod at the neck the same cross-section 
 as the piston rod and tapering from the point. 
 
 Theoretically the taper need not extend beyond a point 
 at o.6l, but in high-speed engines there is some chance for the 
 crank pin to seize and then the connecting rod becomes a canti- 
 lever beam loaded at the end which might snap off. Hence in 
 this class the rod taper is usually carried out to the crank end. 
 In slow-speed engines the rod very often tapers each way from 
 the center and is circular in cross-section. 
 
 The brasses or boxes EE, Fig. 307, are next considered in the 
 light of the above rod design and an endeavor is made to make 
 the distance I" of Fig. 318 equal to d". This result would make 
 the forging of the rod simpler, and if it can be done by making 
 the thickness of the flange slightly greater or less by any 
 amount than that shown in Fig. 318, these should be changed 
 to accomplish this end. 
 
 CONNECTING ROD ENDS 
 
 The connecting rod end is next considered. The sides of the 
 box end carry tension; the end, bending; the rod end, compress- 
 ion. The bolts of the marine end carry tension; the keeper, 
 bending, and the club end, compression. The strap end has to 
 be designed for tension and bending in the strap, compression 
 in the rod end, and shear and compression in the gib and 
 key. 
 
 Box end. (Figs. 307 308.) Assume the diameter of the 
 bolt attached to the wedge " or "=^ iv . 
 
STEAM END DETAILS 387 
 
 Find b from brass and if possible make it equal to the thick- 
 ness of the rod d". Then 
 
 The rod end could be designed as a beam from the formula 
 
 6 
 
 Since the bending of this rod end would send the load to the 
 comers it may be well to neglect the design of this as a beam, 
 determining only the thickness at the corners for shear and 
 making r = i^ vi and using a circular arc for the ends, as 
 shown, for appearance, 
 
 The wedge should be sufficiently high for crushing and also 
 about three- fourths the height of the opening of the rod in length 
 so that the brass will be well supported. 
 
 The bolt at the end of rod shown in Fig. 308 supports all of 
 the load, hence 
 
 ^p 7 
 
 -d v . 
 
 ** r : 
 
 This should be designed for crushing and the fork end should 
 carry tension 
 
 P' 
 
 P' 
 
 2S,(6-^)' 
 
 This may require a thickening because of the large bolt used . 
 Such a result is shown in Fig. 308. 
 
388 PUMPING MACHINERY 
 
 p' 
 Marine end. Bolt area at root of thread, = ; 5- 
 
 4 
 Breadth keeper, b l =di = amount from box design. 
 
 /i = d' + 2t + di (t from brass ). 
 
 C Z, / 9 
 
 Thickness keeper, |P'/i = ' * 1 . 
 
 Length of keeper, l,+2di =/. 
 Crushing, P' <SJ // b 1 . 
 Strap end. Key and gib design: 
 
 '"1'" = 2S/V" = 25, (ft - 6"'X' =P'. 
 
 c 
 
 sr f -*| (in general) 
 
 Taper of key f " per foot. 
 
 The key and gib are each made /'" in width at top, when key 
 is first entered. 
 
 In the rod end the height t is usually much greater than 2t, and 
 hence there is no need of investigating for crushing between the 
 key and the rod end. If this were not so an investigation would 
 be made. 
 
 The end of the strap is designed in the same manner as the 
 end of the box. 
 
STEAM END DETAILS 389 
 
 BEARINGS 
 
 The bearings of the shaft are designed for sufficient bearing 
 area to care for the load. Using the vertical shears at each side 
 of the bearing in the continuous beam design the algebraic 
 difference' of these would be the reaction, and if this is found 
 in the horizontal and vertical planes the resultant R is given 
 
 by 
 
 As an example in Fig. 317, the shears at the left of the 
 supports have been found as Vi, V 2 , V$ y etc. On the right 
 of these supports the shears will be 
 
 P v ' at right of first. 
 
 V l at left of first. 
 
 Fi + TF at right of second. 
 
 F 2 at left of second 
 
 V 2 +P V " at right of second, etc. 
 
 The first reaction from the vertical forces will be P' V\. 
 The various .reactions from the vertical forces will be found in 
 the same manner; in like manner the horizontal reactions and 
 from them the resultant. 
 
 In Fig. 316 the reactions are: 
 
 Right horizontal, PhT- 
 Right vertical, -?A 
 Left horizontal, Ph , . 
 Left vertical, P v f ^~ 
 
 Right resultant, =R = ^P h ' 2 ~ + ( -P v ' % 
 
 T TJ 1 i *- ' * *-. * '' I v \ i -r-k * ^ 1 ^ * -r-r-r \ 
 
 Left resultant, 
 
390 
 
 PUMPING MACHINERY 
 
 Having the pressure allowable, the diameter of the shaft 
 at the journal (from stress design) and the load on the bearing, 
 the length is given by 
 
 R_ 
 
 ~ pd"' 
 
 For p Unwin recommends from 150 to 450, the larger 
 
 FIG. 321. Main Bearing. 
 
 values being used with slow-running engines, in which there 
 may be a reversal of load as near the stroke end. For fast engines 
 with shafts carrying a steady load in one direction the low 
 
STEAM END DETAILS 
 
 391 
 
 value should be used. 200 pounds is an average value which 
 may be used. 
 
 The bearings of large pumps are shown in Figs. 321 and 
 322. They are usually made in four parts: one bottom 
 brass A, two cheek pieces B, and one top piece C. This is 
 done so that wear may be taken up horizontally and vertically. 
 
 In building the boxes for the main bearing it is well to 
 make the lower section circular at the bottom, so that when 
 necessary to renew this section it may be taken out by blocking 
 
 
 
 \ 
 
 o 
 
 FIG. 322. Bearing and Frame. 
 
 up the shaft and twisting the box around. Of course such a 
 section does not resist the effect of friction to turn the 'brass, 
 nor can a shim or liner be introduced beneath the circular bottom 
 as it could be with the square base. If the box can be moved 
 along the axis of the shaft by the removal of bolts or caps this 
 would enable one to remove the box easily, and in this case a 
 flat base could be used, as in Fig. 322. 
 
 The proportional parts of these boxes and bearings are 
 shown in Figs. 321 and 322. The boxes are to be grooved so 
 that oil will be distributed and in many cases it will pay to 
 deliver oil under pressure to the main bearings. In any case 
 
392 PUMPING MACHINERY 
 
 a steady oil stream should be used, the oil being filtered after 
 using. 
 
 The thickness and projection of flanges on bearing brasses 
 are made J/. 
 
 If desired the surface of the central portion of the box which 
 supports the box may be cut away throwing the load to the 
 ends of the box. When this is done the width of the removed 
 portion is J/. 
 
 Where collars or crank bosses wear against the flange of the 
 box a slight amount of play is given to the shaft by leaving a 
 small amount of clearance. 
 
 The bolts holding the top cap should be strong enough to 
 hold the cap in place, and also in the case of horizontal engines, 
 to carry part of the thrust of the piston across the top of the 
 opening in the frame. This may be done by projections on the 
 side of the top cap, as shown in Fig. 322, when the bolts are 
 only used to keep the cap in place. The exact amount of this 
 force required to hold the cap in place cannot be computed, so 
 that this part of the design will be empirical, and the following 
 number of bolts will be used. 
 
 Shaft Diameter. No. of Cap Bolts. Diameter. 
 6" 2 ij-'* 
 
 8 2 I| 
 
 12 4 If 
 
 18 4 If 
 
 24 42 
 
 These bolts are often made T-headed and fitted into pockets 
 in the sides of the spaces left to receive the boxes. 
 
 The wedge bolts for lifting are usually two in number with 
 one or two set screws in addition, so that the bolts may be used 
 to raise the wedge while the set screws are used to force it down 
 after the bolts are loosened, or they may be used to hold the 
 cheek piece in place by tightening them up against the pressure 
 from the bolts. These bolts and set screws may be made f" 
 in diameter for boxes of shafts up to 12" and i" beyond that. 
 
STEAM END DETAILS 393 
 
 FLY WHEEL 
 
 The fly wheel is usually made as shown in Figs. 323 and 324. 
 That shown in Fig. 323 is in two sections while that of Fig. 324 
 is in five sections. 
 
 The two-section wheel is bolted together at the shaft, and 
 at the rim it is held by a link A which passes around projections 
 BB on each end of the sections. There are two, three or four of 
 these rings used at each joint. These projections are usually 
 
 FIG. 323. Two-part Fly Wheel. 
 
 made by forming a ring shaped cavity in the fape of the rim 
 so that the ring is flush with the surface. 
 
 In Fig. 324 the wheel consists of five sections bolted at the 
 center to two flanges A A, called spiders or centers. The sections 
 are held together at the rim by T-headed links B which are 
 placed in cavities in the rim. These links and those of the 
 previous figure are usually placed into position by heating, 
 the length, when cooled, being enough less than the length 
 between ends of projection that the extension to the proper 
 length would produce the force desired to hold these together 
 under the centrifugal load. 
 
 The fly wheel is used at times as a driving wheel for gears 
 and belts, but in any case it is subject to stresses of a complex 
 nature. These will be examined in parts. 
 
394 
 
 PUMPING 
 
 Consider the rim of a wheel as solid and free from spokes; 
 on each square inch of area of this rim there is a centrifugal 
 normal force of 
 
 where t= thickness in inches, 
 is)= weight of one cu.in., 
 V = velocity of center of this rim, 
 r= radius of center of rim [%(r ou 
 
 in ft. 
 
 
 FIG. 324. Sectional Fly Wheel. 
 
 Now the rim undor this load may be, considered as a thin 
 cylinder subject to a hydrostatic pressure, and hence the formula; 
 pd = 2tS holds. 
 
 In this case 
 
 p=F, 
 
 5= unit stress. 
 
STEAM END DETAILS 
 
 ;95 
 
 Hence 
 
 twV* 
 -- 
 
 or for the total breadth 
 
 V 2 i2b=tbS = T 
 g 
 
 When T is the total pull causing rupture. 
 The unit stress 5 becomes 
 
 I2WV 2 
 o = . 
 
 
 
 That is, the unit stress in a rim under centrifugal force when 
 rotated at the velocity V is 
 independent of the radius of the 
 wheel and size of the rim, and is 
 equal to 
 
 I2WV 2 
 
 - 
 
 6 
 
 For cast iron w =0.26 Ib. per 
 cu.in. 
 
 S=-i\V 2 approximately. 
 
 This stress of %wV 2 produces 
 an elongation in the radius 
 
 wV 2 r 
 of ! ,and if spokes are used 
 
 FIG. 325. Rim Shape. 
 
 the rim will take the exaggerated form shown in Fig. 325. The 
 spokes will be stressed by the rim pull and in turn this will cause 
 a distortion, as shown, so that at points A A there will be no 
 bending, but at the other portions of the rim bending forces 
 will result. These bending forces have maximum values at b 
 and c, the positive maximum being at c, while at b the negative 
 maximum occurs. 
 
 The variation of these stresses has been investigated theo- 
 
396 PUMPING MACHINERY 
 
 retically by Unwin in his "Machine Design," part II, pp., 293 
 301, and he derives the following formulae: 
 
 
 
 FR FR cos - 
 
 M = - - 
 
 2a 2 sin OL 
 
 _ 
 
 2 sn a 
 
 (a- <ft) 
 
 2 sn a 
 
 Where .F=pull on one arm in pounds. 
 
 M= bending moment at any section of the rim 
 
 in pound feet. 
 
 T = tension at any section of the rim in pounds. 
 5= shear at any section of the rim in Ibs. 
 a=^ the angle between the arms in radians. 
 G= weight of metal per cu.ft. in Ibs. 
 g =32.16. 
 
 A =area of cross-section of rim in sq.ft. 
 AI =area of cross-section of spoke in sq.ft. (mean). 
 V= velocity of center of rim in ft. per sec. 
 R = radius to center of rim in ft. 
 <= angle in radians to any point measured from 
 center line of spoke. 
 
 The experimental investigation of the resistance of fly wheels 
 to breaking or exploding, as it has been called, has been made 
 by Dean Chas. H. Benjamin of Purdue. In these investigations 
 reported to the American Society of Mechanical Engineers, 
 (Vols. 20 and 22) wheels of different designs, of reduced size, 
 
 were run until rupture occurred. At this point the speed was 
 
 72 
 noted, and from it the coefficient , derived above, was 
 
 10 
 
 computed. 
 
STEAM END DETAILS 397 
 
 In the case of a 1 5-inch solid wheel, the speed at rupture was 
 
 F 2 
 390 feet per second (6000 R.P.M.) and the value of =14,500. 
 
 The strength of the cast iron from which this wheel was made 
 was 19,000 Ibs. per square inch in direct tension and 39,000 was 
 the modulus of rupture. With a flanged jointed wheel of 24 
 mches diameter, the speed at rupture was 190 feet per second, 
 
 V 2 
 with =3600 Ibs. per square inch, and S t was 19,600. With 
 
 a linked joint similar to Fig. 323 the speed at rupture was 305 
 
 72 
 feet per second and the value of was 9300. 
 
 From these and other tests it is seen that the solid rim wheel 
 will develop slightly less than its tensile strength, while the 
 linked wheel will have about two-thirds the strength of the 
 solid wheel, and the flanged jointed fly wheel but one-quarter 
 the strength. 
 
 It may be assumed then that the rim is designed to take 
 the centrifugal tension. 
 
 There are other stresses brought on the rim and arms when 
 the speed changes. Consider a wheel in which the arms and 
 rim sections are not united. 
 When the engine is suddenly 
 speeded up the inertia of these 
 sections acts as a load on the 
 sections and bends the arms, 
 as shown in Fig. 326, where the 
 rim sections are joined. This 
 same action causes the rim to 
 take the form shown. This 
 action complicates the form- 
 ula given by Unwin, and to 
 cover them low stress should 
 
 FIG. 326. Effect of Bending. 
 
 be used, and then ' the rim is 
 
 designed for the centrifugal load, while the arms are designed 
 
 to take the bending due to driving or to inertia. 
 
 The twisting moment or tangential effort diagram, Fig. 283, 
 
398 PUMPING MACHINERY 
 
 gives the moment for which the arm must be designed. Let 
 h be the greatest height of the combined tangential effort diagram 
 which shows the amount of effort given to the two fly wheels. 
 The moment exerted by this is 
 
 JAX scale XL.P. piston area X crank arm=M. 
 
 This is the maximum moment during one revolution after 
 the engine has been started, but if steam is admitted suddenly 
 in starting the pump, the entire energy from the piston will 
 be transmitted to the fly wheels, and it is well to design the arms 
 for that contingency. This represents the maximum which 
 could ever come on the arms. This gives 
 
 nd 2 
 
 Mmax=i 4. P . />boiierX crank arm. 
 4 
 
 This moment is divided between m arms, hence 
 
 m 32 ' 
 
 for the elliptical arms of diameters b and d. 
 
 This is really the moment and size at the center of the shaft 
 and the arms are tapered on each diameter to the rim to two- 
 thirds these dimensions. 
 
 The rim of a solid wheel is sufficiently strong, provided V 
 be determined by the formula 
 
 V 2 
 
 < safe stress. 
 
 10 
 
 There is no way of making this stronger if the material 
 is fixed. For high speeds steel may be used to increase the 
 right hand side and so permit a higher velocity. 
 
 The links used to join the sections are subject to a load 
 
 fze'F 2 Xarea of section. 
 
 Hence, if there are n sections of link, each of area a at the 
 section, 
 
 =naS t . 
 
STEAM END DETAILS 399 
 
 From this a can be found, and from it the cross-section 
 dimensions of the link. 
 
 The bolts at the center of the wheel, as in Fig. 323 or 324, 
 should be large enough, if possible, to carry the section if the 
 links should break. To find the radial force produced b}' a 
 section of the wheel, it is necessary to find the center of gravity 
 of the section since 
 
 CF-^ F2 
 
 C - F '-g R> 
 
 where W is the weight of the rotating body, V the velocity of 
 the center of gravity and R the radius to that point. 
 
 The center of gravity of a section of a ring of angular extent 
 a and of thickness t and mean diameter D, is 
 
 This then gives the centrifugal force from the rim of a section, 
 to which may be added that of the arm by considering the 
 center of gravity to be at the center of the length of the arm. 
 Calling these forces JF c .f., the following gives the size of the n' 
 bolts in tension or the n" bolts in double shear to resist this, 
 
 c ,. 
 
 The center discs for the built up wheel and the hub of the 
 split wheel are usually of empirical design, and in the figure the 
 unit has been taken equal to x/rim area. 
 
 The fly wheels should be keyed to the shaft. The force coming 
 
 M- 
 max 
 
 on the key is, - , f , ,. - . 
 ' shaft diam. 
 
 M 
 
 f 
 shaft diam. 
 
 /= length of key in inches, 
 b = breadth of key in inches, 
 7 -- depth of key in inches. 
 
400 
 
 PUMPING MACHINERY 
 
 The balancing of fly wheels of pumps is a simple matter, for 
 the speeds are low, and all that is necessary is to have the wheel 
 in static balance. The balancing of the reciprocating parts 
 is not important in the slow-running pumps. For a detailed 
 discussion of balancing of the reciprocating and rotating parts 
 the reader is referred to "The Balancing of Engines," by W. 
 E. Dalby. 
 
 FRAMES 
 
 The frames of pumps are of different forms. Fig. 327 shows 
 a type in which the A frames supporting the cylinders are 
 
 FIG. 32% Frame for Vertical Engine: 
 
 carried on bed plates. The bed plate is carried on the air chanv 
 bers on onp side and on masonry on the other. The design ol 
 such structures is from experience. The thickness of the metal 
 usually being about one inch and the sizes of the columns are 
 
STEAM END DETAILS 401 
 
 fixed by what has been done. The load coming on the frame, 
 if inertia and friction are neglected, is equal to the net steam 
 pressure multiplied by the area of the piston. This force has to 
 be transmitted down in case of the upstroke, and on the down- 
 stroke the pressure is upward and is resisted by the weight of 
 this section of the engine and the weight of the masonry when 
 the pump is carried below on the foundation. In the case of a 
 single-acting pump the major part of the force on the upstroke 
 is transmitted to the main bearing and this causes a bend in 
 the frame of the overhung crank engine. This bending moment 
 is equal to 
 
 PX (distance a, Fig. 316). 
 
 The moment is resisted by the frame. On the downstroke 
 the pressure from the steam is balanced by the water pressure, 
 which is central, and so there is no tendency to bend. To cut 
 down vibrations in the frame cross pieces or knees are put in 
 at frequent intervals. 
 
 Fig. 346 shows a type of frame in which cast iron columns 
 are used to carry the weight to the lower foundations, while in 
 Fig. 349 the weight is carried by the air chambers entirely, 
 a method which gives good results, as it clears space around the 
 base, facilitating repair, examination or operation. By examin- 
 ing the various figures of the horizontal and vertical pumps, the 
 methods of .general practice may be seen. 
 
 The operating floors are carried at different levels from 
 beams bolted to the frames. These should be designed to carry 
 about two hundred pounds per square foot. 
 
 The guide surfaces are carried in the frame, and if not a 
 portion of the main casting the connections should be made 
 to carry the load 
 
 P' 
 n 
 
 The foundations are designed from experience. There are 
 two principles to consider: First, the foundation should be large 
 enough to properly distribute and carry the load, and second, 
 
402 PUMPING MACHINERY 
 
 there should be sufficient mass to reduce vibrations of the 
 machine. 
 
 The area of the base of the foundation should be such that 
 the weight carried from the engine and foundation is 3000 to 
 6000 pounds per square foot. The first value is for good clay 
 soil, the latter for a compact gravel. To get this it may be 
 necessary to spread the foundation, and to do this a batter of 
 about one foot in two or three feet of height will be found proper 
 to develop the strength of the projecting part. The foundations 
 may be made of concrete of proportions: one part cement, 
 two parts sand and five parts broken stone; or hard burned 
 bricks, set in cement mortar, may be used. These brick should 
 be thoroughly wet when laid and the joints should be grouted 
 with a watery mixture of cement mortar. 
 
 The amount of material to be placed in the foundation to 
 reduce vibrations is a matter of experience. In a slow running 
 machine there is practically no need of heavy foundations, 
 and when the engine is even of high speed and properly balanced, 
 there is no need of making the foundation very thick. A thick- 
 ness of three or four feet is sufficient for horizontal machines, 
 and for vertical machines the thickness will be determined by 
 the amount of height necessary to give the proper spread to 
 the base. 
 
 When the engine is to be placed over a very soft soil, it is 
 necessary to drive piles for the support of the masonry. These 
 are laid out so as to support a given load, say forty tons apiece, 
 and when driving, the pile is driven until the amount of pene- 
 tration under a given blow of the hammer shows that it will 
 carry the desired load. If this load can not be obtained the 
 piles must be placed close together. At times pier holes may 
 be sunk over the engine base area to good soil or rock and 
 the main foundation can be carried on the piers built in 
 them. 
 
 A formula recommended by Baker for bearing power of a 
 pile is 
 
 P = 100 (VWh + (5<><*) 2 ~5od), 
 
STEAM END DETAILS 403 
 
 where P = supporting pressure in tons, 
 W = weight of ram in tons, 
 A=height of fall in feet, 
 d= penetration of pile in feet at last blow. 
 
 The piles should be of good quality, of sound white oak, not 
 less than 10 inches in diameter at the smaller end and 14 inches 
 at the larger. They should be straight grained and have all 
 bark removed. After driving they are cut off below the per- 
 manent water line, and then concrete is put around the end, 
 making a solid bed. 
 
 SPECIAL STEAM PIPING AND VALVES. 
 
 The steam piping used on the engine should be designed so 
 that the velocity of the steam is 6000 feet per minute. This is a 
 simple method and gives good results. There is a method in 
 which the pipe is designed to give a certain discharge when the 
 drop in pressure is assumed. The formula developed by Prof. 
 Carpenter for this (A. S. M. E., vol. xx, p. 342) is 
 
 / 3 
 V f d' 
 
 2o.66 3 V ' r/Af* 1 
 
 where p =loss of pressure in Ibs. per sq.in., 
 k = a constant =0.0027, 
 d' = diameter of pipe in inches, 
 ze>'=flow of steam per minute in pounds, 
 L = length of pipe in feet, 
 D= weight of one cu.ft. of steam. 
 
 This may be used when long pipes are employed, but the 
 rule above for the area will give satisfactory results in pumping 
 stations, and even 8000 feet per minute may be used when neces- 
 sary for large pipes. 
 
 The steam lines should be made of heavy full weight pipe 
 and tested to stand 250 pounds per square inch. In this work 
 most of the pipes are large and the joints are made by flanges. 
 
404 
 
 PUMPING MACHINERY 
 
 Fig. 328 shows the proportion of the standard screwed flange, 
 while Fig. 329 shows a similar flange with a caulking recess on 
 
 FIG. 328. Flange. 
 
 FIG. 329. Screwed Flange 
 with Caulking Recess. 
 
 FIG. 330. Shrunk Flange. 
 
 FIG. 331, Rolled Joint. 
 
 the back which may be filled with a metal for caulking. This 
 is not always advisable. These two figures show the pipe screwed 
 on the flange. This is done so that 
 the pipe projects a slight distance 
 beyond the flange when it is tight- 
 ened to its full extent, and then the 
 projection is turned off flush with the 
 flange. The thread of the pipe is 
 cut deep to accomplish this result. 
 
 FIG. 332. Welded Flange. FIG. 333.- Globe Valve. 
 
 Flanges are now shrunk on the pipe, Fig. 330, after which 
 the end of the pipe is peened into a cavity left in the flange, and 
 
STEAM END DETAILS 
 
 405 
 
 then the end is turned flush with the flange. The Crane Company 
 sometimes roll the pipe into grooves turned in the flange, as 
 shown in Fig. 331. This operation is similar to the expanding 
 of boiler tubes into the heads of boilers. 
 
 The welded flange, Fig. 332, is a new form of flange conneo 
 
 M-. 
 
 FIG. 334. Ludlow Gate Valve. 
 
 tion for joining pipes, and according to the manufacturer it is 
 meeting with favor for high pressure work. 
 
 The dimension of these flanges, the number and sizes of the 
 bolts, the size of various fittings, bends and other specials for 
 pipe work, are to be found in the catalogues and on the dimension 
 sheets of Crane, Walworth, or the other manufacturers of pipe 
 fittings. 
 
406 
 
 PUMPING MACHINERY 
 
 The valves used are of the globe or gate types. The globe 
 valve shown in Fig. 333 is of the general form with an outside 
 yoke, and the type of gate valve shown in Fig. 334 is 
 often employed. The dimensions of these valves are found in 
 tables furnished by the manufacturers, so that in laying out 
 work the engineer may know how much to allow for these fittings. 
 
 FIG. 335. Crane Gate Valve. 
 
 The valves may have screw ends, Fig. 333, or flange ends, Fig. 
 334, while for water pipe bell ends are used. The large gate 
 valves, Fig. 335, used on water lines are built with gearing 
 to turn the spindle. 
 
 CONDENSERS. 
 
 The condensers used are either jet or surface condensers. 
 Wht,n the jet condensers are used, the volume of the con- 
 denser head should be one third of the volume of the low 
 
STEAM END DETAILS 407 
 
 pressure cylinder. The pipe through which the water enters the 
 head should be such that the velocity is 2 \ feet per second. That 
 is, 
 
 _ 
 62.5 X 2 ' 
 
 ^4 =net area pipe in sq.in. 
 W cw . = water required per sec. in Ibs. 
 
 The surface of the surface condenser should be determined 
 so that the condensation may be accomplished by the trans- 
 mission of 500 B.T.U. per hour per sq. ft. of surface per 
 degree difference in temperature. 
 
 This gives 
 
 W(H-q Q ') 
 
 5ooU- 
 
 where 
 
 5= surf ace in square feet, 
 W = weight of steam condensed per hour, 
 H = heat content of exhaust steam 
 
 =q + xr or approximately q+r, 
 t s = temperature of steam in condenser, 
 t Q = temperature of condensing water leaving condenser, 
 ti = temperature of condensing water entering condenser. 
 
 In the Journal of the American Society of Mechanical Engi- 
 neers for November, 1910, Mr. Geo. A. Orrok gives a resume 
 of a greater number of experiments and theoretical papers on 
 the transmission of heat by condenser tubes as well as the results 
 of a great number of experiments of his own and the deductions 
 drawn from all of these. In this paper Orrok derives the 
 equation, 
 
 S= square feet of condenser surface. 
 W = steam condensed per hour in pounds. 
 
408 PUMPING MACHINERY 
 
 G = condensing water per pound of steam in pounds. 
 
 t s = temperature of steam. 
 
 ti = temperature of condensing water at inlet. 
 
 to = temperature of condensing water at outlet. 
 
 c= cleanness factor, varying from i.oo with clean tubes 
 
 to 0.5 with dirty tubes. 
 fi= material coefficient. 
 
 = r.oo for copper. 
 
 =0.98 for Admiralty tubes. 
 
 =-0.97 for Admiralty aluminium lined. 
 
 = 0.92 for Admiralty oxidized (black). 
 
 = 0.87 for aluminium bronze. 
 
 = 0.80 for cupro-nickel. 
 
 =0.79 for tin. 
 
 =0.75 for zinc. 
 
 = 0.74 for Monel metal. 
 
 =0.63 for Shelby steel. 
 
 =0.55 for Admiralty badly corroded. 
 
 =0.47 for Admiralty vulcanized inside. 
 
 =0.25 for glass. 
 
 ^0.17 for Admiralty vulcanized both sides. 
 
 p-~- ratio of the steam pressure corresponding to the tem- 
 
 p 
 perature to the total vacuum pressure = *. 
 
 -* c 
 
 V w = velocity of water in tubes in feet per second. 
 Now 0=mean temperature difference in condenser 
 
 r 
 
 In practice, when ^ = 70 F., ^=90 F., t 8 =--g8 F., = 16.5. 
 If there is a 28-inch vacuum, and the tubes are medium clean 
 Admiralty tubes in which the water is flowing at 4 feet per 
 second, the following value is found for S for each 1000 pounds 
 of cooling water per hour: 
 
 9437 
 
al Pressure 
 
 q. in. 
 
 Sq. ft. 
 per i H.P. 
 . I.7T 
 
 
 I ^7 
 
 
 * O/ 
 . I ^O 
 
 
 .1 .43 
 
 
 I 37 
 
 
 .1.30 
 
 STEAM END DETAILS . 409 
 
 Sometimes the surface is determined by the I.H.P. of the 
 engine. The following table is taken from Peabody's " Ther- 
 modynamics of the Steam Engine": 
 
 20 
 15 
 12* 
 
 10 
 
 8 
 6 
 
 The air pump of the jet condenser is made sufficiently large 
 to care for the condensed steam, the condensing water and the 
 air. For this purpose it is made so that its displacement is 
 about forty times the volume of the condensed steam. For wet 
 air pumps used with surface condensers, the pump handles only 
 the condensed steam and air, and therefore the volume displaced 
 is reduced to one-half of the former amount or twenty times 
 the volume of the condensed steam. A better method of deter- 
 mining the volume will be given in Chapter XI. 
 
 The quantity of cooling water to be supplied these condensers 
 and to be cared for by the pumps is given by the equation: 
 
 W(H-qo') 
 
 *>-?< ' 
 W c . w . = amount of cooling water to be used with W Ibs. 
 
 of steam; 
 H = heat content of exhaust steam =q+xr assumed 
 
 to be q + r; 
 
 <?o /i= heat of liquid of cond3nsed steam; 
 <7o=heat of liquid of condensing water leaving; 
 <?j=heat of liquid of condensing water entering. 
 
CHAPTER IX 
 TEST OF PUMPING ENGINES 
 
 DUTY TRIAL 
 
 ENGINES are usually tested for steam consumption; that 
 is, the number of pounds of steam per horse-power-hour, but 
 with pumps, their duty is usually determined. Duty is the 
 number of foot pounds of useful work done by one thousand 
 pounds of dry steam or one million British thermal units used 
 by the pump. The problem, then, is to find the number of foot 
 pounds of work done by the pump, and the number of heat units 
 used by the pump. Duty meant originally the amount of work 
 per hundred pounds of coal, so that different pumps could be 
 compared. As coals varied this did not have a definite meaning, 
 and so a more definite unit was desired. The value 1,000,060 
 was taken because on an average one pound of coal gives 10,000 
 B.T.U. to the water in the boiler, and hence 1,000,000 B.T.U, 
 is equivalent to an average of TOO pounds of coal, while 1,000 
 pounds of steam can be assumed to be evaporated by this coal, as 
 each pound of water requires about 1,000 B.T.U. to evaporate it. 
 
 In considering the useful work done, the work necessary 
 to bring the water to the pump and that necessary to send it 
 from the pump to the reservoir must be included. The work 
 necessary to overcome the friction in the pump itself, i.e., 
 through valves and losses of sudden contraction, etc., is not 
 included, for these are due to the pump. With this in mind it 
 is seen that the useful work is 
 
 W 2g 
 
 where W is total useful work; w the weight of one cubic foot of 
 water; Q the total number of cubic feet pumped; pi the pres- 
 
 410 
 
TEST OF PUMPING ENGINES 411 
 
 sure head in the force main in pounds per square inch; p 2 the 
 vacuum in the suction pipe in pounds per square inch, and h 
 the distance in feet between the force gauge and the suction 
 gauge; or 
 
 W = Q ( i44p i + wh + 144^2 ) 
 
 where the symbols mean the same as before. 
 
 The value of w can be found from the temperature of water 
 pumped, and in important cases a known volume of water 
 should be weighed to find the effect of suspended matter. 
 
 The quantity of water is usually determined in one of five 
 ways: first, by displacement; second, by weirs; third, by 
 meters; fourth, by weighing; fifth, by measuring. 
 
 First, the diameter of the plunger is measured and also the 
 length of the stroke. These, with the number of the strokes, 
 give the displacement, or the quantity of water pumped if there 
 is no leakage, or slip, as it is sometimes called. This slip occurs 
 around the piston or plunger in an inside packed piston; but with 
 the method of outside packing used to-day this leakage is 
 nothing and the only other place for leakage to occur is past the 
 valves. In order to determine the leakage around the piston 
 when inside packed, and the valves when seated, standard 
 methods are used. (See Carpenter, " Experimental Engineering," 
 page 560, and Kent, "Pocket Book," page 611.) The leakage 
 past the valves in seating cannot be ascertained, except by 
 determining Q by one of the other methods. 
 
 In some pumps the stroke of the plunger is not definite, 
 and hence the stroke should be measured continuously. If 
 conditions are nearly constant the stroke will not vary much, and 
 hence readings of length taken at regular intervals are made. 
 This is not absolutely correct, and hence some form of registering 
 apparatus should be employed. (See A.S.M.E., vol. 12, page 
 981.) 
 
 Having now the areas, stroke and leakage, the amount 
 pumped may be computed 
 
 100 
 
412 PUMPING 
 
 Second, for this method the constants for the weir and 
 also the hook gauge must be known. Attempt should be made 
 to reduce the fluctuation of the water surface, and the gauge 
 should be read in a box on one side of the flume. 
 
 Third, since the introduction of the Venturi meter this 
 instrument can be used to advantage in pump tests, as its error 
 is small. The Pi tot tube also serves as a most accurate method 
 of determining the quantity if a traverse is made by a tube of 
 small size pointing in direction of the current when the static 
 pressure is taken from holes in the pipe walls. 
 
 Fourth, for small pumps the water pumped can be actually 
 weighed, thus giving the exact amount pumped. 
 
 Fifth, cisterns and reservoirs may be used when their 
 contents are known, but here allowance must be made for the 
 evaporation from the surface and the effect of wind. 
 
 The quantity p\ is determined by means of a pressure gauge 
 which must be tested. (Be sure to take maker's number of all 
 instruments used and position where used.) Since the pressure 
 in the force main varies considerably, the gauge cock is throttled 
 off, or an air chamber is introduced between the gauge and the 
 force main which will reduce *the fluctuations. p 2 is usually 
 determined by a mercury tube, and hence must be reduced 
 from inches of mercury to pounds per square inch. This pressure 
 is greater than the head between the points where measured and 
 the water in the forebay, as it is equal to the velocity head and 
 the head of friction loss, in addition to the lift. This should be 
 included, as it is not on account of the pump that such losses 
 occur, but on account of the location of the pump demanding 
 a certain length of pipe. 
 
 The difference of level between the suction gauge and force 
 gauge must be taken, because the water is raised through this 
 distance, but the distance is not given by either of the pressure 
 readings. 
 
 Having these, the useful work can be found. If indicator 
 cards are taken from the cylinder, they give the work done on 
 the water, and the difference between this and the useful work 
 
TEST OF PUMPING ENGINES 413 
 
 just determined, would give the losses due to leakage and the 
 friction of the water in the pump. 
 
 The readings of pressure and the revolution counter are 
 taken at regular intervals, and the average result used. 
 
 To determine the quantity of heat used by the engine, the 
 quantity of heat in each pound of steam used in the engine 
 must be found as well as the total number of pounds of steam 
 used. The best way to determine the quantity of steam used is 
 to condense the steam in a surface condenser and weigh it. To 
 this must be added the amount consumed in the steam jackets 
 and the superheaters. 
 
 The water coming from the jackets and superheaters is at 
 such a temperature that it would partially evaporate if allowed 
 to discharge into the air before being cooled. There are two 
 methods of weighing this: one consists of allowing the hot water 
 to discharge into a tank partially rilled with cold water; the 
 other is to conduct it through a coil surrounded by cold water 
 and discharge into tanks. 
 
 As condensers attached to pumping engines are often jet 
 condensers the above method is impossible, and so the following 
 method must be used: A number of boilers sufficient to run 
 the pump to be tested are separated from the other boilers and 
 all cross connections of the water or steam pipes are broken 
 and blanked off. Then no steam may be used for any other 
 purpose and no water can be brought into the boilers, except 
 by the one feed pump, which draws its supply from a tank, the 
 water in which has been weighed. 
 
 Now if there is any water collected in a separator between 
 the boilers and the engine, or used for any purpose whatsoever, 
 other than for the engine, it must be carefully determined. 
 
 Suppose W pounds of water are pumped into the boiler, 
 w pounds used in a calorimeter to determine the quality of the 
 steam, w' pounds taken from the separator, and w" pounds used 
 in the jackets and reheaters. The amount of water used then by 
 the main pumping engine is W w-w' pounds. Care must 
 be exercised to investigate the piping to see that there is no 
 
414 PUMPING MACHINERY 
 
 chance of leakage of the water or steam in any way, and all 
 joints into other lines should be broken and blanked off. A 
 leakage test of several hours' duration should be made at the 
 end of the test, with the main engine shut off to determine the 
 loss from the boilers and piping. The quantity so found is 
 subtracted from W. 
 
 By taking readings of the steam gauge, barometer, and on 
 the calorimeter, the absolute pressure and the quality of the 
 steam can be determined, and hence the quantity of heat above 
 32 F. supplied per pound. The total heat supplied above 
 32 F. is then 
 
 H = (W-w-w')(q+xr) 
 where 
 
 <7=heat of the liquid; 
 
 x = quality of steam; 
 
 r=heat of vaporization. 
 
 For superheated steam q+xr becomes q + r + I c p dt. 
 
 In general, the steam jackets and reheaters drain through a 
 trap into the boiler feed, and hence return some heat to the 
 boiler. To determine this the weight of the water from them 
 is found as before described and the temperature of that water 
 before it is cooled. The temperature is found by placing a 
 thermometer cup in the pipe and taking temperature readings 
 on a thermometer placed therein. The heat then returned is 
 
 'V 
 
 The feed pump, under actual running conditions, takes 
 water from the hot well, hence the temperature of this must 
 be observed and the heat supplied by this source determined. 
 If the temperature be 3 , the quantity returned is 
 
 (W-w-w'-w")q3 
 The total heat chargeable to the engine is 
 
 then 
 
 Work 
 
 i, 000,000= Duty 
 
TEST OF PUMPING ENGINES 415 
 
 If the duty to be found is on the basis of 1,000 pounds of dry 
 steam, the steam used must be expressed in terms of equivalent 
 dry steam. The equivalent dry steam is given by 
 
 (W-w-w')(q+xr-q 3 ) ., 
 
 W =- : TT-: , if saturated, 
 
 (q+r-qs) 
 
 or 
 
 / /-/sup. \ 
 
 (W -w-w'}\Q+r+ I cdt-qzj 
 
 We -- _^_ _L if superheated. 
 
 (q+r-q 3 ) 
 
 Work 
 Then duty = , 1,000. 
 
 There are other quantities which are reported. The mechan- 
 ical efficiency of the pump is determined by taking indicator 
 cards from the water and steam cylinders and determining the 
 horse-power of each end, and the ratio of these gives the mechan- 
 ical efficiency. The first subtracted from the second gives the 
 frictional loss. 
 
 Call the horse-power of the steam end I.H.P. and of the 
 
 ' D H P 
 
 water end D.H.P. Then mechanical efficiency ~ T TT W"- 
 
 l.rl.r . 
 
 I.H.P. -D.H.P. =H.P. of engine friction. 
 
 The quantity of steam used per hour, divided by the I.H.P., 
 is the steam per I.H.P. hour, a result useful for comparison with 
 other kinds of engines. This is 
 
 W w w' 
 
 = Steam per I.H.P. hour. 
 
 I.H.P. 
 
 or better 
 
 W e 
 
 prrp- =Dry steam per I.H.P. hour. . 
 
 The quality of the steam, or in other words the amount 
 of condensation, in each cylinder at any time after cut-off in 
 that cylinder can be determined, first, by finding the volume 
 (including clearance volume) and pressure at that point from 
 
416 PUMPING MACHINERY 
 
 the indicator cards, and second, from the table of the properties 
 of steam, the volume occupied by one pound of steam at this 
 pressure. From the quantity of steam used in the cylinder, the 
 number of pounds per stroke is found; and on adding to this 
 the steam in the clearance space the total steam in the cylinder 
 may be ascertained. The clearance weight is approximated by 
 considering the steam dry at compression and then finding 
 weight by the formula: 
 
 Volume 
 
 Wt 
 
 Specific Volume' 
 
 The product of the specific volume and the total weight per 
 stroke gives the volume which should be occupied. The ratio 
 of the actual volume at the point to this amount represents the 
 percentage steam present at that point, the remainder being 
 water, the volume of which has been neglected. In this way 
 the initial condensation may be determined. 
 
 The amount of heat removed by the surface condenser is 
 W'(q'-q"}+W"q'", where W is the weight of the condensing 
 water, and q' is the heat of the liquid leaving the condenser and 
 q" the heat of the liquid in the entering condensing water. W" 
 is the weight of the condensate and q'" its heat of the liquid. 
 
 This quantity of heat removed by the condenser may be 
 called #'; then 
 
 H-H'- Work = Radiation, 
 
 Radiation ^ , , ,. .. 
 
 or 100 = % loss by radiation. 
 
 H 
 
 The following precautions are to be noted in making obser- 
 vations : 
 
 First, carefully arrange observation sheets and become 
 familiar with readings before beginning test. . 
 
 Second, after noting observations make a check observation 
 when possible. 
 
 Third, keep a vigilant watch for leaks of any kind. 
 
 Fourth, note everything in book that happens out of the 
 ordinary. 
 
TEST OF PUMPING ENGINES 417 
 
 Fifth, make note of number and position of all instruments 
 used. 
 
 Sixth, calibrate and check all instruments and measure- 
 ments. 
 
 Seventh, in weighing coal or water, take the weight of vessel 
 empty and full, together with the time, in each case. 
 
 Eighth, note the time of all observations, except when 
 several different ones are made in quick succession, then note 
 the time of the first only, but take them always in the same 
 order. 
 
 Ninth, always have instruments ready for next operation 
 before leaving them. 
 
 The data and results called for in duty trials are given in 
 Kent's Pocket-book, and the following additional results and 
 data are also advisable for investigation of the pump action: 
 Plunger displacement per stroke. 
 Amount pumped, cubic feet. 
 Quantity pumped, computed from displacement. 
 Quantity of water used in condenser. 
 Quantity of heat absorbed by condenser. 
 Quantity of heat absorbed by condenser, per square foot 
 
 of surface. 
 
 Percentage loss by radiation. 
 
 These tests are usually made to find whether or not the 
 guaranteed duty is met, but an investigation in the same manner 
 may be made to find whether or not a given device or arrange- 
 ment is of value. For instance, in many cases it is found that 
 the use of the reheating coils in the receivers are not a source 
 of gain and at times a test without certain jackets will give 
 better results than a test made with the jackets in service. 
 
 Superheated steam can be used and its effect determined 
 in this manner. The use of sufficient superheat to cut down 
 the initial condensation and give dry steam at cut-off will pay, 
 but beyond this the expense in making the superheated steam 
 will be more than the gain in the engine. The only reason why 
 it should be of value when the initial condensation is eliminated, 
 is the fact that when the steam is dry or superheated the walls 
 
418 
 
 PUMPING MACHINERY 
 
 do not take up as much heat as when covered with a film of 
 
 moisture. 
 
 One reason why superheated steam should not add to the 
 
 efficiency, the amount apparently shown by the increase of the 
 upper temperature T\ of the expression 
 for the Carnot efficiency, 
 
 -" T y 
 
 1 1 
 
 is the fact that this heat is not supplied 
 at a fixed temperature 7^, but at a grad- 
 ually increasing temperature, and hence 
 although the efficiency of the cycle is 
 increased a slight amount the increase 
 is very small. This may be shown by 
 Fig. 336, which is a T<j> diagram. In 
 
 the figure, i, 2, 3, 4 represents the cycle of an engine using wet 
 steam, while i, 2, 3^,4' is one for an engine supplied with super- 
 heated steam. The efficiencies are 
 
 5 6 G 
 
 FIG. 336. Entropy 
 Diagram. 
 
 and 
 
 ~ 
 
 5,1,2,3,6 
 
 i, 2. 3'. 4' 
 
 1,2,7,4' 
 
 1274 
 51276 
 
 is slightly greater than 
 
 T ^ Q A 
 
 ' ' 
 
 so that sa and E 
 
 sat 
 
 su 
 
 are almost the same. 
 
 The use of jackets on the cylinder heads and barrels is 
 recommended although tests do not always show an increase 
 of economy. 
 
 The overall efficiency of the engine is given by dividing 
 the duty per 1,000,000, B.T.U. by 778,000,000, which is the 
 number of foot pounds equivalent to 1,000,000 B.T.U. ; this 
 number has reached a value of 0.23. This means that of the 
 heat supplied to the test pumps 23 per cent is utilized. 
 
TEST OF PUMPING ENGINES 419 
 
 The preparation of a pump for a duty test consists in much 
 adjustment of valve-gearing and other steam apparatus which 
 consumes time. Many preliminary runs are made to discover 
 the effect of changes and this can be done only by a systematic 
 method of keeping data and making alterations. One change 
 should be made at a time and its effect determined before 
 proceeding to another change. In all of this work a positive 
 circulation must be maintained in the heating coils and jackets, 
 for if the water of condensation is not removed these cease to 
 operate properly. 
 
 To illustrate the data and results of a duty test, the following 
 has been taken from a test on a 20,000,000 gallon pump at the 
 Lardner's Point Pumping Station in Philadelphia, conducted by 
 Francis Head, mechanical engineer of the Bureau of Water 
 of the city of Philadelphia, and Edgar G. Hill, mechanical 
 engineer representing the Holly Manufacturing Company. 
 
 "At Lardner's Point are three pumping stations, No. i, No. 
 2 and No. 3. Station No. i was formerly known as Frankford 
 station, and contains boilers and pumping engines that have 
 been in service a great many years, which are now retained 
 only as a reserve, the machinery being entirely too expensive 
 to keep in regular service. 
 
 "Stations No. 2 and No. 3 are new pumping stations through- 
 out; the buildings and equipment therein were constructed 
 for the new nitration distribution system. Each consists of an 
 engine house and boiler house. The two stations are placed 
 contiguous and for all practical purposes are one station. 
 
 "In engine houses No. 2 and No. 3 are twelve vertical 
 triple expansion self-contained pumping engines of 20,000,000 
 U. S. gallons daily capacity each, designed for a normal head of 
 225 feet, but capable of operating economically against heads 
 ranging from 180 to 280 feet. 
 
 "These twelve pumping engines, each substantially a 
 duplicate of the other, designed, constructed, and installed by 
 the Holly Manufacturing Company of Buffalo, N. Y., make an 
 installation the largest and most comprehensive of its type, not 
 only in the United States but in the whole world. 
 
420 PUMPING MACHINERY 
 
 "The water which these engines pump is supplied from the 
 nitration plant of the city of Philadelphia at Torresdale, between 
 two and three miles from Lardner's Point pumping station, 
 through an underground conduit leading from the Torresdale 
 filtration plant to the Lardner's Point pumping station. It 
 is pumped from the Lardner's Point pumping station through 
 a number of delivery mains, ranging in diameter from 48 to 60 
 inches, to various parts of the city of Philadelphia. 
 
 "Ordinarly, eight or nine of these engines discharge the water 
 through two 60 inch delivery mains, which are so connected 
 together as to form substantially one main of twice the capacity 
 of a 60 inch pipe, against a head of approximately 185 feet. The 
 other three or four engines ordinarily pump through a separate 
 pipe system against a head of approximately 275 feet. To 
 show the action of the governor the following is mentioned: 
 
 "At 8 o'clock, the morning of December 10, 1909, at a point 
 about one thousand feet from the pumping station, a 12 foot 
 length of the 60 inch delivery main, into which eight of these 
 pumping engines were at that time delivering water at the rate 
 of over 160,000,000 gallons per day, against a head of approx- 
 imately 185 feet, split from end to end, instantaneously reducing 
 the work from full load to almost no load. Five of these pump- 
 ing engines were in Station No. 2, and three in Station No. 3. 
 
 "All of the pumping engines were under such perfect control 
 of the governors and automatic safety devices that only one 
 of the eight attained sufficient speed to make the automatic 
 shutting-down device operative, the result being that one 
 engine stopped automatically, and the other seven ran at a 
 uniform speed until gradually closed down by the employees 
 at the station. Absolutely no damage was done to any of the 
 engines or to any of the machinery in the pumping stations. 
 
 "The broken 60 inch delivery main was promptly repaired 
 and the engines were all in service again the afternoon of the 
 same day the accident occurred. 
 
 "This is undoubtedly the first time the governing apparatus 
 of so many pumping engines has been subjected simultaneously 
 to a test of this character. That none failed to work properly 
 
TEST OF PUMPING ENGINES 421 
 
 and satisfactorily, speaks louder than words for the reliability 
 of the governing apparatus. 
 
 "One of the twelve engines was officially tested March 
 9-10, 1910, and developed a new high duty record for pumping 
 engines fitted with attached jet condenser, on a 24-hour trial. 
 
 "During the test, readings of the water pressure were taken 
 every five minutes from a correct pressure gauge checked by a 
 mercury column. A complete round of observations was taken 
 every fifteen minutes. All readings and weights were checked 
 by two observers, one representing the city and one the company. 
 
 "An eight-hour boiler leakage test was made before the duty 
 trial, also immediately after the duty trial. 
 
 "A summary of the principal results is given below, by the 
 makers: 
 
 Size of engine 30", 60", 9o"x33"x66" 
 
 Total water pressure, Ibs 95-74 
 
 Average total feed water per hour, Ibs 9,265.625 
 
 Feed water corrections, average per hour, Ibs. : 
 
 Drips from steam header and steam separator. . . . 135-27 
 Boiler leakage, average of two tests 196.39 
 
 Total.... ' -__. 33166 
 
 Average total. steam per hour delivered to engine, Ibs. ... 8,933.07 
 
 Entrainment, per cent . i .13 5 
 
 Average total dry steam per hour delivered to engine 8,832.56 
 
 Water pumped per 24 hours, gallons 21,218,788 
 
 Duty per 1,000 Ibs. steam delivered to engine 182,382,200 
 
 Duty per 1,000 Ibs. dry steam 184,476,200 '* 
 
 In the specifications for pumping engines definite statements 
 should be made of the manner of conducting the test and of the 
 working out the results. As an example, the following quota- 
 tions are made of clauses from the specifications for the Lard- 
 ner's Point pump: 
 
 "Section 119. Duty Test. During the period of probation, 
 and before its final acceptance, each engine shall be subjected 
 to a duty test of twenty-four (24) hours' duration. The test 
 shall be conducted by two (2) engineers, one (i) to be selected 
 by the director, and the other by the contractor, etc. 
 
422 PUMPING MACHINERY 
 
 "Section 120. Determination of Head. The nead for com- 
 putation of duty will be the sum of the head in feet indicated 
 on a gauge attached to the discharge main beyond the last pump, 
 and the vertical elevation of the center of this gauge above the 
 level of water in the pump well. The level of water in the pump 
 well shall be determined by a suitable float gauge, and no 
 correction or allowance will be made for any friction losses 
 between the water in the pump well and in the discharge pipe 
 just beyond the pump. 
 
 "Section 121. Total Head. The total head for the contract 
 duty test shall be not less than two hundred and fifteen (215) 
 and not more than two hundred and twenty-five (225) feet. 
 
 "Section 122. Capacity. The capacity of the engines shall 
 be in the duty test not less than twenty million (20,000,000) 
 gallons per twenty-four (24) hours, at a speed of not more than 
 twenty (20) revolutions per minute. The capacity of the pumps 
 during the duty tests will be determined by plunger displace- 
 ment, and no correction will be made for slip unless leakage 
 from plungers and valves is found by test to exceed two and 
 one-half (2^) per cent. 
 
 "If the leakage or slip is found by Pitot meter to be more 
 than two and one-half (2j) per cent, the capacity and duty 
 shall be computed from the measured capacity. 
 
 "Section 123. Stuffing Box Leakage. The leakage from all 
 the water stuffing boxes of the engine shall not exceed five 
 hundred (500) gallons per hour. This shall not be charged 
 against the engine. 
 
 "Section 124. Feed Pumps. The direct-connected feed 
 pumps shall be operated during the duty trial, supplying water 
 to the measuring tanks, and running against the usual discharge 
 pressure. One auxiliary feed pump shall be operated to remove 
 condensation. No allowance will be made for these pumps. 
 
 "Section 125. Steam Measurement. The water fed to the 
 boilers shall be weighed and any condensation in the lines and 
 drains from the live steam separator shall be deducted. All 
 steam passing through the separator to the throttle valve shall 
 be charged against the engine as dry steam. 
 
TEST OF PUMPING ENGINES 423 
 
 "Section 37. Uniformity of Steam Diagrams. The con- 
 struction and adjustment of the pump valves and steam valve 
 gear, and the balancing of the plungers and other moving parts, 
 shall be such as to give approximately uniform engine indicator 
 diagrams from each of the steam cylinders on the up and down 
 strokes." 
 
 In making the report on the test of this pump, the following 
 statements are made: 
 
 "The level of water in the pump well was determined by a 
 gauge glass placed near the pump suction instead of a float, 
 as this gave better results. 
 
 "Under Section 122, regarding correction for slip, as before 
 this was assumed to have been under 2-|- per cent, and no cor- 
 rection was made for it. 
 
 "Under Section 123, stuffing .box leakage was measured and 
 found to be less than allowed. 
 
 "Under Section 124, which calls for the operation of one 
 auxiliary feed pump to remove condensation, the results of the 
 test made on July 2oth, 1909, reported in the test of No. 16, 
 were used, and the amount so determined charged against the 
 engine. 
 
 "Under Section 125, the same six inch steam line was used 
 to convey steam from boiler No. 26 to the pump. As before, 
 the drip was trapped, condensed, weighed and deducted from 
 the feed water. 
 
 "The correctness of the feed water scales was checked by 
 Fairbanks Standard weights, and the steam and water gauges 
 were corrected by a Crosby dead weight tester, which had 
 been carefully calibrated. 
 
 "The water pressure gauge was further checked by a mercury 
 column which was set up beside the pressure gauge. The 
 density, of the mercury used was determined in the testing 
 laboratory of the Department of Public Works. 
 
 "On March 8th, the suction and discharge valves were 
 tested and made tight. 
 
 "On March 8th, a leakage test of the boiler and piping was 
 made from 9.47 A.M. to 5.47 P.M. On March loth, another 
 
424 PUMPING MACHINERY 
 
 test was made from i.oo to 7.55 P.M. The average net loss 
 was 196.392 pounds per hour. 
 
 "The capacity of the engine, as actually run during the test, 
 from plunger displacement, was 21,218,481 gallons per 24 hours 
 at a speed of 20.13125 R.P.M., 
 
 "The engine worked smoothly and satisfactorily during 
 the entire test, and the requirements of the specifications as to 
 economy and capacity on the duty test have been fully complied 
 with." 
 
 The following data was obtained during test from average 
 readings : 
 
 DATA AND RESULTS 
 
 Engine tested, contractor's number 604 
 
 Water Bureau number 14 
 
 Date of test March 8, 9, and 10, 1910 
 
 Duration of test 24 hrs. 
 
 CAPACITY 
 
 Revolutions during test 28,989 
 
 Average revolutions per minute 20.13125 
 
 Average diameter of plungers, inches % 32.985 
 
 Average stroke, feet 5-496 1 
 
 Number of plungers 3 
 
 Displacement per revolution, gallons 73 1.9494 
 
 Displacement per 24 hours, gallons 21,218,481 
 
 Displacement per 24 hours, at contract speed 21,080,144 
 
 Water used to lubricate plungers, per hour, gallons 360 
 
 WORK DONE 
 Head pumped against 
 
 Pressure, corrected gauge 86.918 Ibs. =200.50 ft. 
 
 Suction lift to center of pressure gauge 8.822 Ibs. = 20.35 ft. 
 
 Total 95-740 Ibs. =220.85 ft. 
 
 Work done per hour 1,629,397,423 ft.lbs. 
 
 Duty 
 
 DUTY 
 Foot-lbs. per hr. X 1000 
 
 Net steam charged to pump per hour 
 1,629,397,423,000 
 
 8991.713 
 = 181,211,013 ft.lbs. 
 
TEST OF PUMPING ENGINES 425 
 
 This is sEghtly different from the builders' computations given 
 
 above. 
 
 \ 
 
 PRESSURES 
 
 Throttle gaug-e reading, Ibs. per sq.in 190.82 
 
 Corrected for height 180. 19 
 
 First receiver, Ibs. per sq.in 33.8 
 
 Corrected for height 25.4 
 
 Second receiver, inches vacuum 9.5 
 
 Vacuum in condenser (reading of mercury column) inch 27.68 
 
 Average barometer at 32?. sea-level 14-789 Ibs. =30.06 
 
 Average barometer at floor level and room temperature 14.868 Ibs. =30.20 
 
 JACKET AND RECEIVER DATA 
 
 High-pressure jacket, Ibs. per sq.in 180.19 
 
 Intermediate jacket, Ibs. per sq.in 34.5 
 
 Low-pressure jacket, Ibs. per sq.in 0.5 
 
 First receiver drip, per hour 422.16 
 
 Second receiver drip, per hour 360.87 
 
 Jacket drip, per hour 660.62 
 
 Total drip, per hour 1443.65 
 
 Per cent of steam passing throttle -.- 16.16 
 
 TEMPERATURES 
 
 In exhaust pipe, 4 ft. below L.P. cylinder 113 F. 
 
 Water pumped 40 
 
 Condensing water '. 40 
 
 Water leaving condenser 7 
 
 Water after passing feed heater in exhaust pipe 85 
 
 Air in engine room at mercury column 80 
 
 Outside average 3 
 
 CALORIMETER 
 
 Taken after test, instrument between throttle and cylinder: 
 
 Pressure, Ibs J 7 6 -7 
 
 Temperature 297.25 F. 
 
 Moisture, per cent I - I 33 
 
426 
 
 PUMPING MACHINERY 
 
 EVAPORATION 
 
 Water pumped to boiler per hour, Ibs 9265.625 
 
 Boiler leakage per hour, Ibs 196.392 
 
 Water evaporated per hour 9069.233 
 
 Engine room drip per hour 135.270 
 
 Steam required to run drip pump per hour. 
 
 + 57-75 
 
 Net steam charged to engine per hour 8991.713 
 
 Boiler horse-power developed 324 
 
 Boiler rated horse-power 500 
 
 Head End L.P. 
 10.07 
 
 Crank End L.P. 
 10.11 
 
 To accompany Report of Daty rrui ut 11. 
 Point Pumping Station dated March 21th 1910 
 Cards taken 10.30 P.M. 3-9-10 
 
 H.P. Water 
 C2.C2 
 
 Int. Water 
 61.00 
 
 L.P. Water 
 63.40 
 
 FIG. 337. Indicator Cards from Lardner's Point. 
 
TEST OF PUMPING ENGINES 
 
 427 
 
 " The boiler used was an Edgemoor water tube boiler No. 26 
 of the same size and type as used on the previous tests. 
 
 ' ' It was connected to the engine by a special six inch steam 
 pipe which was well covered. 
 
 ' ' Indicator cards are arranged to correspond to the position 
 of the cylinders from which they were taken. These are shown 
 in Fig. 337. 
 
 "The distribution of -power between the different cylinders 
 as determined from the cards is shown on the following table: 
 
 Steam. 
 
 Head. 
 
 Crank. 
 
 Total. 
 
 Per cent of 
 Total. 
 
 High-pressure cylinder . . 
 Intermediate cylinder. . . 
 Low-pressure cylinder. . . 
 
 163.47 
 !35-35 
 130.10 
 
 168.55 
 J33-75 
 
 130. 12 
 
 332.02 
 269 . 10 
 
 26O . 22 
 
 38.6 
 3!-3 
 3 1 - 1 
 
 428 . 92 
 
 432.42 
 
 861.34 
 
 100 . 
 
 Mechanical efficiency, ratio of net delivered water horse- 
 power to indicated steam horse-power 95-54 
 
 Horse-power from volume and head of water 822 . 93 
 
 Water horse-power indicated . 839 . 8 
 
 Excess of indicated over delivered water H.P 2.1% 
 
 Steam passing throttle per indicated H.P. per hour, Ibs... 10.37 
 
 Steam passing throttle per delivered water H.P. per hour, Ibs . 10 . 86 
 Heat units from steam pressure to vacuum temperature used 
 
 per I. H.P. per minute 193 .04 
 
 Heat units passing throttle from steam pressure to vacuum 
 
 temperature used per delivered water H.P., per minute. 202.05 
 Efficiency from heat in steam delivered to engine above tem- 
 perature in exhaust pipe to work done in discharge main . 20 . 99 
 
 NOTE. In the test the reheating coils in both receivers were out 
 of service." 
 
 This test represents one of the highest duties obtained on 
 pumping engines and should be studied, as many results ^an be 
 obtained from the data. 
 
CHAPTER X 
 HIGH DUTY PUMPS AND WATER WORKS STATIONS 
 
 To study the modern form of the water works high duty 
 pump and its installation a number of examples will be given. 
 
 The Western Pumping Station of the Water Works of Cincin- 
 nati, Ohio, is shown in the Figs. 338 to 341. The station 
 at present contains three 25,000,000 gallon pumps for lo;v 
 service and three 12,000,000 gallon pumps for high service, 
 built by the Holly Manufacturing Co. of Buffalo, N. Y. The 
 station, Fig. 339, contains space for an additional pump of each 
 kind. The water enters through a gravity tunnel to a well, and 
 from this point a conduit of pipe extends to the suction of the 
 various units. Two suction lines are taken off the main 
 conduit at each pump, one of them passing through the con- 
 denser, and these enter the suction valve chambers, as will be 
 seen later, and finally connect together in a suction air chamber 
 at the end. The two discharges from each side of the pump 
 are connected into the discharge main, which is connected to two 
 service mains leaving the station on each side of the center. 
 
 The branches from the sides of the pump may be cut off 
 by valves which are shown by the conventional cross, and by 
 valves in the service lines; any of these may be cut off from 
 the station. Valves in the main discharge line of the station 
 make it possible to use certain of the pumps on one service main. 
 Such arrangements are valuable at times for service and for 
 experimental work. Attention is called to the fittings left for 
 the installation of the new pumps. 
 
 The steam is generated in the boiler room in which the 
 boilers are placed on the side away from the pump room. This 
 makes a longer steam line, but with it the chimneys can be 
 located in a better manner. The open space in front of the 
 
 428 
 
HIGH DUTY PUMPS AND WATER WORKS STATIONS 429 
 
 boilers is so wide that tubes may be removed from the boilers. 
 This space forms a convenient place for firing. The boilers are 
 
 FIG. 338. Western Pumping Station 
 
 connected to a steam header carried on the wall between the 
 boiler room and pump room and branches lead to the engine. 
 
430 
 
 PUMPING MACHINERY 
 
FIG. 340. Front 
 
of Holly Cincinnati Pump. 
 
 (To face page 431) 
 
HIGH DUTY PUMPS AND WATER WORKS STATIONS 431 
 
 Each branch from the boiler to the steam main has two valves, 
 and the steam main has a number of valves distributed along its 
 length so that certain boilers may be -used with a given pump if 
 desired. The branch leading to each pump is controlled by a 
 valve just inside the engine room for emergency while the 
 main stop valve is placed at the point where the steam enters 
 the high pressure cylinder. The former valve is not very acces- 
 sible, while the other valve used in starting or stopping the 
 pump is controlled from all platforms or the main floor. It is 
 better to place the high-pressure cylinder next to -the boiler 
 room wall, but conditions may arise so that this is not advisable. 
 With the water suction pipe" in the position shown in the plan of 
 the station the condenser must be placed on this end, and hence 
 the high-pressure cylinder is placed away from the boiler room 
 and the reason for the use of the longer pipe is evident. 
 
 The feed water for the boilers is taken from the condenser 
 and then pumped through economizers by the pumps shown in 
 the boiler house. These economizers are placed behind the 
 boilers in the flues leading to the stacks. The valves controlling 
 the supply of water to these and the method of discharging 
 around the economizers may be seen. 
 
 If forced draft is required to make up the draft reduction 
 by the economizers, blowers may be placed in the boiler room, 
 forcing air beneath the boilers. This keeps the air in the boiler 
 room in circulation, renewing if from the outside. 
 
 The small electric generators placed in the center of the 
 building are used for lighting and for power for the cranes and 
 shops. A crane, the runway of which is seen in Fig. 338, serves 
 to handle the parts in erection or repair. 
 
 Fig. 339 shows clearly the arrangement of offices, machine 
 shop, store room, tool room and toilet rooms. There are two 
 rooms, one for the engineers and oilers and one for the firemen. 
 These rooms should be large, airy and well ventilated. 
 
 A study of this plan will reveal the many good features of 
 the design. 
 
 The pumps shown in Figs. 340 and 341 for this station 
 illustrate a type of high duty pump. 
 
432 
 
 PUMPING MACHINERY 
 
 END ELEVATION-LOW SERVICE 25 MILLION GALLON ENGINE. 
 
 FIG. 341. Holly Cincinnati Pump. 
 
 Water enters from A, passing the valves D on each side of 
 the pump and entering the pipes B, which are continued through 
 
HIGH DUTY PUMPS AND WATER WORKS STATIONS 433 
 
 the valve chambers HH to the end of the pump, where they 
 combine and enter the suction air chamber E. The water is 
 forced out by the plungers, as described on page 286, and is 
 passed through the pipe / which is made a portion of the upper 
 part of the valve chamber. The water from each side passes 
 through check valves F and finally enters the discharge pipe at 
 G. The discharge check valve is provided with a small by-pass 
 valve for priming the pump. 
 
 The tops of the valve boxes form air chambers and these 
 are supplied with compressed air when necessary through the 
 pipe K. The steam end is triple expansion, with the three 
 cylinders arranged in succession. The 
 cross heads are equipped with four 
 rods extending over the crank and 
 shaft to the plunger cross head below. 
 The end cranks are overhung but 
 carry return cranks which are con- 
 nected to the end cranks on the 
 shaft M. These return cranks are 
 so placed on the pins that they turn 
 the shaft M positively, that is, when 
 the cranks on one end are passing a 
 dead point for the connecting rod, the 
 others are at maximum throw. The 
 shaft M is used to operate eccentrics, 
 the motions of which serve to oper- 
 ate the Corliss valves of the cylinders. 
 
 In high duty engines the cylin- 
 ders are usually jacketed and in 
 many cases the receivers between the 
 cylinders are heated by coils. Fig. 
 342 shows the internal arrangement 
 of a reheating receiver. The coil is 
 
 drained by means of a trap and this 
 
 . . , FIG. 342. Reheater Receiver. 
 
 must act positively. 
 
 Fig. 343 is the interior of the Central Park Avenue Pumping 
 Station in Chicago, which is similar to the Springfield Avenue 
 
434 
 
 PUMPING MACHINERY 
 
 FIG. 343. Central Park Avenue Station. 
 
 FIG. 345. Worthington Pumps at Fall River, Mass. 
 
N 6'Steam Pipe to Frest pt Header 
 
 42 Delivery Elev.+20.5J| 
 
 6 Steam Pipe to Present Engine 1202 
 
 FIG. 344. Word 
 
Gallery Elev.+27V 
 
 ijrh Water Kiev. + 5 
 
 Bottom of Pit Elev.-18'O" 
 
 
 n High Duty Pump. 
 
 (To face page 435) 
 

 
 
HIGH DUTY PUMPS AND WATER WORKS STATIONS 435 
 
 station. The pumps at the latter, shown in Fig. 344, developed 
 a duty of 174,735,801 foot pounds per 1,000 pounds of steam. 
 Water is brought to this station through a long tunnel from 
 Lake Michigan and discharges into a well built in the room. 
 The suction pipes are provided with foot valves and the surface 
 
 FIG. 346. Allis Pumps at the Baden Station. 
 
 condenser uses water which is taken into the suction for condens- 
 ing water. The pump, as described earlier, is a duplex, triple 
 expansion, direct-acting pump, supplied with compensators. 
 The cylinders are all in tandem and the weight is carried by a 
 balancing piston so connected to the discharge main, as shown in 
 Fig. 344, that should a break occur in the main the pump will 
 come to rest. These pumps may be placed in a small floor area. 
 
436 PUMPING MACHINERY 
 
 Fig. 345 shows a station equipped with a Worthington 
 horizontal, triple-expansion, direct-acting engine. This pump 
 was described on page 100. 
 
 FIG. 347. Allis Pumps at Bissel's Point. 
 
 Fig. 346 shows the appearance of the Allis-Chalmers pump, 
 installed at the Baden station of the St. Louis water works, 
 while Fig. 347 shows their pump at the Bissel's Point station 
 
HIGH DUTY PUMPS AND WATER WORKS STATIONS 437 
 
 of th-j same system. In Fig. 347 the return crank driving the 
 valve shaft on the first balcony is clearly seen. The valves of 
 both these pumps are of the Corliss type on the high-pressure 
 
 FIG. 348. Holly Pump at Boston. 
 
 and intermediate-pressure cylinders, but on the low-pressure 
 cylinder poppet valves are used. 
 
 Fig. 348 shows a pump built by the Holly Company for the 
 Spot Pond Station of the Boston water works. In this pump 
 the valve or eccentric shaft is driven by bevel gears. The exhaust 
 valve of the intermediate-pressure cylinder of this pump and 
 
438 
 
 P UMPI NG M A CHI NER Y 
 
 both valves of the low-pressure cylinder are of the poppet type. 
 The figure shows the various gauges and valves as well as the 
 governor which actuates the cut-off on the high-pressure and 
 
 FIG. 349. Allis-Chalmers Pump. 
 
 intermediate-pressure cylinders. In all these pumps the bracing 
 of the frame is to be noted. 
 
 Fig. 346 shows clearly the type of complete frame often 
 used with these pumps and in them the valve chambers are 
 separate from the frame. In Fig. 349, showing an Allis-Chalmers 
 pump, the valve chambers are used for supports. In the Holly 
 
HIGH DUTY PUMPS AND WATER WORKS STATIONS 439 
 
 pump of Fig. 340, the separate air chambers are used, the bed 
 plate at the floor line being carried by the inclined frames. 
 A small 2,500,000 gallon pump built by the Holly Manu- 
 
 FIG. 350. Holly Pump at Washington, D. C. 
 
 facturing Company for the Trumbull Street Pumping Station of 
 Washington, D. C., is shown in Fig. 350. This pump gave a duty 
 of 164,644,000 foot-pounds per 1,000 pounds of dry steam. 
 
440 
 
 PUMPING MACHINERY 
 
 The valve shaft in this engine is oscillated by an eccentric on the 
 main shaft. A governor controls the cut-off on the high-pressure 
 
HIGH DUTY PUMPS AND WATER WORKS STATIONS 441 
 
 cylinder, but the cut off on the other cylinders is controlled by 
 small hand wheels seen beneath the upper platform. All of the 
 valves are of the Corliss type. 
 
 Fig. 351 represents the arrangement of a station in 
 Brooklyn, N. Y. The boiler 
 houses are placed at the ends of 
 the pump house and the steam is 
 taken to the engines by the main 
 steam supply. The water ends 
 are connected to a suction main 
 on one side of the building and 
 discharges are carried from each 
 pump separately. In the rooms 
 on the side of the pump house, 
 as indicated by outline only, are 
 placed the offices, shops, toilets 
 and lockers, while the houses 
 beside the boiler rooms are 
 intended for coal storage. 
 
 The storage of coal is import- 
 ant in all plants for the supply 
 of water or for any other pub- 
 lic service. This matter cannot 
 receive too much attention. A 
 supply for a week's consumption 
 should be provided for and if 
 conditions of coal delivery are 
 poor this should be increased. 
 In some cases a month's supply 
 is none too much. The coal 
 should be stored beneath cover 
 as the alternate wetting and dry- 
 ing when uncovered causes the 
 coal to deteriorate. With soft 
 coal the pile should be ventilated. FlG ' 35 2 -- Mem P his Station - 
 
 The arrangement of the station with the boiler room at one 
 end is not advisable if conditions are such that the arrangement 
 
442 PUMPING MACHINERY 
 
 shown in Fig. 339 is possible. When possible an arrangement 
 with the pump room and boiler room beside each other will 
 make a better plan for economy and for ease of superintending. 
 In this case the plant is more compact and in a short time the 
 chief engineer may have a complete view of the whole plant. 
 
 Fig. 352 shows a plant at Memphis, Tennessee, where the first 
 vertical Worthington high duty pumps were installed. The 
 plant was limited in ground area as a given capacity had to be 
 placed in this space. The vertical high duty pump installed 
 was of the compound duplex type, and on test in 1891, this pump 
 gave 117,325,000 foot-pounds per 1,000 pounds of steam. The 
 design was well thought out and as shown four 10,000,000 gallon 
 pumps were placed in a pit 38 feet in diameter. The steam 
 cylinders were 30 and 60 by 48 inches, while the water cylinders 
 were 28 by 48 inches. The suction well is shown in the figure. 
 The pit is founded on a heavy concrete base and the wall is 
 built circular so that it need not be as thick as an ordinary 
 letaining wall. % 
 
 Fig. 353 shows the arrangement of the water works at 
 Zurich, Switzerland, as described in the Engineering News, Vol. 
 32, page 34. This plant is operated by turbines, the vertical 
 shafts of which drive the bevel wheels E, which arc placed on 
 short counter-shafts. These counter-shafts drive the main shaft 
 B by means of the gears D, and the gears C drive the shafts of 
 the pump which are blocked in at A A. The pumps are of the 
 two-crank type, a crank being at each side of the pump with the 
 pins quartering. These pumps are connected to four different 
 services so that any pump can be used for any pressure. Of 
 course this means that the pump power will change and this 
 fact explains the reason for the long jack-shaft. A pair of pumps 
 which would be operated by a pair of turbines properly at one 
 pressure would not operate if it were desired to use a higher 
 pressure. When such is the case, the additional power is obtained 
 from other turbines. Although the use of the jack-shaft means 
 a certain amount of friction, the flexibility of the station is 
 increased so much by it that the installation is proper. 
 
 A small auxiliary boiler and engine may be used when the 
 
HIGH DUTY PUMPS AND WATER WORKS STATIONS 443 
 
 water wheels are out of commission. These are shown at H and 
 G, and at WW are high pressure water wheels which may be 
 
444 
 
 PUMPING MACHINERY 
 
 run from the high-pressure reservoir when needed to operate 
 electric generators 00 for lighting at night. The station also is 
 connected to a rope transmission at R. 
 
 Fig. 354 shows the arrangement of air compressors for the 
 Harris pump described in Chapter XIII. The steam cylinder 
 
 FIG. 354, Harris Air Pump Station at High Bridge. 
 
 A receives steam from a boiler house not shown, while the air 
 cylinder B discharges its air into one of the pipes CC and draws 
 its supply from the other. The action of the pump is described 
 later, but the plant is shown at this point so that the arrange- 
 ment of machinery may be seen. 
 
 The station shown in Figs. 355 and 356 is located in Mil- 
 
HIGH DUTY PUMPS AND WATER WORKS STATIONS 445 
 
 waukee, Wis., and is used for flushing the Milwaukee river. 
 In the tower room are the "cylinders of the steam engine, which 
 is used to drive a screw pump which lifts 30,000 cubic feet of 
 water (224,000 gals.) per minute against a head of 3^ feet. The 
 water is drawn through a tunnel from Lake Michigan and 
 forced into Kinnickinnic River. The tunnel passes under the 
 power house and turns after passing the screw of the pump. 
 
 FIG. 355. Kinnickinnic River Station. 
 
 The boiler house is equipped with two return tubular boilers 
 furnished with superheaters. The engine is a tandem compound 
 engine and the screw which was shown in Fig. 107 is I2-J- feet in 
 diameter. 
 
 The boiler room and coal room are seen in the plan together 
 with toilet room and locker room. 
 
 The Lardner's Point Pumping Station of the city of Phila- 
 delphia consists of three separate pumping stations, Figs. 357, 
 358 and 359. " The first is an old station formerly termed the 
 " Frankford Pumping Station " and was used in the old system 
 to pump water from the Delaware River to the Frankford dis- 
 
446 
 
 PUMPING MACHINERY 
 
 tribution system. It is now termed No. i House, and the con- 
 nection with the river has been closed and a new connection 
 made to the filtered water conduit leading from the outlet 
 shaft of the Torresdale conduit from the filter beds. 
 
 " No. i pumping station, or the old ' Frankford,' consists 
 ^f one compound vertical Cramp pump of ten million gallons 
 capacity, one Wetherill horizontal, ten million gallons capacity, 
 
 FIG. 356. Kinnickinnic River Station. 
 
 one Southwark vertical, twenty million gallons capacity and one 
 Southwark vertical horizontal, fifteen million gallons capacity. 
 For this station there are twelve marine type boilers of 200 H.P. 
 capacity each. 
 
 ' ' Two entirely new stations were constructed and contain 
 twelve (12) vertical triple expansion Holly pumping engines of 
 twenty million gallons daily capacity each. The engine rooms 
 are built separate from the boiler houses and are 171 feet long by 
 87 feet wide, constructed of gray standard size brick trimmed 
 
HIGH DUTY PUMPS AND WATER WORKS STATIONS 447 
 
448 
 
 PUMPING MACHINERY 
 
 FIG. 358. Lardner's Point Station. 
 
 FIG. 359. Lardner's Point. 
 
 with granite and terra cotta. All the roof coverings are of red 
 tile. The water ends of the pumps are set in the basement under 
 
HIGH DUTY PUMPS AND WATER WORKS STATIONS 449 
 
 the floor of the engine room, and the entire steam ends are all 
 above the floor level. The pump well is located under the 
 basement floor in the center of the engine houses, extending 
 their full length. It is constructed of reinforced concrete, horse- 
 shoe shaped in section, 14 feet in width and height. 
 
 "Between engine houses Nos. 2 and 3 a gate chamber is 
 located which controls the discharge from the larger connection 
 to the outlet shaft of the Torresdale conduit. It is connected 
 to the pump well of both houses, and gates have been installed 
 for connecting the pump well of a future house to be located 
 west of the present plant. 
 
 * ' The boiler houses of the new stations are of the same general 
 architecture and contain the following: 
 
 House No. 2\ 
 
 Six Edgemoor water tube boilers, 500 horse-power 
 
 capacity each, equipped w.ith Wetzel stokers. 
 Twelve Internal-fired tubular boilers, 200 horse-power 
 capacity each. 
 
 House No. j: 
 
 Eight Edgemoor boilers, 500 horse-power capacity each, 
 equipped with Wetzel stokers and two Green econ- 
 omizers. 
 
 " In the annex of boiler house No. 2 are three 50 K.W. 
 generators each, furnishing light for the entire station and 
 current for the coal handling machinery, electric crane, etc. 
 
 ' ' The smoke flues of the boiler houses are connected to four 
 brick chimneys 150 feet high and 7 feet internal diameter two 
 for No. 2 and No. 3 houses of the Custodis and Heinicke type 
 respectively. 
 
 "The coal is delivered to overhead pockets of 3,000 tons 
 capacity in the boiler houses of Nos. 2 and 3 stations by means 
 of a tower and belt conveyor with capacity for handling 50 tons 
 per hour. Coal may be received either by rail or by boat, and 
 the general design of this equipment is similar to the one at 
 Torresdale. 
 
 "Data relating to engines and boilers in Nos, 2 and 3 houses, 
 Lardner's Point pumping station; 
 
450 PUMPING MACHINERY 
 
 ENGINES 
 
 Nominal capacity of each engine 20,000,000 gallons daily. 
 
 Number of revolutions per minute 20 
 
 {Stroke si ft. 
 
 Piston speed, feet per minute 220 
 
 High. Intermediate Low. 
 
 Cylinder diameter 32" 60" go" 
 
 Diameter piston rod i\" 7$" 7$" 
 
 First. Second. 
 
 Receiver volume 205 cu.ft. 304 cu.ft. 
 
 Receiver heating surface 166 sq.ft. 304 sq.ft. 
 
 Diameter. Length. 
 
 Cross-head pins 12" n" 
 
 Crank pins 12" n" 
 
 Shaft bearings 17$" 32" 
 
 Shaft at center 2o" 
 
 Distance rods. Four (4) each, 5 inches diameter. 
 
 Air pump. One (i) 28-inch diameter, 66 inches stroke. 
 
 Feed pump. One (i) 3^-inch diameter, 66 inches stroke. 
 
 Feed water heater. One (i) in exhaust, 308 sq.ft. 
 
 Flywheels. Two (2-) 20 feet diameter and weighing 32 tons each 
 
 (approximate) . 
 
 Throttle valve. 8 inches diameter. 
 Exhaust pipe. 24! inches diameter. 
 
 Suction pipe. Main 42 inches diameter, branch 30 inches diameter. 
 Discharge pipe. Main 42 inches diameter, branch 30 inches diameter. 
 Suction injection. 8 inches and 10 inches diameter. 
 Force injection. 3 inches and 3^ inches diameter. 
 Overflow. 18 inches diameter. 
 Diameter of plungers. 33 inches. 
 
 No. 2 House. No. 3 House 
 Number of pump valves 960 864 
 
 BOILERS 
 Furnace Flue Tubular: 
 
 Number of boilers 12 
 
 Diameter of shell 108 ins. 
 
 Length of shell 20 ft. 
 
 Thickness of shell If in. 
 
 Number of fire boxes . 2 
 
 Diameter in inches 41 
 
 Length of fire boxes 8 ft. 
 
 Number of tubes 195 
 
 Length of tubes 9 ft. 3 ins. 
 
 Diameter of tubes 3 \ ins. 
 
 Length of grate 5 ft. 5! ins. 
 
 Area of grates 40 J sq.ft. 
 
HIGH DUTY PUMPS AND WATER WORKS STATIONS 451 
 
 Water Tube Boilers: 
 
 Number of boilers 14 
 
 Stoker Wetzel. 
 
 Number of tubes 252 
 
 Length of tubes 18 ft. 
 
 Diameter of tubes 4 ins. 
 
 Steam drums (two) 36 ins. diam. 
 
 Length of steam drum 2 1 ft. 
 
 Length of grate * 8$ ft. 
 
 Area of grate 102 sq.ft. 
 
 Fig. 360 is an exterior view of the Queen Lane station of 
 the Philadelphia water works and is shown to- illustrate the 
 appearance of a station. 
 
 Fig. 361 shows a Riedler water works pump used in the 
 
 FIG. 360. Queen Lane Station. 
 
 water works of the city of Berlin, Germany. The large positively 
 closed Riedler valves are seen and the tubes beneath the suction 
 valves prevent the suction from breaking and make the upper 
 part of the suction chamber an air chamber. The plunger is 
 passed through a long sleeve and although different from the 
 Worthington plunger and ring, it is the equivalent of it. The 
 large air chambers on the discharge reduce the variation in 
 water pressure. 
 
452 
 
 PUMPING MACHINERY 
 
HIGH DUTY PUMPS AND WATER WORKS STATIONS 453 
 
 The steam cylinders show the poppet valve form of cylinder 
 so -extensively used in Europe. 
 
 One of the Jersey City water works stations is equipped 
 with this type of pump. The station is shown in Fig. 362. 
 
 The pumps are driven by turbines or steam cylinders placed 
 in tandem with the water ends. There are three units installed 
 in the station, two being high-pressure pumps and one a low- 
 pressure pump. The suction is taken from the fore bay by 
 pipes on each side of the unit and the two discharge air chambers 
 are connected to the discharge main, which is carried to the 
 end of the building. 
 
 Name. 
 
 Station. 
 
 Date. 
 
 Duty per 1,000 
 Ibs. Dry Steam. 
 
 Savery 
 
 
 
 5,000,000 
 
 Smsaton Newcomen . . 
 Watt 
 
 
 
 
 
 I2,OOO,OOO 
 20,000,000 
 
 Cornish engine 
 Simpson's 
 
 
 
 
 
 60,000,000 
 8^,000,000 
 
 Holly Quadruplex . , . 
 Gaskill 
 
 
 
 
 
 85,000,000 
 1 1 7,936,698 
 
 Worthington 
 
 
 
 120,000,000 
 
 Reynolds 
 
 
 
 1 18,327,041 
 
 Reynolds-Hannibal. . 
 Reynolds 
 Reynolds 
 Worthington (Direct) 
 W 7 orthington (Direct) 
 Allis-Chalrners . . 
 
 North Point. 
 St. Louis. 
 Chicago. 
 Chicago. 
 St. Louis. 
 
 Nov. 3-4, 1903 
 Aug. 27-28, 1902 
 
 154,048,700 
 154,048,704 
 179,454,250 
 161,676,942 
 I74,734,8oi* 
 181,000,000 
 
 Holly 
 
 Washington, D.C. 
 
 Nov. 2-3, 19015 
 
 162,644,000 
 
 Holly 
 
 Cleveland, Ohio. 
 
 Dec. 5, 1907 
 
 164,642,226 
 
 Holly 
 
 Brockton, Mass. 
 
 Oct. 14-15, 1909 
 
 170,000,000 
 
 Holly 
 
 Louisville, Ky. 
 
 May 1-2, 1909 
 
 195,020,000* 
 
 Holly 
 
 Albany, N. Y. 
 
 May 29-30, 1909 
 
 182,281,000 
 
 Snow 
 
 Mahanoy City, Pa. 
 
 Nov. 12, 1906 
 
 141,369,439 
 
 Allis-Chalmers 
 Allis-Chalmers 
 Holly 
 
 St. Louis. 
 Boston, Mass. 
 Philadelphia, Pa. 
 
 Feb. 26, 1900 
 May 1-2, 1900 
 Mar. 9, 1910 
 
 179,454,255 
 178,497,000 
 184,476,200 
 
 Worthington (Direct) 
 Worthington (Direct) 
 
 Fall River. 
 Montreal. 
 
 Sept. 23, 1909 
 Nov. 27, 1909 
 
 136,500,000 
 177,538,000* 
 
 *Superheat 
 
 The 12 inch steam pipe and the 24 inch exhaust pipe are 
 carried in a trench on one side of the station and the exhaust 
 from all the engines is cared for by one large jet condenser. 
 
454 
 
 PUMPING MACHINERY 
 
HIGH DUTY PUMPS AND WATER WORKS STATIONS 455 
 
 In such a plant the engines are only used in case of low water, and 
 at other times the pump rods are disconnected from the steam 
 piston rod so that the piston does not act as a resistance to 
 motion. 
 
 To give some idea of the value of the duty of modern pump- 
 ing engines, see table on page 453. 
 
CHAPTER XI 
 SPECIAL PUMPING MACHINERY 
 
 THERE are several special pumps used to handle liquids 
 which should be discussed. These, although the same in principle 
 as those which have been described, have particular features 
 which make them noteworthy. 
 
 Condenser Pumps are of two kinds: Circulating pumps and 
 air pumps. The circulating pump is used to force the condensing 
 water through the condenser, while the air pump is used to 
 remove the condensed steam and air from the steam side of a 
 condenser, or, in the case of a dry air pump, the pump handles 
 air and water vapor only, the condensed steam being removed 
 from the condenser by a wet pump or by gravity as in a barome- 
 tric condenser. The circulating pump in many cases lifts the 
 water a short distance only, and for that reason it is built as a 
 tank pump, while in other cases it may be used to overcome 
 the friction of a long pipe line to and from the cooling tower 
 as well as the lift to the top of the tower, in which case it must 
 be heavier. 
 
 The circulating pump most often takes the form of a centrif- 
 ugal pump, as it has to lift large quantities of water, and space 
 is limited. 
 
 Fig. 363 shows a combined air and circulating pump of the- 
 Wheeler Condenser and Engineering Company. In this pump 
 the two water cylinders are placed in tandem with the steam 
 cylinder which takes the form of any of the simplex pumps. 
 
 In Fig. 363 the Kno\vles pump is shown. The circulating 
 pump forms one end support for the condenser. The water is 
 discharged through A into one set of tubes and then it retunis 
 through B and the upper set of tubes to C, where it discharges. 
 The air pump forms the other support for the shell. It takes the 
 air and water from the condenser and discharges it through 
 /), The suction space F is connected to G. 
 
 456 
 
SPECIAL PUMPING MACHINERY 
 
 457 
 
 To find the size of the water and air ends of the pump, suppose 
 that W pounds of steam per hour at a pressure p are to be 
 condensed. If r is heat of vaporization of the steam, x its 
 quality, 4 the temperature of the condensed steam, and q 
 the heat of the liquid, and if G pounds of water entering at 
 
 
 
 FIG. 363. Combined Air and Circulating Pump. 
 
 and leaving at t Q F are to be used, G is given by the equation: 
 
 qt -qa 
 
 If the number of revolutions or double strokes N are assumed, 
 the displacement of the water end will be 
 
 G 
 
 The air end of the pump is made in many cases of empirical 
 design. Some authors give ratios of volume displaced by the 
 pump per minute to the volume of the condensed steam or to 
 the volume of the low-pressure cylinder of the engine which is 
 discharging into the condenser. Several of these are mentioned. 
 
458 PUMPING MACHINERY 
 
 RATIO OF Am CYLINDER DISPLACEMENT TO LOW-PRESSURE 
 
 CYLINDER. 
 
 Single-acting vertical pump surface condenser i 9 
 
 Single-acting vertical pump jet condenser i 13 
 
 Double-acting horizontal pump surface condenser. ... i 15 
 
 Double-acting horizontal pump jet condenser i 12 
 
 Double-acting horizontal pump-compound engine sur- 
 face condenser i 26 
 
 Single-acting horizontal pump-compound engine sur- 
 face condenser i 16 
 
 RATING OF AIR CYLINDER DISPLACEMENT TO VOLUME OF 
 CONDENSED STEAM. 
 
 Surface condenser 20 : i 
 
 Jet condenser 40 : i 
 
 This may be a satisfactory way, but it is better to estimate 
 the volume from the air probably present. Water usually con- 
 tains air to about one-fifteenth of its volume. This amount of 
 air is at atmospheric pressure p a and it must be cared for by 
 the air pump at a reduced pressure. In addition to this there 
 are small leaks in the pipe line which allow more air to enter. A 
 small hole will destroy the proper acting of the air pump. The 
 author at one time desired to run a 2000 H.P. condensing engine 
 with no vacuum and still use the condenser to operate so as to 
 measure the steam consumption, and he found that a 2-inch plug 
 removed from a 30 inch exhaust main was sufficient to destroy 
 the vacuum completely with the air pumps running at full 
 speed. To find the volume of air per minute the following 
 formula will be used, allowing 100 per cent for leakage. 
 
 p= absolute pressure in the condenser, Ibs. per sq.in. 
 p s = vapor tension or absolute steam pressure correspond- 
 ing to T c 
 
 T c = absolute temperature in condenser 
 T a = absolute temperature atmosphere. 
 W = weight of steam per hr. or weight of water and 
 steam per hr. in a jet condenser. 
 
SPECIAL PUMPING MACHINERY 
 
 459 
 
 This equation shows the importance of making p t as different 
 from p as possible. The terms p and p s do not differ much, and 
 by taking the mixture of air and vapor on its way to the air 
 pump, through as cold a passage as possible, the term p a is 
 made smaller and the denominator is increased, making V small. 
 This is the reason for the great advantage in counter current 
 for condensers, and even in the condenser, shown in Fig. 363, 
 the coldest water should enter directly over the air pump inlet 
 so as to cool the mixture going to the pump. 
 
 From the volume thus com- 
 puted the displacement of the 
 air pump is given by: 
 
 Knowing the displacements 
 of these pumps a stroke may be 
 assumed, and from it the area 
 determined. 
 
 I 1 
 
 Aap 
 
 D, 
 
 ap 
 
 FIG. 364. Cards from Air Pump and 
 Circulating Pump. 
 
 The cards from the water end 
 are shown in the lower part of 
 Fig. 364, while those for the air end are shown above. The 
 combination of these or the addition of them when reduced to 
 the proper scale, on account of the difference in piston area, 
 will give- the total work, and from this the size of the steam 
 cylinder is given, if the M.E.P. be found for a given boiler 
 pressure. Allowing 33 per cent for friction, which is made large 
 to give a certain driving power, the following results: 
 
 (M.E.P. 
 
 (M.E.P. ) P A P 
 
 (i.oo-o. 33 )(M.E.P.) 8 
 
 Separate air pumps are often used. Fig. 365 shows a steam 
 driven pump used in the U. S. Navy. This air pump is 
 
460 
 
 PUMPING MACHINERY 
 
 FIG. 365. Air Pump. 
 
SPECIAL PUMPING MACHINERY 461 
 
 made with two air cylinders driven through gears from a 
 steam cylinder placed on one side of a pump barrel. The 
 pump is of the bucket type with foot valves A A and head 
 valves at B. These with the valves in the bucket at C are 
 all spring-controlled metal valves. The foot valves are 
 placed on an inclined partition for the purpose of making it 
 easier to discharge the air when the piston rises and forms 
 a vacuum. The lip around the discharge valve matyes a dam 
 and covers the valve with water. This makes thenr air tight. 
 The other valves are in that condition, sirice all of the water 
 on the bucket or that over the foot valves can not be driven 
 out, as the valves limit the motion of the bucket. On the down 
 stroke of the bucket the pressure in the space above it soon 
 falls to a low vacuum because it had been completely filled 
 with water; this, then, causes the valves to open and take air 
 from the lower portion of the cylinder. The air in the water 
 also separates and rises to the top of the cylinder. Finally 
 trie bucket reaches the water below, and this is driven through 
 the 'valve openings whicli are uncovered. It is seen that the 
 air leaves first in this case; the water is struck by the bucket 
 surface aftd will cause considerable shock if the pump is running 
 too, rapidly. 
 
 To do away with shock and to decrease valve resistance, the 
 Edwards Air Pump, Fig. 366, was introduced. In thisvair pump 
 water and air enter the space A at the bottom of the pump 
 which is made conical in form. The piston B which is driven 
 from the steam piston in / by two rods CC extending over 
 the shaft and crank, is" provided with a conical bottom. As 
 this piston descends there is a vacuum produced:, so that when 
 the top of the piston uncovers the openings EE, air enters from 
 the space A around the cylinder barrel, and as the conical bottom 
 enters the water in the bottom of A, this is forced around the 
 curved passage and discharged into the openings at E. This 
 continues even after the piston starts up, as the momentum of 
 the water continues its motion. This discharge of water into 
 the openings as the piston is moving upward acts as a valve to 
 keep the air from coming ouTas the piston ascends. In a short 
 
462 
 
 PUMPING MACHINERY 
 
 time, however, the piston covers the ports or openings E and 
 then the air and water are compressed until the pressure is 
 sufficient to overcome the atmospheric pressure on the head 
 valves at H, which are drowned by the use of a lip around the 
 valve deck. The piston rods CC are carried through long-sleeve 
 
 FIG. 366. Edwards Air Pump. 
 
 s tuning boxes so arranged that the point H, at which leakage 
 could occur, is water sealed, leaving only one stuffing box at 
 the plate K to care for. This is a simple matter. 
 
 The Mullen Valveless Air Pump, Fig. 367, is somewhat 
 similar to the Edwards Pump. In this case, the deep piston 
 C, provided with a number of packing grooves, uncovers ports 
 
SPECIAL PUMPING MACHINERY 
 
 463 
 
4G4 PUMPING MACHINERY 
 
 at the center of the cylinder. The vacuum formed by the motion 
 of the piston from the end draws in the air and water from 
 the space A around the cylinder barrel and the return of the 
 piston cuts these ports off and compresses the air against the 
 spring valves at the end, which open after the atmospheric 
 pressure is reached. The spaces BB unite and lead to the hot 
 well. 
 
 The piston of this pump is of such a length that when the 
 ports of A are completely uncovered, the piston just reaches 
 the end of its travel, leaving a small amount of clearance which 
 is filled with water, so that as soon as the piston leaves the end 
 of its stroke the pressure falls. . This action is the same in all 
 wet air pumps. The piston rod of the Mullen Air Pump is 
 attached to a cross head E which is connected to the cross 
 head -F of the steam end by two rods, so that the shaft and crank 
 may be cleared. The stuffing box at D is water sealed to cut 
 down the leakage of air. Air leaks, even the smallest, are to be 
 avoided on account of the low pressure in the pump. 
 
 Dry air pumps have become quite common. They were 
 first introduced for barometric condensers and afterwards for 
 use with surface condensers. As a type of this class, the Alberger 
 Air Pump, Fig. 368, is shown. In it the piston A travels from 
 end to end and is so finished on the ends that its clearance is 
 small. A rotary valve below the spring discharge valve D 
 directs the air to the suction chamber C or to the discharge. 
 This valve B is positively driven. As shown in the left hand 
 figure, the air on the right is being drawn from C. The air on 
 the left is not discharged until it has reached a pressure slightly 
 above the atmosphere, when it can raise the spring valve D and 
 escape. From this point the air is driven out as the piston 
 continues to move to the left. When the end of the stroke is 
 reached the pressure on one end is atmospheric and that on 
 the other is that of the maximum vacuum carried on the con- 
 denser. The piston will draw no more on this stroke and the 
 vacuum of the condenser would not be effected if air was allowed 
 to enter this side of the piston. The other side of the piston 
 has air at atmospheric pressure, filling the clearance volume, 
 
SPECIAL PUMPING MACHINERY 
 
 465 
 
 including the passage E. If this air remained in the clearance 
 volume, it would cut down the amount of air drawn in, as none 
 would enter until it had reached the pressure of the condenser, 
 if the suction had a valve; or, it would change the vacuum if 
 allowed to discharge back into the condenser. To obviate this, 
 the Alberger Company introduced a small cross connection 
 passage in their rotating valve which connects the two ends of 
 the cylinder when the piston is just at its dead point. This 
 arrangement is shown on the right of Fig. 368, where the piston 
 
 FIG. 368. Alberger Air Pump. 
 
 is at the right hand end of its stroke. In this position the space 
 to the right of the piston, as far as the valve B, is filled with 
 air at atmospheric pressure, while the space on the left of 
 the piston is at condenser pressure and has reached the point 
 where it will not receive any more air from the condenser. 
 If now the valve cuts off both of these ends from the air and 
 condenser but connects them through the small passage in the 
 valve, the pressure on the right will be reduced very materially, 
 as the volume of the left hand end of the cylinder is so large 
 compared with the volume of the clearance. The pressure in 
 
466 
 
 PUMPING MACHINERY 
 
 the clearance is then reduced to vacuum pressure when the 
 valve opens the right side to the condenser, and the left hand 
 end of low vacuum has received a little more air to be 
 discharged. 
 
 This valve passage therefore makes the volumetric efficiency 
 of the pump greater, although the power required to drive the 
 
 pump for a given quantity is 
 not changed. Fig. 369 shows 
 the card from such a pump 
 compared with one without 
 this arrangement. The solid 
 . curve shows how the clearance 
 FIG. 369. Alberger Air Pump Card, effect is reduced and how the 
 
 ' quantity of air handled is 
 
 made greater. The curve also shows that the power required 
 in compressors of the same displacement is increased. The 
 volume of air handled per stroke is increased from AB to CB. 
 
 FIG. 370. Sewage Pump. 
 
 The small curve at C is due to the pressure not falling to the 
 vacuum of the condenser when the cross connection is made, 
 while the rise in pressure at B, giving a starting point of com- 
 
SPECIAL PUMPING MACHINERY 467 
 
 pression above the vacuum pressure, is due to this cross 
 connection. 
 
 Fig. 370 illustrates a pump built by the Laidlaw, Dunn, 
 Gordon Co. for the pumping of sewage. The valves are made 
 large on account of the solid matter in suspension, and they are 
 placed in large valve boxes on the sides of the pump. The 
 pump is otherwise similar to any duplex compound pump. 
 
 The Underwriters' Pump, or better, the " National Standard 
 
 FIG. 371. Underwriters' Pump. 
 
 Fire Pump," to which attention has been called in Chapter II, is 
 illustrated again in Fig. 371. This figure illustrates the form 
 used in accordance with the specifications of the National 
 Board of Fire Underwriters, and to give the important parts 
 of these specifications the following has been taken from their 
 pamphlet. 
 
468 PUMPING MACHINERY 
 
 " SPECIFICATIONS OF THE NATIONAL BOARD OF 
 FIRE UNDERWRITERS FOR THE MANUFACTURE 
 OF STEAM FIRE PUMPS," EDITION OF 1908. 
 
 UNIFORM REQUIREMENTS. 
 
 The following specifications for the manufacture of Steam Fire Pumps, 
 developed from those originally drawn by Mr. John R. Freeman, are now 
 used throughout the whole country, having been agreed upon in joint confer- 
 ence by representatives of the different organizations interested in this class 
 of work. They will be known as "The National Standard," and have been 
 up to this time adopted by the following associations: 
 
 Associated Factory Mutual Fire Insurance Companies. 
 
 National Board of Fire Underwriters. 
 
 National Fire, Protection Association,,' ' 
 
 NATIONAL STANDARD SPECIFICATIONS FOR THE MANU- 
 TURE OF STEAM FIRE PUMPS 
 
 1. Workmanship, a. The general character and accuracy of foundry 
 and machine work must "throughout ' equal that of the best steam-engine 
 practice of the times, as illustrated in commercial engines of similar horse- 
 power. 
 
 2. Duplex Only. a. I Only " Standard Duplex pumps" are acceptable. 
 
 So-called "Duplex" pumps, consisting of a pair of pumps with 
 "steam-thrown valves" actuated by supplemental pistons, are not 
 acceptable. 
 
 Further, the direct-acting duplex has the great advantage over a 
 fly-wheel pump of not suffering breakage if water gets into steam cylinder. 
 
 3. Sizes of Pumps, a. Only the four different sizes given in the table on 
 page 469 will be recognized for "National Standard" pumps. 
 
 b. The tabular sizes of steam and water cylinders and length of stroke 
 have given general satisfaction and will now be considered as standard. 
 
 A steam piston relatively larger than necessary is a source of weak- 
 ness. It takes more volume of steam, and gives more power with which 
 to burst something if the throttle is opened wide suddenly during 
 excitement. 
 
 It has been common to make all fire pumps with water plunger of only 
 one -fourth the area of steam piston, with the idea that the pump could 
 thereby be more readily run at night, when steam was low. The capacity 
 in gallons is thus reduced 25 per cent as compared with a 3 to i plunger 
 en the same steam cylinders. 
 
 Often, especially with large pumps, "4 to i" construction is a mis- 
 take, and -gives no additional security, although the pump might start 
 and give a few puffs with 30 Ibs. of steam on banked fires; because, 
 if any pump of whatever cylinder ratio draws 50 or 100 horse-power of 
 steam from boilers with dead fires, it can run effectively only a very 
 short time (ordinarily, perhaps, 3 to 5 minutes), unless fires are first 
 aroused to make fresh steam to replace that withdrawn. 
 
 Steam pressures stated above must be maintained at the pump, 
 
SPECIAL PUMPING MACHINERY 
 NATIONAL STANDARD PUMP SIZES 
 
 469 
 
 Pump Sizes. 
 
 1 
 
 Capacity at 100 Ibs., 
 at Pump. 
 
 Boiler Power 
 Required. 
 
 Full Speed. 
 
 
 
 
 c 
 o 
 tfi 
 
 
 
 d 
 
 
 
 
 1 
 
 
 
 
 ft 
 
 d 
 
 X 
 
 a 
 
 
 tl 
 
 
 1 
 
 
 
 
 o 
 
 
 
 CD 4 
 
 
 g 
 
 o3 
 
 **"! . 
 
 
 
 
 
 
 M 
 
 J2 
 
 a ^j 
 
 u - 
 
 M 75 
 
 a 
 
 13 
 
 
 
 
 ~ct 
 
 "o w ' 
 
 t& . 
 
 
 
 
 
 
 
 
 
 & 
 
 t-. a 
 
 , . ^ 
 
 M i-, 
 
 o 
 
 PL! d 
 
 Id 
 
 -2 
 
 Steam. 
 
 Water. 
 
 Stroke. 
 
 About 
 
 jl 
 
 Jl 
 
 I 3 
 
 1 
 
 |J 
 
 02 
 
 Revolu 
 minu 
 
 Piston ' 
 minu 
 
 14 X 7 XI2 
 
 4 to i 
 
 Two 
 
 500 
 
 483 
 
 IOO 
 
 40 
 
 70 
 
 140 
 
 14 x 7^x12 
 
 
 
 
 520 
 
 
 
 
 
 i 6 xg xi 2 
 
 3 to i 
 
 Three 
 
 75 
 
 806 
 
 US 
 
 45 ' 
 
 70 
 
 140 
 
 l8 X2O XI2 
 
 3 to i 
 
 Four 
 
 loop 
 
 999 
 
 15 
 
 45 
 
 70 
 
 140 
 
 i8ixiojxi2 
 
 
 
 
 1050 
 
 
 
 
 
 20 XI2 Xl6 
 
 2j tO I 
 
 Six 
 
 1500 
 
 1655 
 
 200 
 
 5 
 
 60 
 
 160 
 
 to give full speed and 100 Ibs. water pressure. Pressure at boilers 
 must be a little more, to allow for loss of steam pressure between boiler 
 and pump. Pumps in poor order, or too tightly packed, will require 
 more steam. 
 
 c. Two hundred and fifty gallons per minute is the standard allowance 
 for a good i inch (smooth nozzle) fire stream. 
 
 A so-called "Ring Nozzle" discharges only three-fourths as much 
 water as a smooth nozzle of the same bore, and is not recommended. 
 
 From fifteen to twenty automatic sprinklers may be reckoned as 
 discharging about the same quantity as a i^-inch hose stream under the 
 ordinary practical conditions as to pipes supplying sprinkler and hose 
 systems respectively. 
 
 4. Capacity, a. Plunger diameter alone will not tell how many gallons 
 per minute a pump can deliver, and it is not reasonable to continue the old 
 time notion of estimating capacity on the basis of 100 feet per minute piston 
 travel. 
 
 b. The capacity of a pump depends on the speed at which it can be run, 
 and the speed depends largely on the arrangement of valves and passageways 
 for water and steam. 
 
 c. It is all right to run fire-pumps at the highest speed that is possible 
 without causing violent jar, or hammering within the cylinders. Considerations 
 of wear do not affect the brief periods of fire service or test, hence these speeds 
 are greater than allowable for constant daily duty. 
 
 d. Careful experiments on a large number of pumps of various makes 
 at full speed, show that in a new pump with clean valves, and an air-tight 
 suction pipe, and less than 15 feet lift, the actual delivery is only from i| 
 to 5 per cent less than plunger displacement. This slip will increase with 
 
470 PUMPING MACHINERY 
 
 wear, and for a good average pump in practical use, probably 10 per cent is 
 a fair allowance to cover slip, valve leakage, slight short-stroke, etc. 
 
 e. Largely from tests, but partly from " average judgment," and recogniz- 
 ing that a long stroke pump can run at a higher rate of piston travel in lineal 
 feet per minute than a short stroke pump, and that a small pump can make 
 more strokes per minute than a very large one, the speeds given in the preceding 
 table have been adopted as standards in fire service for direct-acting (non- 
 fly-wheel) steam pumps, which have the large steam and water passages 
 herein specified. 
 
 /. Rated capacity is to be based on the speed in the preceding table, correct- 
 ing the plunger displacement for one-half the rod area and deducting 10 per 
 cent for slip, short -stroke, etc. 
 
 5. Capacity Plate, a. Every steam fire-pump must bear a conspicuous 
 statement of its capacity securely attached to the inboard side of air chamber, 
 thus: 
 
 NATIONAL STANDARD FIRE PUMP 
 16 X 9 X 12 
 
 CAPACITY 
 
 750 GALLONS PER MINUTE, OR THREE GOOD l-IN. SMOOTH NOZZLE 
 
 FIRE STREAMS 
 
 FULL SPEED 
 
 70 REVOLUTIONS PER MINUTE 
 
 Never let steam get below 50 pounds, nights, Sundays, or at 
 any other time 
 
 b. This plate must have an area of not less than one square foot, and 
 must be made of an alloy at least 85 per cent aluminum and the remainder 
 zinc. The letters must be at least one-half inch in height, plain and distinct, 
 with their surfaces raised on a black background and buffed off to a dead 
 smooth finish. 
 
 c. A smaller plate of composition must be attached to steam chest bearing 
 the size of pump, the shop number, and the name of shop in which the pump 
 was built. 
 
 6. Strength of Parts, a. The maker must warrant each pump built under 
 these specifications to be, at time of delivery, in all its parts strong enough 
 to admit of closing all valves on water outlet pipes while steam valve is wide 
 open and steam pressure eighty pounds, and agree to so test it before shipment 
 from his works. 
 
 b. The pump must be warranted so designed and with such arrangement 
 of thickness of metal that it shall be safe to instantly turn a full head of steam 
 on to a cold pump without cracking or breaking the same by unequal expansion. 
 
 7. Shop Inspection. A systematic shop inspection must be given to 
 each pump to ensure complete workmanship, and to prevent the use of defective 
 parts, improper materials, or the careless leaving of foreign matter in any 
 part of the cylinders or chests. 
 
SPECIAL PUMPING MACHINERY 471 
 
 THE STEAM END. 
 
 8. 'Steam Cylinders, a. These must be of hard close iron with metal 
 so distributed as to insure sound castings and freedom from shrink cracks. 
 The following are the minimum thicknesses acceptable: 
 
 14" Diam. I" thick. 
 
 1 6 
 
 " 
 
 1 8" Diam. i" thick. 
 
 20 
 
 b. The inside face of the steam cylinder heads and the two faces of the 
 piston must be smooth surfaces, fair and true, so that if the piston should hit 
 the heads it will strike uniformly all around, thus reducing to a minimum 
 the chances of cramping the piston rod or injuring the pump. 
 
 c. All flanged joints for steam must be fair and true and must be steam 
 tight under 80 pounds pressure if only a packing of oiled paper i-ioo inch 
 thick covered with graphite were used. Jenkins, " Rainbow" or equivalent 
 packing of not exceeding ^ inch original thickness is acceptable. Oiled 
 paper is not acceptable as a final packing, as it burns out. 
 
 For size of steam and exhaust pipes, standard flanges and bolting, see 
 Art. 39. 
 
 d. Heads at both ends of cylinder must be beveled off very slightly over 
 a ring about one inch wide, or equivalent means provided to give steam a 
 quick push at piston, should it stand at contact stroke. 
 
 The specifications originally required machine facing for all these 
 surfaces. The art of machine molding from metal patterns with draw 
 plates, etc., has, however, attained such excellence in certain shops, 
 that in regular practice "foundry faced" cylinder heads and piston 
 faces can be made true and fair, and steam joints can be made tight 
 under 80 Ibs. pressure with a packing of oiled paper only, i-ioo-inch 
 thick. 
 
 11. Steam Ports, a. The area of each exhaust steam passage, at its 
 smallest section, must not be less than 4 per cent of the area of the piston 
 from which it leads. 
 
 b. Each admission port must be not less than 2^ per cent of area of its 
 piston and to avoid wasteful excess of clearance, these passages should not 
 be bored out larger in interior of casting than at ends or passage. 
 
 c. The edges of the steam-valve ports must be accurately milled, or chipped 
 and exactly filed to templets, true to line, and the valve seat must be accurately 
 fitted to a plane surface, all in a most thorough and workmanlike manner and 
 equal to high-grade steam-engine work. 
 
 d. To guard against a piston ring catching in the large exhaust ports, 
 these ports must have a center rib with cylinder at cylinder wall. See also 
 Art. 13 d. 
 
 12. Steam-clearance Space, a. Clearance (including nut-recess, counter- 
 bore, and valve passages) must not exceed 5 per cent for contact stroke or 
 about 8 per cent for nominal stroke (i.e., contact stroke should overrun nominal 
 stroke at each end about one-half inch). 
 
 b. The clearance space between face of piston and cylinder head must 
 be reduced to smallest possible amount, and these contacting surfaces be 
 
472 PUMPING MACHINERY 
 
 flat, without projections or recesses other than the piston rod nut and its 
 recess. 
 
 Some makers, with the idea that a fire pump need not be economical, 
 have not taken pains to keep these waste spaces small. 
 
 Securing small clearance costs almost nothing but care in design, and 
 is often of value, since at many factories boiler capacity is scant for the 
 large quantity of steam taken by a fire pump of proper size. 
 
 13. Steam Pistons, a. May be either built up or colid, as maker thinks 
 best. 
 
 It is believed that "solid" (cored) pistons with rings "sprung in," 
 are for fire pumps much preferable to built-up pistons, since follower 
 bolts do sometimes get loose. 
 
 b. The thickness of piston should be about one-fourth of its diameter. 
 If solid, walls should be not less than \ inch thick, and special care should 
 be given to shop inspection to insure uniformity of thickness. 
 
 This will demand, for the four sizes of pumps, pistons as follows: 
 5oo-gal. 75o-gal. xooo-gal. isoo-gal. 
 
 Diameter 14 in. Diameter 16 in. Diameter 18 in. Diameter 20 in. 
 Thickness 33 in. Thickness 4 in. Thickness 4^ in. Thickness 5 in. 
 
 Manufacturers desiring to use existing patterns approximating 
 these thicknesses may be allowed to do so after due consideration of 
 working drawings. 
 
 c. If built-up pistons are used, involving follower bolt3, such bolts must 
 be of best machinery steel, with screw thread cut for nbout twice the diameter 
 of the bolt and fitting tightly its whole length. 
 
 d. The width of each piston ring must exceed the length of the large 
 exhaust port by at least \ inch. 
 
 This is to avoid the possibility of piston ring catching in the port. 
 See also Art. n d. 
 
 14. Steam Slide-valves, a. Slide valves must be machine fitted on all 
 four of the outer edges, the exhaust port edges, and the surfaces in contact 
 with rod connections. 
 
 b. The slide valve itself must have its steam and exhaust edges fitted up 
 " line and line" with their respective steam and exhaust ports. 
 
 The adding of lap to these edges in lieu of lost motion is not accept- 
 able further than a possible $$ of an inch to cover inaccuracies of edges. 
 
 c. The valves must be guided laterally by guide strips cast in steam chest, 
 and these strips must be machine fitted. The lateral play at these surfaces 
 should not exceed & inch. The height of these guide strips should not 
 be less than \ inch, measuring from valve seat. 
 
 The construction must be such as to absolutely preclude the possibility 
 of the valve riding up on top of this guide strip. 
 
 d. The valves must be guided vertically by the valve-rod itself, the inside 
 end of which must be kept in alignment by the usual form of tail-rod guide. 
 
 The vertical play at these parts should not exceed | of an inch. 
 
 e. The surface of valves must be machine faced and accurately fitted to 
 a plane surface, and be steam tight when in contact with the seat of steam valve. 
 
 15. Steam Slide-valve Adjustment, a. The lost motion at the valves and 
 
SPECIAL PUMPING MACNINERY 4>3 
 
 the settling of them must be determined by a solid hub on the rod, finished 
 in the pump shop to standard dimensions, so that no adjustment is possible 
 after the pump is once set up. 
 
 It is recognized that the practice of making adjustable valve tappets 
 located outside of the steam chest is a good thing in a large pump in 
 constant service and operated by a skilled engineer, but for the infre- 
 quently used ordinary fire pump, the utmost simplicity is desirable, 
 and it is best not to tempt the ordinary man to readjust the valve gear. 
 
 The common form of lost motion adjustment consisting of nut and 
 check nut at each end of the slide valve is not acceptable, as these nuts 
 are liable to become loose and may be incorrectly reset by incompetent 
 persons. A long rectangular nut in the center of the valve is also not 
 acceptable, as it can be moved out of adjustment. A solid hub made as 
 a part of the rod is required, as it absolutely avoids the possibility of 
 the hub becoming loose, an accident possible with a separate hub attached 
 to the rod. 
 
 The amount of lost motion should generally be such that admission 
 takes place at about f. of the stroke of the piston, i.e., for 1 2-inch stroke 
 R.H. valve will be about to open when L.H. piston has moved 7$ inches 
 to 8 inches from the beginning of stroke. When piston is at end of 
 stroke the ports should bs full open. 
 
 1 6. Rock Shafts, Cranks, Links, etc. a. Rock shafts must be either 
 forged iron, forged steel, cr cold rolled steel. Cast iron is not acceptable: 
 The following are th^ minimum diameters acceptable: 
 
 500 gallon pump i^ in. 
 
 750 gallon pump if in. 
 
 1000 gallon pump 2 in. 
 
 1500 gallon pump 2 to 2 J in. 
 
 b. The rock-shaft bearings must be bushed with bronze and the bushings 
 pinned firmly in place. The length of each of these non-corrosive bearings 
 must be not less than 4 inches. 
 
 c. Rock-shaft cranks, valve-rod heads, valve-rod links, and piston-rod 
 spools or cross heads may be wrought iron or steel forgings, or steel castings. 
 If of a heavy, strong pattern, these parts, with the exception of valve-rod 
 links, may be of semi-steel or cast iron. 
 
 d. The sectional area of all connections between rock-shaft cranks and 
 valve rod must be such as to give a tensile or compressive strength substantially 
 equal to that of the valve rod. 
 
 ' 17. Valve-motion Levers, a. The valve-motion levers must be steel, 
 wrought iron, or steel castings. Cast iron is not acceptable. Steel castings, 
 if used, must be deeply stamped with the name of the makers, with letters 
 one-eighth inch high, near the upper end of each lever, where it can easily 
 be seen, thus " Steel Castings." 
 
 Cast-iron arms, if bulky enough to be safe against external blows, 
 are awkward in shape. The sectional area necessary for any arm 
 depends upon the means provided for preventing a sidewise stress on 
 the lever, due to rotation of piston or friction of its connection to piston 
 rod. The spool or cross head on the piston rod should be so designed 
 that no sidewise strain can be thus produced in the lever. 
 
 b. The levers must have a double or bifurcated end at cross head. 
 
 The double end is less likely than a single end to put undue stress 
 
474 
 
 PUMPING MACHINERY 
 
 on the lever as the rod turns, and is also less likely to give trouble from 
 lack of lubrication or from a loosening of any small parts, and has 
 proved to be the most satisfactory arrangement. 
 
 18. Valve-motion Stand. a. The valve motion stand must be securely 
 dowel -pinned to the yoke castings, to prevent any movement after once adjusted. 
 
 19. Cushion Valves, a. Cushion -release valves regulating the amount 
 of cushion steam retained at ends of stroke must be provided. 
 
 b. The cushion release must be through an independent port, as shown 
 
 FIG. 372. Cushion Valve. 
 
 in Fig. 372, so located as to positively retain a certain amount of cushion steam. 
 
 The old form of cushion release through bridge between ports is 
 not acceptable. This form, while leading into the exhaust passage as 
 formerly, differs by starting from a small independent port (about 
 ^-inch wide by z\ inches long) through the cylinder wall, located about 
 | or \ inch back from the cylinder head. (The exact position for affording 
 the best action has to be determined by experiment with each different 
 make of pump, as it depends somewhat on the extent of clearance space 
 and on the point of closure of exhaust by piston and somewhat on the 
 weight of reciprocating parts) . 
 
 This style of cushion port makes the pump safer in case cushion 
 valves are unskilfully left open too wide, and tends to prevent a pump 
 from pounding itself to pieces in case of a sudden release of load, as by 
 a break in suction or delivery mains, or by a temporary admission of 
 air to suction pipe. 
 
 Pumps made with this form of cushion release have given very 
 satisfactory results, and if the ports are properly 
 located, there will be no rebound of piston. 
 
 c. Cushion valves must be . always provided 
 with hand-wheels marked as per sketch, Fig. 373, 
 for the reason that a very few 
 men in charge of fire -pumps are 
 found to clearly understand or to 
 remember their use. 
 
 The lettering must be very 
 open, clear and distinct, not liable 
 to be obscured by grease and dirt, 
 and of a permanent character. 
 
 It is desirable that spindle or wheel be so formed that 
 a monkey wrench can get a grip to open a jammed 
 valve. Fig. 374 shows the stem flattened for this purpose. p IGt 
 
 d. The valve and stem of cushion valve must be in 
 one piece without any swivel joint. 
 
 Swivel joints are apt to come apart and make it impossible to operate 
 the valve. 
 
 FIG. 373. Hand Wheel. 
 
 _ Flat- 
 tened Stem. 
 
3PECIAL PUMPING MACHINERY 
 
 475 
 
 20. Piston Rods. a. Piston rods for their entire length must be of solid 
 Tobin Bronze, and the distinguishing brand of the manufacturers of this 
 metal must be visible on at least one end of each rod. 
 
 b. The sizes must be not less than in table below. 
 
 Size of Pump. 
 
 500 gal. 
 
 75 gal. 
 
 1000 gal. 
 
 1500 gal. 
 
 Diameter of rod 
 
 2 inch 
 
 2\ inch 
 
 2f inch 
 
 2| inch 
 
 c. The size and form of connection of rod to "piston plunger and cross 
 head must be such that the stress in pounds per square inch at bottom of 
 screw thread, or at such other point of reduced area as receives the highest 
 tensile stress, shall not exceed 8000 pounds per square inch, when the steam 
 pressure acting on the piston is 80 pounds per square inch. 
 
 d. Piston rod nuts, in both steam and water ends, must be tightly fitted, 
 and preferably of a finer thread than the United States Standard. This is to 
 
 FIG. 375. Lock Nut. 
 
 avoid as much as possible the unnecessary weakening of the rod at the bottom 
 of the thread, and to reduce the tendency of the nut to work loose. 
 
 In practice eight threads per inch has been found to give good 
 satisfaction. 
 
 e. In addition to a tightly fitting nut, some reliable device must be provided, 
 in both steam and water ends, for absolutely preventing these nuts from 
 working off. 
 
 Fig. 375 shows one form of such a locking device and illustrates 
 the kind of security desired. 
 
 This device combines the advantage of a taper key and a split pin, 
 and the elongated key-slot gives sufficient leeway to always insure that 
 the key can be driven up tight against the nut and thus prevent it from 
 even starting to work off. Other methods will be approved in writing, 
 if found satisfactory. 
 
 21. Valve Rods. a. Valve rods for their entire length must be of solid 
 Tobin Bronze, with sizes not less than in table below. 
 
 Size of Pump. 
 
 500 gal. 
 
 750 gal. 
 
 1000 gal. 
 
 1500 gal. 
 
 Diameter of rod 
 
 i inch 
 
 i inch 
 
 i inch 
 
 i -J- inch 
 
 
 
 
 
 
 b. The net area of valve-rod at its smallest section subject to tensile stress, 
 
476 PUMPING MACHINERY 
 
 must not be smaller than at bottom of U. S. Standard screw thread on rod 
 of diameter given above. 
 
 The construction of this rod as affecting lost motion at slide valve 
 is specified under Art. 15. 
 
 22. Stuffing Boxes, a. Ail six stuffing boxes must be bushed at the bottom 
 with a brass ring with suitable neck and flange, and the follower or gland 
 must tc either of solid brass, or be lined with a brass 
 shell ^-inch thick, having a flange next the packing, 
 as shown in the sketch. 
 
 The bottom of stuffing boxes and the end of the 
 glands should taper slightly towards the center, as per 
 sketch. 
 
 b. These glands should be strong enough to withstand 
 considerable abuse, so as not to break from the unfair 
 Bot tU treatment of unskilled men. 
 
 23. Pressure Gauge, a. A pressure gauge of the 
 
 Lane double tube spring pattern with 5-inch case must be provided and 
 attached to the steam chest inside the throttle valve. 
 
 The dial of gauge should be scaled to indicate pressures up to 120 pounds 
 and be marked "STEAM." 
 
 24. Drain Cocks, a. Four brass drain cocks, each with lever handle 
 and of one -half inch bore, are to be provided, and located one on each end 
 of each steam cylinder. 
 
 25. Oiling Devices, a. A one-pint hand oil pump, to be connected below 
 the throttle, and a one-pint sight feed lubricator, to be connected above the 
 throttle, must be furnished with each pump. 
 
 b. Oiling holes must be provided for all valve motion pins, and for each 
 end of both rock shafts. 
 
 26. Stroke Gauge, a.' A length -of -stroke index must be provided for 
 each side of pump. These must be of simple form for at all times rendering 
 obvious the exact length of stroke which each piston is making, and thus 
 calling attention to improper adjustments of cushion valves or stuffing boxes. 
 
 b. The gauge piece over which the index slides must have deep, conspicuous 
 marks at ends of nominal stroke, and also light marks at extreme positions; 
 it need contain no other graduations. 
 
 c. This stroke index must be rigidly secured to cross head in such a way 
 that it cannot get loose or out of adjustment. 
 
 THE WATER END. 
 
 27. Water Cylinders, a. These must be of hard close iron with metal 
 so distributed as to ensure sound castings, and freedom from shrink cracks. 
 
 b. The design should be along lines best calculated to resist internal 
 pressures so as to avoid as much as possible the need of ribs for stiffening. 
 
 c. They must be capable of withstanding, without showing signs of weak 
 ness, the pressures and shocks due to running under the conditions mentioncu 
 in chapter "Tests for Acceptance," Arts. 48-54. 
 
SPECIAL PUMPING MACHINERY 
 
 477 
 
 The suction chamber should be able to withstand a water pressure of 100 
 pounds. 
 
 Although suction chambers are not regularly subject to a pressure, 
 it is sometimes desired to connect them to public water supplies, and 
 where foot valves are used there is a chance of getting pressure on the 
 suction, so that ample strength is necessary. 
 
 Foundry finish may be permitted on the joints at water cylinder 
 heads and at hand-hole plates, provided surfaces are so true that a 
 rubber packing not over ^ of an inch in thickness is sufficient to secure 
 perfect tightness. 
 
 d. Conveniently placed hand -holes of liberal size must be provided for 
 the ready examination and renewal of valve parts at the yoke, end of water 
 cylinders and in the delivery chamber. 
 
 This will necessitate holes not less than 6x8 inches, or its equivalent, 
 for the two largest-size pumps, and holes proportionately as large for 
 the 500- and 75o-gallon pumps. The easy access to the valve parts is 
 of vital importance, and must receive careful attention. 
 
 e. The thickness of metal for cylinder shell, valve decks, partitions, ribs, 
 etc., will depend largely upon the form of construction, but, in a general way, 
 to establish safe minimums for the average water cylinder, of nearly cylindrical 
 form, whose flat surfaces are stiffly ribbed, we submit the table belo.w: 
 
 Size of Pump. 
 
 500 gal. 
 
 750 gal. 
 
 1000 gal. 
 
 1500 gal. 
 
 Thickness of cylinder shell when 
 of nearly cylindrical form 
 Thickness of valve decks when 
 well ribbed 
 
 Inches 
 1 
 ll 
 
 Inches 
 I 
 ii 
 
 Inches 
 I* 
 
 il 
 
 Inches 
 
 *i 
 
 ii 
 
 Thickness transverse partition, 
 depending on ribbing 
 
 i i to i $ 
 
 il to ii 
 
 i to 2 
 
 l tO 2 
 
 Thickness of longitudinal parti- 
 tion depending on ribbing 
 
 T 1 tO I i 
 
 1 1 to 1 5 
 
 I i to 2 
 
 I i to 2 
 
 Thickness of ribs 
 Thickness of suction chamber. . . . 
 Thickness of delivery chamber 
 
 1 
 
 1 
 1 
 
 i 
 I 
 
 i 
 
 I 
 I 
 'I 
 
 T 
 I 
 
 'i 
 
 Lighter construction than herein specified will not be regarded as satis- 
 factory, and any construction will be finally passed upon on examination of 
 drawings. 
 
 /. The bolting of all parts of the water end is to be of such strength that 
 the maximum stress at bottom of screw thread will not exceed 10,000 pounds 
 per square inch (disregarding for the moment the initial stress due setting 
 up nuts) for a water pressure of 200 pounds per square inch, computed on 
 an area out to center line of bolts. 
 
 No stud or tap bolt smaller than f inch should be used to assemble parts 
 subject to the Ltress of water pressure, as smaller bolts are likely to be twisted 
 off. This does ret apply to standard flanges where through bolts are used. 
 
 Although these pumps are not expected to be designed for a regular 
 
478 PUMPING MACHINERY 
 
 working water pressure of 240 or 320 Ibs., it is expected that bolts, 
 shells, rods, etc., will be figured to stand this comparatively quiet, 
 temporary, high pressure, exclusive of further allowance for initial 
 stress due setting up of bolts, with a factor of safety of at least four. 
 
 This high test pressure is analogous to the custom of proving all 
 common cast-iron water pipes to 300 Ibs. and all common lap-welded 
 steam pipes to 500 Ibs. per square inch, and common water- works gate 
 valves to 400 Ibs., even though these are to be regularly used at much 
 less pressure. 
 
 We are assured that castings no heavier than at present used by the 
 best makers will stand this test, if properly shaped and liberally bolted. 
 
 g. For requirements for stuffing boxes, see Art. 22. 
 
 28. Water -Plungers and Bushings, a. The "inside plunger and bushing" 
 is preferred for all situations where the water is free from grit or mud. 
 
 b. Water-plungers must be of solid brass or bronze, and the bushing in 
 which they slide must also be of brass or bronze. The composition of the 
 plunger and its bushing should be of very hard, though dissimilar alloys, 
 to ensure good wearing qualities. 
 
 For material and size of piston rods and lock for nuts, see Art. 20. 
 
 With poor alignment or bad workmanship or lack of skill in mixing 
 the alloys, brass plungers are liable to score and give trouble; but with 
 proper . selection of alloys and true cylinders accurately aligned, they 
 can be made to run all right wherever iron ones can. It is quite a fine 
 point to get these wearing surfaces just right; and this is wherein the 
 experience, skill, and shop practice of one maker is likely to be much 
 superior to that of another working under the same specification. 
 
 c. The length of machined cylindrical bearing within the partition must 
 be not less than two inches. The plunger bushings must have a faced seat 
 transverse to its axis against partition, forming a water-tight ground joint 
 not less than one-half inch wide. 
 
 Any rubber gasket or other compressible packing for making this joint 
 water-tight is not acceptable. 
 
 d. The construction of bushing and hole in partition must be such that 
 a cylindrical shell for use with a packed piston can be interchangeably inserted 
 in its place and secured by the same bolts. 
 
 This can readily be arranged, and enables a packed piston to be 
 inserted in place of a plunger subsequent to the installation of the pump 
 with a minimum of expense, should this become desirable from change 
 of conditions at any future time. 
 
 e. Small transverse grooves cut within the sliding surface of the plunger 
 bushing, with a view to lessen the leakage, are not acceptable. 
 
 Although a slight advantage in this respect for clean water, they are 
 a disadvantage on the whole, as dirt catches in them in the ordinary 
 situation and cuts the plungers. 
 
 29 Standard Dimensions of Plungers and Plunger Bushings. a. To 
 bring all these expensive parts to the same standard of weight and bearing 
 surface, the following dimensions are specified as the least that will be accept- 
 able. These are based on a length of plunger which uncovers the bushings 
 one inch at end of nominal stroke; 
 
SPECIAL PUMPING MACHINERY 
 
 479 
 
 SOLID BRONZE PLUNGERS AND BUSHINGS 
 
 Size of Pump. 
 
 500 gal. 
 
 750 gal. 
 
 1000 gal. 
 
 1500 pril. 
 
 Plunger 
 
 Inches. 
 
 Inches. 
 
 Inches. 
 
 Inches. 
 
 Diameter 
 
 7 or i\ 
 
 9 
 
 10 or loj 
 
 12 
 
 Length 
 
 17 
 
 1 7 
 
 18 
 
 2.4. 
 
 Thickness of transverse parti- 
 
 
 
 
 
 tion 
 
 | 
 
 
 
 4. 
 
 3 
 
 Thickness next to partition .... 
 
 I 
 
 1 
 
 I 
 
 I 
 
 Thickness next to end 
 
 ft 
 
 1 
 
 f 
 
 1 
 
 Number of ribs . . 
 
 4 
 
 4 
 
 6 
 
 6 
 
 Thickness of ribs 
 
 A 
 
 A 
 
 | 
 
 a 
 
 Bushing 
 
 
 
 
 
 Length 
 
 7 
 
 7" 
 
 8 
 
 10 
 
 Thickness at end .... 
 
 A 
 
 
 
 | 
 
 1 
 
 Thence tapered evenly to a 
 
 
 
 
 
 thickness next to bearing of 
 
 
 
 
 
 not less than 
 
 i 
 
 f 
 
 f 
 
 1 
 
 Thickness at the center bearing 
 
 
 
 
 
 not less than 
 
 I 
 
 . I 
 
 1 
 
 H 
 
 30. Water Pistons and Bushings. a. The " water piston with fibrous 
 packing" is preferred for many situations in the West or South, or for water 
 containing grit or mud, like that of the Ohio River; and, for the comparatively 
 few cases where pump pressure governors are used, the packed piston will 
 give better service and longer wear. 
 
 b. The removable bushing or cylinder in which this piston works must be 
 of solid bronze. 
 
 c. As stated in Art. 28 d, this bushing should be so constructed as to be 
 readily interchangeable with the bushing of the inside plunger type. 
 
 d. The length of cylindrical bushing must be such that the outer edge of 
 packing will come short of the edge of bushing at contact stroke about \ inch 
 and not uncover. 
 
 e. The thickness of the cylindrical bushings must be not less than that 
 given in the following table: 
 
 BUSHINGS FOR PACKED WATER PISTONS 
 
 Size of Pump. 
 
 500 gal. 
 
 750 gal. 
 
 1000 gal. 
 
 1500 gal. 
 
 Solid Bronze 
 
 Inches. 
 
 Inches. 
 
 Inches. 
 
 Inches. 
 
 Thickness at extreme end 
 
 A 
 
 \ 
 
 i 
 
 ft- 
 
 Tapered evenly from end to a 
 
 
 
 
 
 thickness next to bearing of 
 
 
 
 
 
 not less than 
 
 A 
 
 f 
 
 H 
 
 f 
 
 Thickness at center bearing, at 
 
 
 
 
 
 least 
 
 1 
 
 
 
 f 
 
 ri 
 
 
 
 
 
 
480 PUMPING MACHINERY 
 
 f. In other respects, the specifications for plunger bushings, already given 
 in Art. 28, will apply to the above. 
 
 g. The water piston used in the shell described above must expose not 
 less than 2 inches in width of fibrous packing, and must be of bronze, with 
 disc and follower accurately turned to a sliding fit, so that the leakage past 
 it will be a minimum, even when no fibrous packing is in place. There must 
 be at least 2 inches in length of metallic bearing on both disc and follower. 
 
 The follower must be accurately centred, and fitted to hub of piston, so 
 that alignment will not be disturbed if taken apart. 
 
 h. The water piston must be of simple and strong construction, with 
 follower bolts tightly fitted, and with fibrous packing so cut as to prevent 
 by-passing. 
 
 i. All materials used in construction of piston, except packing, must be 
 brass, bronze, or other non-corrosive metal. 
 
 j. Bushing studs must be of Tobin bronze, and of such size and number, 
 that the maximum stress at the bottom of the screw thread shall not exceed 
 10,000 pounds per square inch, in the event of plunger becoming fast in the 
 bushing with 80 pounds of steam in the steam cylinders. 
 
 k. For each bushing stud there must be provided a composition nut and 
 check nut. 
 
 /. All minor parts exposed to the action of water in water cylinder, that 
 are not herein specified, must be of brass, bronze, or other non-corrosive 
 material. 
 
 31. Pump Valves, a. All the suction and discharge valves in any one 
 pump must be of the same size and interchangeable. 
 
 b. There must be a clear space around each rubber valve, between it 
 and the nearest valve, equal to at least one-fourth of the diameter of the valve, 
 or between it and the wall of the chamber of at least one-eighth of the diameter 
 of the valve. 
 
 c. These valves must be of the very best quality of rubber, of medium 
 temper with a face as soft as good wearing quality will permit. 
 
 They must be double-faced, so they can be reversed when one face is 
 worn. 
 
 The quality of rubber is almost impossible of determination by 
 brief inspection or by chemical analysis. The relative amount of pure 
 gum and of cheaper composition may vary, or good material may be 
 injured by defective vulcanization. The only safe way to secure excel- 
 lence and uniformity is for the pump manufacturer to test samples 
 of each new lot under severe duty (as by a week's run in a small special 
 pump, with say 150 pounds pressure and heavy water hammer, or by 
 some equivalent means) and to furthermore require the rubber manu- 
 facturer to mould a date mark as " (Name of pump manufacturer, 
 lot 201 April 3, 1904) " on the edge of every valve, by which the pump 
 manufacturer can keep track of those which prove defective. 
 
 32. Size and Number of Pump-Valves, a. The diameter of the disc 
 of rubber forming the valve must not be greater than 4 inches or less than 
 3 inches. 
 
 Three and five-eighths inches diameter appears to be the size best 
 *neeting all the conditions, and has been adopted by several manufac- 
 turers but is not insisted upon. 
 
SPECIAL PUMPING MACHINERY 
 
 481 
 
 There is some confusion between different shops about designating 
 size of valves. The practice is here adopted, which is much the most 
 widely used, of naming the diameter of the disc of rubber which covers 
 the ports, and it is hereby specified that this shall be about one-half 
 inch greater than the diameter of the valve-port circle which it covers, 
 thus affording about one-fourth inch over-lap or bearing for the rubber 
 disc all around its edge. 
 
 If valves are larger than four-inch there is an increased tendency 
 to valve-slam at the very high speed at which the pump is designed 
 to run, and if valves are smaller than three inches diameter the greater 
 number tends to unnecessary multiplication of parts, and the ports 
 being so small are a little more liable to become obstructed by rubbish. 
 
 b. The thickness of the rubber valve must in no cases be less than f -inch. 
 
 33. Suction Valve Area. a. The total lift of suction valves must not 
 exceed |-inch. 
 
 b. The net suction valve port area and the total suction valve outlet area 
 under valves lifted J inch high must not be smaller than the figures given in 
 the table below: 
 
 
 
 
 Approx. Actual 
 
 (6) 
 
 Net 
 
 
 
 
 
 
 Max. Piston 
 
 Suction 
 
 (7) 
 
 
 (l) 
 
 Length 
 of Stroke 
 (in 
 inches) . 
 
 (2) 
 
 Greatest 
 No. 
 Revolu- 
 tions per 
 Minute. 
 
 Corre- 
 sponding 
 Piston 
 Travel per 
 Minute. 
 
 Velocity at Full 
 Speed per Column 
 
 (3)X2.2. 
 
 Valve-port 
 Area Re- 
 garded 
 Necessary 
 for this 
 Speed 
 Per Cent 
 of Plunger 
 
 Suction 
 Valve-out- 
 let Area 
 Under 
 Valves 
 Lifted * 
 Inch High. 
 
 & 
 
 charge 
 Valve 
 Area. 
 
 (4) 
 Feet 
 
 (s) 
 
 Feet 
 
 
 
 
 per Min. 
 
 per Sec. 
 
 Area. 
 
 
 
 12 
 
 70 
 
 140 ft. 
 
 308 
 
 5-i 
 
 56% 
 
 56% 
 
 1 of 
 
 Suction 
 
 16 
 
 60 
 
 1 60 ft. 
 
 35 2 
 
 5-9 
 
 64% 
 
 64% 
 
 Valve 
 Area. 
 
 By " valve-outlet area," we mean the vertical cylindrical surface 
 over the outer edge of the valve ports, i.e., the distance L multiplied 
 by the circumference at the outer edge of the valve ports C. Thus for 
 a four-inch valve, with ports inscribed in a three and one-half-inch 
 circle, whose circumference is 3.5x3.1416 = 11 inches; the valve "out- 
 let area" for one-half-inch lift would be 5^ inches. 
 
 The actual velocity of piston during the middle portion of stroke 
 is from 2.0 to 2.4 (average 2.2) times as great as the piston travel per 
 minute (as determined in experiments by Mr. J. R. Freeman on several 
 duplex pumps of different manufacture). This is because each piston 
 stands still nearly half the time, or while its mate is working, and, more- 
 over, moves more slowly near start and finish of stroke. The words 
 "piston speed " are commonly incorrectly used, and refer to "piston 
 travel." A clear understanding that the actual piston speed is more 
 than twice as great leads to more generous valve design. 
 
 Large aggregate valve areas are necessary for pumps designed to run 
 as fast as these, and experience has shown that to prevent valve slam 
 at high speed and to accommodate high suction lifts, it is just as important 
 to have a large " valve outlet area " as to have a large area of valve port. 
 
 It is valve slam or water hammer which commonly limits the highest 
 speed at which a pump can be run. This water hammer may originate 
 from the pulsations in a long or small suction pipe. The vacuum chamber 
 
482 
 
 PUMPING MACHINERY 
 
 lessens it, but there is commonly some point of high water in the vacuum 
 chamber that will give much smoother action than any other. 
 
 Valve slam in this style of pump is caused chiefly by the short rebound 
 of the steam piston against the elastic steam cushion at the end of the 
 
 stroke. This in turn snaps the valves 
 down with a jump when the speed is high. 
 Dividing this impact or slam on numerous 
 valves of low lift, tends to break up and 
 lessen the shock, therefore with valves of 
 the size and style used in fire pumps, other 
 things being equal, the less they have to 
 rise and drop to let the water through 
 them, the less will be the valve slam. This 
 height of rise and drop is governed by the 
 circumference rather than the port area. 
 Experience and practice have shown that a 
 -inch limit of lift is reasonable and does 
 insure a smooth working pump under all 
 ordinary conditions. 
 
 c. The following table gives minimums for 
 
 FIG. 377. Water Valve. 
 
 aggregate valve port area and aggregate valve outlet area for the different 
 size plungers, figured on a basis of 56 per cent of plunger area for a 1 2-inch 
 stroke, and 64 per cent for a i6-mch stroke. 
 
 
 Size of Pump. 
 
 500 gal. 
 
 75o gal. 
 
 1000 gal. 
 
 1500 gal. 
 
 I 
 
 Diameter of plunger. Inches. . . . 
 
 7i" 
 
 9" 
 
 10" 
 
 12" 
 
 2 
 
 Area of plunger in sq inches 
 
 41 28 
 
 63 62 
 
 78 S4 
 
 1 T 7 TQ 
 
 
 
 
 
 
 
 3 
 
 56% of plunger area, or minimum 
 aggregate valve-port area al- 
 lowed per section. Sq. inches. . 
 
 23.11 
 
 35-63 
 
 43-98 
 
 64%- 
 72.38 
 
 4 
 
 Minimum aggregate valve-port 
 circumference, allowed per sec- 
 tion. Inches. . . 
 
 46 . 22 
 
 71 . 26 
 
 8? 06 
 
 144. 76 
 
 
 
 
 
 
 
 5 
 
 Minimum aggregate valve outlet 
 area allowed per section for 
 valves lifted \ inch high. Sq. 
 inches 
 
 22 II 
 
 7 r 6 3 
 
 A -I 08 
 
 72 -?8 
 
 
 
 
 
 
 
 d. If we consider using any one of the three sizes of valves below, whose 
 port areas may be assumed approximately as given, then the necessary number 
 of valves per section will be as in the table following: 
 
 Diam. Valve. 
 
 Diam. of Valve- 
 port Circum 
 
 Circum. of V. C. 
 Circle. 
 
 Valve-port Area (Net). 
 Square Inches. 
 
 3" 
 3t" 
 
 4" 
 
 4" 
 
 3t" 
 3 1" 
 
 7-8S" 
 9.82" 
 10.99" 
 
 3-5 
 
 5- 1 
 
 6.3 
 
SPECIAL PUMPING MACHINERY 
 
 4S3 
 
 Size of Pump. 
 
 500 gal 
 
 750 gal. 
 
 1000 gal. 
 
 1500 gal. 
 
 Size of valves, inches 
 
 3 
 
 3i. 
 
 4 
 
 3. 
 
 3l 
 
 4 
 
 3 
 
 3f 
 
 9 
 
 4 
 8 
 
 3 
 19 
 
 21 
 
 3f 
 14 
 *5 
 
 4 
 14 
 
 12 
 
 Necessary number of valves 
 to satisfy (4) under c 
 
 6 
 
 S 
 
 6 
 
 9 
 
 8 
 
 7 
 
 ii 
 
 
 Necessary number of valves 
 to satisfy (3) under c. 
 
 7 
 
 5 
 
 4 
 
 10 
 
 8 
 
 6 
 
 *3 
 
 9 
 
 7 
 
 
 The exact number and size of valves will, however, not be insisted upon 
 provided the aggregate valve area and the aggregate valve outlet area for 
 each section is not less than that given in the table under c for the limiting 
 lift of \ inch. 
 
 Manufacturers will note that with the established lift of -inch, 
 the 3 f -inch valve will permit a valve outlet area more nearly equal 
 to its port area than will either the 3 -inch or 4-inch valves, and a 
 relatively less number of valves will satisfy the specifications. 
 
 34. Delivery Valves, a. The total lift of delivery valves must not exceed 
 \ inch. 
 
 This is to avoid valve slam, as explained in Art. 33. 
 
 b. The aggregate valve-port area should be restricted to about two-thirds 
 the suction-valve area. 
 
 A small restriction of water-way through the delivery valves steadies 
 the action of the pump and tends to prevent undue pulsations of pres- 
 sure in the delivery pipe or fire hose. Fewer delivery valves than suction 
 valves are, therefore, preferred, and if extra holes in the delivery deck 
 are cast for shop purposes these had better be plugged than fitted with 
 valves. 
 
 The suction valves require more generous port-circumference and 
 port-area than delivery valves because when a pump has to suck its 
 supply through a considerable height or through a long pipe there should 
 be the least practicable waste of the atmospheric pressure in getting 
 the water into the plunger chamber, or in retarding it from following the 
 plunger in full contact. With the water once in the plunger chamber 
 there is plenty of steam pressure available to force it out through the 
 delivery valves. 
 
 35. Valve Springs, Guards and Covers, a. All valve springs must be 
 of the best spring brass wire, and must be coiled on a cylindrical arbor. 
 
 Conical valve springs are not approved because the stress is not 
 uniform throughout spring, thereby increasing the liability to breakage 
 and the chance of their getting out of center and becoming " hooked up." 
 
 b. The valve spring must be held centrally at its top by resting in a groove 
 in valve guard, substantially as shown in Fig. 378. 
 
 c. A light, rustless, metallic plate must be interposed between the bottom 
 of the spring and the rubber valve, and must be the full area of the valve. 
 This plate must also be formed with a raised bead to guide the spring at the 
 bottom. 
 
 The weight of this plate should be small, for the inertia of the lifting 
 
484 
 
 PUMPING MACHINERY 
 
 parts of the valves should be the least possible, to permit quick action 
 and to avoid pounding. 
 
 d. For the average condition of a 10- or 15 -foot lift, the stiffness of suction 
 valve springs should be such that a force of about one pound per square inch 
 of net port area will lift valve J-inch off its seat. 
 
 The springs on the delivery valves should ordinarily be from two to three 
 times as stiff as just specified, but any other reasonable degree of stiffness 
 which is proved to work well in practice will not be objected to. 
 
 For suction under a head, the greater snap with which water enters 
 the plunger chamber when thus pushed in by say twice the atmospheric 
 pressure renders it difficult to avoid water hammer 
 at high speed. Extra stiff suction valve-springs 
 will commonly aid in controlling this and should 
 be used wherever pumps are to work under a 
 head. 
 
 An approved type of indicator water gate on 
 the suction pipe near the pump, which can be 
 partly closed, will enable the pump to run quietly 
 at high speed. Such a gate is an extra not in- 
 cluded in price of the pump. 
 
 36. Sticking of Valves, a. Steam fire-pumps 
 should be started to limber them up at least once a 
 week. 
 
 Although vulcanized India rubber is much 
 the best material yet used for fire-pump valves, 
 unfortunately the brass is sometimes corroded by 
 the free sulphur contained in the rubber, so that 
 if the pump is left standing for several weeks the 
 rubber valve discs may become stuck to their 
 brass seats, and, if suction has a high lift, there 
 may not be vacuum enough to tear all the suction 
 valves open when pump is started. 
 
 37. Valve Seats, a. All water valve seats muct 
 be of bronze composition. They may be either 
 
 FIG. 378. Valve. 
 
 screwed into the deck on a taper or forced in on a smooth taper fit. With 
 either arrangement, the seat must be either flanged out on the under side all 
 the way round or be provided with a substantial lug opposite each rib, these 
 lugs being expanded out after the valve is inserted. 
 
 If the valve seats are not expanded after being put in place, there 
 is a possibility that now and then a valve seat will work loose and come 
 out, thus crippling the pump. 
 
 b. The under side of the valve deck must be rounded over to give good 
 bearing for the expanded part of the seat. 
 
 c. Three-inch valves must have four or five ribs, 3^-inch valves five or 
 six ribs, and 4-inch valves six ribs. 
 
 Enough ribs must be provided to give proper support to the rubber 
 valve, but too many are objectionable, as small ports would be liable 
 to obstruction by refuse. 
 
 d. The edges of the valve-seat ports must be moderately rounded over 
 to remove such sharp edges and points as would be liable to cut or damage 
 the rubber valve when under pressure. 
 
SPECIAL PUMPING MACHINERY 
 
 485 
 
 38. Valve Stems, a. All valve stems must be of f-inch Tobin bronze 
 and of the fixed type, and must have the guard fastened on by one of the 
 methods shown by Figs. 378 and 379. 
 
 Other methods may be approved, in writing, if found by test and 
 experience to have especial merit. 
 
 b. These stems must be screwed into the seats on a straight, tightly fitting 
 thread, and the lower end then well headed over 
 into a countersink. The valve guard and nut must 
 be of composition. 
 
 In Fig. 378 the upper part of the stem is slab- 
 bed off on two opposite sides and fits a correspond- 
 ing hole in the guard. 
 
 The guard, therefore, cannot turn. The out- 
 side of the special nut is fitted on a taper to the 
 inside of the guard, and the nut tapped out to fit 
 the five-eighths U. S. thread on the stem. 
 
 The action of the valve, whether with the spring 
 or without, tends to drive these taper fits together, 
 producing a frictional lock similar to that of a 
 friction clutch ; and although the nut may be loose 
 on the thread, it cannot possibly work off. 
 
 It will be apparent that the taper fit on the 
 nut must be so made as to always bear on the 
 taper fit in the guard, and not bottom in the guard. 
 
 It is believed that with the present screw 
 machine practice in shops of to-day these small 
 parts can readily be turned out accurately and 
 cheaply in large quantities. The nuts and guards 
 made in any one shop must be exactly of stand- 
 ard dimensions, so that the product of different 
 periods will be interchangeable. 
 
 FIG. 370. Valve. 
 
 The taper should be about one inch to one foot. With this taper 
 the nut can be readily turned in or out, but there is friction enough to 
 hold the guard and nut together even if the spring is off. 
 
 In Fig. 379 the top of the guard is recessed in the form of a hollow 
 inverted pyramid of six sides, to correspond to a hexagonal nut. The 
 angle of two opposite sides of this recess, which should be about 75 degrees, 
 will both surely lock the nut and still permit of its being turned with a 
 wrench. 
 
 The guard is kept from turning by slabbing off the stem in the same 
 manner as described and shown in Fig. 378. 
 
 To facilitate the removal of the nut, the edges should be slightly 
 chamfered. An unfinished nut simply drilled and tapped is all that is 
 desired. Any hexagonal or square nut within the size of the tapered 
 recess will be' locked. 
 
 With this construction, the nut cannot turn in either direction without 
 compressing the spring and is therefore locked, and, in the event of 
 the spring breaking or being left off, the nut is well protected in its 
 recess, from the possible turning effects of water currents, and experi- 
 ments have shown that it will stay in place. 
 
 With machine molding it will be possible to make these guards 
 complete in foundry, requiring no machine work further than a pos- 
 sible broaching out of hole to fit the stem, as a fairly good fit is 
 necessary. 
 
 While both of these devices are effective even though not tightened 
 down to a shoulder, they should be so tightened for greater safety and 
 to fix the lift at the half -inch limit. - 
 
PUMPING MACHINERY 
 
 39. Pipe Sizes. . a. Water and steam pipe connections must have standard 
 flanges to connect with pipes of the sizes given below. 
 
 
 1 
 
 
 
 
 Size of Pump. 
 Gal. per Min. 
 
 Diameter of 
 Suction Pipe. 
 Inches. 
 
 Diameter of 
 Discharge Pipe. 
 Inches. 
 
 Steam 
 Pipe. 
 
 Exhaust 
 Pipe. 
 
 500 
 
 8 
 
 6 
 
 3 
 
 4 
 
 750 
 
 10 
 
 7 or 8* 
 
 Si 
 
 4 
 
 IOOO 
 
 12 
 
 8 
 
 4 
 
 5 
 
 1500 
 
 14 
 
 10 
 
 5 
 
 6 
 
 *Eight-inch preferred, this being the more common size for valves, fittings and pipes. 
 
 These suction pipe-sizes, although larger than common for trade 
 pumps of the same size, are believed to be amply justified by experience, 
 and exert a powerful influence toward enabling the pump' to run smoothly 
 at high speed with water cylinders filling perfectly at each stroke. No 
 defect is more common than restricted suction pipes. 
 
 b. A single suction entrance at the end of the pump is to be provided 
 unless otherwise specified by the purchaser. 
 
 Some situations render desirable side suction entrances, for permitting 
 drafting water from two different sources of supply. These additional 
 openings are to be considered as extras. Ordinarily, the purchaser 
 can provide for such situations by proper piping at the single end suction 
 entrance. 
 
 If there is to be but one suction opening on casting, this had best 
 be at center, for the reason that, if suction pipe ever gets to leaking air, 
 this air stands a better chance of being distributed equally to the two 
 plungers, and has less tendency to make the pump run unevenly. 
 
 c. Standard flanges and standard bolt layouts as adopted by the Master 
 Steam Fitters, July 18, 1894, must be used on all the above pipe connections, 
 as per table given below. 
 
 SCHEDULE OF STANDARD FLANGES 
 
 Size of Pipe 
 X Diam. 
 of Flange. 
 Inches. 
 
 Diameter 
 of Bolt Circle. 
 
 Inches. 
 
 Number 
 of Bolts. 
 
 Size of 
 Bolts. 
 
 Inches. 
 
 Flange 
 Thickness 
 at Edge. 
 Inches. 
 
 3 X 7i 
 
 6 
 
 4 
 
 IXrt 
 
 '. 
 
 3*X 8} 
 
 7 
 
 4 
 
 IXai 
 
 I 
 
 4X9 
 
 7l 
 
 4 
 
 |Xt| 
 
 if 
 
 4iX 9 t 
 
 7l 
 
 8 
 
 1X3 
 
 it 
 
 5 Xio 
 
 H 
 
 8 
 
 |X 3 
 
 if 
 
 6 Xn 
 
 9* 
 
 8 
 
 |X 3 
 
 I 
 
 7 Xiai 
 
 ioi 
 
 8 
 
 iX3f 
 
 *& 
 
 8 Xi 3 i 
 
 iif 
 
 8 
 
 1X3* 
 
 i| 
 
 9 Xis 
 
 ji 
 
 12 
 
 iX 3 l 
 
 il 
 
 10 Xi6 
 
 14} 
 
 12 
 
 lX3f 
 
 'A 
 
 12 X 19 
 
 J 7 
 
 12 
 
 |X3i 
 
 ii 
 
 14 X2i 
 
 i8i 
 
 12 
 
 i X4i 
 
 if 
 
SPECIAL PUMPING MACHINERY 487 
 
 Do not drill bolt holes on center line, but symmetrically each side of it. 
 On steam and exhaust openings loose flanges threaded for wrought-iron 
 pipe must be provided. 
 
 Where the situation will not permit of a standard flange on exhaust 
 opening for lack of room, a special flange threaded to fit the proper size 
 wrought-iron pipe may be used. 
 
 40. Air and Vacuum Chambers, a. Air and vacuum chambers in accord- 
 ance with the sizes given in~ the following table must be provided with all 
 pumps. If the air chamber is cast iron, the pump manufacturers must warrant 
 that it has been subjected to a hydraulic test of 400 pounds per square inch 
 before it is connected to pump. 
 
 It is to be thoroughly painted inside and out to diminish its porosity. 
 
 SIZE OF VACUUM AND AIR CHAMBERS 
 
 
 Vacuum Chamber 
 is to Contain 
 
 Air Chamber 
 is to Contain 
 
 coo gallon pump 
 
 13 gallons. 
 
 1 7 gallons. 
 
 7 co ' ' 
 
 18 
 
 25 ' ' 
 
 
 24 
 
 30 ' ' 
 
 I ZOO ' ' 
 
 ?o 
 
 40 
 
 i^WVJ 
 
 
 
 The air chamber, combined with connections for discharge pipe, 
 relief valve, and hose valves, should be carefully designed to make 
 the whole weight as small as possible. Keeping this weight down makes 
 the pump run steadier and brings less stress on the flanges at high 
 speeds. 
 
 An air chamber of hammered copper and warranted tested under a 
 hydraulic pressure not less than 300 pounds per square inch is a little 
 better than cast iron, as it holds air better, and being lighter it wrenches 
 and strains the pump less when running fast and shaking, but because it 
 costs from $25 to $50 more than cast iron, it is not often adopted. 
 
 b. The vacuum chamber must be attached to the pump in the most direct 
 way practicable, but provision must be made for attaching it in such manner 
 as not to prevent readily taking off the cylinder heads. 
 
 c. Every vacuum chamber should be provided on one side near the top 
 with a |-inch pipe tap plugged. This to be used for attaching a vacuum 
 gauge if desired. 
 
 41'. Pressure Gauge, a. A pressure gauge of the Lane double tube spring 
 pattern with 5-inch case, must be provided with the pump, and connected 
 near to inboard side of air chamber, as shown in Fig. 381, by a J-inch cock, 
 with lever handle. 
 
 The dial of this gauge should be scaled to indicate pressures up to 240 
 pounds, and be marked " Water." 
 
 This kind of gauge is used on locomotives and is the best for with- 
 standing the vibration which causes fire-pump gauges to be often unre- 
 liable, Moreover, this double spring form is safer against freezing. 
 
488 
 
 PUMPING MACHINERY 
 
 42. Hose Valves, a. Hose valves must be attached to the pump (and 
 included in its price) as follows: 
 
 For the 2 stream or 5oo-gal. pump, 2 hose valves. 
 
 For the 3 stream or 75o-gal. pump, 3 hose valves. 
 
 For the 4 stream or looo-gal. pump, 4 hose valves. 
 
 For the 6 stream or i5oo-gal. pump, 6 hose valves. 
 
 These are to be 2^-inch straightaway brass valves, without cap, 
 and similar and equal in quality to those made by the Chapman Valve 
 Company, the Ludlow Valve Company, or the Lunkenheimer Company. 
 
 The hose-screw at end of these valves is to be fitted to a hose coupling 
 furnished by the customer, or where this cannot be procured may be 
 left with the thread uncut. 
 
 To accommodate locations where all the lines of hose must lead off 
 from one side of the pump makers can furnish a spool piece or special 
 casting to which the hose valves can be attached but this is an extra 
 not included in the regular price. 
 
 43. Safety Valve, a. A safety or relief valve of the Ashton, Crosby, 
 American, or other make agreed upon in writing with this office, is to be regu- 
 larly included in the price, and is to be attached to each 
 pump; preferably extending horizontally inboard from 
 base of air chamber, as shown in Fig. 381, so that its 
 hand-wheel for regulating pressure is within easy reach. 
 This hand-wheel must be marked very conspicuously, as 
 shown in sketch. 
 
 b. This valve is to be set ordinarily at a working 
 pressure of 100 pounds to the square inch, and is to be of 
 such capacity that 
 when set at 100 
 
 pounds it can pass all the water dis- 
 charged by the pump at full speed, 
 at a pump pressure not exceeding 125 
 pounds per square inch. 
 For 5oo-gallon pump, a 3-inch valve. 
 For 75o-gallon pump, a 3^-inch valve. 
 For icoo-gallon pump, a 4-inch valve. 
 For i5oo-gallon pump, a 5 -inch valve. 
 The relief valve must discharge in 
 a vertical downward direction 'into a 
 cone or funnel secured to the outlet of 
 the valve. (See Art. 44.) 
 
 The valve must be so attached to 
 
 FIG. 380. Hand 
 Wheel. 
 
 FIG. 381. Safety Valve. ' 
 
 the delivery elbow and discharge cone by flange connections as to permit of 
 its ready removal for repairs without disturbing the waste piping. 
 
 44. Discharge Cone. a. This cone should be so constructed that the 
 pump operator can easily see any water wasting through the relief valve, 
 and its passages should be of such design and size as to avoid splashing 
 water over into the pump room. 
 
 b. The cone must be provided with an opening to receive the air vent pipe 
 
SPECIAL PUMPING MACHINERY 
 
 489 
 
 required by Art. 45, and the arrangement must be such that the pump operator 
 can easily tell whether water is coming from the air pipe or is wasting through 
 the relief valve. 
 
 c. The cone should be piped to some point outside of the pump house 
 where water can be wasted freely, the waste pipes being as below. 
 
 Size of Pump. 
 
 Diameter of Waste 
 Pipe from Cone. 
 
 500 gallon 
 75 
 
 IOOO 
 I <OO ' ' 
 
 5 inches. 
 6 " 
 
 7 
 8 
 
 
 
 The waste pipe can pass down to floor between the yokes at middle 
 of pump. It should be piped in such a way that steam and gases from 
 other drains or waste pipes will not work back through it, and, by being 
 troublesome in the pump room, suggest the covering of the cone in any 
 way, as it is desirable that the pump operator should always be able to 
 see instantly any waste from the relief valve or air vent. 
 
 This cast-iron cone, connected to the safety valve and air vent, 
 is included in price of pump, but the waste pipe beyond it is not. 
 
 45. Air Valve, a. An air vent with a brass gate valve and brass pipe for 
 connecting up must be provided and connected with delivery elbow and 
 discharge cone. 
 
 b. The size of this air vent should be i inch for 5oo-gallon and 75o-gallon 
 pumps, and i\ inches for the looo-gallon and i5oo-gallon sizes. 
 
 c. The hand-wheel of this valve must be marked as per sketch. The 
 lettering must be very open, clear, and distinct, not 
 
 liable to be obstructed by grease and dirt, and of a 
 permanent character. 
 
 The object of this valve is to reduce the pres- 
 sure above force valves and secure a prompt riddance 
 of all air that may come through the water cylinders 
 when first starting up. 
 
 This valve, of course, should be closed, when 
 once pump is under way, to prevent waste of water. 
 
 46. Priming, a. Each pump must be fitted with 
 a set of brass priming pipes and valves, according to 
 either one or the other of the following methods : 
 
 b. For looo and i5oo-gallon pumps, the priming pipes must be i| inch. 
 For the 500 and 75o-gallon pumps, the pipes must be i inch. Pump-makers 
 are to furnish these pipes and the fittings called for below, and are to connect 
 them up providing a 2-inch outlet, looking upwards, ready for the supply 
 from the priming tank. 
 
 The pipe from the priming tank to this outlet should be at least 
 two-inch, and may be of iron, and is to be furnished by the purchaser. 
 All parts furnished by the pump-maker are to be of brass, and are to be 
 included in the price of the pump. 
 
 Controllable Valve Arrangement, c. Four two-seat controllable valves, one 
 for each pulsation chamber, and of the general type illustrated in Fig. 383, 
 
 FIG. 382. Valve 
 Handle. 
 
490 
 
 PUMPING MACHINERY 
 
 must be provided. In these the inlet of water and the outlet of air are simul- 
 taneously opened and closed by the pump operator. 
 
 This valve can preferably be provided with a flange connection in 
 place of the threaded one, and secured to water cylinder with three 
 five-eighth bolts. This will permit of easier fitting up as to pipe con- 
 
 FIG. 383. Valves. 
 
 nections. Objection has been raised to this double-seated valve from 
 the possible difficulty of keeping both seats tight. If desired, the stem 
 between the two seats may be somewhat enlarged and provided with a 
 suitable spring, thus giving flexibility between the two seats and pre- 
 venting all trouble from uneven wear. 
 
 d. The hand-wheel of each of these valves must be marked as per sketch 
 so that the pump operator may clearly understand 
 their use. The lettering must be very open, clear, and 
 distinct, not liable to be obscured by grease and dirt, 
 and of a permanent character. 
 e. There must be provided 
 and fitted to each combined 
 valve a check and umbrella- 
 top air vent, as shown in Fig. 
 385. This fitting must have a 
 clear passageway through it, 
 the full equivalent of a half -inch bore. 
 The check valve is to permit the outflow of air, but to prevent the 
 influx when the plunger is sucking. 
 
 FIG. ^84 Valve 
 Handle. 
 
 FIG. 385. Check Valve. 
 
SPECIAL PUMPING MACHINERY 491 
 
 This method is preferred to the one using rubber priming checks, 
 as now and then a rubber valve will stick on its seat and thus prevent 
 priming of one of the chambers. In this arrangement the pump operator 
 has absolute control over the priming water into each chamber. 
 
 Another advantage is that the connection of the air vent with the 
 priming valve ensures that the air vents will be opened ; and further, by 
 the vigorous spurting out of water as soon as the pump is primed, the 
 pump operator is reminded that the priming valve should be closed. 
 
 Should the pump operator, however, through a mistaken idea of the 
 proper method of operation, think that the priming should be con- 
 tinued until all air was exhausted from the suction pipe, and the pump 
 running in normal condition, there would be some by-passing between 
 chambers, but as there is a free vent for the air, the main result would 
 be simply to limit the amount of air exhausted per stroke, from the main 
 suction, by the amount of water which entered a chamber in this way. 
 The amount of water thus entering, however, would not be appreciably 
 greater than that which would enter from the priming tank with the 
 check- valve arrangement. 
 
 If, in spite of the warning given by the spurting air vents, the pump 
 operator should neglect to close the priming valves when the pump was 
 running normally, the priming tank would eventually be overflowed; 
 but this would not be as serious as the drawing in of air from an exhausted 
 priming tank, which would result with the check-valve method, were the 
 main two-inch valve similarly neglected. 
 
 Rubber Check Valves, f. Four rubber check valves, one for each pulsation 
 chamber, and similar to ordinary pump valves, must be provided. The 
 chambers for these should preferably be made as a part of the pump cylinder, 
 thus securing a compact arrangement. 
 
 g. The valve seat should have three ribs to the central hub, supporting 
 the rubber valve. The net port area through the valve should not be less 
 than one and one-half square inches. 
 
 This valve seat should- rest in an inverted position, and can be so 
 fitted up as to be readily removed. The valve 
 stems can be of the removable type screwing into 
 the seat, but must be made long enough to receive 
 a check nut on the opposite side of seat. This 
 will effectually lock the stem in place. 
 
 h. Care must be taken to arrange the water pas- 
 sages through and about these priming checks, so as 
 to avoid all air pockets and so as to reduce to a mini- 
 mum the possibility of the valves becoming choked 
 up by refuse. 
 
 i. The valve seats, stems and all parts must be 
 of composition and of strong rugged design, so fitted 
 up that there is the least chance for the rubber valves 
 to stick, and with all parts securely put together the 
 valves must be readily accessible. 
 
 ;'. The valve springs must have only sufficient strength to keep the valves 
 on their seats, so that they will freely open even with the low head of priming 
 water often existing. 
 
 k. There must be provided, and attached to the top of each plunger cham- 
 ber, a brass check valve and air cock with umbrella top, as shown by Fig. 
 
492 PUMPING MACHINERY 
 
 386. This cock and valve must have a clear passageway through them the 
 full equivalent of a half -inch bore. 
 
 The check valve is to permit the outflow of air, but to prevent the 
 influx when the plunger is sucking. Cocks with lever handles, are used, 
 as these show clearly whether they are open or shut. 
 
 /. There must also be provided a two-inch brass gate valve for the general 
 control of the water to the four check valves. The hand-wheel of this valve 
 must be marked as per Fig. 384. The lettering must be very clear, open, and 
 distinct, not liable to be obscured by grease, and of a permanent character. 
 
 It is essential for a properly working pump that the main two-inch 
 priming valve should be closed as soon as the pump is primed. Other- 
 wise water will be drawn from the priming tank, lessening the lifting 
 power of the pump through the main suction, and if this is continued 
 the priming tank will often be exhausted and air drawn into the pump, 
 interfering with its proper action. It is for this reason that the marking 
 on the priming valve is required. 
 
 For all average situations, either method of priming permits of getting 
 the pump under way in a very few minutes, but, for cases where the 
 suction pipe is over 300 or 400 feet in length, or sometimes where the 
 lift is over 18 feet, or where there is a combination of long length and 
 lift within these limits, so much time is consumed in exhausting the air 
 from the suction pipe that it becomes desirable to supplement this 
 method. 
 
 For such situations, a steam ejector connected to the suction pipe 
 near the pump is advised, and may be required in addition to the regular 
 priming pipes and tank. The size of the ejector should be sufficient 
 to exhaust the suction pipe within about three minutes. Such ejectors 
 will be considered as extras not included in the ordinary pump fittings. 
 
 For cases where pump can only take its suction under a head, if 
 absolutely certain that the level of the suction water will never fall 
 below level of center of pump, these priming pipes may be omitted, but 
 openings for them, into the pump shell must be provided and capped 
 or plugged. 
 
 A foot valve on a fire- pump suction is not advised except in very rare 
 cases, as with a lift of 18 feet or a suction pipe 500 feet or more long. 
 A foot valve is not needed when there is a good efficient set of priming 
 arrangements, as described above, and it is commonly found it gives a 
 false sense of security, and that with a fire pump left standing several 
 days the water will often be found to have leaked back, so that it is no 
 better than if no foot valve had been used. 
 
 A foot valve must of necessity generally be located where it is inac- 
 cessible for quick repairs, and as they grow old, foot valves are often 
 a source of trouble. Where a suction pipe is exposed even slightly to 
 frost, a foot valve is especially objectionable. 
 
 A priming tank is provided by the purchaser in all cases where there 
 is ever to be any lift on the suction. It is generally advised that this 
 tank have a capacity of one-half of what the pump can throw at full 
 speed in a minute. This means 250 gallons for a 5oo-gallon pump, and 
 500 gallons for a icoo-gallon pump, etc. It is the intention to make 
 the pump a truly " independent source " of supply, therefore the need 
 of a special priming tank. 
 
 Older Priming Arrangements. The form of priming arrangements here- 
 tofore used, with metal check valves, one main two-inch priming valve, and 
 one- inch priming pipes, separate controllable air cocks, may be retained on 
 all pumps at present in service, and will be considered satisfactory, if kept in 
 good order. 
 
SPECIAL PUMPING MACHINERY 493 
 
 If in any case such checks give trouble the priming arrangement may be 
 changed and valves like Fig. 383 or rubber checks, as described in sections 
 /-/, made up in detachable form, may be put on if desired, where the con- 
 nections on the pump permit them. 
 
 Where neither method is desired or where neither is feasible the faulty 
 checks may be replaced by a special type such as are now made for this use 
 by the Locke Regulator Company, of Salem, Mass. These are one-inch 
 check valves, adapted to use a small disc of medium hard rubber, similar to a 
 pump valve. 
 
 These fittings are very near the dimensions of the commercial check 
 valve, so that with sight shortening of piping connections they will fit into 
 the present arrangements, and give satisfaction. 
 
 47. Drain Cocks, a. Five brass drain cocks, each with a lever handle 
 and of half-inch bore, are to be provided, and located one on each end of 
 each water cylinder, and one above the upper valve deck. 
 
 Care should be taken to select a pattern of cock whose passageway 
 is the practical equivalent of a half-inch hole. Some patterns of half-inch 
 commercial cocks although threaded for half-inch pipe thread have but 
 a quarter-inch hole through them. Such are not acceptable. 
 
 TESTS FOR ACCEPTANCE. 
 
 48. Test for Smoothness of Action, a. Provide outlets for the water; 
 start the pump slowly, gradually open steam-throttle to bring the pump to 
 full speed. The pump should run smoothly at the rated full speed of 70 
 revolutions per minute (or 60 revolutions if a i5oo-gallon pump) with full 
 length of stroke, and meanwhile maintain a water pressure of 100 pounds 
 per square inch. 
 
 If the hose lines are short, or discharge is too free, partly close the water 
 outlet valves, thus throwing an extra back pressure on the pump equivalent 
 to that which would be produced through a greater length of hose. 
 
 During this trial it is preferable to discharge the water through 
 lines of aj-inch cotton rubber-lined hose, preferably each 150 feet long, 
 each connected directly to the hose outlets on the pump, and each line 
 having a i^-inch smooth nozzle at its outer end. Two lines should .be 
 connected for a 5oo-gallon pump, three for a 750, and so on, having 
 as many lines as rating of pump requires. 
 
 A hose Iine"i5o feet long, with. an inside surface of average smoothness, 
 and .with a i^-inch nozzle attached, will require about 80 pounds pressure 
 at the pump to discharge 250 gallons per minute, and the nozzle pressure 
 will be about 45 pounds. Therefore, with lines attached as above, q, 
 pressure at the pump of about 80 pounds should represent a discharge 
 about equal to the rated capacity of the pump, and would ordinarily 
 correspond with the rated full speed revolutions. 
 
 If the pump runs smoothly under these conditions, it is well to open 
 the throttle somewhat further, and bring the pressure at the pump up 
 to 100 pounds. This will give a discharge of about 280 gallons per stream, 
 or about 12 per cent in excess of the rated capacity. The revolutions 
 will, of course, correspondingly increase, and under all ordinary conditions 
 a pump should run smoothly at this higher capacity, though a little 
 more vibration and pounding would be expected than when running 
 simply at its rated speed. 
 
494 
 
 PUMPING MACHINERY 
 
 After cushion valves are adjusted there should be no noteworthy 
 water hammer or valve-slam. Sometimes valve-slam is not the fault 
 of the pump, but arises from an obstructed suction pipe. It is objection- 
 able to doctor water hammer in a pump by shifting air into the suction, 
 as this cuts down the efficiency and is a poor expedient. 
 
 The quietness of that part of the hose near the pump, or its freedom 
 from rubbing back and forth crosswise an inch or more with each pulsa- 
 tion of the pump, is a good index of the pump maker's skill in securing 
 uniform delivery. Bad pulsation quickly wears holes in the hose, and 
 to reveal this is the object of testing with hose connected directly to the 
 pump. 
 
 49. Test of the Internal Friction, a. This is shown by the reading of 
 steam gauge compared with water pressure gauge at air chamber. 
 
 Tests have generally run about as follows, for pumps running at full 
 rated speed: 
 
 Size Gallons 
 per Minute 
 Capacity. 
 
 Ratio of 
 Steam Piston 
 Area to 
 Water Piston 
 Area. 
 
 Water 
 Pressure Ibs. 
 per Sq.in. 
 
 Steam 
 Pressure 
 Theoretically 
 Necessary, 
 Disregarding 
 Friction. 
 
 Excess of 
 Steam Pres- 
 sure Needed 
 to Overcome 
 Friction Back 
 Pressure, etc. 
 
 Actual 
 Steam Pres- 
 sure Found 
 Necessary 
 at the Pump. 
 
 500 
 
 4 Times 
 
 IOO 
 
 2 5 
 
 15 
 
 40 
 
 75 
 
 3 
 
 IOO 
 
 33 
 
 12 
 
 45 
 
 IOOO 
 
 3 
 
 IOO 
 
 33 
 
 12 
 
 45 
 
 1500 
 
 2f 
 
 IOO 
 
 36-5 
 
 13-5 
 
 5 
 
 b. The steam pressure needed will vary slightly with the freedom of the 
 exhaust pipe and with the tightness of the packings, etc., but a steam pressure 
 of 45 pounds at the steam chest should suffice for 100 pounds water pressure 
 on pump in proper adjustment. 
 
 50. Test of Strength and Tightness, a. First, shut the main valve between 
 the pump and the fire system lest a sprinkler head be burst, then shut all 
 water outlets nearly, but not quite tight, so pump will move very slowly. 
 Screw safety valve down hard. Slowly and carefully admit steam pressure 
 sufficient to give 240 pounds per square inch water pressure. 
 
 b. With this extreme pressure all joints should remain substantially tight, 
 and the slow motion of the pump should be tolerably smooth and uniform. 
 (The leakage of a few drops here and there and a little unsteadiness of motion 
 are to be expected.) 
 
 c. If boiler pressure is above 85 pounds, the safety valve on pump should 
 be attached and screwed down only enough to hold the required pressure. 
 For with 100 pounds or more of steam the water pressure might be carried 
 too high. 
 
 After completing the above test slack off on safety valve, setting it so that 
 it will begin to open at about 100 pounds pressure. 
 
 51. Test of Capacity of Safety Valve, a. The relief valve may next be 
 tested by first adjusting it to pop at 100 pounds, then shut the main outlet 
 to pump, and then shut the hose gates one by one, and thus force all the dis- 
 charge through the relief valve, meanwhile opening steam throttle, so as to 
 
SPECIAL PUMPING MACHINERY . 495 
 
 run pump first at two-thirds speed or about fifty revolutions per minute, and 
 finally at full speed (seventy revolutions). The safety valve (relief valve) 
 should carry all this and not let the pressure rise above 125 pounds. 
 
 The pressure in a quick-moving fire-pump necessarily fluctuates 5 
 to 15 pounds at different points in stroke, and an air chamber of reasonable 
 size cannot wholly remove this. Therefore the safety valve must be 
 set at about 1 5 pounds higher than the intended average working pressure; 
 otherwise it will get to jumping with almost every stroke. 
 
 52. Test of Internal Leakage or Slip. a. Set safety valves at 115 pounds, 
 shut all water outlets, admit steam enough to give 100 pounds water pressure, 
 then pump will move very slowly under the influence of the leakage past 
 plungers; about one revolution of pump per minute shows a proper accuracy 
 of fit. Anywhere from one-third to two revolutions per minute is satisfactory. 
 
 Too tight a fit is bad, as if not exceedingly uniform it induces scoring 
 or fretting of the metals. Moreover, should pump happen to be run 
 dry for a few minutes before catching its suction a slight warming and 
 expansion of the plunger may cause it to stick and fret. 
 
 53. Test With Maximum Working Pressure, a. For this, alternately 
 shut down the main outlet gate and adjust the hand-wheel of the safety valve, 
 and open up on the throttle as may be required, running pump at say one-half 
 speed (or, in experienced hands, at full rated speed), and note the greatest 
 water pressure which the full boiler pressure (unless boiler pressure is above 
 85 pounds) will yield with pump at full speed. 
 
 Sometimes it may be necessary to force water through very long 
 lines of hose, or to an unusual height. 
 
 Steam fire engines are not infrequently called on to give 200 pounds 
 per square inch water pressure. 
 
 To test short hose lines with anywhere near so high a pump-pressure 
 is dangerous, lest nozzle kick and pull itself away from the man holding 
 it and thresh around; but the ability of the pump may be tested by 
 puttjng this high-pressure delivery mainly through the safety valve, 
 or in part through the partially closed main outlet gate. 
 
 It is not advisable to carry this water pressure above 200 pounds in 
 this test at the factory, although in the shop test the water pressure is 
 carried to 240 pounds, and engine driver should stand with his hand on 
 the throttle. 
 
 '54. Test for Maximum Delivery, a. This can best be tried by adding 
 one or, in some cases, two more streams than the pump is rated to deliver 
 by attaching the extra lines of hose to some hydrant near, and then speed 
 up the pump gradually, to see how fast it may be run before violent pounding 
 or slamming of valves begins. 
 
 Sometimes the increased delivery can be drawn off through an open 
 hydrant-butt, meanwhile holding sufficient back pressure to show 100 
 pounds on the water gauge by partly closing the discharge valve. 
 
 The engine driver should stand with his hand on or neai the throttle 
 when thus speeding the pump. 
 
 It is all right to run a fire-pump up to the utmost speed possible 
 before water hammer begins, and very often a pump, while new and if 
 favorably set up, can deliver 25 to 50 per cent more than rated capacity: 
 nevertheless, although expert treatment can force 1000 gallons from a 
 16x9x12 pump we can rate it as only a 75o-gallon pump. There 'must 
 be some margin to allow for wear and for the possible absence of the expert 
 at time of fire. 
 
496 PUMPING MACHINERY 
 
 The main points of difference between the "National Standard" and the 
 "Trade Pump" are: 
 
 Brass plungers instead of cast-iron plungers. 
 
 Wrought-iron side levers instead of cast-iron; 
 
 Bronze piston rods and valve rods instead of iron or steel. 
 
 Pump has brass lined stuffing boxes instead of cast-iron. 
 
 Rock shafts are brass bushed. 
 
 Area of water valves is 25 to 50 per cent greater. 
 
 Steam and exhaust passages 20 to 50 per cent greater. 
 
 Suction pipe connections two or four inches greater diameter. 
 
 Cushion valves better arranged. 
 
 Air chamber is made much larger. 
 
 Shells and bolting are warranted especially strong. 
 
 The following necessary fittings are included in the price, and regularly 
 furnished as a part of this pump, viz: 
 
 A capacity plate. 
 
 A stroke gauge. 
 
 A vacuum chamber. 
 
 Two best quality pressure gauges. 
 
 A water relief valve of large capacity. 
 
 A cast-iron relief valve discharge cone. 
 
 A set of brass priming pipes and special priming valves. 
 
 From two to six hose valves. 
 
 A sight feed cylinder lubricator connected above throttle. 
 
 A one-pint hand oil pump connected below throttle. 
 
 The hydraulic pressure pumps are built for very high pres- 
 sure. In these pumps, as shown in Fig. 387, the large steam 
 cylinders are connected with small water cylinders GG. . In 
 the pump shown steam is used in the cylinder E and. then 
 discharged into cylinder F. The piston rods A A are connected 
 to the cross head B and from this the plungers extend into the 
 two water cylinders. The connecting rod C is forked at DD 
 so that it clears the right hand pump G. The use of the fly- 
 wheel permits of the expansive use of steam. Such pumps 
 are often connected to a cylinder closed by a weighted ram 
 (an accumulator), so that the water may be stored under pres- 
 sure which may amount to several thousand pounds per square 
 inch. The use of an accumulator permits the pump to run 
 continuously even though the use of water is intermittent. 
 When there is such a load that the pump will have a number 
 of periods of inaction, a direct-acting pump is better than a 
 fly-wheel pump with a fixed stroke, as there is then little or no 
 

 FIG. 387. Pres 
 
3 ump. 
 
 (To face page 496} 
 

 
SPECIAL PUMPING MACHINERY 
 
 497 
 
 danger of the pump being wrecked from a collection of con- 
 densation in the stearn cylinder. The valve boxes HH are 
 similar to the pressure valve boxes shown earlier. 
 
 At times triplex pumps are used for this purpose. Such 
 pumps, Fig. 388, are made of heavy parts. The water is pumped 
 directly into the ram cylinder and from this it is discharged to 
 the reservoir. A small by-pass valve is set to discharge from 
 
 FIG. 388. 5oo-ton Press and Pump. 
 
 the pump to the reservoir, when the hand valve at the press 
 is shut off or as soon as the pressure in the system reaches a 
 certain amount. The 15 horse- power motor used with this 
 pump of the Waterbury Farrel Foundry and Machine Co. is 
 directly connected to the pump, although at times belted con- 
 nections are used. The proportions of the pump and press 
 are given in the table below, which is taken from the catalogue 
 of the builders. 
 
498 
 
 PUMPING MACHINERY 
 
 Press. 
 
 Pump. 
 
 Capacity to 
 
 ns 500 
 i. 16 
 
 5 
 
 22 
 
 40 
 
 18^ by 12^ 
 26 
 
 165 by .36 
 
 82i 
 S. I7I7S 
 
 Diameter of plunger 
 
 ...in ij 
 6 
 
 Maximum stroke of ram 
 Diameter top of ram 
 
 Approximate capacity, 
 inches 
 Pressure, Ibs. per sq.in . 
 
 cubic 
 1060 
 4975 
 
 Top platen to ram when down 
 Opening between rods 
 Floor to top of ram 
 Size of supply pipe 
 Floor space, with pump .... 
 Extreme height from floor . . . 
 Total weight, including motor. It 
 
 R.P.M. of crank-shaft. . . 
 Ra f io of gearing 
 
 61* 
 
 Floor space 
 
 . . . in. 91 by 32 
 
 
 
 Figs. 389 and 390 show two air pumps for beer racking. 
 These pumps are mounted on the air reservoir which acts as a 
 
 FIG. 389. Worlhington Beer-racking Pump. (Size 5^X6X5.) 
 
 bed plate. This gives proper storage capacity. The cylinders 
 are lined with a composition. The regulator shown on each 
 pump is connected with the air reservoir and keeps the pressure 
 constant. The pump of Fig. 390 shows how a fly wheel can be 
 used between the cylinders of the steam and air ends. A yoke 
 is placed in the rod between the two cylinders and a crank pin 
 of the shaft operates between these. This is the equivalent 
 
SPECIAL PUMPING MACHINERY 
 
 499 
 
 of a connecting rod of infinite length, and by its use the fly wheel 
 may be applied so that steam can be employed expansively. 
 
 FIG. 390. Beer-racking Air Pump. 
 
 The table below gives the sizes of this type of pump as made by 
 
 Worthington : 
 
 1 
 
 * 
 
 0) 
 
 1 
 
 
 Sizes of Pipes for 
 Short Lengths 
 to be Increased as 
 Length Increases. 
 
 Approximate 
 Space Occupied. 
 Feet and inches. 
 
 *o 52' 
 
 "o 
 
 t/2 
 
 Annticil C/3,pa,city of 
 Brewery. 
 
 
 
 
 
 
 
 s! 
 
 1.1 
 
 .3 
 
 
 <u 
 
 j fl) 
 
 g 
 
 
 
 jA 
 
 
 S^ 
 
 B7, 
 
 1 
 
 
 IK 
 
 | 
 
 .s a 
 
 ^^ 
 
 M 
 
 D 
 
 1 
 
 Q 
 
 Q 
 
 j 
 
 
 
 
 w 
 
 Q 
 
 JS 
 
 'P 
 
 5l 
 
 si 
 
 5 
 
 20, ooo to 5o,ooobbls. 
 
 f 
 
 i 
 
 !l 
 
 ! 
 
 3 7 
 
 1 9 
 
 6 
 
 8 
 
 6 
 
 50, ooo to ioo,ooobbls. 
 
 i 
 
 ii 
 
 2 
 
 1^ 
 
 4 9 
 
 2 4 
 
 7i 
 
 9 
 
 10 
 
 100, ooo to i5o,ooobbls. 
 
 J i 
 
 2 
 
 3 
 
 3 
 
 6 7 
 
 2 5 
 
 9 
 
 12 
 
 10 
 
 1 50,000 to 200,000 bbls. 
 
 2 
 
 *\ 
 
 4 
 
 3 
 
 6 10 
 
 3 o 
 
 To designate the sizes, give the diameters of the steam and air 
 cylinders and the length of stroke. 
 
500 
 
 PUMPING MACHINERY 
 
 The hydraulic ram in one of its modern forms is shown in 
 Fig. 391, while Fig. 392 illustrates the method of connecting 
 it. The drive pipe F connects the source of supply A with the 
 ram at B at a lower level. The water can escape from the ram 
 by an opening or waste valve K. The velocity set up at the 
 discharge opening is sufficient to lift the valve G and suddenly 
 
 FIG. 391. Hydraulic Ram. 
 \ 
 
 close the passage. The momentum of the water in the pipe line 
 produces an increase in the pressure near the valve, as shown 
 from the rise from a to b in Fig. 393. This figure was drawn 
 by the pencil of an indicator attached to the pipe F, known as 
 the drive pipe, while the drum was moved uniformly. The 
 heights, therefore, represent pressure while length represents 
 time. The increase of pressure will cease as soon as the pressure 
 is sufficiently great to open the ball valve H when the water 
 
SPECIAL PUMPING MACHINERY 
 
 501 
 
 flows into C. The inertia and friction of the valve being over- 
 come, the pressure drops with a backward wave to c and returns 
 to d, after which the ball valve closes; the pressure drops in the 
 space around G and a backward wave may even reduce this to a 
 pressure below the atmosphere. In any case the weight / is 
 
 FIG. 392. Arrangement of Ram. 
 
 sufficient to force G down against the water pressure due to 
 the elevation of A, after which the operation is repeated at 
 g, h, i, etc. The discharge is taken from L by the discharge 
 pipe D which conducts it to the tank house. 
 
 In making and installing the ram, the design of pipe and 
 size is fixed by practise, although there are theoretical deductions 
 made at times. (Zeitschrift des Vereines Deutscher Ingenieure, 
 
 h 9 e a- 
 
 FIG. 393. Card from Hydraulic Ram. 
 
 Jan. 15, 1910.) The drive pipe F should be of ample size 
 and should be laid as straight as possible. It should be of a 
 length at least equal to three-quarters the head against which it 
 is to pump, or five times the head causing the flow through the 
 drive pipe. The inclination of the drive pipe should never be 
 over 30 at any place, and when the length or proper angle 
 cannot be had it is well to coil the drive pipe to a large radius so 
 
502 PUMPING MACHINERY 
 
 as to accomplish this. The length of the pipe is necessary to 
 produce the momentum for driving. The length of delivery 
 pipe should not be over twenty times the lift and when a longer 
 pipe must be used less lift will be obtained. 
 The efficiency of the hydraulic ram is 
 
 w(H-h) 
 ~Wh~' 
 where 
 
 #=head from ram to reservoir in feet; 
 /j=head from supply to ram in feet; 
 / = time in seconds; 
 w water pumped in Ibs; 
 W = water discharged through waste valve in Ibs. 
 
 This efficiency varies from o . 40 to o . 60, and by assuming o . 50 
 for its value, the equation may be used to find W when w, H 
 and h are known. The drive pipe should then be made large 
 enough to give a velocit}^ of three feet per second and the 
 delivery pipe two feet per second. 
 
 62.5X/X3 ' 
 ^ del - = 62.5X/X2' 
 
CHAPTER XII 
 
 . 
 
 INJECTOR AND PULSOMETER 
 
 THE Sellers injector, Fig. 394, has been taken as a type of 
 the injector. In starting, the handle A is drawn back a short 
 distance permitting steam to enter the space B between the 
 two tubes. The pressure of this steam between the two edges 
 produces so high a velocity of discharge across the upper end 
 of space C that it entrains the air at that point, driving it out 
 through the openings D E into the space G, and from this 
 
 FIG. 394. Sellers Injector. 
 
 into the atmosphere. This produces a partial vacuum at C 
 and water enters through the suction pipe H. This water is 
 taken up by the steam discharge, driven through the combining 
 tube / and is afterwards discharged through Z), and E, finally 
 appearing at the overflow. 
 
 When water appears at the overflow, the handle A is drawn 
 back completely and steam then discharges through the main 
 nozzle K. The steam in this acquires a high velocity due to 
 
 503 
 
504 PUMPING MACHINERY 
 
 the drop in pressure in the nozzle; this velocity is many times 
 greater than 'that acquired by water under the same drop in 
 pressure on account of the small density of the steam. 
 
 The steam is then condensed by the water, but when the 
 steam particles draw together to form a drop of water their 
 velocity is maintained and drops of water, moving with a 
 very high velocity, result. This condensed steam can then 
 strike the body of water drawn through H and by impact, 
 impart to it a velocity greater than that which would be acquired 
 by water discharging from the same boiler. If this is the case, 
 this mixture of water and condensed steam could enter the 
 boiler against steam pressure. 
 
 The combining tube 7 is made convergent since the steam 
 is gradually condensed in the passage and the water gradually 
 increases in velocity. When the throat L is reached the steam 
 is condensed and the water is supposed to have its maximum 
 velocity. From this point the delivery tube M diverges, making 
 the velocity less, and with this comes a change of velocity head 
 into pressure head until at the end there is sufficient pressure 
 to force open the valve when the mixture will enter the boiler. 
 
 The pressure in the space within the end of the combining 
 tube reaches a point of low vacuum of about 25 " while at 
 the end of the steam nozzle it is at atmospheric pressure. At 
 the small part of the nozzle known also as the throat, the pres- 
 sure is found to be 0.58 of the boiler pressure. The pressure 
 at the throat of the delivery tube is usually about atmospheric 
 pressure. 
 
 If how the pressure of the steam is p pounds per square 
 inch, absolute, and the barometer reads p a , the pressure against 
 which the water must enter is (p pa) and if ^ is the density of 
 the water or weight per cubic foot at the throat, and v-2 the veloc- 
 ity entering the boiler, the head in feet of water to be overcome is 
 
 2g 
 
 A may be taken as 50 pounds for the first approximation. 
 JTie reason for this low density is the fact that steam particles 
 
INJECTOR AND PULSOMETER 505 
 
 are mixed with the water. Such a result has been found from 
 experiment. The velocity at the point E will be 
 
 V d = 
 
 The velocity of water drawn through the suction tube is 
 produced by the pressure p c within the combining tube, and 
 if the water is lifted h feet and the friction is h/ feet the. head 
 causing flow is 
 
 62.5 
 
 the velocit of the water is 
 
 The steam discharging from the nozzle is under a pressure 
 of p originally, and in the combining tube it is under a pressure 
 of p e . To find the velocity, the heat content H of the initial 
 steam is found, considering the quality of the steam and then 
 the heat content H c at the lower pressure, assuming that the 
 steam has the same entropy at the two points. If 15 per cent 
 friction loss is assumed, the velocity V of the steam is given 
 by the equation: 
 
 S5 2 -[H-Hc] where H = q+xr and A = --. 
 
 Having now the velocities of the steam, water and mixture, 
 the equation for impact may be used to find the amount of 
 water per pound of steam. 
 
 Since impact is not perfect, the sum of the first products has to 
 be multiplied by a factor K. Mr. Strickland L. Kneass, in his 
 " Theory of the Injector," has shown that a value of 0.50 
 should be used for K in the formula: 
 
 KV-V d 
 
506 PUMPING MACHINERY 
 
 With M known, the condition of the mixture can be found by 
 equating the heat on each side of the throat in which q d is the 
 onl unknown. 
 
 of liquid in suction, 
 2^= heat of liquid of discharge, 
 H = total heat content of entering steam. 
 
 The density of the water in the discharge depends on the 
 temperature of the water and the amount of steam left in the 
 water at this point. Kneass has shown how this may be found, 
 and in some cases its value is o.2iw, where w is the weight of 
 a cubic foot of water at the temperature of discharge. This 
 result was determined for a small discharge. The value 0.7520 
 is more nearly the value to be used in practice. This would 
 give 45 pounds with water weighing 60 pounds per cubic foot. 
 
 If the weight of water wanted per hour is W pounds, the 
 amount of steam per second is 
 
 
 6ox6oM 
 
 hence the orifice must be of area sufficient to cany w'(i+M) 
 pounds of density A at the speed of V d feet per second. The 
 area at the throat is therefore 
 
 The diameter of this expressed in millimeters gives the 
 nominal size of the injectors according to some makers. 
 
 The area of the throat of the steam nozzle is sometimes 
 made two or three times the area of the delivery throat, but 
 this may be designed by considering the velocity set up by 
 the fall of pressure from p to 0.57/7. Let H be the heat content 
 for the steam under the conditions at entrance and H t that 
 at 0.57/7, but with the same entropy, then 
 
 778' 
 
INJECTOR AND PULSOMETER 50? 
 
 The factor 0.98 is used here because of the short easy curve. 
 
 A -1 
 V t 
 
 S t is the specific volume of the steam at the throat. 
 
 If He is the heat content of the steam at the end of the 
 nozzle where the pressure has fallen to about atmospheric 
 pressure, the following may be used to find the area at the end: 
 
 V, 
 
 H e is taken from the same entropy column as H but in finding 
 V e , the volume is that which corresponds to a heat content 
 H' e =H e +Q.i5(H-H e ). The reason for this is the fact that 
 the loss due to friction has been changed into heat and remains 
 in the steam increasing the heat content over that resulting 
 from true adiabatic expansion. 
 
 The nozzle is usually rounded to the throat area with quad- 
 rants and then the remainder of the nozzle is made along a 
 straight line to the end. The combining tube slopes from the 
 delivery throat to an area somewhat larger than the nozzle end. 
 
 The delivery tube should be made of gradual convergence, 
 and Kneass recommends ("Theory of Injectors," p. 37) to use 
 a curve such that the negative acceleration along the tube 
 is constant. This gives the equation: 
 
 4 /" 
 
 v V 2 -2ay 
 
 Where R is the radius of the tube at a distance y from the throat 
 of radius r and a is the acceleration while V is the velocity at 
 the throat. 
 
 Suppose that the velocity is to be destroyed in / feet of 
 tube, then 
 
 F 2 
 
508 
 
 PUMPING MACHltf&tV 
 
 This may be put in the above equation and the following 
 results: 
 
 from this the various radii may be found along the tube. 
 
 The shapes of the tubes actually used in practice have been 
 determined by experiment, and the above serves as a guide for 
 certain leading dimensions. 
 
 The pulsometer described in Chapter II is built under a 
 number of names and with various shapes, but of the same 
 principle. Fig. 395 shows the Emerson Pump. The cylindrical 
 vessels A A are for the same purpose as the pear-shaped vessels 
 of the pulsometer. 
 
 The pulsometer is operated by the action of the steam on 
 the water direct. The quantity of water to be handled by a 
 given size is fixed from practice and the table below gives these 
 sizes taken from a catalogue. 
 
 TABLE OF SIZES OF EMERSON STEAM PUMPS 
 
 
 e 
 
 
 
 
 
 
 
 *i2 
 
 Dimensions Over 
 
 2 
 
 ^ 
 
 
 T3 
 C 
 
 1 
 
 i 
 
 
 
 " 
 
 8, 
 
 
 
 "d 
 
 1 
 'rj 
 
 II 
 
 all in Inches. 
 
 1 
 
 
 
 
 
 O 
 
 ">. 
 
 B 
 
 . 
 
 rt 
 
 ,c 
 
 Ccj 
 
 
 
 O J 
 
 
 
 
 
 
 
 a 
 
 o 
 
 i 
 
 ts 
 
 jj 
 
 C3 
 
 c jj 
 
 S -=: 
 
 
 
 
 ^-> 
 rt 
 
 Number. 
 
 Diameter 
 Inches. 
 
 Length of 
 Feet. 
 
 i SizeofStc 
 Inches. 
 
 |l 
 
 u 
 
 ft 
 
 P 
 
 1! 
 |l 
 
 y 
 
 Js 
 
 o 
 
 *** rS 
 
 &3 
 
 u 
 
 O 
 
 Breadth. 
 
 Width. 
 
 ^3 
 I 
 
 Approxim 
 in Ibs. 
 
 I 
 
 6 
 
 6 
 
 a 
 
 4 
 
 3 
 
 1 
 
 225 
 
 i3.5oo 
 
 324,000 
 
 i6i 
 
 18 
 
 97* 
 
 95 
 
 2 
 
 8 
 
 6i 
 
 I 
 
 4 
 
 3 
 
 415 
 
 24,900 
 
 597,600 
 
 21^ 
 
 21 
 
 104 
 
 1370 
 
 3 
 
 10 
 
 7 
 
 II 
 
 5 
 
 4 
 
 7 2 5 
 
 43.50 
 
 1,044,000 
 
 26 
 
 24 
 
 "3 
 
 J 905 
 
 4 
 
 12 
 
 8 
 
 I* 
 
 6 
 
 5 
 
 1200 
 
 72,000 
 
 1,728,000 
 
 29! 
 
 27i 
 
 127 
 
 3100 
 
 5 
 
 16 
 
 8 
 
 2 
 
 8 
 
 6 
 
 2IOO 
 
 126,000 
 
 3,029,000 
 
 43 i 
 
 33 
 
 132 
 
 4400 
 
 6 
 
 20 
 
 8 
 
 l 
 
 10 
 
 8 
 
 3275 
 
 196,500 
 
 4,7l6,OOO 
 
 5*1 
 
 36J 
 
 !35 
 
 5400 
 
 7 
 
 24 
 
 8 
 
 3 
 
 12 
 
 10 
 
 4700 
 
 282,000 
 
 6,768,000 
 
 
 
 
 7000 
 
 
 
 Capacities in gallons per minute, stated in table, vary with the 
 steam pressure and height of lift. Special sizes above those listed, 
 on application. Pumps made entirely of bronze, when called for. 
 at special prices. 
 
INJECTOR AND PULSOMETER 
 
 509 
 
 IP 
 
 r 
 
 FIG, 395. Section of the Emerson Steam Pump. 
 
510 PUMPING MACHINERY 
 
 To compute the quantity of steam used, suppose that steam 
 is supplied at a pressure pi and is used within the chamber 
 at pressure p 2 . Assume that there is no radiation and if this 
 is the case the steam in passing through the valves is reduced 
 in- pressure but the heat content remains the same. Hence 
 
 HI =H 2 . 
 
 From HZ at the pressure of p 2 , the specific volume S 2 of the 
 steam can be obtained and then for a given quantity of water, 
 V cu.ft, the weight of steam used would be 
 
 V 
 
 = wi. 
 
 Oo 
 
 ' , -. 
 
 The heat in each pound of this steam is H = q+Xiri, and if 
 the temperature of discharge is t d the heat chargeable per pound 
 of steam is Hq d . 
 
 q d and the temperature of the mixture is given by the equation 
 
 Mq a +m(qi +x^i ) = (M + m)q d + A 144^2 (M + 
 
 V 
 
 M =weieht of V cu.ft. of water = 7 , 
 
 62.5' 
 
 of liquid in suction. 
 
 = volume of I Ib. of w;iter= z . 
 
 62.5 
 
 The efficiency of the pump is 
 work 
 
 heat 7?8m(H-q d Y 
 
 The above equations have considered no loss due to initial 
 condensation nor radiation, and these depend on many con- 
 ditions. It may be said that the steam will probably be increased 
 to $m or more and this will reduce the efficiency to about one 
 
INJECTOR AND PULSOMETER 511 
 
 third its former value. The term q^ will be changed as its 
 equation becomes 
 
 Mq s + 3m(qi +xtfi ) -= (M +^m)q -{-Radiation +Work. 
 Radiation =K(A )(T - T 2 )t\ 
 
 K = 300 = amount of heat radiated per sq.ft. per hour 
 
 per degree difference in temperature; 
 A area of outside of vessels; 
 TI --= temperature on inside of pulsometer; 
 
 __ J- steam "r J- d 
 2 
 
 T2 = temperature air; 
 ^ = time in hours in which V cu.ft. of water is pumped. 
 
CHAPTER XITI 
 AIR LIFT PUMPS AND PNEUMATIC PUMPS 
 
 THE air lift pump, as was mentioned in Chapter II, was 
 patented in 1880 by James P. Frizell, and Pohle took out a 
 
 FIG. 396. Air Lift. 
 
 patent in 1886, although these two are by no means the first 
 records of this type of pump. 
 
 Fig. 396 .shows the general arrangement of a plant for an 
 
 512 
 
AIR LIFT PUMPS AND PNEUMATIC PUMPS 513 
 
 air lift. In this figure the air pump is attached to a storage 
 tank from which the air is conducted to the well. The air 
 compressor is shown to be of a single stage type although in 
 many cases two stage compressors are used to increase the 
 efficiency. The air cylinder A is in tandem with the steam 
 cylinder B. The air from the storage tank C, with its gauge and 
 safety valve, is conducted to the head of the discharge pipe 
 at E through the pipe D. This pipe is continued down to a point 
 F near the bottom of the delivery pipe. The well is usually 
 cased outside of the delivery pipe through all earthy matter 
 to solid rock if certain ground waters are to be kept from the 
 well. The water in the well stands at a height h below the 
 discharge and the depth of immersion from G to H is called h'. 
 
 The figure shows several methods of introducing the air. 
 In the first place the pipe is carried down inside of the main 
 pipe and opens 'directly into the delivery pipe. There are two 
 other figures showing the small pipes carried down into the 
 well and introduced into the side of the delivery pipe, while the 
 fourth figure shows the method of carrying the air down in 
 an annular space between two pipes, the inner one of which 
 is the delivery pipe. A slide valve, controlled from the top 
 of the well, admits air at some height above the bottom for the 
 purpose of introducing air at a higher point when starting the 
 apparatus. The air in this case is introduced at I into a head 
 at the well top. 
 
 There are several methods of arranging the well tops for 
 the reception of the discharge and for the introduction of the 
 air pipes. These are shown in Fig. 397. In the first a concrete 
 head A receives the discharge from the deflecting cap G. It 
 is carried away from this by the conduit leading to the reservoir 
 or irrigation ditch. The handle at H controls the admission 
 of the air by a rod which extends to the lower end of the air 
 pipe. When the water is to be carried to a higher level, the 
 S bend shown at B is used, but the control valve entering the 
 pipe at the ground level is similar to that used in A. When 
 an elbow is used, the connection for the air pipe is shown 
 at C and the forms at D and E are those used with the "i* 
 
514 
 
 PUMPING MACHINERY 
 
 supplied in the annular space outside of the delivery pipe. 
 In E the discharge is caught in a tank before being delivered, 
 while in D the discharge is directed into the irrigation ditch 
 or cistern. 
 
 Such installations may be arranged at considerable distance 
 from a central air compressing station. One station is used 
 to furnish air to a number of wells. The pipe lines should 
 be arranged so that the air travels at a rate of from 2000 to 
 4000 feet per minute. The area of this pipe is usually cne- 
 sixth the area of the delivery pipe. Some pump makers claim 
 that this method of raising water should not be applied when 
 the water has to be raised over 80 feet, but others name 180 to 200 
 feet as the limit, and when the water is to be carried by the air 
 
 FIG. 397. Air Lift Well Tops. 
 
 pressure in addition to a small lift the distance is limited to 
 700 or 800 feet. Be this as it may, greater lifts are used, although 
 the cost of lifting may be excessive. 
 
 When greater lifts are required it might be well to use 
 multiple lifts discharging through a portion of the height to 
 a deep pipe reservoir in the main well and from this to 
 another. 
 
 The compressors should in all cases be arranged to compress 
 
AIR LIFT PUMPS AND PNEUMATIC PUMPS 
 
 515 
 
 the air with as nearly isothermal compression as possible. The 
 only way to do this is to use an inter-cooler between the stages. 
 When the pressure is from 60 to 300 pounds, two stages 
 should be used, and above this three or more stages. The air 
 pipe should be introduced near the bottom of the discharge 
 pipe and should be immersed so far that the ratio of h' 
 to h is 3 to i at the start and 2.2 to i in operation, according 
 to a test on a particular well to find out what depth of im- 
 mersion gave the best results. This result, however, should 
 not be used as a general law as there are many conditions 
 affecting it. Some claim that the ratio should be 1.2:1 to 4: i. 
 These will do as guides. The following table gives a series of 
 tests showing what may be expected: 
 
 Place. 
 
 h' 
 
 Immersion -=- 
 ft 
 
 Efficiencies. 
 
 Tunbridge, England 
 
 3:1 to 2.2:1 
 
 ?6% 
 
 Grinnell, Iowa 
 
 
 2Q 6% 
 
 San Francisco 
 
 0.6 : i 
 
 16% 4-1% 
 
 San Francisco 
 
 i ' i 
 
 IO% 42% 
 
 San Francisco 
 
 1.4 : i 
 
 74% 41% 
 
 San Francicso 
 
 2 A. ' I 
 
 T r% 24% 
 
 San Francisco 
 
 } Q ' I 
 
 2% 
 
 
 j e 
 
 CO 
 
 
 
 o" 
 
 
 o 66 to i 
 
 
 
 o ^o 
 
 JV 
 
 2 C. 
 
 
 
 
 
 0.43 to i 
 
 
 The efficiency shown in this table is the ratio of the work 
 done on the water to that done in compressing the air. 
 
 The pressure to be carried on the air system is greater than 
 the water pressure at the lower end of the discharge pipe or 
 p =0.434/^+5. This quantity diminishes after the pump starts 
 to act as the head h' decreases, due to the removal of water. 
 The difference between the head at start and while running, 
 represents the head causing flow into the well. This quantity 
 will vary in different localities, depending entirely on the form- 
 ation and the nature of the water bearing rock. 
 
 The amount of free air varies according to different authors. 
 
516 
 
 PUMPING MACHINERY 
 
 One gives the amount as 3.9 cubic feet to 4.2 cubic feet of 
 free air per cubic foot of water, or 
 
 cu.ft. of free air per min. = f . 
 
 16.824 
 
 L = lift of water above surface in well in feet. 
 Q = cu.ft. of water per min. 
 
 A pump manufacturer recommends the following: 
 
 cu.ft. air = , 
 while another uses: 
 
 = cu.ft. of air per min. 
 
 The air is transmitted at 2000 feet per minute, and the 
 c~rea of water pipe should be about six times that of the air 
 pipe. The immersion should be i times the lift. The table 
 below has been given in "The Engineer" for June i, 1906: 
 
 Size of Well. 
 
 W*ater Pipe. 
 
 Air Pipe. 
 
 Gallons per Minute. 
 
 4" 
 
 ii 
 
 I 
 
 25 
 
 44" 
 
 2 
 
 I 
 
 5 
 
 5" 
 
 2 4 
 
 I 
 
 75 
 
 6" 
 
 3 
 
 it 
 
 IOO 
 
 7" 
 
 3i 
 
 J 4 
 
 150 
 
 8" 
 
 4 
 
 li 
 
 200 
 
 9" 
 
 5 
 
 2 
 
 300 
 
 10" 
 
 6 
 
 2 
 
 45 
 
 This same article gives the following relation for the cubic 
 feet of free air per cubic foot of water: 
 
 Lift. 
 
 Cu.ft. Air per 
 Cu.ft. Water. 
 
 . Submergence 
 if 1.5. 
 
 Air Pressure in Ibs. 
 
 25 
 
 2 
 
 37-5 
 
 17 
 
 5 
 
 2 
 
 75 
 
 33 
 
 75 
 
 IOO 
 
 I2 5 
 150 
 
 4-5 
 6 
 
 7-5 
 9 
 
 112.5 
 
 !5 
 
 187.5 
 225 
 
 49 
 65 
 82 
 98 
 
 J 75 
 
 IO -5 
 
 202 . 5 
 
 JI 5 
 
 200 
 
 12 
 
 300 
 
 130 
 
AIR LIFT PUMPS A21D PNLUI.IATIC PUMPS 
 
 517 
 
 As an example, the following data may be mentioned: 
 A lift of 129 feet was obtained from a well 300 feet deep; the 
 water was 44 feet from the top of the well, leaving an 85-foot 
 lift above the well. The well was 8 inches in diameter and 
 the discharge pipe was 3^ inches with a ij-inch air pipe. The 
 pump gave 82.5 gallons per minute and required 7.43 cubic 
 feet of air per cubic foot of water. The air pressure required 
 was 107 pounds and the loss in pressure was 9%. 
 
 As another example of the use of the air lift pump, 
 the plant at Redlands, California will be mentioned. This 
 was described in the "Engineering Record," Vol. 51, p. 8. 
 The plant was built to replace machinery at four wells, from 
 450 to 570 feet deep, which were separated some distance. The 
 first pump was of centrifugal form and was operated by a motor; 
 the second, a plunger pump driven 
 by a steam engine; the third, a 
 centrifugal pump with a steam 
 engine, and the fourth had not 
 been used. It was decided to 
 operate these wells by air from a 
 central station. The wells were 
 to be piped with 4-inch and 7- 
 inch discharge pipes, extending 
 down from 306 to 360 feet, while 
 the air pipes were ij to 2 inches 
 in diameter. 
 
 The station was a brick build- 
 ing 40X46 feet and equipped 
 with return tubular boilers 66 
 inches in diameter and 16 feet 
 long. These used oil for fuel. A 
 13 and 26x30 Cross compound 
 
 engine with 14 and 22X30 two-stage air cylinders in tandem, 
 supplied 1124 cubic feet of free air per minute to 125 pounds 
 at 85 R.P.M.; at this time it developed 190 I.H.P. The boiler 
 feed pumps at the station were a 2fx 4-inch duplex pump and 
 a 4! X3X 4-inch triplex pump. 
 
 FIG. 398. Wheeler System. 
 
518 PUMPING MACHINERY 
 
 The 48-hour test on this .plant gave the following results: 
 
 
 Mean for 48 
 Hours. 
 
 Maximum. 
 
 Minimum. 
 
 Boiler pressure 
 
 146 . 7 
 
 I SO 
 
 14.2 
 
 Air compressor pressure 
 
 QO 08 
 
 04. 
 
 8c 
 
 Vacuum in inches 
 
 24.83 
 
 2 C . 2 C 
 
 o 
 
 24. 
 
 Well No. i : 
 Air pressure 
 
 8? A 
 
 80 
 
 ^ 
 SA 
 
 Depth water level 
 
 101 44 
 
 I OQ 
 
 08 
 
 Well No. 2: 
 Air pressure 
 
 Q-2 . 6 
 
 S 
 
 
 Depth water level 
 
 QO . 77 
 
 Q-? 
 
 86 
 
 'Well No. 3: 
 Air pressure 
 
 87 .01 
 
 QO 
 
 8s 
 
 Depth water level 
 
 I IO 67 
 
 116 
 
 
 Well No. 4: 
 Air pressure 
 
 8c 4. 
 
 86 
 
 SA 
 
 Depth water level 
 
 IO2 2 
 
 JOS 
 
 IOI 
 
 Fuel consumed in barrels per 24 hrs . 
 Rate of pumping in gals, per 24 hrs . 
 
 I6. 3 
 
 3,157,622 
 
 17.2 
 3,280,824 
 
 15-7 
 3,i3i.395 
 
 When sufficient immersion can not be had, the Wheeler 
 system, Fig. 398, may be used. Water is admitted into 
 one of the two vessels A by drawing out the air, when water 
 will flow in from the well; then, on allowing air pressure to 
 enter through B, this forms an equivalent head on the water. 
 
 The action of the pump and compressor will now be exam- 
 ined. Suppose that v cubic feet of air are admitted per cubic 
 foot at the bottom of the discharge pipe, under a pressure of 
 feet. The specific gravity of the mixture of i cubic 
 
 foot of water and v cubic feet of air will be - , neglecting the 
 
 I Til 
 
 weight of air. 
 
 The pressure of the mixture in the pump pipe is h' +34 feet 
 of water at the bottom of the pipe and 34 feet at the top after 
 rising through h+h' feet of the pipe. This pressure is equal 
 to the weight above this point of the column of mixture of one 
 square foot cross-section plus the atmospheric pressure. It 
 
AIR LIFT PUMPS AND PNEUMATIC PUMPS 519 
 
 will vary according to some law which must be found. It is 
 evident, however, that, if it be assumed that the air is not 
 absorbed by the water, the volume of the air will vary inversely 
 as the pressure at any point considered, because there is so 
 much water in contact with the air that the expansion of the 
 air is isothermal. At any point x the pressure is that due to 
 the weight of a column of mixture x feet high plus the atmospheric 
 pressure of 34 feet of water. Call the weight of the column 
 wh x if expressed in feet of standard water. 
 
 , C x wdx 
 
 wh x = 
 
 -'-, 
 
 Jo 
 Now v x 
 
 Differentiating the expression for whx, 
 
 j 7 wdx 
 wdh x = = , 
 
 dx 
 
 dh x = 
 
 I+ ,^34)' 
 (A.+34) 
 
 fe-f34 
 
 x 
 When x = o, h x = o', and when x = h'+h y h x = ti. 
 
 .'. c = o-v(h'+34) Iog e 
 and 
 
 h'+h = h'+v(h'+ 34 ) log, 
 
 34 
 from this 
 
 34 
 or 
 
 h 
 
 v = 
 
 34 
 
520 
 
 PUMPING MACHINERY 
 
 Given h and //, v can be found. This is the amount of com- 
 pressed air per cubic foot of water pumped. To change this 
 to free air per cubic foot of water the following is used : 
 
 _ 
 
 V a 
 
 34 
 
 If water is desired to be delivered at a given rate at the 
 bottom of the main discharge the velocity of entrance is given 
 by the equation: 
 
 Now 
 
 or 
 
 If h" is found this amount should be subtracted from h' to 
 get the effective head on the inside. In that case the head 
 inside of the casing is h" feet less than that outside and this 
 
 FIG. 399. Compressor Diagram. 
 
 will give w the velocity of approach. This method gives the 
 rational manner of designing 
 
 To find the power and size of the compressor consider Fig. 
 399, in which the curve of compression ab is of the form 
 
AIR LIFT PUMPS AND PNEUMATIC PUMPS 
 
 pv n = K. It is desirable to have this curve a rectangular 
 hyperbola, but that is impossible, as even with water jackets 
 around the cylinder the exponent is rarely below 1.2 and 1.4 
 is usually found. The work done in compression is 
 
 Cpdv- 
 
 Jv, 
 
 now 
 Hence 
 
 pdv~K ('v-n 
 
 _P\V\ -piV'< 
 
 i-n 
 The work becomes: 
 
 W =p 2 V 2 -{- 
 
 nI 
 
 This is the theoretical power, in which pi is the suction pres- 
 sure and FI the volume taken in if there is no clearance or 
 leakage. There is no effect of clearance on the work of a com- 
 pressor as the work obtained on the expansion of the clearance 
 air is equal to the work required to compress it. There is an 
 effect due to leakage and this is to change the work by the 
 
 factor : 7^-. The volumetric efficiency is the ratio of the 
 vol. eff. 
 
 air delivered to the amount which should be delivered. This is 
 about 95 per cent. The friction of the compressor increases 
 the amount of work done, so that the net H.P. required to 
 apply to the V\ if this is the amount of free air required per 
 minute is: 
 
 33,000 X H.P. 
 
 mech. eff. vol. eff. n 
 
522 
 
 PUMPING MACHINERY 
 
 The clearance effects the displacement by cutting out the 
 amount of air cf, Fig. 400, and caring for the amount fa only 
 
 FIG. 400. Compressor with Clearance. 
 
 The factor which gives this if cl= per cent clearance is; 
 
 C/\/>! 
 
 The displacement then is 
 
 D 
 
 from the formula 
 
 vol;eff.X{/)' 
 
 FIG. 401. Multi-stage Compression. 
 
 for a double acting compressor the size of the cylinder and 
 the number of revolutions may be found to give the necessary 
 free air V a . Allowance should be made for piston rods. 
 
AIR LIFT PUMPS AND PNEUMATIC PUMPS 523 
 
 Since it is not possible to bring the curve of compression 
 close enough to the isothermal, a method of multi-staging 
 the compression has to be devised. In this, Fig. 401, the com- 
 pression is carried to a pressure p 2 ' when the air is discharged 
 into a series of tubes or around them, while on the other side 
 of the tube surface cold water is circulated. In this manner 
 the air can be left in contact a sufficient time to be brought 
 to the original temperature, and when taken into the second 
 cylinder at h f it is on an isothermal from a. In this way the 
 work of the area h W h' is saved. The work in this case is: 
 
 =p2'V 2 ', since h' is at the same temperature as a, hence 
 
 dW 
 This is a minimum when -j f = 0, 
 
 np p n p 2 ' 
 
 i\~r/:i* 
 
 pj \pip 
 
 is the condition for a minimum amount of work. Then 
 
524 PUMPING MACHINERY 
 
 For a three-stage compressor the condition for a minimum is 
 
 or in general for m stages 
 
 p2 f " = 
 
 etc. 
 
 The Harris Pneumatic pump is shown in Fig. 402. In 
 
 this case to start the pump, 
 air is sucked out of one cyl- 
 inder A until it is filled with 
 water, while the pump line 
 on the other side is filled 
 with water and air com- 
 pressed into the cylinder B 
 until the water is driven 
 from the chamber. At this 
 time the turn-over valve or 
 switch M is shifted so that 
 the compressed air on one 
 side may fill the pipe on the 
 other side. That is, if the 
 volumes of the tanks are 
 each represented by V and 
 the volume of the pipe lines 
 by NV where N is usually a 
 FIG. 402. Harris Pneumatic Pump. fraction, the air of volume 
 
 (i+N) V in the discharging 
 
 side at a pressure P is connected with a volume NV in which 
 the pressure is called p (when filled with water). The temper- 
 
AIR LIFT PUMPS AND PNEUMATIC PUMPS 525 
 
 ature is constant in these pipes since they are exposed to the 
 atmosphere and have much surface, hence when the valve M 
 is switched 
 
 
 1+2^ 
 
 After the pressure has fallen in the system to PI the com- 
 pressor draws air from the tank just discharged and builds 
 up the pressure in the pipe line leading to the full tank. After 
 the pressure in that tank becomes equal to the discharge head 
 P the water begins to discharge. The foot valve on the empty 
 tank does not open until sufficient air has been drawn out to 
 bring the pressure to pi. At that time water just begins to 
 enter and when the tank is full, this pressure has become p Q 
 These two pressures, pi and p- , do not differ by much since the 
 head difference is the height of the tank, and this may be very 
 small especially if the tank is on its side. 
 
 To find the quantity of air to be pumped from one tank 
 to another there are three periods to consider. 
 
 ist. That in which the pressure in the pipe line leading to the 
 full tank is brought from P x to P . 
 
 2d. That to reduce the pressure in the suction tank to pi 
 while water is discharging from the full tank. 
 
 3rd. That to fill the suction tank. 
 
 To find the pressure in the suction tank at the end of the 
 first period the following equation may be used: 
 
 P NV+PnV(i +N) =PiV(i +2N). 
 
 _P l (i+2N)-P N 
 
 (i+N) 
 
 Since N is small, this quantity PI i will be found to be almost 
 the same as PI, but for exact work the formula must be used as 
 stated and the displacement found. 
 
 Suppose the displacement of the compressor is D and this 
 volume is abstracted from the volume (i+N)V of the pipe 
 and tank and discharged into the volume NV. 
 
526 PUMPING MACHINERY 
 
 The ratio of - ~ - is M, and after the first stroke the 
 
 M 
 pressure PI is reduced to ^ PI, after the second to 
 
 - PI\ or ( 1 r^-\ PI, and after the t strokes the 
 Af + i \M + i V \M + i/ 
 
 ) Pj. 
 
 pressure is 
 
 This, however, equals PH. hence 
 
 or 
 
 logs Pi i- logs Pi 
 
 tD=Q, the quantity taken through the compressor during this 
 first stage. During the second period the compressor compresses 
 against a fixed discharge pressure but the suction is variable. 
 In this case 
 
 where dV v is the variation of volume in the delivery tank, dV c 
 that in the compressor, and Pj the pressure in the tank being 
 emptied of air. Now at any time 
 
 P*V(L +N) +P (V V + NV) =P V(i +N) +p NV. 
 Hence 
 
 PO __ V(i+N) _ 
 
 Px V(i+N)+-NV-V v -NV 
 YQ 
 
 caUing 
 
 Then 
 
 dVc= P~ K 
 
AIR LIFT PUMPS AND PNEUMATIC PUMPS 527 
 
 f* vl Pt\ C VI V(i-\-N^ 
 
 Q2 = Volume displaced = / -dV v = I . v . . x dV v 
 
 r^-dv = r r^+w 
 .'- '- 
 
 Fi is the volume of the discharge tank when the water is 
 just entering the other tank, or the pressure is pi in that tank, 
 hence 
 
 P V l = V(P G 
 = V[P 
 
 Now N is a fraction and p piis very small, hence 
 
 F!=^-(P O -/>I).- 
 ^o 
 
 Hence 
 
 _ 
 
 Po 
 
 i 
 
 logs 
 
 Neglecting AT^o in comparison with P 
 
 Now, during the third period, or that of the filling of the 
 suction chamber, the pressure in this chamber changes from 
 
528 PUMPING MACHINERY 
 
 pi to pQ when it is being filled with water, and since this change 
 is so small the pressure may be assumed constant and the quan- 
 tity removed during this time will be V. 
 
 The total displacement of the pump will then be 
 
 In this, V is the volume of the water lifted by the displace- 
 ment of Q cubic feet of air by the compressor piston. 
 
 If the displacement of the compressor per revolution is 
 D and it makes 2V revolutions per minute the time taken to 
 fill one tank is: 
 
 ~^ =/' minutes. 
 
 The effect of friction of the pipes, according to Harris, is to 
 increase the pressure required by the amount. 
 
 Loss = KP = - - -7 v 2 P in pounds per sq.ft. 
 
 /= length of pipe in feet; 
 d = diameter in inches; 
 v= velocity in ft. per sec.; 
 P= pressure in pounds per sq.ft. 
 
 The effect of this is to multiply variable pressures of discharge 
 by i+K, or what is the same thing, to multiply the volume 
 which passes through the compressor by i +K, and therefore 
 the time is increased in the same manner. 
 
 The quantity K is seen to vary as the velocity of the air in 
 the pipe varies, but for general use an average value of o.i is 
 recommended by Harris. The quantity of air passing per 
 second as the pressure changes is a variable, but the velocity 
 of this air from the suction tank remains about the same because 
 its pressure changes; while that entering the discharge tank 
 varies with the weight, since the pressure is constant. 
 
 The size of the compressor can be determined when the 
 
AIR LIFT PUMPS AND PNEUMATIC PUMPS 529 
 
 quantity of water to be lifted per hour is known, or the equation 
 for Q may be used to get the quantity of water which may be 
 handled by a given compressor. 
 
 The work done in the compressor when V x cubic feet are 
 compressed has been shown to be 
 
 n-i 
 
 where n is the exponent of the compression curve. This quantity 
 W changes in value as the air is drawn out. To find when it is a 
 maximum express it in terms of the variable p x , as V x is con- 
 stant, depending on the displacement of the compressor while 
 P is a constant, 
 
 nV 
 work= - 
 
 This is a maximum when 
 
 d 
 dp 
 
 dW nV fi tit*- TJE "1 
 
 == _ _ n p n _j 
 
 n-iln r 
 
 For ^ = 1.4; p x = 
 Putting this in for p x the work for V cubic feet passing 
 through the compressor becomes: 
 
 If the displacement of the compressor is ND where D is 
 the discharge per revolution, the horse-power which must be 
 cared for is: 
 
 Max ,000 
 
 This is divided by the efficiency of the compressor and 
 driving motor to get the applied power at this instant. 
 
 UP AV-'^VV> 
 
 x. H.P. = 1 -I 
 
 \nl 33,ooo 
 
530 PUMPING MACHINERY 
 
 To get the total power employed, the following method 
 will be used, for the second stage: 
 
 Now although the compression within the cylinder is accord- 
 ing to the law pV n =K the relation between the pressures and 
 volumes in the tanks and pipe lines is isothermal, so that at 
 any time 
 
 px(T.+N)V+Po(V x +NV)=p 9 '(L+N)V+P Q (V 2 +NV), 
 
 or considering a small amount, dV taken from the volume 
 V(i +N) the change becomes 
 
 = V(i+N)dP x , 
 
 then 
 hence 
 
 n^^P 9 ^ fp^P,-^^n C dPl 
 ni J n-i J 
 
 The work done during the first period is 
 
 and the work on the last period is 
 
 nV F T-i - 1 
 W 3 = -[P a " P 1 '-p 1 \. 
 
AIR LIFT PUMPS AND PNEUMATIC PUMPS 
 
 The useful work is 
 
 V(P (} -pi) 
 
 Hence the efficiency of this method of 
 pumping is 
 
 Eff.= 
 
 X 
 
 eff. compressor and motor. 
 
 Harris has designed for this 
 method of pumping a special switch 
 which will reverse the turnover valve 
 after the compressor has made the 
 necessary number of turns t'N, or by 
 a system of diaphragms the valve is 
 changed when the pressure p Q is 
 reached. Leakage from the system 
 is made up from the atmosphere and 
 allows a supply of air when the pres- 
 sure in the suction is below the 
 atmosphere before the valve re- 
 verses. If an excess of air is drawn 
 in this will do no harm, as it will be 
 blown out of the discharge before the 
 valve reverses the action. 
 
 It is recommended by the build- 
 ers of this pump to make the air pipe 
 of such a diameter that the maxi- 
 mum velocity is not over 5000 feet 
 per minute when it is assumed to 
 have the volume of the free air. 
 That is 
 
 Area air pipe = 
 
 /XI4-7XI44X50QO 
 
 The tanks are usually made of 
 such a volume that N=, or tanks 
 are four times the vclumes of the 
 
 i 
 
 FIG. 403. Harris Pneumatic 
 Pump. 
 
532 
 
 PUMPING MACHINERY 
 
 pipes. The discharge pipes for the water are made 2 . 8 diame- 
 ter of air pipes and they should discharge into a head in the 
 
 FIG. 404. Separate Stage Pump. 
 
 same manner as that shown for the air lift pumps. Fig. 403 
 shows the installation of a Harris pneumatic pump for draining 
 the inverted syphon at High Bridge. The tanks AB are con- 
 
AIR LIFT PUMPS AND PNEUMATIC PUMPS 
 
 533 
 
 nected to the compressor by the pipes CD which extend to the 
 power-house. 
 
 Fig. 404 illustrates a method of using the system for high 
 lifts where it is desired to use an air pressure less than that cor- 
 responding to the total lift. The air forced into A lifts the water 
 against the pressure P into the 
 suction box B while at this time 
 the tank C is being filled. When 
 the pump reverses the water in C 
 is discharged while A is filled by 
 suction. 
 
 The method shown in Fig. 405 is 
 intended to use compressed air under 
 a moderate pressure to raise water 
 a considerable height. Air is delivered 
 into B through A at sufficient pres- 
 sure to force water through the foot 
 valve C into the tank D. When the 
 air which rises to the top of B has 
 driven the water from B so that the 
 bottom of the pipe E is exposed, the 
 air supply is cut off and the air in 
 B escapes and passes up to the next 
 tank and lifts the water from it 
 through the foot valve in F. The air 
 which filled tank B will fill tanks B 
 and D finally and hence the pressure 
 will be half the pressure of the supply 
 if the two tanks are of equal size. 
 This means the second lift can 
 
 FIG. 405.- 
 
 Differential Air Lift 
 Pump. 
 
 only be made half of the lift between the first and second 
 tanks. If a third tank is used the next lift will be one third 
 the original lift. 
 
 In this method the air is finally discharged and the 
 only use of the internal energy of the air is in the 
 successive lifts. 
 
 After the air has been driven out from the upper tank the 
 
534 PUMPING MACHINERY 
 
 lower tank is allowed to fill as it is open to the atmosphere 
 through E and D and water will flow in by gravity. 
 
 There are other compressed air pumps which act the same 
 as these, but the principles here brought out should serve to 
 make their action clear. 
 
CHAPTER XIV 
 CENTRIFUGAL PUMPS 
 
 THE last few years have seen many improvements in cen- 
 trifugal pumps; much higher efficiencies have been obtained 
 and greater lifts have been overcome. These gains have resulted 
 
 FIG. 406. Section of Centrifugal Pump. 
 
 from a careful study of the principles of the pump and from 
 the theoretical design of pumps rather than from the empirical 
 design which was customary in earlier times. To understand 
 the theory of these pumps an examination will be made of 
 Figs. 406 and 407. Fig. 406 is a longitudinal section through 
 
 535 
 
536 PUMPING MACHINERY 
 
 a single-stage pump, while Fig. 407 is a cross-section through 
 the impeller or runner A and the diffuse* vanes B. With water 
 in the discharge tank D at the level shown in the figure, the 
 whole system will be filled with water and this water will tend 
 to flow from the tank through the pipe E, the volute casing C, 
 
 FIG. 407. Section through Impeller, Diffuser and Casing. 
 
 the diffuser B, the runner A and the suction pipe F. The 
 runner A is mounted on the shaft H which is supported by the 
 foot bearing I. If the shaft is horizontal two bearings take the 
 load resulting from the weight of the parts. 
 
 If now the shaft H be driven by a belt on the pulley G or 
 by a motor directly connected with the shaft, the water is 
 
CENTRIFUGAL PUMPS 
 
 537 
 
 driven by the runner and forced to move outward by the shape 
 of the vanes and by centrifugal force; moreover this motion 
 takes place in such a manner that the water leaving the outer 
 tips of the runner enters the vanes in the diffuser with no shock. 
 In the diffuser the direction of motion is changed to one more 
 nearly the direction to be obtained in the volute chamber and 
 also the velocity is decreased so that velocity head is changed 
 into pressure head. If the speed of the runner is sufficiently 
 great the water will acquire enough velocity that when the 
 ends of the vanes B are reached the pressure will be greater 
 than that due to the head of the tank D. 
 
 As the water is driven outward on the wheel a partial 
 
 FIG. 408. Single and Double-flow Impellers. 
 
 vacuum is produced at the center and water is drawn up to 
 fill the space. At times this water is drawn in on two sides of 
 a disc giving a double-suction pump. The shaft is often placed 
 in a horizontal position, but for simplicity in the derivations 
 of equations the shaft has been shown here in a vertical position. 
 In some cases the vanes of the diffuser are omitted and in this 
 event the space around the wheel is sometimes called the 
 whirlpool chamber. There may be guide vanes placed in the 
 suction at the point of entrance to the runner for the purpose 
 of making the direction of the water entering the runner more 
 definite. Fig. 408 illustrates the two forms of runner. 
 
 A study will now be made of the pressures in the pump 
 

 538 PUMPING MACHINERY 
 
 and of the action of the pump in which the following symbols 
 will be used: 
 
 r\ = radius of point of entrance = ; 
 
 r% = radius of point of exit = ; 
 
 ;= coefficient of lost head (see Chapter V); l ^ 
 
 c= absolute velocity of water at entrance; 
 
 u= absolute velocity of water at exit from impeller or 
 
 entrance to diffuser; 
 
 c\ = velocity at entrance relative to impeller; 
 c 2 = velocity at exit relative to impeller; 
 Vg = velocity in suction pipe; 
 Vd= velocity in discharge pipe; 
 v v = velocity in volute; 
 u 3 = velocity at exit from diffuser; 
 v = velocity in outlet or discharge pipe; 
 A =area perpendicular to c; 
 A with subscript =area perpendicular to velocity with same 
 
 subscript; 
 
 Wi = velocity of wheel or impeller at entrance; 
 w 2 = velocity of impeller at exit; 
 <*j = angular velocity; 
 AT = R.P.M.; 
 
 hi = suction head to center of wheel; 
 h 2 = discharge head from center of wheel; 
 H=hi+h 2 = total net head; 
 
 Pa, pb, PC, pd, etc., pressure at various points in feet head; 
 w= weight of i cu. ft. of water; 
 a = pressure of the atmosphere in feet; 
 g= acceleration of gravity; 
 All the above are expressed in feet and pounds. 
 At the point of entrance into runner, the pressure in the 
 casing p a is given by the equation, 
 
 Vs 2 C 2 
 
CENTRIFUGAL PUMPS 
 
 539 
 
 When the water enters the runner, the portion of p a which 
 remains to overcome the friction of the passage of the runner 
 and to accelerate the water in so far as the section of this pas- 
 sage may change is called />&. This is given by: 
 
 - 
 
 (2) 
 
 is the pressure head due to the action of centrifugal 
 
 o 
 
 force.* This action, which exists in all rotating bodies con- 
 taining fluids, is to produce a hydrostatic pressure equal 
 to the above expression. This, like any other hydrostatic 
 pressure, acts in all directions. Where there is a free surface 
 as in a cup, this manifests itself by causing the water to pile 
 up where the velocity is the greatest in such a manner that the 
 free surface is a paraboloid of revolution. Where there is no 
 free surface, there is a force equal to the height to which the 
 water would rise above the point considered were a free surface 
 
 possible. This force is equal to . The actual force on a 
 rotating body which would cause flow and overcome friction 
 
 * In Fig. 409, the particle of water in 
 the free surface of the rotating vessel at a 
 distance x from the center is moving with a 
 velocity xa), if o> is the angular velocity. It 
 
 w 
 is subject to the centrifugal force a) 2 x ex- 
 
 g 
 
 pressed in pounds, while the force w acts 
 vertically, due to gravity. The resultant of 
 these two forces acts normal to the free sur- 
 face, hence 
 
 dy 
 dx 
 
 C.F 
 
 IV 
 
 _ Cx^'X 
 
 Jo g 
 
 dx= = 
 
 FIG. 409. Revolving Vessel. 
 
540 PUMPINC MACHINERY 
 
 w x 2 w\ 2 
 
 is not the external head acting at a point, but that force minus 
 \ 2 
 g 
 
 , Ci 2 ^vCl 2 C 2 2 , W 2 2 Wl 2 . 
 
 Ate, *-*+-- aF - 2g V7,' ^) 
 Atd, pa = pc~. . . . . ./.. . . (4) 
 
 (5) 
 
 If Ws = Vv there is no loss at discharge into the volute, but 
 if this is not the case, then there is a sudden change in velocity 
 and there is a loss 
 
 3 2 
 
 (6) 
 
 V v * 
 
 - (7) 
 
 but' ' p,+f g =h 2 + a + g +- g . ., i ,. i ;, ; . (8) 
 
 Substituting for p v , p e , pd, p c , pb, etc., their values from the 
 preceding equations the following results after collecting and 
 rearranging: 
 
 V, 2 VQ 2 U 2 W 2 2 C^ C 2 2 Wj C 2 
 ~r ** i ^ -- 1 -- T - - - ( Q I 
 
 *2g 2g 2g 2g 2g 2g 2g 2g' 
 
 The passages of the pump are completely filled with water, 
 hence 
 
 Ac=AiCi=A 2 c 2 =A e xU=A 3 u 3 , etc. . . . (10) 
 
 From these equations it is seen that all of the velocities may 
 
CENTRIFUGAL PUMPS 
 
 541 
 
 be expressed in terms of the areas and any other velocity. For 
 example: 
 
 2-jy2 
 
 , 
 
 The Eq. (9) now becomes 
 
 
 S 
 
 If there is to be no impact at entrance and exit, the velocities 
 
 FIG. 410. Velocities at Entrance a positive, a.^ negative. 
 
 FIG. 411. Velocities at Discharge 2 negative, /? positive. 
 
 at those points must have certain relations which are seen from 
 the parallelogram of velocities. Calling the angle between 
 the radius and c, a; between the radius and c\ t a\\ between 
 the radius and c 2 , a 2 ] and finally that between u and the radius, 
 /?; the following relations may be seen from Figs. 407, 410 and 
 411. (The angles are taken as positive when they are measured on 
 the side of the radius toward which the motion w^ is taking 
 
542 PUMPING MACHINERY 
 
 place, provided the lines be drawn away from the point con- 
 sidered and in direction of the velocity. ) 
 
 Ci cos a\ =c cos a (14) 
 
 c sin a Ci sin a\ =Wi; . .. . . . (15) 
 
 Ci 2 =c 2 +Wi 2 2cwi sin a or c 2 +Wi 2 c 1 2 =2cz# 1 sin a; (16) 
 
 C 2 cos a 2 =u cos /?; .' ; . '. '. . .- . (17) 
 
 u sin /? c 2 sin a 2 =w 2 ; ... , (18) 
 
 c 2 2 =u 2 +w 2 2 2uw 2 sin /?or u 2 +w 2 2 c 2 2 =2uw 2 sin /?. (19) 
 
 Substituting (16) and (18) in Eq. (13) there results: 
 
 u 2 v 2 i _ 
 
 h + =-[uw 2 sm 3 cwism a]. . . (20) 
 
 2g 2g g 
 
 n 
 
 Calling St; the lost head, IH, and the residual velocity 
 
 2g 2g 
 
 head, r#, the left-hand side of the equation represents the total 
 head to be accounted for. This when multiplied by one pound 
 is the number of foot-pounds of work done on each pound of 
 water passing from the suction forebay to the discharge tank. 
 
 #(i+/+r)=work per pound of water 
 = (uw 2 sin p cwi sin a). 
 
 o 
 
 To get the total work this expression must be multiplied by 
 the quantity of water passing through the pump. All of the 
 water taken into the pump at a does not leave the discharge 
 point of the diffuser, for a certain amount leaks past the clear- 
 ance space between the wheel and the diffuser or casing. The 
 quantity leaking from this point will be computed, but for the 
 
CENTRIFUGAL PUMPS 543 
 
 present it will be called Qe. The amount desired is Q cubic 
 feet per second, and so the amount Qp, to be pumped is 
 
 Qp=Q+Qe-, ....... (21) 
 
 QpwH(i+l+r)= Work per sec.; .... (22) 
 
 Qpw- (uw 2 sin /? - cwi sin a ) =Work per sec. (23 ) 
 
 6 
 
 The expressions above are those which give the work required 
 to do the hydraulic work, but in addition to this there must be 
 energy supplied to overcome the resistance of the friction of the 
 bearings, stuffing boxes, and the leakage water against the 
 runner. This latter quantity may be quite large, while the 
 second one depends on many variables. The first may be 
 made very small by proper lubrication. These will be dis- 
 cussed later, but for the present they may be written as f w W u , 
 f b W u and f 8 W u . The total work then becomes 
 
 wO 
 =p(uw<> sin ff ci&i sin a)(i +f. 
 
 ; (24) 
 
 wQpH(i+l + r)(i+f w +f b +f 8 ) 
 w(Q+Q e )H(i+l+r)(i+f w +f b +f 8 ) 
 
 The useful work is wQH or in some cases, wQH(i+r), if 
 the residual velocity may be utilized in any manner. The 
 efficiency then becomes 
 
 wQH i 
 
 If ~ TI7 /T^L/7\/T_L.;j^\/-rl^ 1 4 \ -t \' ' (25) 
 
 The efficiency is seen to vary with the values of these coef- 
 ficients; they should all be as small as possible. The deter- 
 mination of these quantities will be considered after a discus- 
 sion of the action of the water in the wheel. 
 
 There are several relations which may be determined from 
 
544 
 
 PUMPING MACHINERY 
 
 Figs. 407, 410 and 411. Fig. 412 gives the two diagrams of 
 Figs. 409 and 410 in one. 
 
 a is often made o so that the work equation becomes 
 
 now 
 
 (26) 
 
 tl. 
 
 =d sin 
 
 A, ' 
 
 c=c 
 
 FIG. 412. Diagrams at Entrance and Exit. 
 Hence by combining these 
 
 eK . 
 
 kH= -- sin/? sin 
 r\A l g 
 
 or 
 
 u 
 
 2g(kH) 
 
 * A 
 
 T 2 ** ex 
 
 2 -r- sin /?sm 
 
 r l A l 
 
 (27) 
 (28) 
 
CENTRIFUGAL PUMPS 545 
 
 If it is desired to express this in terms of the effective head 
 j Eq. (20) may be written: 
 
 2' 
 
 ' (30) 
 
 If either of these equations be used, the losses must be 
 expressed in terms of H or u before u can be found. The ratios 
 
 ^ and may be assumed from practice as well as the ratio 
 A r\ 
 
 For convenience it is well to take the components of u and 
 c in a radial direction; these will be called u r and c r . 
 
 w r = wcos/?=c 2 cos a: 2 , ..... (31) 
 
 c r =c cos a=Ci cos a\ ..... (32) 
 
 Neumann uses the total head kH in finding the velocities 
 w, u r and c 2 in the following manner: 
 
 -uw 2 sm{3=kH, . . . ^ -v * (26) 
 
 o 
 
 From Fig. 412, Eq. (18) and Eq. (31), 
 
 ugkH 
 w 2 =u sin /? c 2 sin a2 = -- w r tan 
 
 =- -- w r tan 2 , . . . . ..... (33) 
 
 u r tan 
 
 . (34) 
 
546 PUMPING MACHINERY 
 
 For a 2 =o this becomes 
 
 Since tf r =wcos/?=wsin/?cot/? = , . . (34^) 
 
 W 2 
 
 - (33) may be written from (340): 
 
 2COt/? ' - (35) 
 
 , . . (36) 
 
 tan 
 calling i. . .,, ,_^- . ,,. (37) 
 
 Eq. (33) may be used to determine the angle 2 in case ze> 2 , 
 and u r are known, and Eq. (35) may be used if w 2 and /? are 
 known. Thus 
 
 gkH 
 
 ; (38) 
 
 U r 
 
 gkH w 
 
 ~ W2 *-&B 
 
 = ~- - (39) 
 
 For the relative velocity c 2 , Neumann uses the equations, 
 
 U r 
 
 C 2 = 
 
 COS a 2 
 
 _wcos/?_ u sin ft _ gkH 
 2 ~~ cos 2 . tan ft cos a 2 ~w 2 tan ft cos 2 * * '^ ' 
 For w: 
 
 --^ ( 
 
 z^ 2 sin 8 
 
 gkH I <>kH 
 
 tan 
 
 sin " 
 
CENTRIFUGAL PUMPS 547 
 
 These equations are used when certain quantities are known 
 or assumed. The assumptions fix the equations which would 
 be used. These quantities are also obtained in another manner 
 by Neumann. He uses the expression 
 
 w 2 =x\ // gkH 
 
 and for u r he proceeds as follows: 
 
 cot 8 
 
 (42) 
 
 when a 2 =o, w 2 = A vgkH. . . .... -. . (43) 
 
 Hence u ro =\ / gkH cot p Q =c 2 cos o=c 2 , . . . (44) 
 
 calling cot ft = \ // 2 ^, 
 
 
 
 , ........ . (45) 
 
 Using the value of w 2 = 
 
 ^77 cot ft VgkH cot 
 
 / -r== , ... (46) 
 
 xVgkH 
 
 and if in general cot ft = \zX or X =% c t 2 /?, the result follows: 
 
 VtegkH 
 "r = f , (47) 
 
 from Eq. (41), 
 
 (48) 
 
 
 The author has re-computed and constructs the curves 
 shown in Figs. 413 and 414, which are those used by Neumann. 
 By assuming various angles for 2 and ft the values of u, w 2 , 
 and u r may be found in terms of kH. 
 
 A series of curves drawn by the use of the equation, 
 
 shows how w 2 varies with different values of ft and a 2 , but 
 
548 
 
 PUMPING MACHINERY 
 
 90 
 
 FIG. 413. Values of x. (After Neumann.) 
 
 60 
 
 
 .06 0.10 0.15 0.20 0.25 0.30 0.35 0.40 0.15 
 
 FIG. 414. Values of X. (After Neumann.) 
 
CENTRIFUGAL PUMPS 
 
 549 
 
 since w 2 =x when VgkH = i, these curves need not be drawn, 
 as the x curve shows this variation to some scale. 
 
 It is seen from an examination of Fig. 413 that when 
 /?=go, x = i for all values of 2 except 90 and that for 0:2=0, 
 
 
 FIG. 415. Combined Entrance and Discharge Diagram. 
 
 all values of x for different values of ft are unity. The varia- 
 tion in x and consequently w 2 is greater for changes in 2 for 
 smaller values of ft than for values of ft near 90. When ft 
 has a value of 85 there is little change, as 2 changes from 50 
 to -50. When considerable variation is desired for different 
 values of a 2 > ft must be taken smaller, say 60. 
 
550 PUMPING MACHINERY 
 
 When a is not o the equations just found will be of differ- 
 ent forms. The entrance velocities will now be considered. 
 
 w\ 
 A new figure, using the construction for w 2 = r 2 , is given 
 
 in Fig. 415, showing these other relations. In this figure, 
 
 A 2 
 
 Cl =c 2 -j-. Now c r =c cos <x=Ci cos i, 
 -*1 
 
 c sin a =c r tan a. . : \ . . '. . (49) 
 2 =w 2 +u r tan a 2 . . . (50) 
 
 Now (uw 2 sin 8cwi sin a) =kH, 
 
 o 
 
 or w 2 2 +w 2 u r tan a 2 Wic r tan a=gkH. . . . (51) 
 
 Wi=w 2 (52) 
 
 Hence 
 
 r tan a w r tana 2 l 2 
 
 Now U r 2nr 2 b-2=c r 2xr 1 bi, hence 
 
 tan * ~~~ tan "-+g^. (53) 
 
 In one of these, c r must be assumed as well as - 7^, a 2 . 
 
 r 2 b 2 
 
 and a, while in the other c r , u r , , a 2 , and a are assumed. 
 
 If the velocity u is desired in terms of the effective head H, 
 the following results; 
 
 
CENTRIFUGAL PUMPS 551 
 
 n/- C w 1 sn a _-; 
 
 ze? 2 =wsin/? c 2 sin 0:2; 
 c 2 cos 2 =w cos /?; 
 
 w 2 =u sin/? w cos /? tan 2 =w(sin /9 cos /? tan 2 ); . (54) 
 ze>i =c sin a c\ sin i =c (sin a cos a tan a) ; 
 
 2gH = aw 2 sin ^5 2 cos ft tan a 2 sin ft 
 
 [sin a -cosa tan aj sin a - S ; - - 
 
 A/ 2[sin/? cos/3tana 2 ] sin/9 ( |^) [sin a cos a tanajsin a 2 ( J 
 
 (55) 
 
 and substituting the value of w above, w 2 is found. Another 
 method gives the following: 
 
 sn WC sn a 
 
 u sin 5 2 -- + c r tan a; 
 w 2 r 2 
 
 __ > 
 ~cos/? ; 
 
 T 
 
 tan 
 
 Now z# 2 = u sin /? w r tan 2 , 
 
 gkH n 
 W 2 =- -- \ c r tan a u r tan 
 
 (56) 
 
552 PUMPING MACHINERY 
 
 Substituting the value of u r from Eq. (56), 
 
 gkH\ tan 0:2! r\ tana 2 l 
 
 z0 2 = - i T f- +-c r tana I - -~\. 57 
 w 2 L tan /? J r 2 L tan /? J 
 
 PI tan a f tan a 2 l 
 
 These equations may be used in determining the velocities 
 of various parts of the wheel for a given set of conditions. 
 
 LOSSES 
 
 The first loss to be considered in Eq. (9) is the loss in the 
 suction pipe ^,. This coefficient includes the losses due to 
 entrance, friction in pipe and bends. These losses have been 
 discussed in Chapter V, and from what has been given there 
 the following may be written: 
 
 The equation shows that ^ s varies directly with / and 
 inversely with d, and moreover the total loss varies with v 8 2 . 
 
 It is an important matter to determine just how much ^ 
 may be. In any case 
 
 p^a-h.-^-^pt, .:. . . (60) 
 
 where p t in feet is the steam pressure corresponding to the tem- 
 perature of the water in the suction pipe. If this inequality is not 
 true, the suction column will part. Of course, the term hi is 
 the one which is usually responsible for the inequality not 
 holding, and a change in it may make this true. It may be, 
 
CENTRIFUGAL PUMPS 553 
 
 though, that v s 2 is so great and that -7 is so great that the 
 
 v 2 
 term , is responsible for this inequality not holding. In 
 
 such a case the diameter of the suction must be increased, 
 
 I 
 
 giving a smaller value of -7 and v s 2 . The bends in the pipe 
 
 must be gradual to keep the value of m as small as possible. 
 Although the velocity in the suction pipe should be such 
 that there is no danger of the column of water parting, the loss 
 should be reduced to a low value. It is difficult to tell just 
 what velocity to allow to give the loss in the suction pipe a 
 value consistent with the other losses. A method used by some 
 is to make the velocity in the suction pipe equal to the velocity 
 caused by a certain percentage of the total head. Neumann 
 uses 2 per cent of the head kH. If the suction pipe is short 
 and direct, this value will be used while with a long pipe (say 
 50 diameters) a i per cent value will be employed. This gives 
 
 v 8 = 8. 02*/- (short pipe), . . . . . (61) 
 
 ' +J 
 
 8 02 
 
 =-^-VkH (long pipe) (62) 
 
 In any case a discussion similar to that given under the discharge 
 pipe could be used at this point. 
 
 c 2 
 The loss ^ent is due to the incorrect angular relation 
 
 between the fixed vanes (if used) and the movable ones at 
 entrance and also to the interference between the moving or 
 fixed vanes with the flow of water, causing impact. The same 
 
 u 2 
 kind of losses occur at exit in the term i; ex . For this reason 
 
 these two will be considered at the same time. 
 
 In Fig. 416 a series of movable vanes is placed opposite 
 a set of fixed vanes. If the velocities are as shown in the 
 diagram, the water as it leaves the moving vanes is traveling 
 
554 
 
 PUMPING MACHINERY 
 
 through space as shown in the figure. When this water strikes 
 the blunt vane, there is a certain amount of loss due to sudden 
 contraction, and moreover when the water leaves the blunt end 
 of the moving vane there is a loss due to sudden enlargement. 
 
 Diffuser 
 
 FIG. 416. Interference of Blunt Vanes. 
 
 To cut down these losses the edges of the vanes are sharpened 
 as shown in Fig. 417, and there should be a certain amount of 
 clearance between the two sets of vanes. This clearance permits 
 the 'water leaving one set of vanes to come together as one 
 
 FIG. 417. Pointed Vanes. 
 
 mass of water before entering the other set of vanes. The 
 action of one set of vanes on the other, if the vanes are so close 
 that there is a space between the masses of water discharging 
 from each channel, should be such that much noise and shock 
 would result. 
 
CENTRIFUGAL PUMPS 
 
 555 
 
 By cutting the edges of the vanes and by separating the 
 various sets, these coefficients ^ en t and ^ ex are made small. 
 
 Now these losses are those which occur when the relations 
 between the various velocities are such that there is no impact. 
 In many pumps, however, there is a change in the pressure or 
 quantity discharged while the pump is driven at a constant 
 speed. This means a change in the velocities c 2 and u on the 
 outlet side, since these are given by the equations: 
 
 If Q is increased c 2 and u will be increased, while a decrease 
 
 C 2 (Decreased quantity) 
 
 FIG. 418. Effect of Changing Quantit}*- with Fixed Speed. 
 
 in Q will change c 2 and u in the opposite manner. These 
 changes are shown in Fig. 418. It is seen that the resultant 
 of c 2 and w 2 is not in the direction of u nor of the value of u. 
 For this reason there is a loss due to a sudden change in velocity 
 of 
 
 1/^2 
 
 =loss, 
 
 and a loss due to the impact against the side of the vane which 
 may not be very large, as has been proven by experiments on 
 the loss due to the impact of a jet against a flat plate at right 
 angles to it. A similar loss to this occurs at entrance when 
 the quantity changes from the normal amount. 
 
556 PUMPING MACHINERY 
 
 The losses in the vanes, C~~i and that in the diffuser,^ , 
 
 are similar to loss in pipes and bends. They depend on the 
 
 
 length of the passage /, the hydraulic radius * (A =area of 
 
 passage, />=the perimeter of passage), the smoothness of the 
 passage, and its radius of curvature. These losses are of the 
 form 
 
 S.-fcfc'j. 
 P 
 
 k, being the coefficient for straight pipes, = : =0.005, an d k 
 
 4 
 being the coefficient for curvature. 
 
 The loss - -- on entrance to the volute casing around 
 
 2 # 
 the diffuser may be eliminated by making v =u 3 . In any 
 
 case it is a small quantity. 
 
 o 
 
 The loss ^ is the loss which occurs in the discharge pipe 
 
 and is, therefore, similar to the loss in the discharge pipe of 
 any pump, 
 
 In the design of an installation a number of pipes should 
 be investigated for loss and for cost. That one should be taken 
 which gives a result such that the use of a larger pipe would 
 increase the yearly cost of interest, depreciation, insurance 
 and taxes on the investment, more than the saving in the 
 decreased cost of power by the reduction in the lost head, 
 while the use of a smaller pipe would, on account of the greater 
 loss in head, increase the cost of power more than the saving 
 in the yearly cost for interest, depreciation, insurance and taxes. 
 A velocity may be assumed in terms of the total head kH as 
 in the case of the suction pipe, using the same values; however, 
 the method given above is the better one. 
 
CENTRIFUGAL PUMPS 557 
 
 When time will not permit of this investigation, the same 
 values may be used as for the suction, giving 
 
 Q 
 
 ^0=8.02--, 
 
 or, 
 
 In this work for design / and V may each be taken as 
 : -, and these terms will include the losses at entrance to 
 
 nmner and to diffuser. 
 
 LEAKAGE 
 
 The leakage of water through the clearance space is due to 
 the difference in pressure within the runner and on the outside. 
 Fig. 419 shows several forms of impellers. The pressure on the 
 inside of the runner is p c (Fig. 406), while on the outside, the 
 pressure is p a . If the clearance is t feet and the circumference 
 is 27iT2, the area through which leakage may occur is 
 
 2 X 2xr 2 t. 
 The velocity is 
 
 Vi=c v 8.o2\ / p c pa, ...... (63) 
 
 c v = coefficient of velocity. 
 FromEqs. (i)to(8)," 
 
 u 2 u 2 vi 2 v 2 v 2 
 
 W c 2 
 
 ----; p e =pt, 
 
 U 2 Vn 2 C 2 U 2 
 
 + ^ + ^ + ^+^}~+---. . (64) 
 
 c 2 u 2 
 
 Neumann proposes to assume = ( Cent + V) . 
 
 o o 
 
558 
 
 PUMPING MACHINERY 
 
CENTRIFUGAL PUMPS 
 
 559 
 
 Then 
 Hence 
 
 Qe=c v 47tr 2 t 8.02* kH 
 
 (66) 
 
 From this it is seen that the leakage is a function of the total 
 head, but it decreases with an increase in the absolute velocity 
 
 FIG. 420. Non-aligning Ring Oiling Bearing. 
 
 of discharge from the pump. If u is small, the velocity of dis- 
 charge through the clearance space is practically that due to the 
 total head acting on the pump. For this reason it is necessary to 
 make the clearance between the edge of the runner and the casing 
 as small as possible. At times the leakage is diminished by 
 placing rings on the casing and runner as at 6, Fig. 419. This 
 increases the pressure in the space a to p a ' and so decreases 
 the flow. Such an arrangement will increase the end thrust if 
 the runner is not a double one. 
 
 Friction at Bearings and Stuffing Boxes. The friction at 
 a bearing may be made very small by the use of ring oiling 
 bearings and by the use of spherical pivoted bearing boxes. 
 Such boxes are shown in Figs. 420 and 421. With the proper 
 
560 
 
 PUMPING MACHINERY 
 
 lubrication the coefficient of frictioji will be between o.oi and 
 0.004, so tnat tne work of friction will be 
 
 ..... (67) 
 
 W= weight of impeller or shaft for a motor-driven pump; 
 = resultant of weight and belt pull for a belt-driven pump; 
 r= radius of shaft in feet; 
 F = foot-pounds of work per minute due to bearing friction. 
 
 The coefficient /JL for lubricated journals varies with the 
 pressure and with the velocity. This must be taken into account 
 
 FIG. 421. Self-aligning Ring Oiling Bearings. 
 
 in the determination of the friction. Fig. 422 shows the vari- 
 ation of JJL with p a , the pressure per square inch of projected 
 area and with velocity. These curves are for a good grade of 
 spindle oil. 
 
 In fixing the length of the bearing, the designer would 
 assume the length from practice, as the allowance of 50 to 200 
 Ibs. per square inch of projected area would not give a bearing 
 long enough for a motor-driven pump. In the belt-driven 
 pump the length is determined by the formula below, using 50 
 Ibs. as the allowable bearing pressure, 
 
 (68) 
 
CENTRIFUGAL PUMPS 
 
 561 
 
 This friction may be increased if the bearings bind in any 
 way, and for that reason self-aligning bearings should be used. 
 The plain ring oiling bearing shown in Fig. 420 is often used. 
 With the bearings in good condition the friction at these points 
 will be a very small part of the total power, especially with 
 motor-driven pumps. The end thrust due to the pressure action 
 of the entering water in a single-flow pump is carried by a 
 thrust collar bearing as, shown in Fig. 423. The pressure is 
 
 
 0.04 
 
 0.03 
 
 0.02 
 
 0.01 
 
 Values of Ibs. per Sq. in. or Ft. per Sec. 
 
 50 100 Ibs. 150 
 10'ft. 
 
 200 
 
 )' 
 
 250 
 
 300 - '6M 
 30 ' 
 
 400 
 40' 
 
 450 
 
 FIG. 422. Curves of Variation of Coefficient of Friction. 
 
 represented by P and the necessary area is found by allowing 
 a bearing pressure of 60 Ibs. per square inch, 
 
 (69) 
 
 The friction from this source is found by using a coefficient 
 of friction /* which has the value 0.035 for properly lubricated 
 thrust bearings: 
 
 F. = 
 
 N 
 
 (70) 
 
 F 7/ =work of friction per minute in foot-pounds; 
 di = outer diameter of collars in inches; 
 d= diameter of shaft in inches. 
 
562 PUMPING MACHINERY 
 
 The value of /* for thrust bearings does not vary much with 
 the speed or pressure. The value of p given above is a mean 
 value determined by Tower in his noted experiments. 
 
 The friction from stuffing boxes depends on the amount, 
 kind and arrangement of the packing. From experiments 
 performed by Professor Benjamin, the formula below may be used 
 as a guide for the friction to be expected from each stuffing 
 box when the nuts are tightened by the application of 16 pounds 
 at the end of a 7-inch wrench. (The bolts are spaced 6 inches 
 apart on the pitch circle.) 
 
 1= length of packing in feet; 
 
 v= velocity of rubbing in feet per minute. 
 
 To cut down the amount of friction from the stuffing boxes, 
 w r ater is sometimes allowed to enter the space around the shaft 
 on the suction side as shown in Fig. 430, so that air is not 
 allowed to enter the suction space. This water may be drawn 
 in without interfering with the action of the pump and at the 
 same time the stuffing box at the end of the shaft does not have 
 to be tight enough to be air tight. The stuffing box on the other 
 side of the pump is under high pressure and water tends to 
 flow outward. This should be allowed to happen, as the small 
 amount of leakage tends to cool the stuffing box. In operating 
 pumps the stuffing boxes on each side should be allowed to 
 drip, as this not only cools the shaft but indicates that the pack- 
 ing is not too tight. 
 
 Friction of Water on Back of Impeller. The friction of 
 water on the back of the impeller is an important item and 
 may amount to a large part of the loss in the pump. 
 
 Let r be the radius to a particular part of the impeller and 
 2V the revolutions per minute. If ^>=the friction on unit area 
 at this point, the resistance of an elementary ring is 
 
 Now, p is proportional to the square of the relative velocity 
 
CENTRIFUGAL PUMPS 563 
 
 between the water and the surface of the impeller, and since 
 this would vary from the center to the outside, 
 
 p=k(2nNr) 2 , 
 hence, 
 
 /Vo 
 
 Total W /t/ = (27r) 4 &V3 I r*dr 
 
 J'c . ... (71) 
 
 Professor F. G. Hesse performed a series of experiments at 
 the University of California on a step bearing in which the 
 friction is about the same as that occurring on the two sides 
 of an impeller, and derives the following formula: 
 
 T^ = i 7 6Xio-W 3 Z) 5 ft.lb. per minute, . . . (72) 
 from this &'=58Xio- 5 . 
 
 D= diameter in feet; 
 
 Stodola gives the loss from the rotation of steam turbine discs 
 in steam as 
 
 / \ ^ 
 
 N' =0.0721 D i 2 ' 
 
 where N'=HP due to friction; 
 DI = diameter in feet; 
 
 HI = peripheral velocity of disc in feet per second; 
 f= weight of i cu.ft. of medium. 
 
 He gives this as an empirical formula from the results of Odell 
 and from his own experiments as well as the experimental work 
 of Lewecki. 
 
 If this is changed to a form similar to that given above with 
 762.5 it reduces to work per minute, 
 
 .... (73) 
 
 Of course this is not intended to be used with water as the 
 medium in which the disc is revolving, but the form is quite 
 similar and the formula is given here for reference. 
 
564 
 
 PUMPING MACHINERY 
 
 Unwin in his hydraulics describes experiments for the deter- 
 mination of this friction and reduces the following equation: 
 
 , ; . . (74) 
 
 or 
 
 FIG. 423. Sectional View of Worthington Volute Pump. 
 
 In order to cut down this friction loss the leakage water 
 may be drawn off so that it will not completely fill the space 
 between the impeller and the casing. This, however, means 
 increased leakage, as the pressure between the wheel and the 
 space around it at the discharge edge will be greater in this 
 case. At times air under pressure is put in this- space to keep 
 the water back and thus reduce friction. It is well to remem- 
 
CENTRIFUGAL PUMPS 565 
 
 ber that since the friction varies as N 3 D 5 it is more important 
 to use a small diameter and large N than a large D and small N 
 when the value xD 2 N =w% is fixed. This fact appears in practice 
 where small impellers are found. 
 
 FORMS OF CENTRIFUGAL PUMPS 
 
 There are many forms of centrifugal pumps. When the 
 guide vanes are omitted the water passes directly into the 
 volute casing through a filling -ring. This was the original form 
 of centrifugal pump. By some, manufacturers it is called a 
 volute centrifugal pump. Fig. 423 shows the section through 
 a single -suction Worthington Standard volute pump. Fig. 424 
 shows a double-flow volute pump of one of the earlier forms. In 
 Fig. 423, water enters the suction head A at the point B where the 
 suction) pipe is attached. From the head it enters the impeller 
 C, meeting the vanes D, which force the water outward into the 
 filling ring E and finally into the volute casing F. The impeller 
 is of the form shown : in Fig. 408. The vanes are enclosed on 
 each side by a disc so that the water does not come in contact 
 with the bearing head G on one side or the suction head H on 
 the other. The form of impeller in Fig. 424 is not of the enclosed 
 type and consequently there is considerable water friction. 
 Moreover, the mechanical friction is greater in the unenclosed 
 type, and as the vanes fit so closely to the heads a slight amount 
 of end play will cause rubbing. 
 
 In the impeller there is an increase of energy due to the 
 increase in the peripheral speed. This increase of energy may 
 take the form of increase in pressure if u is the same as c, or 
 it may be in the form of increase in kinetic energy if p a =pd- 
 In general there is an increase in both the kinetic energy arid 
 the potential energy or pressure. In general p a <pd, so that 
 there is leakage from the outlet of the impeller into the space 
 around this part of the pump. 
 
 The bushing rings /, /, prevent excessive leakage from the 
 space around the impeller into the suction space. The space 
 within the right-hand bushing ring is connected to the suction 
 
566 
 
 PUMPING MACHINERY 
 
 space by the openings in the central part of the impeller disc, 
 so that any leakage at the ring will be taken into the suction 
 space. These openings also equalize the pressure at the center 
 of the wheel and reduce the end thrust. The stuffing boxes 
 at the two sides of the pump are equipped with a cage J between 
 the two sets of packing rings for the introduction of 'oil or grease 
 from cups attached to the stuffing boxes K and to connect water 
 pipes from discharge casing to farm a water seal. The outboard 
 bearing L and the suction bearing M are provided with oil rings 
 
 FIG. 424. Double Flow Unincased Impeller. 
 
 N and the oil cellar 0. The oil is introduced through the opening 
 capped by the cover P. The thrust bearing Q, supporting the 
 small end thrust, is principally used to keep the impeller in 
 a central position. The cap R on the end of the shaft is an oil 
 thrower. The shaft S is supported by two bearings and turned 
 with a small shoulder at the center so that the impeller may 
 be secured in a given longitudinal position by the nut. The 
 impeller is keyed to the shaft and the coupling used to attach 
 this to the motor is also keyed. The ribs around the volute 
 casing are used to strengthen that casting, which is not a com- 
 
CENTRIFUGAL PUMPS 
 
 567 
 
 plete pipe on account of the opening from the filling ring, and 
 for that reason the bending action on the casting requires addi- 
 tional ribs. 
 
 It is the object of the whirlpool chamber to reduce the 
 absolute velocity of discharge, changing it into pressure. It 
 does not accomplish this efficiently, as there is shock here and 
 a sudden change in velocity which means a loss. By having 
 this chamber gradually increase in area the velocity of the 
 water decreases as it passes through, and with this action the 
 
 FIG. 425. Worthington Belted Volute Pump. 
 
 pressure increases. To make this change in an effective manner, 
 however, the diffuser is used, in which there is no shock, and the 
 velocity head is changed into pressure. 
 
 Fig. 425 illustrates one of the Worthington single-flow 
 volute pumps for a belt drive, while Fig. 426 illustrates a double- 
 flow volute pump, motor driven. These figures show the 
 external appearance of the pumps; the pipes for water cir- 
 culation in the stuffing boxes and thrust bearing; the stiffen- 
 ing webs; the method of supporting the casing and the method 
 of attaching the heads to the casing. The double-flow pump 
 
568 PUMPING MACHINERY 
 
 of Fig. 426 is not equipped with a thrust bearing, as in this type 
 of pump there should be no end thrust. 
 
 These pumps are intended primarily for heads up to 65 
 feet, and in most cases when this head is exceeded the turbine 
 type of pump with a set of diffuser vanes would be used. The 
 impellers built by Worthington are specially designed to suit 
 the conditions of the particular service required by the pur- 
 chaser, and this should always be done, as each speed and head 
 
 FIG. 426. Motor Driven Worthington Double Flow Volute Pump. 
 
 requires a definite form of impeller. The suction head shown 
 in Fig. 423 is so arranged that water may enter from all parts, 
 giving an adequate supply to the impeller. 
 
 These pumps are built in different sizes, as shown by the 
 tables below; the first table giving the capacities, speeds, heads 
 and power; the second, the dimensions of the casing. 
 
 The pumps shown so far have been arranged with the suc- 
 tion below, while the discharge has been taken off vertically. 
 This arrangement may be changed to suit as shown in Fig. 
 427, which illustrates a few of the different arrangements pos- 
 sible with their designating numbers. 
 
CENTRIFUGAL PUMPS 
 
 569 
 
 
 o m ro 
 
 moovcooom mmmmm 
 
 PH " 
 
 VO 
 
 sti^ 
 
 OONco^vovOin inmTTto 
 
 u 3 > 
 
 & C 2 
 
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 m o i'" 
 
 00 00 NO 
 O r^ m 
 
 ONininmoom OOOOO 
 
 II 
 
 
 W H M 
 
 M ON 00 
 
 c rt -5^ 
 
 in 
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 c? 2) 
 
 moommmo mommo 
 ^OOOMvCt-N vOmCNTO 
 
 TOOOWMVOM O\vO'-'r-fo 
 
 ^p JJ 
 
 
 M M w 
 
 t- 
 
 
 
 1 S 
 
 in o Q 
 
 ^ >> 
 
 
 
 
 H^^f r o"^poor*'T'i 500 ToNr^r^fM^s" 
 
 8^1 
 
 in 
 
 
 
 ^ o '2 
 
 
 
 
 
 
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 moo 
 
 o. ^.S'So 
 
 4J 
 
 oo m rr 
 
 ??0in2r- ^HNCM. 
 
 c S*" fc 
 
 T 
 
 g^vg^g 
 
 %^-^^-ITi ^ ^s-N^^-g- 
 
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 M M M 
 
 
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 m 
 
 rt S-ti o 
 
 4-5 
 
 
 Hi-,inNt^Min MO-TONin 
 
 J'ftJ 
 
 O 
 
 | r* | Hjj | 
 
 ^Iml^lolTl^T^-l 00 1* To r "5 r^- i'm 
 
 8211 
 
 
 
 ONOf^t^coir.M ONt^fOOvo 
 
 -Sfe^c 
 
 
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 <N T O 
 
 
 lljl 
 
 -*J 
 
 
 
 O M O O 
 
 io in 
 pOMininomO inino"~m 
 
 fill 
 
 
 
 
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 I- N 00 
 
 moomo m OmOOO 
 
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 O\ t*~ \c 
 
 -r^n^PP^ iTk s ? 2 r 
 
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 l.l.?l 
 
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 ^"i'^^js^H" rrrr,? M 
 
 Ili5 
 
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 Of, 10 NO 
 
 Ommmmoi^. omom 
 
 M| N | col * 1 'vO 1 ^r^l c3 1 roTTl vol'xO ON 
 
 
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 Moomfooxoovo inTr>N"5 
 
 <D^^*W 
 
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 Hill 
 
 
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 5 8 8, 
 
 OOOOO O. O OOOOO 
 minOOOOO OOOOO 
 Mt^-Ooccoo inmioOO 
 
 fijl 
 
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 OOOOOOO OOOOO 
 M<sro inN^cooin 
 
 frl ^ >> ^ o 
 
 p.^S ^^> 
 
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 8 & a 
 
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 rt ag 
 
570 
 
 PUMPING MACHINERY 
 
 a 
 
 o 
 
 "I "~, O OO O O O 
 
 00 O ioOO M vo t> o t^- rf 
 
 OO O M ro 10 
 
 00 OO fO ^t 00 
 
 1"- OO O O **) <OO OO M 
 
 *4"* N> H* "-H 1 Hoc ' 
 
 O N r*VO OO * N OO 
 MHMMhHCICICSCSO 
 
 W 
 
 O 
 
 fo 
 
 O fO M OOO 
 
572 PUMPING MACHINERY 
 
 A recent design of volute pump for a low head but large 
 capacity is shown in Fig. 428. This pump is known as a tri- 
 rotor volute pump. It was installed in the Interborough Rapid 
 Transit Station for circulating the water through the surface 
 condensers. It is in reality three double-flow volute pumps 
 with four vertical suction pipes rising to the center of the casing 
 and threading among the three horizontal discharge pipes 
 which unite and discharge from one large opening at the front 
 of the bed plate. The head being small, there are no stiffen- 
 ing ribs found on the exterior of the casing. The ring oil 
 bearings, thrust bearing, water-supply pipes for stuffing boxes 
 
 FIG. 428. Worthington Turbine Driven Tri-rotor Volute Pump. 
 
 and thrust bearing as well as the general arrangements for casting 
 the casing and heads are clearly seen. The figure also illustrates 
 the method of driving the pump by a steam turbine. 
 
 The section, Fig. 429, is given not only to show the method 
 of bringing the water to the various suction chambers and taking 
 it from the separate volute casings into the one discharge 
 opening, but the arrangement of the impeller is to be noted. 
 The impellers here are the typical double-flow volute runners 
 and the bushing rings, filling sleeves and bearings are all in 
 evidence. These two figures illustrate the type of split-casing 
 pumps, which type is usually employed in sugar factories, or 
 gas plants where the pump may require cleaning frequently. 
 
HI 
 
 
 
FIG. 429. Worthington Tri-i 
 
c r Volute Pump. 
 
 (To face page 572} 
 
CENTRIFUGAL PUMPS 
 
 573 
 
 When higher heads than 60 or 70 feet are to be overcome 
 the volute pump is replaced by the form of pump used in the 
 development of the theory of centrifugal pumps, one with 
 diffuser vanes. This type of pump is known as a turbine pump. 
 The rotative speed of these pumps is often quite high, as the 
 diameter of the impeller may then be made small. The dis- 
 charge velocity in these pumps is quite high and hence to change 
 this over to pressure head efficiently the diffuser is used. At 
 times the head is so great that the losses would be large, and 
 
 DISCHARGE 
 
 SUCTION 
 FIG. 430. Section of Worthington Standard Turbine Pump. 
 
 in such a case two pumps would be placed in a series, the first 
 one caring for part of the pressure and the second for the remain- 
 ing part. In some cases as many as ten steps are used. These 
 separate pumps may be combined in one casing, giving a multi- 
 stage turbine pump, each stage caring for a head of from 75 
 to 200 feet. Fig. 430 is a section of a Wortliington two-stage 
 pump. The action of this pump is similar to the one described 
 earlier, except that as the water leaves the impeller A it enters 
 the diffuser B, which has channels formed as shown in Fig. 
 431. The diffuser is made up of two parts, the diffusion ring 
 C with vanes and the diffusion ring without vanes D, In this 
 
574 
 
 PUMPING MACHINERY 
 
 the velocity is reduced so that water passes through the channel 
 ring E with a low velocity but with considerable pressure, 
 entering the second impeller F, which discharges into a second 
 
 FIG. 431. Diffuser. 
 
 diffusion ring from which it discharges into the final discharge 
 channel G in the casing H of the pump, leaving the dis- 
 charge /. The casing is cast with the outboard head solid and 
 it is machined on the inside so that the diffusion rings and 
 
 FlG. 432. Four Impellers and Shaft. 
 
 channel rings can be introduced from one side with the suction 
 head bolting into position on account of the turned projection 
 on its inner face. All parts are thus held in their proper posi- 
 tions. Such a construction makes it possible to dismantle 
 the pump readily for examination and repair and yet reassemble 
 
CENTRIFUGAL PUMPS 575 
 
 with all parts coming into their proper positions. From Fig. 431 
 the manner of forming the diffusion ring with vanes is seen and 
 it is evident that it is possible to polish these smooth to reduce 
 the friction losses. Fig. 432 shows the four impellers of a four- 
 stage pump. The entrance channels are seen on the right-hand 
 sides of the discs. At the end of the shaft are seen the collars 
 of the thrust bearing. In Fig. 430 the stuffing boxes are shown 
 attached to the heads by flanges and the bearings are seen on 
 heavy brackets. The thrust bearing is made 
 in halves and bolted to the end of the sue- 
 tion bearing. This construction, shown in 
 Fig. 433, is necessary for the introduction of 
 the shaft. The cooling of this bearing is 
 
 ' FIG. 433. Worthmgton 
 
 accomplished by a current of water through Thrust Bearing, 
 the shell, while continuous lubrication is 
 effected by the oil thrower /. The manner of introducing 
 water into the suction stuffing box is evident from the figure. 
 The bushing rings at K prevent the excessive leakage of water 
 into the suction. 
 
 The method of filling the space between two impellers by 
 means of the distance bushing M and the collar N of the chan- 
 nel ring may be seen in Fig. 430. The construction of a stage 
 pump of greater number of stages would be similar to this. 
 
 The materials used for the various parts of the pump depend 
 on the kind of liquid handled. The impellers are of cast iron 
 or bronze and are finished smooth on the outside and inside. 
 They are balanced. The diffusion rings are made of the same 
 materials as the impeller. The shaft is made of steel or Tobin 
 bronze. 
 
 Fig. 434 illustrates the method of constructing a two-stage 
 pump as made by the Alberger Pump Company. The arrange- 
 ment of the diffusing vanes, the channel ring, the bearings, 
 the stuffing boxes and the method of uniting the various parts 
 together are all evident from the figure. The pumps built by 
 this company are also designed for the actual conditions of 
 operation and the surfaces of the impeller and diffuser vanes are 
 highly finished all over to reduce the losses. 
 
576 
 
 PUMPING MACHINERY 
 
CENTRIFUGAL PUMPS 
 
 577 
 
 Fig. 435 illustrates a ten-stage motor-driven pump while 
 Fig. 436 is an eight-stage pump. These pumps are for high 
 heads. 
 
 When high heads are to be overcome large diameters of 
 
 FIG. 435. Worthington Motor Driven Ten-stage Pump. 
 
 impellers could be used to get the necessary velocity, but the 
 friction of the water in the impeller passages and on the im- 
 peller back would be very great and the width of passages 
 
 FIG. 436. Worthington 
 
 Eight-stage Turbine Mine Pump with 450 H.P. 
 Motor. 
 
 at discharge would be very small, and moreover there would be 
 erosion of the vanes by the water under high velocity. For 
 these reasons about 150 feet is the limit for each stage and for 
 great heads a large number of impellers is used. Since these take 
 
578 
 
 PUMPING MACHINERY 
 
 up considerable room on the shaft there is a limit to the number 
 which can be put between bearings without making the shaft 
 very large. The shaft must be designed with regard to the 
 critical speed. Ffve or six stages are found within one casing, 
 but when more are required two casings are used connected in 
 series. 
 
 It is well to think of the first wheel increasing the kinetic 
 energy and the potential energy, then the first diffuser changes 
 
 FIG. 437. Worthington Fire Boat Turbine Pump. 
 
 the increase of kinetic energy into pressure, so that when the 
 second impeller is entered the water has the same kinetic energy 
 as it has at entrance to the first impeller, but with a great 
 increase of pressure. This is repeated in each stage and at the 
 discharge from the pump at the last stage there may be less 
 kinetic energy than at entrance, but the potential energy in the 
 form of pressure is very great. 
 
 When heads of several hundred feet are required for fire 
 service the pump takes the form shown in Fig. 437. This type 
 
CENTRIFUGAL PUMPS 579 
 
 of unit is installed on the New York and Baltimore fire boats. 
 
 FIG. 438. Alberger rooo-gal. Underwriter Fire Pump, Turbine Type. 
 
 FIG. 439. Worthington Four-stage Boiler Feed Pump. 
 
 These pumps are also used as fire pumps for factories and the 
 Board of Fire Underwriters has issued specifications of the 
 
580 
 
 PUMPING MACHINERY 
 
 same nature as those given in Chapter XI for reciprocating 
 pumps Covering the equipment and construction of centrifugal 
 underwriter fire pumps. Fig. 438 illustrates a looo-gallon pump 
 of this type. The fixtures of this pump are quite similar to 
 those discussed earlier in the work. 
 
 For boiler feeding a motor driven or steam turbine driven 
 stage pump may be used. When turbine pumps are used 
 
 there is an absence of knocking and 
 shock common with reciprocating 
 pumps because of the steady dis- 
 charge, and moreover, as will be seen 
 later, there is no danger of excessive 
 /Y\ pressure, even if all feed valves are 
 
 HL . closed on the boilers. Fig. 439 
 
 illustrates a four-stage boiler feed 
 pump. Pumps similar to this could 
 be used for house or elevator service, 
 being controlled by a switch operated 
 by a float or pressure regulator. 
 
 For mine sinking or for use where 
 space is limited, vertical pumps may 
 be installed, using electric' motors or 
 steam turbines to operate them. 
 These pumps have enormous capacity 
 for their size, and operate successfully. 
 Fig. 440 illustrates a pump built to 
 operate under 1250 feet head in a 
 single casing. Should the water to 
 be handled be acid the pump is built 
 of a composition to resist this. 
 
 When high-speed pumps are built 
 these are made with double runners 
 
 to eliminate end thrust. Fig. 441 shows a section through 
 one of these a built by Worthington and shown in Fig. 
 437. In this pump the suction water for one side is carried 
 through the open spaces between the channels of the diffuser 
 ring of Fig. 431 and enters a well-rounded cavity leading to 
 
 FIG. 440. Worthington Turbine 
 Sinking Pump. 
 
FIG. 441. Worthington 
 
>eed Pump. 
 
 (To face page 580) 
 

 
CENTRIFUGAL PUMPS 
 
 581 
 
 FIG. 442. Alberger Two-stage Volute Pump, Engine Driven. 
 
 FIG. 443. Alberger Standard Volute Pump, Vertical Shaft. 
 
582 PUMPING MACHINERY 
 
 the impeller. In this way the impeller receives water on each 
 side. The water from the diffuser is then delivered to the suc- 
 tion of the second stage and is delivered from the diffusion 
 ring of this stage into the discharge channel. The velocity 
 head in this channel is so small compared with the pressure 
 head that there is not the need of forming the usual volute to 
 reduce the losses and so the discharge is taken up on each side 
 of the diffuser to a discharge flange in the center in a concentric 
 channel. 
 
 Fig. 442 illustrates a two-stage volute pump built by Alberger. 
 In the figure the discharge is taken from the casing at an angle 
 to the horizontal. The general lines of the pump, the feet for 
 attachment to bed plate, the water seal for the suction stuffing 
 box and the well-supported bearings are to be noted. The 
 volute pump is not used so often in stages, but there is no reason 
 why this cannot be done, as is shown in the illustration. The 
 volute pump may be applied with the shaft in a vertical posi- 
 tion as shown in Fig. 443. Such a construction may be neces- 
 sary on account of the lack of room, or the motor may be placed 
 at considerable distance above the pump on account of flooding. 
 
 Another use for the centrifugal pump in the last few years 
 is in connection with the water removal from jet condensers. Fig. 
 444 shows a steam-engine driven centrifugal pump attached at 
 the lower end of an Alberger condensing head. The condensed 
 steam and condensing water are removed and discharged into 
 the atmosphere by the pump. The air is removed from the 
 head by means of a dry air pump, similar to that described in 
 Fig. 368. 
 
 In order to eliminate the end thrust with turbine pumps of 
 several stages Mr. C. W. Larner has recently patented the method 
 shown in Fig. 445. In this arrangement the water from the 
 second diffusion ring is passed between the channels of the first 
 channel ring on its passage to the second channel ring. In the 
 figure there are four stages. The method used here is some- 
 what similar to that shown in the high-speed pump, Fig. 
 441, although in that figure each impeller is balanced by being 
 a double-flow impeller and half of the water for the second stage 
 
CENTRIFUGAL PUMPS 
 
 583 
 
 is carried through the channels in the diff users. In Fig. 445, 
 the impellers are single-flow impellers; consequently with the 
 same openings at the suction side the pump will handle less 
 water through a greater head than that for the pump shown 
 in Fig. 441. 
 
 The method used by Sulzer to accomplish this is shown 
 
 FIG. 444. Alberger Centrifugal Condenser. 
 
 in Fig. 446. This has been used for many years. The method 
 is to take the discharge from the first stage and pass it through 
 the open spaces in the two diffusers, Fig. 431. The water is 
 then brought in on the opposite side of the second impeller 
 from that ,used on the first. This balances the pressure due 
 to impact, since the velocity of the water in an axial direc- 
 
584 
 
 PUMPING MACHINERY 
 
 tion is the same at each of these points on the second im- 
 
 peller. 
 
 Rateau eliminates the end thrust for one condition of running 
 
 FIG. 445. Larner Pump. 
 
 FIG. 446. Sulzer Pump. 
 
 by cutting off one side of the impeller as shown in Fig. 447, 
 and allowing the unbalanced force on the side remaining to 
 balance the pressure at entrance. This means considerable 
 leakage, however, and moreover the wheel for a part of 
 
CENTRIFUGAL PUMPS 
 
 585 
 
 one side acts as an imencased wheel, which means more fric- 
 tion. 
 
 In the three-stage pump of Worthington, Fig. 448, the Jaeger- 
 Worthington method of holding the impeller in a central posi- 
 tion is illustrated. In this pump the 
 edge of the diffuser at A is cut on 
 each side so that should the impeller 
 move to the right the increased clear- 
 ance space allows more water to leak 
 out on that side than would be cared 
 for by bushing rings at B. As a result 
 the water " backs up " on the right 
 side and the pressure from this forces 
 the impeller to the left. Motion to the 
 left would result in an excess of press- 
 ure on that side. To permit this free 
 action the thrust-bearing grooves are 
 made in a sleeve which has a slight 
 play. There is no pressure on the 
 thrust bearing until the sleeve is 
 brought to one end of its travel or the other. In this way the 
 thrust is carried by water on the backs of the impellers and 
 the impeller s*hould take such a position that the excess leakage 
 on one side would care for the thrust. 
 
 In the pump of Weise and Monski the impellers are divided 
 into two groups, one group with the inlets on the right and one 
 set with the inlets on the left. After passing through the first 
 group the water is discharged through a passage in the casing 
 leading to the other end of the pump, where it enters the inlet 
 of the first impeller of the second group. This is the equiva- 
 lent of having two pumps with their shafts connected, one pump 
 discharging into the other through an outside connection and 
 the inlets to the impellers on one pump turned in the opposite 
 direction from those on the other. Fig. 436 shows such an 
 arrangement. 
 
 In the Schwartzkopff type of pump and in the later Sulzer 
 pumps a balancing piston is used at the end of the pump sub- 
 
 FlG ' 44 '~ 
 
 Centrifu ^ al 
 
586 
 
 PUMPING MACHINERY 
 
CENTRIFUGAL PUMPS 
 
 587 
 
 FIG. 449. Section of a Buffalo Balanced Two-stage Pump. 
 
 FIG. 450. Section through Buffalo Pump. 
 
588 
 
 PUMPING MACHINERY 
 
 ject to the discharge pressure or a portion of it. The pressure 
 on the piston balances the thrust. 
 
 Fig. 449 gives a section through a two-stage Buffalo pump, 
 showing their method of providing for end thrust while the 
 cross-section, Fig. 450, shows the form of passages used to take 
 the water to the second impeller inlet. Fig. 451 shows one of 
 
 FIG. 451. Method of Assembling Buffalo Multi-stage Pumps. 
 
 these pumps partially dismantled. The outer part of the dif- 
 fuser vanes B,B do not fit against the shell, but are bolted by 
 tie bolts G to the return chamber D and the diaphragm C, 
 which reaches to the floating ring around a bushing on the 
 shaft. The return plate D and diaphragm C have a forced 
 fit into the casing. The bolts / are jack bolts used in forcing 
 out the diaphragm C and return plate when necessary to examine 
 the interior. In this arrangement of pump it is not necessary 
 to disconnect the suction or discharge pipe when repairing the 
 interior. Passages cast in the casing deliver water from the 
 
CENTRIFUGAL PUMPS 
 
 589 
 
 left-hand diffuser B to the return chamber D on the right. These 
 are seen in Figs. 449 and 450. 
 
 Fig. 452 shows one of the Buffalo pumps with a vertical 
 
 FIG. 452. Buffalo Vertical Underwriter Fire Pump. 
 
 shaft equipped as an underwriter's fire pump. The pump 
 would be driven by a motor through the coupling shown in 
 Fig. 453 . This type of coupling is one in which pins on one flange 
 
590 
 
 PUMPING MACHINERY 
 
 drive the other flange by means of rubber bushings. Such an 
 arrangement furnishes a yielding medium, so that there is some 
 chance for self alignment when the shafting is deflected in part 
 of the complete machine. In most installations there must 
 be a flexible coupling between the motor and the pump on 
 account of change of alignment at high speeds. 
 
 The Allis-Chalmers pumps built for the high-pressure fire 
 service in New York demand mention as another type of multi- 
 stage pump. These pumps, Fig. 454, are built with the passages 
 in the impellers so formed that the outflow is in an axial direc- 
 tion. This form of pumpi is known as the Gelpcke-Kugel form. 
 These pumps on test gave efficiencies from 72 to 79 per cent. 
 
 FIG. 453. Coupling. 
 
 There is some objection to the discharge as shown, in that the 
 leakage is apt to be excessive. 
 
 Testing Centrifugal Pumps. The centrifugal pump is often 
 designed to run at a fixed speed, being driven by an electric 
 motor. The amount of power used and the efficiency will vary 
 with the quantity of water pumped, and for that reason tests 
 are made to determine the characteristics of the pump. 
 
 The characteristics desired are the curves of power, of pres- 
 sure and of efficiency, plotted against the quantity of water 
 pumped. To obtain these, measurements are made as shown 
 in the diagram, Fig. 455. 
 
 The quantity of water may be determined by a Venturi 
 meter, a Pitot tube, a calibrated nozzle or a weir. The head 
 against which the pump is working is determined by two 
 gauges a suction gauge, and a discharge pressure gauge, or 
 
CENTRIFUGAL PUMPS 
 
 591 
 
592 
 
 PUMPING MACHINERY 
 
CENTRIFUGAL PUMPS 593 
 
 the suction may be attached to one end of a mercury U tube 
 while the discharge is connected to the other end. The power 
 input is determined by knowing the power and efficiency of 
 the electric motor, by indicating the engine used to drive the 
 pump or by the use of a dynamometer. The arrangement of 
 the apparatus for such a test is seen in Fig. 455, in which there 
 are several instruments for measuring the water. The suction 
 gauge at A is usually a mercury gauge. The pressure gauge 
 at B may be a Bourdon gauge. The Venturi meter at C, the 
 Pi tot tube at Z), the nozzle E with its pressure gauge F, or the 
 weir G may be used to determine the water used. The Pitot 
 tube may be applied simply by using a tube flush with the pipe 
 surface for the static pressure tube and sliding the velocity 
 tube across the pipe, making a traverse. With a sharp, small 
 opening in the velocity tube and the tube kept parallel to the 
 axis of the pipe the constant for the tube is unity so that 
 
 h is the difference between the pressures in the litot tube and Lhe 
 static tube. 
 
 q= I 2irrdrV 2gk = S.o2X27r I rvhdr. ;. . (75) 
 Jo Jo 
 
 Hence if \/h is multiplied by r and the product is plotted on 
 a straight base for different values of r from o to ro the area of 
 this curve when multiplied by 16 04 ?r will be the quantity in 
 cubic feet, since all measurements are in feet. 
 
 The formula for the Venturi meter is 
 
 ' Q ~ k VZ?=AJ V * g(Hl - H *>> ' " ' (76) 
 
 where A 1 and A 2 are the areas of the sections of the meter at 
 which the pressures are HI and H 2 feet of water. 
 The formula for the calibrated nozzle is 
 
 Q = k\ / h, 
 
 where h is the head shown by the gauge at F measured in feet. 
 For the rectangular weir the formula of Francis, 
 
 ..... (77) 
 
594 PVMP1NG MACHINERY 
 
 may be used with fair approximation to the correct value. In 
 
 v 2 
 this H is the head on crest of weir in feet and h= is the 
 
 2g 
 
 head of velocity of approach and b the breadth of the weir. All 
 measurements are in feet. The above is applicable to a weir 
 with end contractions. If contractions are suppressed the term 
 o.2H is omitted. 
 
 When the quantity of water is not large a triangular weir 
 may be used, with success. This notch is made with sides at 
 right angles and for small quantities the head is much higher 
 than for the rectangular weir, and for that reason there is less 
 error in measuring the acting head. The formula for this weir is 
 
 ?=o. 3 o5ff 5 / 2 . .. . '. ;.';. .. . (78) 
 
 The constant has been studied, and although it does vary 
 the limits are not far apart, being 0.291 at 0.7", 0.306 at ij" 
 nnd 0.303 at 5". In this formula H is head on crest measured in 
 inches; in all of the above Q is measured in cubic feet per second. 
 
 The head acting not only includes the suction and discharge 
 head, but it should include the velocity head acting at the time. 
 
 Vo 2 
 
 Having the total head H equal to hi +h 2 + and the quantity 
 of water Q, the useful power becomes 
 
 In this w= weight of i cubic foot of water, which is deter- 
 mined by actually weighing a known quantity of water, by 
 the use of a hydrometer, or by taking the temperature and 
 referring to a table of weights of water. 
 
 The efficiency of the pump is determined when the applied 
 power is found by means of a dynamometer or calibrated 
 electric motor. 
 
 HP 
 
 The results of a test are usually plotted with the quantity 
 of discharge as abscissae using head, efficiency, and applied 
 
CENTRIFUGAL PUMPS 
 
 595 
 
 power. as ordinates of three separate curves. Fig. 456 shows 
 such curves from a Worthington 8" volute pump. In this 
 pump the head remains constant for a considerable time as 
 the discharge valve is opened, so that more water is discharged 
 until 1800 gallons per minute is .reached, when the head begins 
 to drop rapidly, accompanied by a decrease of efficiency arid 
 a slight increase of power. It is to be noted that the increase 
 of power begins to fall off at high discharges, due to the decrease 
 
 50 
 .40 
 30 
 20 
 10 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 w, 
 a 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 1 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 'o 
 
 G 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 S 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 ^ 
 
 *&* 
 
 y 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 n 
 
 'HO 
 
 ad 
 
 
 *N 
 
 N 
 
 
 
 
 
 
 
 7U 
 
 
 
 
 / 
 
 ' 
 
 
 
 
 
 "*x 
 
 > 
 
 \ 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 \ 
 
 \ 
 
 V 
 
 \ 
 
 
 
 
 I 
 
 
 
 
 
 
 
 
 <i 
 
 p. 
 
 r ^- 
 
 & 
 
 ^ 
 
 -^ 
 
 
 
 
 
 p- 
 
 1 
 
 
 
 
 
 
 
 ^ 
 
 J^V 
 
 ^** 
 
 
 
 
 
 \ 
 
 
 
 
 o 
 
 
 
 
 
 ^ 
 
 ^^ 
 
 
 
 
 
 
 
 
 
 
 
 
 1 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 200 400 600 800 1000 1200 1400 1600 1800 2000 2200 2400 2600 2800 3000 
 Capacity in Gallons per Minute. 
 
 FIG. 456. Test Curves from an 8" Worthington Volute Pump. 
 
 of head. The head does not increase as the discharge valve is 
 closed and the power decreases, due to the decrease in quantity. 
 The efficiency has a maximum point because when the quantity 
 changes from the designed quantity, the water passes through 
 the wheel with impact. This may be seen when it is remembered 
 that the speed of the pump is fixed and the velocities in the 
 channels are fixed by the equation 
 
 (81) 
 
596 
 
 PUMPING MACHINERY 
 
 In Fig. 457, from -a Worthington class B volute pump, the 
 head is seen to be changed more rapidly than in the preceding 
 figure. The efficiency reaches a higher value while the brake 
 horse power of the motor, which is delivered to the pump, 
 actually decreases as the quantity passes beyond 4000 gallons 
 per minute; this is associated with a head of 44 feet. The 
 power curve is quite characteristic of a centrifugal pump, as 
 
 1000 2000 3000 4000 5000 
 
 Capacity P.M. in Gal's 
 
 FIG. 457. Test Curves of a 10" Worthington Class B Volute Pump. 
 
 it shows how, with a proper design, these is no danger of over- 
 loading the motor if the water is shut off (no discharge) or if 
 the discharge pipe should break (maximum discharge). 
 
 Fig. 458 gives the characteristic curve from a six-stage 
 turbine pump used in the Brooklyn High Pressure Station. In 
 this figure the entire range from zero discharge to maximum 
 discharge under no head is shown. The horse power is seen 
 to have a maximum value at 840 H.P. and the maximum 
 efficiency is seen to reach .the value of 76 per cent. The 
 
CENTRIFUGAL PUMPS 
 
 597 
 
 1000 
 
 90090 
 
 Head 
 
 80080 
 
 700 
 
 600 
 
 :o- 
 
 500% 
 
 
 40040 
 
 
 300 
 200 
 
 100 
 
 30 
 
 -20- 
 
 \ 
 
 <\ 
 
 800 1600 2400 3200 4JJJ 4800 5600 6400 
 
 Gallons per Min. 
 
 FIG. 458. Test Curves of ic" Six-stage Worthington Turbine Pump. 
 
 120 
 
 
 . - 
 
 ^ - 
 
 
 
 
 
 -^ 
 
 
 , 
 
 
 
 
 
 
 
 
 
 
 ^^- 
 
 
 
 
 
 
 
 
 
 
 ^ 
 
 
 
 
 
 
 
 
 100 
 >. 
 
 
 
 
 
 
 
 
 
 
 
 
 X 
 
 ^\ 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 ^*~ 
 
 ^^ 
 
 ^ * 
 
 \ 
 
 
 
 
 
 '3 80 
 
 S 
 
 
 
 
 
 
 
 
 
 ^ 
 
 ^^" 
 
 
 
 
 
 \ 
 
 
 
 
 W 70 
 * 
 g _ 
 
 
 
 
 
 
 v& 
 
 &$ 
 
 v^^ 
 
 
 ^-^*" 
 
 
 
 * 
 
 ^v 
 
 ^ 
 
 \ 
 
 
 
 03 60 
 
 
 
 
 ^ 
 
 ^^ 
 
 s* 
 
 ^^ 
 
 
 
 
 
 
 
 
 
 \ s 
 
 s s 
 
 
 N 50 
 
 a 
 
 
 ^ 
 
 ^ 
 
 > 
 
 X 
 
 
 
 
 
 
 
 
 
 
 
 
 \\ 
 
 
 -s 40 
 
 
 
 ^ 
 
 
 V 
 
 ? 
 
 
 
 
 
 
 
 
 
 
 
 
 
 \ 
 
 \ 
 
 w 30 
 
 
 ; 
 
 ? 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 \ 
 
 20 
 
 
 / 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 r 
 
 10 
 
 / 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 3 
 
 2( 
 
 X) 
 
 4< 
 
 30 
 
 Q 
 
 00 
 
 8( 
 
 K) 
 
 10 
 
 00 
 
 n 
 
 IX) 
 
 1^ 
 
 too 
 
 11 
 
 i(XJ 
 
 180 
 
 Gallons per Minute 
 FIG. 459. Test Curves of an Alberger Single-stage* Pump. 
 
598 PUMPING MACHINERY 
 
 high efficiencies on these three diagrams are important to 
 notice. The characteristic curves from a 10" single-stage tur- 
 bine pump of the Alberger design is seen in Fig. 459. In this 
 the complete range of the pump is given. Fig. 460 gives the 
 excellent results of a pump built by the I. P. Morris Co. 
 
 These curves show how the centrifugal pump is suited to 
 conditions under which the head would vary or where the 
 quantity would vary. In the case of emptying drydocks 
 where the head changes as the dock is emptied, a volute pump 
 would be used. At the start the quantity would be great at 
 a very small head, while as the head increases the quantity 
 decreases, and in some cases the power required increases up 
 to a certain point, after which there is a decrease of power and 
 quantity as the head increases. 
 
 In the case of a fire pump (Fig. 458) the variation of pressure 
 with the change in quantity is evident; with 4 4oo-gallon 
 streams the head is 900 feet, while with 8 such streams the head 
 is reduced to 705 feet. When 12 4oo-gallon streams or 19 250- 
 gallon streams are used, the pressure is reduced to 380 feet. 
 In none of these cases is there danger of overloading the proper- 
 sized motor. In selecting a pump it is better to select one in 
 which the characteristic curve of head gives a maximum value 
 at zero discharge, as there will be no difficulty in obtaining the 
 head for which the pump was designed. If it is desired to keep 
 the head constant at the various discharges, the speed of the 
 pump will have to be increased, and with this the power of the 
 motor. A speed characteristic could be drawn for any varia- 
 tion of head from the head curve at a given speed by remember- 
 ing that the head varies as the square of the speed. 
 
 In order to classify pumps a similar method to that used 
 in turbine classification may be employed. 
 
 Specific Speed. One of the powerful aids in the classifica- 
 tion of turbines for the purpose of design so that a given design 
 may be used for any turbine having a given characteristic, is 
 the specific speed. This is the speed of a turbine of design 
 similar to a given turbine but of different scale, such that 
 under a unit head it would develop one unit of power. A 
 
CENTRIFUGAL PUMPS 
 
 599 
 
 similar characteristic may be used for a centrifugal pump; 
 here the specific speed of a pump will be that speed at which 
 
 8 9 ft a a 9 3 
 
 S * 9 S Jl ft 9 ft 3 
 
 a pump of similar design, but to a different scale, will pump one 
 unit volume of water, through a total head of unity. To derive 
 the formula for the specific speed of a centrifugal pump a 
 
600 PUMPING MACHINERY 
 
 method will be used similar to that used by Professor L. F. 
 Moody for the specific speed of turbines as given in Zeitschrift 
 f. d. gesampte Turbinenwesen, Sept. 10, 1909. 
 
 Suppose that the quantity of water lifted by a pump run- 
 ning at N a revolutions is Q a and the total head is kH. If 
 this wheel is now run under a total head of kiHi the speed 
 will be changed so that 
 
 because Noc Foe VS. 
 
 The quantity is equal to the area multiplied by the velocity, 
 and hence 
 
 (83) 
 
 If now the scale of the pump be changed so that the quantity 
 of water is changed to Q^ the diameter x being changed from 
 D a to ZY. 
 
 Now 
 
 hence 
 
 ' 
 
 (84) 
 
 When D a is changed to ZV the speed in revolutions per 
 minute will be changed in the inverse manner because the veloc- 
 ity remains the same as the head remains constant, hence 
 
 but 
 
CENTRIFUGAL PUMPS 601 
 
 and 
 
 Therefore 
 
 If AT"/ is to be the specific speed, the total head KHi will 
 be unity and the quantity Qi will be unity. 
 Hence 
 
 (87) 
 
 In this N is taken as R.P.M.; Q, cubic feet per section per 
 second; kH, feet head. 
 
 This could be derived from the expression for the specific 
 speed of a turbine by substituting QH for HP in the formula 
 (for turbines), 
 
 N -N 
 iv 
 
 This formula can be used in either the English or French 
 system of units. In the English system Q is in cubic feet per 
 second and H in feet, while in the French system Q is in cubic 
 meters per second and H is in meters. The value N 8 in the 
 French units may be changed to N 8 for the English system by 
 
 multiplying by j. 
 
 The value of the specific speed for a series of different pumps 
 has been plotted in Fig. 461 with the efficiency of the pump as 
 ordinates, and from it the specific speed for the best efficiency 
 may be found. 
 
 The specific speed of a pump is the same for all similar 
 pumps, and for a given quantity, head and number of revolutions 
 the specific speed may be found. This immediately tells the 
 designer the class to which the pump belongs and what assump- 
 tions are to be made. Should the specific speed be too high, 
 the pump could be made of several single stages in parallel, 
 
602 PUMPING MACHINERY 
 
 as the quantity Q in the formula refers to the quantity for one 
 impeller and the head H to the head for that impeller. For 
 a multistage pump the quantity Q is the quantity for the whole 
 pump while H is that for one stage. Should N 8 in a desired 
 plant be too small for high efficiency, the use of a multistage 
 pump would require a higher specific speed and so give a better 
 efficiency. 
 
 DESIGNING CENTRIFUGAL PUMPS 
 
 To illustrate the methods of design, applying the principles 
 of this chapter, a pump will be designed to lift 1000 gallons 
 per minute through 350 feet of pipe against a static head of 
 300 feet, using a motor at 750 R.P.M. or 1000 R.P.M. It is 
 assumed that these are the only speeds possible. The calcu- 
 lations are made with a slide rule. 
 . i. Specific Speed and Number of Stages. 
 
 1000 
 1000 gals, per nun. =g 77=2.23 cu.ft. per sec. 
 
 From (87), 
 
 From the curve, Fig. 461, 75 appears to be a good specific 
 speed, hence 
 
 IV - /J - 
 
 # = 3 6.6 ft. 
 
 This is a very small head for one stage, and consequently 
 a smaller specific speed might be used or a higher actual 
 speed. Decreasing the specific speed means a smaller efficiency, 
 but one which is not reduced very much if proper design is used, 
 using 60 for N 8 and the higher speed of 1000 R.P.M. 
 
 TT . 1000 / 
 
 #* = -rV2.23 = 25.0. 
 
CENTRIFUGAL PUMPS 
 
 603 
 
 This is a good head to use, requiring 3 or 4 stages. Using 
 3 stages and considering the losses in head to be 5%, 
 
 0.90- 
 
 0.80 
 
 0.70 
 
 0.60 
 
 0.50 
 
 30 40 50 60 70 80 90 100 110 120 
 
 0.90 
 
 0.80 
 
 0.70 
 
 0.50 
 
 50 100 150 200 250 300 350 
 R.P.M.-Speciflc Speed 
 
 FIG. 461. Specific Speed Curves. 
 
604 PUMPING MACHINERY 
 
 This is not as good a specific speed as could be obtained if 
 we desired to use another stage. In that case 
 
 looox 2.23 
 
 
 Could the speed of the pump be changed by gears, the 
 actual speed, for A T =75, and 3 stages would be, 
 
 NH l 75 X 
 
 N =-7 =- +=~ = 1650. 
 
 (f v 
 
 This is the speed which should be used with this pump 
 for best conditions but with given data either N g =5& or 
 N a =46 will be taken. To keep down the number of impellers 
 the second of these will be employed. From the curve of 
 Fig. 461, the efficiency is assumed as 78%. 
 
 2. Size of Suction and Discharge Pipe. With a short pipe 
 
 /OQQ 
 
 v 8 -= 8.02-J - approx. from (61) 
 = 19.6 ft. per sec. 
 
 sq - ins - 
 
 <^=4ft in. 
 
 Use 5" =4, 
 
 A 8 = 19.63, 
 
 2.23X144 
 
 v 8 = - - = 16.35 ft. per sec. 
 
 19.63 
 
 Loss assuming length to be 20 ft., 
 
 fy 64.32 
 
CENTRIFUGAL PUMPS 605 
 
 If discharge pipe is of the same size the loss would be 
 
 This loss is over 20% of net lift and hence a larger pipe 
 will be taken for the discharge, say, 8 inches. 
 
 ^=50.26 sq.ins. 
 
 6.38 ft. per sec., 
 
 ^,0.02x330(6.38)^ 
 
 ^zg A 64.32 
 
 This is a permissible loss. 
 If a lo-inch pipe is tried: 
 
 The saving in the use of the larger pipe is 
 
 (6.2-2.1) 2.23 X6 2 . 5=io4HR 
 
 If the pump has an efficiency, from the curve at a specific 
 speed of 46, of 80%, if the motor has an efficiency of 90%, 
 if the pump runs 4000 hours per year and if power cost I cent 
 per K.W. hour, the saving per year is 
 
 .oi =$43.20. 
 
 The additional cost of pipe is 
 
 [n(d d}txl 1 
 
 W- - Xcost per ton [i + installing factor], 
 
 2OOO J 
 
 ro.287r(io-8)X33QXi2 . ] r . 
 
 2000 - $25.00] [2] =$174.60, 
 
606 PUMPING MACHINERY 
 
 Now the interest, depreciation, taxes, and insurance on this 
 pipe line will be taken at 9%, so that the yearly charge will be 
 
 0.09 X$i74.6o = $15.71. 
 
 It is seen that the saving per year from the use of a lo-inch 
 pipe over an 8-inch one is $43.20, while the increased cost 
 is $15.71 per year. The lo-inch pipe will be used and con- 
 nected to the volute casing by a reducer from 8 inches. A 
 larger pipe should be tried in the same manner. If the saving 
 in power is greater than the increased cost on investment, 
 the larger pipe should be used. 
 
 To aid in solving problems of this kind the table on page 607, 
 taken from one of the bulletins of H. R. Worthington, is given. 
 
 In this table the amount of water usually carried in a pipe 
 of a certain size is given together with the loss in head in 1000 
 feet of straight pipe. Since the drop in head varies with the 
 length of pipe this can be used for various lengths. For large 
 pipes the quantities are given in million gallons per twenty- 
 four hours. 
 
 v 2 
 The loss in the wheel is assumed to be 0.3 = 1.2. 
 
 The total head is now 
 
 =310.0 ft. instead of 315 assumed. 
 3. Probable Brake H.P. of Motor and Size of Shaft. 
 
 ' = IO B.H.P. 
 
 The shaft to transmit this power at 1000 will have a 
 diameter given by the equation below if no bending t>- con- 
 sidered: 
 
 (32 
 -V~ 
 
 321,000 H.P. 
 
 P.OOOXIOO 
 
 \ 1000 X 10,000 
 
CENTRIFUGAL PUMPS 
 
 607 
 
 <3a<Da<Bac<cS<oCa<3c!a)G<ua<ie<uC 
 agaoaoaoaoaoaoao a_o a_o a.g a.o 
 
 ^2'o T ^oJ2'y^2cjJ2o r ^o r i2o^2"oJ2 o,23 uJ2 y.22 y 
 ni'C rt'C rt'C rt'C rt'C rt'C 3'C rt'C rt'C rt'C ST C 
 
 ^2 o^2 
 d 4J rt ' 
 
 _ 
 
 ^H fr . ^ fc . * "* fr . 
 
 S S 
 
 o o t o>o o t- o t o ** 
 
 H CM fO t-- "^00 O t- 
 
 t^ \O tr> M HI 00 
 
 T C*OO O to IO t^- to M 10 
 10 M HI O O V 
 
 HI ID 1000 O CvCO tvO OOOOO^O 
 
 N O N 'too 00 M \n\o >o M 
 
 
 oo oo 1000 lONOOOOHO^OOO 
 t^ N l~- TOO C\ 00 -CO rr O 
 
 <5>oONO^OOOOOHiOOOOioO<NOtoOO\OO 
 tO Nl NMO\NOrO<O tooO O\00 > O \O M 
 
 HI er> (f)in 
 
 Mf>vOlH 
 
 O -to-vO -d -00 -tO-O -tO-O -O to 
 Mt OO ONHlt^HiNNt^TlOt^to IOOO Cl 00 N O 
 
 O -lO-O - 
 
 T^8 
 
 a3Ga3cq3ca3cCa3c43aa3ca3ca3Ga3ca)C 
 a.o a_o a_o a_o a_o a_g aoaoaoao a.o a.g 
 
 ^2 w^2 o^2 oJ2 o^j 0^5 o^j 0^3 0^3 oJ5 o^j o^ o 
 3 C d'C rt'C d'C d'C d'C oJ'C d'C d'C d'C d'C 3'C 
 
 to >o\O t >ooO t<5 O O O 
 
 M N (O TT . 10 
 
 00 ro to to N 
 
 fO fO fO 
 
 
608 PUMPING MACHINERY 
 
 On account of bending to be considered later as well as the 
 effect of critical speed, the diameter will be taken as 
 
 = 2 ns. 
 
 4. Angular Arrangement of Vanes and Areas at Entrance 
 and Discharge. From Newmann's formulae (36), (46), 
 
 cot/?. 
 
 The area of outlet measured on the circumference is 
 
 where n= number of vanes; 
 
 t' = thickness of vanes; 
 b 2 = breadth of opening. 
 
 For a given diameter, the breadth 6 2 is given by the formula, 
 
 , & <?2 x 
 
 2== ~7 T f \ = ~7 T f \ 4~o' 
 
 \ 2 cos a 2 / ' \ 2 cos a 2 / 
 
 If now H is small, 6 2 may be of proper size with a small 
 value of x and a large value of cot ft, but when H is large x 
 should be as large as possible and cot ft as small as possible, 
 so as to make 6 2 of proper value. Now 
 
 w 2 =x\ gKH. 
 
 From this it is seen that it is better to keep x as small as 
 possible consistent with other demands in order to make w? 
 small, because the friction of water against the impeller depends 
 on this term. 
 
CENTRIFUGAL PUMPS 609 
 
 On examining Fig. 413 it is seen in the first place that 
 when /?=9o, x = i for all values of a 2 and also that with values 
 of /? near 90 there is little change in x. Consequently for 
 high heads /? should be made as near 90 as possible. It is 
 true that with a negative value of a 2 x is larger for smaller 
 
 values of /?, but cot /? increases so much that z might not 
 
 be large. When the pump has a low head opposing .flow, x 
 may be smaller and /? smaller, giving cot /? a larger value. For 
 high-pressure stages /? may be made 85, while with low-pressure 
 stages /? may be 40 to 50. The value of 2 is usually negative, 
 30 to 60 being used. By using positive values of a 2 , the 
 speed of the pump w 2 may be decreased, giving what may be 
 called bending back. If 2 is made o, all values of /? give 
 x = i, hence w 2 = \ gKH. For pumps in which there is no 
 diffuser /? is made almost o, so that u may be small, giving p d 
 as large a value as possible, since there is no good method of 
 changing velocity into pressure in a volute pump. This will 
 mean that a 2 has a large negative value, say, 70. 
 
 Using in the problem, which is a high-pressure stage problem, 
 a 2 = 60 and /? = 85, # = 1.073, KH = io^\. feet, the following 
 results : 
 
 w 2 = I.073V 32.16 x 104 =62.0; 
 N = iooo per min. =16.7 per sec.; 
 62 
 
 
 = 0.59 ft. =7 ins. 
 
 From (47) and (48), using Fig. 414, ; =0.0038, 
 
 v tegKH ^0.0038 x 64.32 X 104 _ 
 u r = - = =4-75 
 
 \/ 3 2.i6xio4 ,- 
 
610 PUMPING MACHINERY 
 
 As a check: cos /?= r = 53.9X0.08716 =4.7, 
 
 h Q 2.23x144 
 
 = ~T~ nt' \ = ~~T~ 
 
 4.7(71X14 
 
 assuming six vanes & inch thick at point. 
 
 To find the data for entrance certain assumptions must be 
 
 made. varies from ij to 2j. c r may be made equal or 
 
 nearly equal to r . It is better to have u r <c r , as this means 
 that the velocity at discharge is small. 
 
 Let c r = i. 75 u r = 1.75x4.7 =8.22. 
 
 This is the velocity c since =o. If the water entering 
 the impeller have this same velocity the area between the 
 hub and the shrouding or wall of the impeller must be given 
 by the equation, 
 
 _^_<?_2.23Xi44_ Qs ins 
 f h c g 22 39- sq.ms. 
 
 5. Diameters of Impeller. The hub of the impeller may be 
 taken to be inch thick, giving the diameter, 
 
 A h =g.6 sq.ins. 
 The area of circle at inner end of the shrouding will be 
 
 D. =7.82=7!. 
 
 Using 8 inches for the diameter at the edge of the vane 
 of the impeller gives 
 
 r 2 14 
 
 =-- = 1.75, a workable ratio. 
 
CENTRIFUGAL PUMPS 611 
 
 Now c r = i.75w r = i.75X4.7 =8 - 22 ; 
 
 W2 4 r 
 
 i-Ew-y 6 *-*";; . . Jj . 
 
 ^i =8 ins.; 
 
 cos a i/ 0.23 
 
 1.66 ins. 
 
 6. Alternate Graphical Method. The above dimensions may 
 be found in another manner. From Eq. (30), 
 
 
 V A A 
 
 The ratios , ^, ^, and the quantities i, /?, and 
 
 ^1 ^1 ^d 
 
 are assumed from practice. Let 
 
 and ^ for the impeller and for the diffuser each 0.15. Then 
 
 / 64.32x100 
 
 ~\ 2x1.75x1x0.996x0.966-0.3-^" 
 
 In Fig. 462 ABC is taken' of such a length that AB repre- 
 
612 
 
 PUMPING MACHINERY 
 
 sents ri and AC, r ^ == ~ 
 
 and BD is drawn perpendicular 
 
 to AB, and BE is drawn, making an angle of 75 (&i =75) with 
 AB. Lay off BE equal to 37.7 feet per second. A perpendicular 
 
 J 
 
 FIG. 462. Graphical Design. 
 
 ED from E on BD will fix the velocities 
 
 > =36.4 and 
 
 A line is drawn from A through D cutting a perpendicular 
 to AC from C in the point F. FC is equal to w 2 . From the 
 figure this equals 63.7, FG is laid off at 85 to a line parallel 
 to AC and is made equal to 69 feet per second. This gives 
 CG=c 2 =6.2, 2 = -33- 
 
CENTRIFUGAL PUMPS 613 
 
 From the figure ^,=5.26, c r =c=g.8; 
 63.7x12 
 
 2.23X144 
 b 2 = p 7 TH =1.40 ms.; 
 
 2.23X144 
 
 7. Method for Volute Pumps. In the design the problem 
 has been to find the dimensions of a turbine pump, one in 
 which a diffuser lias been used in which u is reduced to u 3 
 by the enlarging channel. As a result of this, u has been 
 made quite large and c 2 has been small. This is due to the 
 assumption of = 85, or some large value, and a 2 = 60. 
 If now there is not sufficient space to reduce u to u 3 , or if the 
 diffuser is to be eliminated, then u is made small by selecting 
 a large negative value for a 2 and a small value for /?, as shown 
 in Fig. 463. Such a pump is usually employed when the head 
 is low. In this case the velocity u from the impeller should be 
 radial and it is to be changed in the whirlpool chamber to u% 
 by gradually enlarging the width and by the increase in the 
 circumference. When the volute is reached the velocity is 
 v r in a tangential direction. The loss then is 
 
 8. Form of Impeller and of Vane Curves. Using the results 
 of the latter method the following data will apply to the 
 pump: 
 
614 
 
 PUMPING MACHINERY 
 
 FIG. 463. Diagram for Volute Pump and 
 Low-head Turbine Pump. 
 
 FIG. 464. Section of 
 Impeller. 
 
 202=63.8 ft. per sec. 
 w\ =36.4 ft. per sec. 
 c=c r =g.8 ft. per sec. 
 Ci=37.8 ft. per sec. 
 bi =1.27 ins. 
 
 C 2 =6.2. 
 
 ^=60.3. 
 
 ns. 
 ins. 
 
CENTRIFUGAL PUMPS 
 
 615 
 
 The data will be first applied in laying out an impeller 
 which has pure radial action, such as used in Figs. 423, 429, 430. 
 The cross-section will be laid out as shown in Fig. 464. The 
 shaft diameter will be investigated after the weight of the 
 impeller has been determined. The next important step is 
 to lay out the vanes of the impeller and diffuser. There are 
 several methods which may be employed. 
 
 In Fig. 465 the vane is drawn as a parabola. At A the 
 
 FIG. 465. Parabolic Vanes. 
 
 angle i is laid off from the radial direction and then the 
 point B is so chosen that the line making the angle a 2 with 
 the radius will intersect the line from A at C, about a mean 
 radius between r and r 2 . These lines are tangents to the vane 
 curve at entrance and exit and to put in the curve between 
 the points, the lines AC and CB are divided into the same 
 number of parts and lines are drawn connecting similar points, 
 the top point of AC being connected to the top of CB, etc. 
 
616 
 
 PUMPING MACHINERY 
 
 These lines axe tangents to the parabola and hence their 
 envelope will be the desired curve. When DE is drawn at 
 one-sixth the circumference from BA it is seen that there is 
 considerable change in angle from A to , and hence an extra 
 vane GF is put in. Since these vanes converge so rapidly as 
 seen at H there is some danger of interference with the flow 
 of water. Consequently these vanes are usually so drawn that 
 
 FIG. 466. Involute Curves. 
 
 there is a portion of the vane just opposite to the corner of 
 the next vane which is parallel to it. 
 
 Such a result may be obtained by using the involute as 
 the curve at entrance and exit when possible. At A, Fig. 466, 
 the angle a\ is laid off as before and a perpendicular AB to this 
 line is drawn. The circle tangent to this perpendicular will 
 be the base circle of the involute AC. Another involute drawn 
 from D will be parallel to the first involute, because the curves 
 at the points C and D have the same centers of curvature. 
 
CENTRIFUGAL PUMPS 
 
 617 
 
 If the involutes were carried further they would be parallel 
 and at the constant distance CD apart. At the point C the 
 water is moving parallel to that entering at D and consequently 
 there is no tendency for the water at the entering corner to 
 be affected by the* interference of the previous vane. At F 
 and H the water is so moving that it is parallel to water entering 
 the impeller directly below it at / and /, when the water. at 
 
 FIG. 467. Involute Curves. 
 
 / and / is entering at an angle ai to the radius at the point 
 in question. 
 
 When the same method is used at outflow, it is found that 
 the circle tangent to the perpendicular MN, perpendicular to 
 the line at an angle of 2 with the radius, has its point of 
 tangency N within the space left for one vane. This would 
 mean that there would be nothing gained by the use of the 
 involute. Intermediate vanes and P are sometimes intro- 
 duced to keep the outflow in the proper direction and by placing 
 so that it falls between N and M the action of outflow is 
 similar to that at inflow. These vanes are only carried in a 
 
618 
 
 PUMPING MACHINERY 
 
 portion of the distance toward inlet as it is not desired to 
 obstruct the inflow. The involutes at inflow and outflow are 
 connected by a curve as at M , or a tangent as at C. 
 
 When 2 has a larger negative value the involute at out- 
 flow takes the form shown in Fig. 467. At A the two involutes 
 are so far apart that a reverse curve has to be used in joining 
 them, while at B a curve of the same form and curvature could 
 be used. At C the involute at outlet is only used at the tip 
 
 FIG. 468. Circular Arc Vanes. 
 
 while a tangent is used between the parts. The form at B is 
 quite common. 
 
 In Fig. 468 positive values of 2 are shown as well as the 
 method of using circular arcs. If perpendiculars to the lines 
 of flow (at angle a\ to radius) are erected at A and B these 
 intersect at C. The point C may be used as a center for the 
 arc from B to D. This curve at D is parallel to the curve A 
 at the point A and hence there is no tendency to interfere 
 with the flow. The perpendiculars at E and F at outflow 
 intersect at G, which is the center for the arc at outflow. The 
 centers at outflow and inflow lie on circles from the center 
 
CENTRIFUGAL PUMPS 619 
 
 of the wheel and these are dotted in. After finding the centers 
 C and G and the radii CA and GE, these other curves are 
 quickly drawn in by using the dotted circles on which the 
 centers must lie. 
 
 One set of vanes shows how the circular arcs are joined by 
 tangents while another shows the use of a curve. At K a 
 parabola is constructed showing the method of using that 
 curve. 
 
 If the line HI is drawn in the direction of flow from' / it 
 appears as if there might be a dead space EHI where no water 
 would flow, but this would be filled with eddies causing loss. 
 If the change in form is too sudden this may happen, but it is 
 to be remembered that this water is under pressure and it 
 would consequently gradually enlarge if the curvature is not 
 too great. 
 
 The curves drawn in the preceding figures are the center 
 lines of the vanes and the half thickness is added to each side 
 of the center line, the end being drawn to a sharp point. 
 
 MIXED FLOW PUMPS 
 
 The figures shown are applicable to pumps in which there 
 is a pure radial action. If it is expedient to have some axial 
 action at the center on account of a desire to keep the outer 
 diameter small, the. vane is carried into the center, as shown 
 in Fig. 469. In this pump the peripheral speed at inlet changes 
 for the various parts of the vane. At a the speed is w\ a at b, 
 Wi b \ at c, wi c , etc. The velocity c should be the same at all 
 points and in a direction normal to the various peripheral 
 velocities. To have this result a\ must vary over the inlet 
 edge of the vane. 
 
 Before carrying out the steps used in design it will be well 
 to consider the action at entrance. Suppose the number of 
 vanes is assumed and the thickness of metal used for them. 
 If in Fig. 469 the inlet edge be divided into four parts by the 
 points i, 2, 3, 4, 5, and the middle points a, b, c, and d be 
 marked, the lines of flow aa" ', bb" ', cc" , dd" may be approx- 
 imated. These stream lines or lines of flow may be considered 
 
620 
 
 PUMPING MACHINERY 
 
 d"c"b"a" 
 
 FIG. 469. Section of Mixed Flow Impeller 
 
 FIG. 470. Velocity Diagram for Various Inflow Points 
 
CEN TRIP VGA L PUMPS 
 
 621 
 
 to be on conical surfaces at the points of entrance, and if tangents 
 are drawn from the points a, b, c, d to the axis of rotation, 
 these lines are the elements of the various conical surfaces. 
 
 The peripheral speed at the points a, b, c, d may be found 
 as shown in Fig. 470 where oa, ob, oc, etc., represent the radial 
 distances to the various points. wi a , w lb , w ic , etc., will repre- 
 sent the various peripheral velocities. These are known from 
 
 w 2 , since Wi x =w 2 -. 
 From Eqs. (3) and (4), 
 
 u< 
 
 C 2 2 
 
 c 2 
 
 ___L?_ Z ^L 2 _Li! ' W__ r ~_ 
 
 2 g 2 g 2g -- <--'*-_. 
 
 FIG. 471. Angular Relations on Developed Conical Surfaces. 
 
 Now in the pump considered p d p b is the pressure difference 
 between that in the discharge space and that in the suction 
 space. This is constant for all of the suction chamber and 
 hence it is constant for all points of entrance. w 2 and c 2 are 
 constants for all divisions of the pump and the losses may be 
 considered as constant. Hence 
 
 Ci 2 Wi r 2 = constant. 
 
 If c is to act at right angles to w lx the expression above 
 is equal to c. c is therefore a constant and the various values 
 of c\ and i at a, b, c and d are found as shown in Fig. 470 
 by making the vertical distances all equal to c. 
 
 Fig. 471 is now constructed by developing the various 
 conical surfaces and drawing on these developments the 
 various vanes. The construction for point c is given as illus- 
 
622 
 
 PUMPING MACHINERY 
 
 trating the method for all of the points. From the center o, 
 Fig. 471, oc is laid off equal to cc' of Fig. 469, and with this 
 as a center the arc cc'" is drawn. The arc cc iv is now drawn 
 of radius cc v from Fig. 469 and c iv is determined by the number 
 of vanes used in the circumference (twelve assumed). If 
 this curve is then rectified and placed on the arc of the conical 
 surface, the pitch distance at the point c (Fig. 471) is found. 
 The angles i c from Fig. 470 are now laid off from the radial 
 lines at c and c"' and then the involute cc vl is drawn in and 
 from this the perpendicular distance between the vanes cf"c vl 
 
 is found. This distance less the 
 thickness of the vanes is the net depth 
 of the passage or c'"c vl t=d c . 
 
 This is done at each point and 
 then the lengths of the outflow edge 
 are rectified as in Fig. 472, giving 
 the line a, b, c, d, as the length of 
 the outflow edge. If now the depths 
 found at the various points be laid 
 off perpendicular to the above line 
 at their respective points a figure 
 aa'd'd is found which represents 
 practically the outflow area. This 
 area is not quite equal to the outflow 
 area as this is a warped surface which 
 cannot be developed. Now the quan- 
 tity of water flowing through any 
 element of length is 
 
 FIG. 472. Curves of Area and 
 Quantity. 
 
 or 
 
 This means that Q is represented by an area in which the 
 ordinates are Cid c and the abscissae are /. Hence if the depths 
 d are multiplied by the corresponding velocities GI from Fig. 
 470 and the product is laid off at the various points of the edge 
 
CENTRIFUGAL PUMPS 623 
 
 in Fig. 472, the figure add" a" is found, the area of wnich repre- 
 sents the quantity of water flowing. 
 
 Now in the formula (when inflow is at o), 
 
 Work = uw 2 sin /?. 
 
 The work per pound of water is the same at all parts of the 
 impeller. Hence when this is put into the form 
 
 Work =Qwu 2 -. sin /? sin a\, 
 
 YI A i 
 
 where ri, ^4i, and i change for different parts of the impeller, 
 the quantities r\, A\ and a\ should refer to a point at the cen- 
 ter of gravity of the outflow area. To show this suppose that 
 the area of discharge is divided into a series of elements JA ex , 
 such that the quantity discharged is the same in each. Then 
 =c\ x AA\ x =K and 
 
 Work = 2wci x AA\ x u 2 . sin /? sin ai x . 
 
 r\ 
 
 Now Ci approximately varies as r, as may be seen from 
 Figs. 470 and 472 and sin ai is practically constant. 
 
 K " ^KI^- 
 
 Now ZrAA lx is the static moment of the inflow area and 
 for that reason it equals r c . g Ai, or the work is that required 
 by the water if it all entered at the radius of the center of 
 gravity of the actual area. 
 
 To find the center of gravity of aa'd'd lay off the curve 
 d lv c lv b lv a lv , found by multiplying each ordinate aa', bb', etc., 
 by the distance from the left-hand end of the figure and using 
 these as ordinates for the new curve, 
 
 Xae, 
 
 The area of this curve is the static moment of the original 
 
524. PUMPING MACHINERY 
 
 area about the left corner. This area divided by the area of 
 the original curve will give the distance of the center of gravity 
 from the left corner, 
 
 ed lv c lv b lv a lv 
 C ' g ' = dd'a'a ~ 
 
 J 
 
 fd c 
 
 xdl 
 
 In designing a mixed flow impeller the method of procedure 
 is as follows: Assume for a given design the radius at the 
 center of gravity, as r Q , Fig. 469, and with u 2 , w 2 , c, and a iy 
 known for the point r , the constructions of Figs. 462 or 463 
 will give the distance d Q and the velocity c io . The length of 
 outflow edge is now found approximately by the formula, 
 
 l= _Q_ 
 
 This is then used in Fig. 469 and a curve drawn equal to 
 this in length. After the curve is assumed the various points 
 are used as shown in Figs. 469, 470, 471, and 472, and the 
 actual area and center of gravity is found. If now the center 
 of gravity is slightly different from that assumed, or if the area 
 
 of the quantity curve is greater than , the length or shape 
 
 of the outflow edge is altered to give the correct values. 
 Several trials will give the desired e.g. and quantity. 
 
 Since there is a slight change in the angle <*i and moreover 
 since the paths are of different projected lengths the surface of 
 the vane is rather complex. The inlet edge is often contained 
 in a radial plane so that the curve, Fig. 469, is seen in its true 
 length. At other times the curve is in a plane which does not 
 pass through the axis and sometimes the inlet edge is a non- 
 planar curve. In the second case the projection of the curve 
 on a plane perpendicular to the axis would give a straight 
 line passing one side of the axis while in the last case the 
 projection would be a curve. 
 
 Since the curves of the vanes at inlet have angles ai differ- 
 ing slightly from each other, and since the paths to the outlet 
 
CENTRIFUGAL PUMPS 625 
 
 are of varying length the outlet line may not be parallel to 
 the elements of the outlet cylinder, although by properly select- 
 ing the shape of the paths this may be accomplished. 
 
 Fig. 473 shows a first set of curves for a wheel. Fig. 469 
 and the cross-sectioned part of Fig. 473 are revolved pro- 
 jections of the vanes as if the complete vane was in a radial 
 plane. The stream lines are not in the position shown. To 
 obtain their correct position the following method is used: 
 If the quantity curve of Fig. 472 be divided into equal parts, 
 say four, by the lines A, B, and C, these lines determine the 
 positions along the outflow edge at which divisions could be 
 placed in the passage, so that each part would carry one-fourth 
 of the quantity passing through the impeller. At the outlet 
 edge the divisions will be spaced equally because c 2 and 2 are 
 the same at all points. If now the constructions of Figs. 470 
 and 471 be made for each of the points a, b, c, d, and e of Fig. 
 473, the true shape in projection of these lines may be found. 
 The points at the half pitch across the bucket have been used 
 to get better results. The points will all lie on a radius 15 
 from OR in the end view of the figure. Consider the part marked 
 C of Fig. 471. The point of the vane on a radial plane 15 
 from the entering edge of the involute vane will be at a dis- 
 tance ab from involute to circle at the edge. This point is at 
 a distance ab=cc f from point c on the element of the cone, 
 but when this cone is wrapped into position the true position 
 of the point will be on an element 15 from the original 
 point in the end view. Hence if c' is projected over to c" 
 in the end view and this is swung to the 15 radius, the 
 point c'" is determined. This is carried over to the side 
 view and c lv is then determined by the intersection of this 
 projecting line and a perpendicular to the axis from c' '. In 
 the same manner cc^ is made equal to c'"c vl and Ci lv is found 
 from the point c\" in the end view, which is determined by 
 swinging c\" over to the 30 radius. The intersection of the 
 two projecting lines gives c\ lv . The same operation is used for 
 the other points. From the first two points of these various 
 lines in the end view, the remainders of these lines are sketched 
 
62G 
 
 P UMPIXG MA CH1NER Y 
 
CENTRIFUGAL PUMPS 627 
 
 in for this view and these are ail brought to the same element 
 of the discharging cylinder and 2 is made the same for each. 
 
 To construct the true shape of the various stream lines in 
 the first view from the curves just sketched in, the following 
 method is used: such a point as c 4 is revolved to c 5 , carried 
 over until it strikes the stream line from c at c&, this is pro- 
 jected down until it cuts the projection line from c 4 , giving 
 the true point C?. This is done for all stream lines, and if the 
 shape of any is too complex or. has too much reverse curvature, 
 a new set of lines is assumed. After these two projections are 
 made, sections of the vane form should be made at various 
 angles for the purpose of seeing that the vane will have_-the 
 proper shape in all directions. Such an operation, similar to 
 fairing ship curves, is important. The curves seen at the right 
 of Fig. 473 are the radial plane intersections. Those shown as 
 heavy dotted lines in the end view are those on planes I, II, 
 III, IV, V, and VI perpendicular to the axis. These latter 
 curves are those used in constructing the core boxes from 
 which the moulds for casting are made. In the above there 
 has been no thickness of vanes shown. After finding the 
 curves for one side of the vane a similar operation could be used 
 for the other face. 
 
 AREA CURVE THROUGH BUCKET 
 
 After the vanes are determined and drawn in, it is well 
 to find the area at various points along the middle line of the 
 bucket (center of gravity of various areas) and if the lengths 
 between various centers be laid off as a base line, Fig. 474, 
 and the areas at these points are used as ordinates, the area of 
 this figure, by the theorem of Pappus, is the volume of the vane 
 bucket. 
 
 The curve so constructed should gradually increase or 
 decrease. A curve with considerable change, as shown by the 
 dotted line, is objectionable, as this means frequent changes 
 
 in velocity \ v x = ~T~) an d changes in velocity are usually accom- 
 
P UMPING MA CHINER Y 
 
 panied by loss. If such a curve as the dotted one is found, 
 the axial width should be changed to bring it into the condi- 
 tions shown by the solid line or the vanes are thickened up 
 
 FIG. 474. Curve of Passage Area. 
 
 by the use of back vanes, as shown in Fig. 475. These are 
 necessary in the figure shown to keep the velocity more nearly 
 constant, as the distance between the vanes is too great at the 
 center. 
 
 FIG. 475. Back Vanes. 
 
 ABSOLUTE PATH OF WATER 
 
 The area of the curve, Fig. 474, between the inlet point and 
 any other point is the volume of the bucket to that. point and 
 since AiCi is the quantity of water entering the bucket per 
 second, this volume divided by A^c^ is the time taken for a 
 particle of water to move from the entering edge to the point 
 considered, 
 
 *~ 
 
 vol. 
 
 x area x 
 
CENTRIFUGAL PUMPS 
 
 629 
 
 To find the absolute path of a particle of water passing over 
 the vane av of the Fig. 476, the times taken to pass to the points 
 v, x, y, z, and a are found, as shown, from a figure similar to 
 Fig. 474. The points v y x> etc., are then carried in radially 
 to v', x', y', etc., and from these points the distances v'vi, 
 
 FIG. 476. Absolute Path of Water. 
 
 #iV, yfyi'i - j equal respectively to w^, w\t x , w\t y , etc., 
 are laid off. These points determine the positions of the radii 
 when the water reaches the point in question. Hence, if from 
 x an arc is drawn intersecting the radial line from x\ this 
 determines the point Xi of the absolute path. In this manner 
 #i> ^i> yi> #ij an d Vi are found, giving the absolute path of the 
 water through the impeller. This path should be a smooth 
 
630 PUMPING MACHINERY 
 
 curve without sudden changes 'of curvature. The tangent to 
 it represents the absolute direction of the water at any instant, 
 hence at outflow the angle formed with the radius is /?, while 
 at entrance the angle is a. 
 
 9. Diffusion Chamber. Having the impeller designed the 
 next step is to design the diffusion chamber. The angles of the 
 vanes of the diffuser are /? at entrance and a d at discharge. 
 Since the water is to discharge into the volute cas'ng and travel 
 in that chamber in a tangential direction it is advisable to 
 make ad as large as possible. The velocity of discharge v^ 
 is given by the equation, 
 
 O 
 
 -^' 
 
 cos a d \ 
 
 cos * 
 
 When ad is made large Dd must be large and td must be 
 made very large if v d is made small. This cannot in general 
 be done, and in many cases ad is made small, being o in some 
 cases. This gives 
 
 Q 
 
 Dd may be so large in this case that if td is increased 
 over the t at exit from the impeller, the net area, even after 
 the enlargement of t r to give a better flow, to care for the 
 supporting bolts or to make a passage as is done in the Worth- 
 ington pump, is so large that v d is quite small. 
 
 In Fig. 477 six vanes of the diffuser have been drawn as cir- 
 cular arcs with a d and /? as the angles at the two ends. On 
 one of the center lines a plain vane has been drawn with its 
 sharpened end, on a second one the vane has been constructed 
 to make a channel for a supply to the next stage for balancing, 
 while the third shows the enlargement of the vane to make a 
 more gradual change in area and to form a place for the sup- 
 porting bolt. The center line may be drawn if desired as an 
 involute on the small base circle shown dotted. 
 
 Although the discharge in the last two cases is at right 
 angles to the direction desired in the volute casing, the velocity 
 
CENTRIFUGAL PUMPS 
 
 631 
 
 has been reduced to such a small value that even if this whole 
 velocity head were lost the amount would be sma 1 !. More- 
 over, the head lift for each impeller of the pumps using diffusers 
 is so great that the percentage loss due to this radial outflow 
 is very small. For this reason the volute casings in these pumps 
 are not made as volutes but are concentric circular paths 
 
 FIG. 477. Diffuser. 
 
 uniting at the top of the pump as shown in Figs. 437, 438, and 
 441. 
 
 In the problem considered assume Z) d = 
 
 2.23x144 
 
 ; =5.3 ft. per sec. 
 
 1.40(71X22-6X4) 
 
 If bd is increased to ij&2 or 2.00" the velocity is changed to 
 v d=3-5 ft. per sec. 
 
632 PUMPING MACHINERY 
 
 If this velocity is entirely lost in impact the loss is 
 
 L d = V -f g =o. 3 i ft., 
 
 or less than J of i% of the head per stage. 
 
 10. Volute Casing. The volute casing of ordinary volute 
 pumps is designed so that the velocity of the water is the 
 
 FIG. 478. Volute Casing. 
 
 same at all portions. The quantity coming off from unit 
 length of whirlpool chamber is 
 
 The quantity passing at any point at the angle 6 from A, 
 Fig. 478, is 
 
CENTRIFUGAL PUMPS 
 
 633 
 
 If the velocity of this water is assumed as v v , the area at 
 the point is 
 
 A = = (if of circular section). 
 
 If A is rectangular and of constant width w in direction 
 of the axis, the dimension in the direction of the radius is 
 
 V v 2nw 
 
 KO. 
 
 /->.! ftftt I 
 
 2nv v WA 
 
 Neumann points out that the first assumption of a circular 
 section gives the limiting curve of Fig. 478 the form of a para- 
 bolic spiral, while in the second case the limiting curve is an 
 involute. 
 
 ii. Shaft Design. The weight of the impellers should be 
 computed from the drawing made, as in Fig. 464, and from 
 previous experience the length of the shaft to care for these 
 impellers is known. A diagram, such as shown in Fig/479, is 
 made, giving weights and positions between supports. 
 
 Bearing 
 
 1 
 
 135 Ibs. 
 D 
 
 135 Ibs. 
 
 135 Ibs. 
 B 
 
 Bearing Coup 
 
 i 
 
 
 
 o" > 
 
 tf 
 
 
 >! i 
 
 
 
 
 ! 
 
 FIG. 479. Load Diagram. 
 
 Considering the beam as a simple beam, the bending moment 
 at the point A is o and the twisting moment T is 
 
 100X33,000X12 . 
 
 T= =6300 in. -Ibs., 
 
 1000 X27T 
 
 from B to C, 
 
 The reactions due to the loads of 135 pounds at each 
 impeller will be 229 pounds at the left and 175^ pounds at 
 
G34 PUMPIXG MACHINERY 
 
 the right. The shear diagram will pass through zero at the 
 second load, hence the moment will be a maximum at this 
 point. 
 
 M = 229^X13 -135X9 -=1768 in.-lbs. 
 
 Under the last impeller the moment is 
 
 M = 175!- X8 = 1404 in.-^bs. 
 The combined moment is 
 
 At B this is 
 
 T c = \ 14042 + ( 6300 ) 2 = 6400 ; 
 at C, 
 
 + ( 4200)2= 4550. 
 For the shaft diameter, 
 
 C 7T^ 3 
 
 6400=^5^ 
 
 As will be seen, the critical speed for such a combination of 
 discs on a shaft demands a larger diameter, and for that reason 
 a diameter of 2 has been assumed before the investigation for 
 critical speed. 
 
 The thrust bearing will be investigated at this point. The 
 thrust is caused by the impact and pressure on the inlet area 
 of the pump and the unbalanced pressure on the shrouding or 
 sides of the impellers. Using the dimensions from Fig. 464, 
 
 Diameter at hub =3J ins. 
 
 " inner clearance ring =8^ ins. 
 
 " " back = 7 ins. 
 
 44 discharge =14.6 ins. 
 
 " outer edge of entrance --=7^ ins. 
 
CENTRIFUGAL PUMPS 635 
 
 5^2=63.8 ft. per sec. 
 
 ^1=36.4 ft. per sec. 
 
 c 2 = 6.2 ft. per sec. 
 
 Ci =37.8 ft. per sec. 
 24 = 604 ft- P er sec - 
 = 9.80 ft. per sec. 
 
 2 ]=62.5 ft- head. 
 
 If the water leaks on each side to the center, as seen in 
 Fig. 419, this pressure difference acts on each side and con- 
 sequently there is no unbalanced pressure from this source. 
 If there was leakage only on one side, say, the back of the 
 impeller, then it would be assumed that the pressure in space 
 would be a portion of the 62. 5. feet, as the water leaks out at 
 center into the suction. Suppose this be assumed to be two- 
 thirds of 62.5, or 40 feet approximately. 
 
 The pressure to the right is then 
 
 40x62.5!" .I4.6 2 7 2 ] 
 p = x -- U-- -TT-- =2180 Ibs. 
 144 L 4 4 J 
 
 The force from the impact of the water to the left is 
 
 . 
 
 Wv wAc 2 J 4 V/8 
 = - = - - = 50 Ibs. 
 g g 144X32.2 
 
 If there are openings at the center of the impeller to relieve 
 the pressure and if leakage occurs on each side, the impact is 
 all that has to be cared for. 
 
 The total pressure to be carried on the thrust bearing in this 
 pump is . . 
 
 XP=3X5o = 150 Ibs. 
 This would require the area, 
 P 150 
 
636 PUMPING MACHINERY 
 
 If there are four collars on the 2-inch shaft the approximate 
 height of the collars to give the proper area will be 
 
 ;. A _J i;. 
 
 *D 4X*X3 
 
 The collars will be made ^ inch high for easier machine 
 
 work. 
 
 CRITICAL SPEED 
 
 If a shaft which is deflected slightly is caused to revolve, 
 it is subject to centrifugal force due to the weights turning at 
 the angular speed a>. When the speed is low the shaft has a 
 chance to bend to the elastic curve due to the weights of the 
 body and the various parts will rotate about their figure centers, 
 or geometrical centers, which are about the same. When, how- 
 ever, the speed is increased this bending cannot occur so rapidly 
 and the shaft, pulleys and weights may be assumed to rotate 
 around the axis of the bearings. Following Reynolds' method 
 as given by Stanley Dunkerley in his excellent paper published 
 in the Philosophical Transactions for 1894, Part A, assume 
 the shaft and disc as shown in Fig. 480. The load per element 
 
 FIG. 480. Whirling Shaft and Pulley. 
 
 of length of the shaft when turning with the angular velocity 
 w at which the weight effect, but not the mass effect, can be 
 
 neglected is 
 
 Awdx 
 Load = (o 2 y = Lax, 
 
 o 
 
 where A =area of shaft in sq.ft. 
 
 w= weight of i cu.ft. 
 g= acceleration of gravity. 
 y= deflect ion. 
 
 Now Shear =^Ldx = V, 
 
 where L = load per foot and V = shear. 
 
CENTRIFUGAL PUMPS 637 
 
 Moment = ( Vdx =M, but M =EI-~ where E equals modulus 
 of elasticity in pounds per sq.ft. and / is the moment of inertia 
 of the cross-section in feet 4 . Hence V = 
 
 , . 
 
 dx 3 dx* 
 
 The equation of the elastic curve then, when the speed is such 
 that the deflection is caused by the centrifugal force and not 
 by the weight, is 
 
 Aw , d 4 y 
 
 co 2 y=EI^ L . 
 g dx 4 
 
 This occurs when the effect of weight is eliminated or the 
 shaft is moving at such a speed that whirling occurs. 
 The equation may be written as 
 
 where m= 
 
 . 
 gEI 
 
 This integrates into 
 
 y =A'e mx +B'e~ mx + C'e mix +D'e~ mix , 
 or y =A cosh mx +B sinh mx + C cos mx +D sin mx. 
 
 A, B, C, D are the constants of integration. 
 
 To eliminate these constants there are several known con- 
 ditions. At bearings, x = o or /, and y = o; for fixed bearings 
 
 j ""P, at a change of loading x, y, and -=- for the equations 
 
 on one side of the load are equal to the respective equations 
 which hold on the other side of this point. The equation 
 
 above has been derived from the condition that L= o> 2 y 
 
 6 
 
 for the curve and this only holds between concentrated loads. 
 At each concentrated load or bearing for a shaft with several 
 bearings there is a new equation. At the concentrated loads 
 
638 
 
 PUMPING MACHINERY 
 
 When there is a disc on the wheel which is deflected by 
 the bending of the shaft into the position shown in Fig. 481 
 the centrifugal action of the disc tends to right the disc and 
 straighten the shaft. 
 
 For an element dm of the disc at distance r from the center, 
 of which the components are v, t and 6, the component of the 
 
 dm 
 centrifugal force tending to right the disc is vco 2 and its arm 
 
 o 
 
 is t. The moment of this is tvco 2 or -^-v 2 a? , 
 
 g dx 8 
 
 t dy 
 since =-T- 
 
 . v dx 
 
 FIG. 481. Rotating Disc. 
 
 The total righting moment is 
 
 j 
 
 g d 
 
 g dx 
 
 y f 
 
 I v 2 dm. 
 xj 
 
 Now 
 hence 
 
 // is very small, 
 
 Ip = l'd (about diameter). 
 
CENTRIFUGAL PUMPS 639 
 
 hence M, the moment due to centrifugal force tending to right 
 the disc, is 
 
 M = !'. 
 g ax 
 
 In passing a disc the difference between the moments on the 
 two sides is 
 
 This is another condition in passing a load which elim- 
 inates another constant. These conditions will furnish suffi- 
 cient equations to eliminate the constants. 
 
 Dunkerley in his extensive article takes up the different 
 cases arising in practice and determines the critical speed. 
 The student is referred to the article for the methods of solving 
 the equations, but the simple case of a plain shaft is given 
 here to show how these are worked out. 
 
 For a simple shaft, the equation for y is 
 
 y =A cosh mx +B sinh mx + C cos mx +D sin mx. 
 
 Now when 
 
 * = oor /, y=o, 
 and when 
 
 d 2 y 
 x = o or 7, M=o, i.e., ~j~^> = - 
 
 -j- =mA sinh mx+mB cosh mx -mC sin mx +mD cos mx. 
 dx 
 
 d 2 v 
 
 - =m 2 A cosh mx-\-m 2 B sinh mx m 2 C cos mx m 2 D sin mx. 
 
 dx 2 
 
 0=A+C. 
 
 0=A cosh ml+B sinh ml + C cos ml+D sin ml. 
 0=A-C. 
 
 0=A cosh ml+B sinh ml - Ccos mlD sin ml. 
 :. A =o, C=o. 
 
 B sinh ml+D sin ml = o. 
 
 B sinh mlD sin ml=o. 
 
 Subtracting Dsmml=c. 
 
 D=o or ml 
 
640 PUMPING MACHINERY 
 
 Adding 
 
 B sinh ml=o. 
 B=o or ml = o. 
 
 The condition which is possible is 
 
 ml = n, 
 
 7 /4 = ?r4 ' 
 
 W == ~jZ t '\ I j == 
 
 I 2 \ Aw 30 
 
 In the case of a disc on a shaft as shown in Fig. 481 the 
 following results if the mass of the shaft is not considered, 
 although the resistance to deflection is taken into account : 
 
 =, 
 
 dx 4 
 
 EIy~x3+-- 
 
 O 2 
 
 x=o, y=o. 
 
 d 2 y 
 
 __ 
 dx = dx~ / ' 
 
 tf~dx~ El g dx 
 /=o. 
 
 v' 
 
 =M =o. 
 
CENTRIFUGAL PUMPS 641 
 
 These give the following equations in order 
 
 Ac 3 A'c* B'c 2 
 
 + - +c'c+D'. 
 62 
 
 is the square of the radius of gyration -. 
 
 0=A'l+B'. 
 
 Eliminating the constants A, B, C, D, A', B', C', D', from 
 these equations Dunkerley obtains the following equation, 
 calling lc=c'\ 
 
 Ji[_3L T l 3/T _ 
 acc'U^c' 2 acc'L 
 
 3/ 
 
 3/ 
 
 but / 
 
 K *= ^t i - a y 
 
 tx _-0-/o / -/ 
 
 Now ^T 2 must be a positive quantity, since it equals 
 
 Hence 
 
 ac 2 c' 2 
 K-T- 
 
642 PUMPING MACHINERY 
 
 w 
 
 Now a=pj.aj 2 and these inequalities give the limiting 
 
 speeds to for the centrifugal force to act to produce whirling. 
 These are only the limiting values beyond which there is 
 certainly no whirling or beyond which it is known to be with- 
 out the range of the critical speed. To get the critical speed 
 the equation for K 2 is solved for a and then a is expressed in 
 terms of w 2 , giving 
 
 4 2 b 
 i-6) 3 J' 
 
 In this fl=^=ratio of the distance from pulley to nearer 
 
 bearing to the rectangular radius of gyra- 
 tion. 
 
 6=y=ratio of distance from pulley to bearing to 
 
 the span (less than J). 
 Calling | of the bracket above 6 2 , Dunkerley gives 
 
 We 3 ' 
 
 In this equation / is the moment of inertia of the shaft 
 section, W is the weight of the disc or pulley, c is the distance 
 from the nearer bearing to the disc, 6 is a factor which depends 
 
 on the ratio y and on the ratio -^ where K is the rectangular 
 
 l> x\. 
 
 radius of gyration. 
 
 Dunkerley then computes a table for 6 for various values 
 of a and b which has been reproduced in the form of a series 
 of curves shown in Fig. 482. 
 
 Dunkerley considers the case of a shaft with three supports 
 and also the case of a span with an overhanging end of length 
 c. If the ratio of the short span to the longer span is 
 
CENTRIFUGAL PUMPS 
 
 643 
 
 oii\-j=a] and if the same symbol a be used for the ratio y, 
 
 for the second case, he gives the specific speed for each in the 
 form 
 
 l=K. 
 
 In these two cases the values of K depend on OL and 
 Dunkerley has computed the values of K for different values 
 of a. These have been plotted in Fig. 482 so that CD may be 
 found and from it N. 
 
 0.25 
 
 0.50 
 
 0.75 
 
 1.00 
 
 1.50 
 
 1.75 
 
 200 
 
 FIG. 482. Dunkerley's Values for Shafts. 
 
 For the case of a disc on the projecting end of length c, 
 the formula 
 
 
 holds for the disc independent of the shaft. 
 
 The value of 0, as in the case of the disc between supports 
 
 c c 
 
 depends on the value of a=^ and b=-j. Dunkerley's values 
 
644 PUMPING MACHINERY 
 
 of 6 for four values of j =b have been plotted for different 
 
 values of a =-^. The value of may be taken from the curves 
 A 
 
 from which aj and N may be determined. 
 
 In the case of the double-span shaft it is seen that the 
 value of K does not vary much from TT, showing that the main 
 effect of using several bearings is to shorten the span only. 
 This shortening has considerable effect, although the continuous 
 beam action is not important. If there must be an overhanging 
 end of a shaft of fixed total length it may be shown that when 
 
 -= or the bearing is the total length from the overhanging 
 If 
 
 end the critical speed will be the greatest. 
 
 With the curves of Fig. 482 the critical speed can be figured 
 for each disc, wheel or pulley on a shaft independently of 
 each other and of the shaft. When these are computed the 
 resultant speed is such that 
 
 JL _* L. , JL , 
 
 Nr 2 Ni*^Nf-N#^ 
 
 where N r is the resultant speed and NI, N 2 , N 3 , etc., are the 
 critical speeds due to any one part independent of the other 
 parts. Hence 
 
 N,N 2 N 3 N 4 . . . 
 
 ...N n 2 +N 1 2 N 2 3 ...N n 2 + ...' 
 
 or for the case in hand of three impellers and a shaft, suppose 
 NI is the critical speed for the shaft, and N 2 , N 3 , and N 
 those for the impellers. The critical speed of the combina- 
 tion will be 
 
 N = 
 
 If these happened to be identical, the following would 
 result: 
 
CENTRIFUGAL PUMPS 
 
 645 
 
 In the design of the shaft the stuffing boxes are con- 
 sidered to be the same as bearings on account of the closeness 
 of fit. This gives for investigation of critical speed a shaft 
 of three spans with an overhung disc. The shaft will be con- 
 sidered alone as a shaft with a 1 2-inch span and a 3o-inch 
 span, as no constants are given for shaft of three spans such 
 as is shown. This gives the lowest speed of any combina- 
 tion for the shaft. All dimensions are in feet. 
 
 7T(2) 
 
 E = 29,000,000 X 144. 
 12 
 
 4-^ft*. 
 
 I2 4 
 
 30 
 
 0.4. 
 
 _30#2 fJEI 
 1 *~ Til 2 \Aw 
 
 0.785 
 32 X 29,000,000 X 144 X 7^4 
 
 2 2 X46o 
 
 4X144 
 
 For the impeller discs the weights are taken as 90 pounds 
 
 D 
 
 each and the radius of gyration =. The separate discs give 
 
 the following: 
 
 Disci, 6 1= =0.13, 0! =-- = !. i, #1=1.05. 
 
 Disc 2, 62 = ^=0.43, 
 
 3.65 
 
 = 3.56, 02=2.1. 
 
 Disc 3, 63 = =0.26, 03= -=2.2, 3 =o.o. 
 3 3-65 
 
 30X1.05 /32 X 29,000,000 X 0.785 _ 
 Wc s -^- ~ /.A a =12,400. 
 
 -J xx^s 
 
646 PUMPING MACHINERY 
 
 3 
 
 =458a 
 
 3760. , . . . 
 
 For the disc on the end # = - = 1.5, b= =J, # = --=4, 
 2.o, c =6, W = i5lbs. 
 
 12 
 
 2.0 //4\ 3 /9\ 
 ,400 --- A/ z~ = 3 1 >3oo 
 
 i.o5\\6/ \i5/ 
 
 The resultant of these various speeds is given by the 
 formula 
 
 The speed N 3 is the factor which has the greatest control 
 of the resultant. These terms depend on \ / for their values, 
 so that if the diameter of the shaft was reduced to ij, the 
 approximate value of N would be 
 
 ~T 
 
 I680, 
 
 a value so close to the actual speed that there might be con- 
 siderable vibration and shock. This makes clear the reason 
 for the increase in the diameter. 
 
 CENTRIFUGAL PUMPS FOR SPECIAL PURPOSES 
 
 The centrifugal pump has been used for many years for 
 the clearing of dry docks. Fig. 483 illustrates the equip- 
 ment for one of the docks at the League Island Navy Yard. 
 The units are 45-inch volute pumps built by Worthington; 
 each is driven by a 450-!!. P. motor and will handle an average 
 quantity of 50,000 gallons per minute against heads varying 
 
CENTRIFUGAL PUMPS 
 
 647 
 
 from i foot to 33 feet. As was pointed out with the test 
 curves, the quantity decreases as the head increases, the speed 
 remaining constant. The power usually reaches a maximum 
 at an intermediate head, so that there is no danger of over- 
 loading the motor in any case. It is this feature which makes 
 the centrifugal pump of value for this service. 
 
 The quantity of 50,000 gallons per minute, which, is the 
 usual way of rating centrifugal pumps, would mean 72,000,000 
 gallons per twenty-four hours, the method used in rating 
 water- works pumps. 
 
 FIG. 483. Dry Dock Units. (Worthing ton.) 
 
 These pumps for many years have been used extensively 
 for this work and for draining the low lands of Holland and the 
 marshes of Italy. Fig. 484 shows one of the stations built by 
 Gwynne in 1876. There are eight pumps placed in pairs 
 driven by compound engines. The boiler houses are located 
 at the ends of the pump room. This arrangement shows a 
 good plan and one well thought out. The pumps were of double- 
 flow volute type handling 57,000 gallons per minute under a 
 head of 7} feet. The impellers were 60 inches in diameter and 
 two of them were driven by a 27f-inch and 46! -inch by 27-inch 
 engine. The plant as shown in Engineering, was built for the 
 Ferrara Marshes in northern Italy. 
 
 The pumps of that day gave efficiencies of 50 to 70. per cent 
 
648 
 
 PUMPING MACHINERY 
 
CENTRIFUGAL PUMPS 
 
 649 
 
 and were very reliable. The higher efficiencies obtained to-day 
 are due to the better method of design. 
 
 One of the largest centrifugal pumps built in 1884 by Simp- 
 son & Co. for the London docks of the East and West India 
 Companies, handled 46,400 gallons per minute. 
 
 The pumps shown in Fig. 485 are remarkable in that they 
 handle 35,000 gallons per minute under the high head of 160 
 feet, requiring a 2ooo-H.P. motor to drive each of them. These 
 pumps were of the turbine type on account of the high head 
 
 FIG. 485. 36-Inch Turbine Pump of Worthington. 
 
 and were used at the St. Louis exposition to supply water 
 to the Grand Cascade. The volute casing of concentric form 
 with an outlet at right angles is common with this type of 
 pump. The figures show how part of the diffuser is cast with 
 the volute casing on one side while the head containing the 
 suction flange and elbow forms the other part of the diffuser. 
 Fig. 486 illustrates a casing of an R. D. Wood pump 
 of 50,000 gallons per minute for condenser work. This is a 
 45-inch pump. Many pumps of this type are used with verti- 
 cal axes. The figure illustrates the method used in forming the 
 casings of pumps of large diameter. Much ingenuity is dis- 
 played at times in the methods of separating the parts. 
 
650 
 
 PUMPING MACHINERY 
 
 The pump shown in Fig. 487 is one recently built by the 
 Alberger Company for the Standard Oil Company. It is a 
 turbine pump and consists of three units connected in parallel, 
 as was the case in the tri-rotor volute pump. This pump is 
 to handle 20,000,000 gallons per twenty-four hours. The form 
 of the casing with the discharge outlets at right angles to 
 the main direction of flow is seen here as in the other turbine 
 pumps. . 
 
 FIG. 486. R. D. Wood 45-Inch Centrifugal Pump. 
 
 Another important use for the centrifugal pump is that of 
 dredging channels. In this case, silt, sand and even rocks are 
 raised with the water by the centrifugal pump and delivered 
 into scows or settling basins on land through long flexible 
 pipes carried on pontoons. The solid material settles out 
 from the water and the water is returned to the stream. This 
 method forms a very cheap and effective manner of dredging 
 where possible. The pump has to be so constructed that 
 the blades of the impeller may be easily replaced when broken 
 or worn out. It is quite evident that the solid matter will 
 
CENTRIFUGAL PUMPS 
 
 651 
 
 cause rapid wear of these parts. It is also important to have 
 all passages direct and of ample size to pass large pieces of 
 solid matter which may be carried through the pump. 
 
 Mr. F. B. Malt by, in the Transactions of the American 
 Society of Civil Engineers, Vol. 54, gives an excellent descrip- 
 tion of dredges used on the Mississippi River and the machinery 
 on them. Mr. Geo. Fowler gives a drawing of the type used 
 in New York harbor in Vol. 31, p. 468, of the Transactions of 
 
 FIG. 487. 20,000,000 Gallon Alberger Multi-impeller Turbine Pump. 
 
 the American Society of Mechanical Engineers. From his 
 description Fig. 488 has been prepared. The figure shows 
 the method of attaching the vanes, tips and the large passages 
 used with these pumps. Mr. Fowler reports that in dredging 
 the New York Ship Channel a piece of shaft weighing 70 
 pounds was lifted and passed by a centrifugal pump, and at 
 Yonkers, N. Y., an 8-inch pump on a wrecking boat lifted 
 and passed a 35 -pound piece of pig iron u^X4f X3i inches. 
 These dredge pumps are driven at a speed of about noo feet 
 per minute at 178 R.P.M. Those used in New York harbor on 
 
652 
 
 PUMPING MACHINERY 
 
 one contract were driven by iga-H.P. engines, delivering 
 10,000 gallons per minute. Mr. Fowler shows the detail of 
 the end of the suction line. This end piece sinks in the 
 soft silt or sand and the solid matter is lifted with the water. 
 Extra openings permit water entering the suction pipe when 
 the silt covers the mouth too much for proper action. 
 
 FIG. 488. Dredging Pump. 
 
 One of the latest uses for centrifugal and piston pumps 
 is that for the high-pressure fire service. The great fire hazards 
 in the congested districts of trade in our large cities has made it 
 necessary to build a separate high-pressure water supply system 
 throughout these districts. The pumping stations may be 
 placed on a river front, where an unlimited supply of water 
 may be had, or in the center of the city, in which case a special 
 supply line of large size is brought from one of the city reser- 
 voirs. The use of salt water in cities such as New York or 
 
CENTRIFUGAL PUMPS 
 
 653 
 
 Doo 
 
654 
 
 PUMPING MACHINERY 
 
 Brooklyn is objected to on account of the damage done by 
 salt water when it touches merchandise. For this reason in 
 New York the supply is taken from the fresh- water mains. 
 The use of water for fire service in such a city is a very small 
 percentage of the total water used, so that this is not a very 
 expensive method, and in the usual system employing fire 
 engines the water is drawn from fire hydrants on the fresh 
 water supply, so that the cost of water is not increased when 
 city water is used with the new stations. 
 
 FIG. 490. Philadelphia High-pressure Station. 
 
 From the pumping stations extra heavy flanged piping is 
 carried in a network over the district. Special hydrants are 
 used and in many cases fire lines are led into the various 
 buildings. Valves must be used at frequent intervals to con- 
 trol lines leading into buildings or sections of the system, so 
 that the waste of water could be prevented in case a wall 
 should fall covering the valve controlling the branch leading 
 to a burning building. 
 
 When a fire occurs the alarm is sent to the pumping 
 station and the pumps are put into commission. One of the 
 earliest of these stations was erected for the city of Philadel- 
 phia. The pumps were of the triplex form driven by gas 
 engines. These are shown in Fig. 489. The gas engines are 
 furnished with gas from a special city main. The engines 
 
CENTRIFUGAL PUMPS 
 
 655 
 
 are started by compressed air held in storage tanks at one 
 end of the building. There are several methods of igniting 
 the charge, so that should one method fail another may be 
 employed. The Westinghouse engines used in this station have 
 always responded to the demands of the service. The Dean 
 Triplex Pumps were built to give 1200 gallons per minute under 
 300 pounds pressure. The gas engines gave 280 H.P. each. 
 
 Fig. 490 shows the seven I2oo-gallon double-acting pumps 
 and the two 35o-gallon pumps originally put in with the two air 
 
 FIG. 491. High-pressure Pumping Station, Brooklyn, N. Y. 
 
 compressors placed at the west end of the building. These 
 air compressors are used to charge the air tanks, which are 
 made up of heavy piping. The suction pipe enters from the 
 river, passing between the two rows of pumps. It is connected 
 to the various pumps by motor-operated valves. The pumps 
 are started under no load and after the engine is operating 
 properly the by-pass valve on the pump is closed, allowing 
 the discharge to enter the pressure main. The pressure main 
 is connected by a check valve with the city mains, so that there 
 is always a pressure of 60 pounds in this line. Provisions 
 are made for connecting fireboats to the pressure line in case. 
 
656 
 
 PUMPING MACHINERY 
 
 there should be a need for it. The gas engine is off centered 
 from the pump to accommodate the gears. 
 
 In 1906 the same kind of system was introduced in 
 Brooklyn, using turbine pumps driven by. means of electric 
 motors. The Brooklyn station is shown in Fig. 491. The five- 
 
 PLAN OF THE STATIONS, SHOWING PUMPS AND PIPING 
 
 FIG. 492. New York High-pressure Station. 
 
 stage Worthington turbine pumps deliver 3000 gallons per 
 minute under a head of 300 pounds. The test curve from 
 these pumps is shown in Fig. 458. The use of electric motors 
 for such stations must be safeguarded by the use of several 
 cross-connected power houses, so that an accident to one will 
 not endanger the reliability of the station. 
 
CENTRIFUGAL PUMPS 
 
 657 
 
 One of the latest installations is that for New York city. There 
 are two stations, one on East River at Oliver Street and one on 
 the North River at Gansevoort Street. Each station has five five- 
 stage Allis-Chalmers centrifugal pumps driven by Allis-Chalmers 
 induction motors. The plan of the station is shown in Fig. 
 
 492, while the interior of one of the stations is shown in Fig. 
 
 493. This photograph shows the appearance of the pump 
 
 FIG. 493. New York High-pressure Station. 
 
 with its outflow casing at the left hand end and the valve on the 
 suction pipe at the right. The 3-phase, 25-cycle, 63oo-volt, 
 74O-R.P.M. induction motors are also seen. The pressure gauges 
 give some idea of the proper operation of the pumps. The 
 extended thrust bearing with its cooling pipes is seen at the* 
 left. A section of this pump was given in Fig. 454. 
 
 There are two suction pipes leading to the river at the 
 top of Fig. 492. And in addition there are two fresh- water 
 suction pipes entering from the city mains on the sides. 
 The two air chambers on the river suction pipes are for the 
 
658 PUMPING MACHINERY 
 
 purpose of steadying the flow. These are kept charged by the 
 suction air pump shown at the upper right-hand corner. The 
 suction pipes form a loop around the station. The discharge 
 mains are also arranged in a loop and both lines are equipped 
 with valves, so that any section of the pipe may be cut out 
 without affecting the operation of the station. Venturi meters 
 are used on the discharge to measure the water used by the 
 system. 
 
 The discharge from each pump is controlled by a valve, 
 so that as soon as the pressure in the main becomes greater 
 than the predetermined amount, due to the decreased use of 
 water, a special valve by-passes a portion of the discharge into 
 the suction. In all of the central .high-pressure stations the 
 pressure is regulated according to the wishes of the fire .chief, 
 as he is in telephonic communication with the station at all 
 times. The pressure is changed by changing the number of 
 units in operation. 
 
CHAPTER XV 
 
 MINE PUMPS 
 
 THE method of mine pumping has been changed very mate- 
 rially of late years. The early method introduced by Newcomen 
 is the one which has been usually followed until recent times. 
 Fig. 494 shows the method of using a long rod extending from 
 the engine house to the pump barrels. When necessary to 
 take off a pump in a side gallery a bell-crank lever was mounted 
 at the side of the rod, one end of which was attached to the 
 rod; the other, to the branch rod or pump. 
 
 When necessary to balance these long rods, beams were 
 mounted above ground or in the rod shaft and on these counter- 
 balance weights were attached. 
 
 By arranging the bell crank lever so that its pump operates 
 on the stroke of the main pump on which no pumping is done, 
 balancing may be accomplished. 
 
 These reduced the load to be moved by the piston, but the 
 inertia of the system was greatly increased. Another method 
 used at times was to place two pumps in the same shaft, 
 balancing one set of rods by the other. Such a pump, 
 built in Aix-la-Chapelle (Fig. 495), illustrates this method 
 very clearly and gives some idea of the complicated system 
 used. The figures illustrate engines with fly wheels although 
 many are used without fly wheels, direct-acting connections 
 being employed. Some of these pumps used in America are 
 remarkable for their size and weight. The Mexican Union 
 Pump of 1880, built with Leavitt jacketed cylinders, had 64X96 
 inches for H.P. cylinders and 100X102 for L.P. cylinder. A 
 36-foot fly wheel was used. The pumps were arranged at various 
 points of the rods, there being fourteen different plungers. 
 The rods were 2618 feet long, and there was so much con- 
 
 659 
 
660 
 
 PUMPING MACHINERY 
 
 FIG. 494. Mine Pump. 
 
MINE PUMPS 
 
 661 
 
 traction in the rods during action that the lo-foot stroke at the 
 engine was reduced to 97 inches at the pumps. This system of 
 
 using a number of pumps in the total 
 lift, in this case 1180 feet, is often 
 used in this type of mine pump. 
 The use of these reduces the amount 
 of head on the various plunger barrels 
 and makes it possible to use lighter 
 parts. 
 
 In the Mexican Union Pump the 
 total weight of the moving parts was 
 1,620,500 pounds. The use of a fly 
 wheel in this pump is not com- 
 mon; many pumps do not have 
 them. 
 
 The Yellow Jacket Shaft Pump 
 of 1880 was one of the best pumps 
 of this time. It was a compound 
 horizontal fly-wheel engine pump. 
 The steam end had 31- and 62-inch 
 cylinders with a lo-foot stroke. The 
 rods were 3055 feet long and the 
 moving weight was 1,510,400 pounds. 
 This pump made only 5^ R.P.M. 
 when giving its maximum dis- 
 charge of 750 gallons per minute. 
 A number of other pumps in this 
 
 FIG. 495 German Mine Pump, district were built with compound 
 
 cylinders 32X129 and 65X96 inches 
 
 with no fly wheels and a total weight of moving parts of 
 
 1,437,900 pounds. 
 
 At the Consolidated California Shaft, a Davey Differential 
 
662 PUMPING MACHINERY 
 
 Pump with steam cylinders 24X90 and 40X96, and pumps of 
 7 feet stroke lifted 500 gallons per minute. The moving parts 
 with 2150 feet of pump rod weighed 860,000 pounds. 
 
 A number of such engines to handle 5400 gallons per minute 
 against a head of 1152 feet cost $1,300,000 without foundations 
 or installation. The cost of operation when 5040 gallons were 
 lifted 1074 feet was $58,210 per month, or $677,440 per year. 
 
 In contrast to this cumbersome method, the method of 
 installing independent pumps, Fig. 496, is to be mentioned. 
 In this case, the pump is installed at a point where needed, 
 and by the use of proper air chambers, the effect of the inertia 
 in the water can be eliminated. 
 
 The pump may be made sufficiently strong and by using 
 the express type of pump, or the centrifugal form, a sufficient 
 capacity may be installed in a small space. The operation of 
 such a pump must be by steam, compressed air, or electricity. 
 The first method is out of the question for great depths, as 
 the condensation of steam means much trouble; moreover, 
 the radiation losses from the steam pipe would be excessive 
 even if the use of high superheat brought the steam to the 
 pump in a dry condition. The use of compressed air may 
 be possible if a high grade compressor is used and the exhaust 
 may be used for ventilation. The loss, however, in this case 
 is one which must be considered carefully. The last method 
 of the use of the electric motor is one which has many points 
 of advantage. 
 
 Power may be brought to the motor in a very simple manner 
 and the location of the apparatus is not difficult. The trans- 
 mission loss with moderately high voltages and proper frequency 
 is not great. 
 
 Fig. 496 shows a shaft in which water is raised from one 
 pump at a great depth and discharged into another at a con- 
 siderable height above it. The water is then lifted by a second 
 pump through the remaining height. The express pump 
 shown in the figure with its motor or a centrifugal pump uses 
 a very small amount of space when compared with the older 
 forms of slow-running pumps even if driven by steam, air or 
 
MINE PUMPS 
 
 663 
 
 FIG. 496. Express Pumps in Series. 
 
664 
 
 PUMPING MACHINERY 
 
 pump rod. In this way considerable expense is saved in pre- 
 paring a pump room. 
 
 The cost of the apparatus in this case has been very 
 materially reduced. Two i6oo-gallon independent high-speed 
 pumps to work against 1500 feet head, and each driven by 
 an 800 H.P. slow-speed induction motor cost $80,000 and 
 they weighed with the motors 600,000 pounds. The same 
 capacity of rod-driven pumps would cost $960,000 and the 
 
 FIG. 497. Knovvles Express Pump. 
 
 moving parts alone would weigh over 5,000,000 pounds. The 
 cost would be twelve times and the weight eight times as 
 much as that of the high-speed electrically-driven pump. 
 
 The express pump of Riedler was explained on p. 163 and 
 a section of it was shown in Fig. 154. After Riedler had 
 shown that high speeds were possible with proper valve and 
 proper air chambers, others took up the construction of these 
 pumps and found that the mechanically-closed valves were 
 not necessary and that self-closing valves could be used. In 
 some cases the valves are of small diameter. 
 
MINE PUMPS 
 
 665 
 
 Fig. 497 is a Knowles express pump intended for mine 
 service. The pump is a duplex double-acting pump, 6iJxi5 
 inches which has a capacity of 1600 gallons per minute, against 
 a head of 1550 feet. The motor is of 300 R.P.M. This speed 
 is very high and the inertia of the water column is so great 
 that there would be a continuous breaking of the column with 
 its accompanying knocking if the air chambers and valves 
 
 FIG. 498. Knowles Express Pump. 
 
 were not properly designed. Poppet valves are the types 
 used with this pump! They have the advantage of being 
 almost balanced and of giving a large area of discharge. 
 
 The figure illustrates the heavy construction necessary on 
 the outlet side, the air piping for charging the chambers, 
 the outside rods to connect the front and back cross-heads, 
 the pressure oiling system and the starting by-pass and 
 priming valves. The general lines shown in the figure are 
 excellent and the design is massive and so arranged as to 
 
666 
 
 PUMPING MACHINERY 
 
 
 
 
 
 
 
 
 
 
 
 
 
 1 
 
 
 
 np 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 K^o 
 
 gte 
 
 HfeKS 
 
 Deled 
 
 E 
 
 ectric Pu 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 DIRECT 'cOfjINEbTE 
 
 DTO 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 I 
 
 i.p 
 
 6 
 
 100H.F|.-?80-2 
 
 M 
 
 or 
 
 rjj 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 Load &.E 
 
 ffipie 
 
 ncy 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 X 
 
 )/y 
 
 le 
 
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 jr 
 
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 il 
 
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 M 
 
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 otor 
 ~SeF 
 
 
 
 
 
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 7 
 
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 600 
 
 
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 / 
 
 
 
 
 
 
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 11, 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
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 5 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
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 SSI 
 
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 e 
 
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 ^ 
 
 ) 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 
 FIG. 499. Test Curves of Express Pump. 
 
MINE PUMPS 667 
 
 care for the strains brought on during the operation of the 
 pump. 
 
 The pump shown in Fig. 498 is a smaller unit, driven by 
 a loo-H.P. direct-current motor. This pump is also built 
 by the Knowles Steam Pump Works. It was intended to 
 operate at 300 R.P.M., pumping 250 gallons per minute against 
 a head of 1000 feet. The plungers were 3^ inches in diameter 
 and of 5j-inch stroke. The machine is self-contained and 
 the frames and bed-plate are constructed to make a rigid 
 structure. The plungers, which are outside packed, are con- 
 nected by outside rods. 
 
 A test of this pump gave the results which are shown by 
 curves in Fig. 499. The result of 94 per cent efficency for the 
 pump and 84 per cent for the combined unit, lifting water 
 1 200 feet is remarkable. The electric line loss in operating 
 such a pump may be rendered small by using a high voltage, 
 when the installation becomes very efficient. 
 
 The types of direct-acting steam pumps used for mine 
 work have been discussed in Chapter III and the forms illus- 
 trated by several figures. Fig. 500 illustrates another one 
 of these pumps built by the Jeanesville Iron Works Co. 
 
 It is intended to lift 1200 gallons per minute against a head 
 of 700 feet. The size of the valve pots, the control valves 
 and gauges, the arms of the rotating steam valves, the method 
 of valve operation, and the forms of the cylinders may all 
 be clearly seen. 
 
 The valve pots are particularly large in this pump because 
 of the large size of valves used by this company. Fig. 501 
 shows two forms of valves used by them for heads up to 750 
 feet. The annular valve is used when the head is not greater 
 than 1000 feet. This figure illustrates the method of lining 
 the valve pot with wood when the water contains acids which 
 will attack the metal. 
 
 The design of mine pumps depends entirely on the type 
 of pump. When the pump is of the fly-wheel type, the method 
 is that used in Chapter V. The great length of discharge pipe 
 would mean considerable inertia, but by the use of air chambers 
 
668 
 
 PUMPING MACHINERY 
 
 the discharge in the pipe is not a fluctuating one and hence 
 the friction in this pipe line is the only part requiring consider- 
 ation in addition to the static head. The problem of the fly- 
 wheel design is slightly different from that considered in Chapter 
 V if an electric motor is used to drive the pump. In this case 
 the tangential effort of the pump end pressure, combined with 
 the friction of stuffing boxes, and forces from inertia and 
 weight, is found. From this the excess and deficiency of tan- 
 gential effort area is determined above and below a line 
 
 FIG. 500. Jeanesville Mine Pump. 
 
 representing the delivered torque from the motor, which, of 
 course, is constant. 
 
 The problem of a direct-acting steam pump without a 
 fly wheel has to be handled in a different manner. The inertia 
 of the moving parts is the important element in this design. 
 The first problem to be considered will be one involving the 
 type of mine pump shown in Fig. 494, but in which the fly 
 wheel is omitted. In this case it is assumed that the indicator 
 cards from the two steam cylinders are those shown at A, 
 Fig. 502. These are combined as at B by taking the height 
 to be 
 
MINE PUMPS 
 
 669 
 
670 
 
 P UMPING M A CHINER Y 
 
 Down 
 
 Up 
 
 ~~~H| Resultant Curve 
 
 W Water Curve 
 C Unbalanced Weight 
 /I Friction 
 / 2 Friction 
 _J Resultant Curve 
 W 2 Water Curve 
 
 Down 
 
 FIG. 502. Forces in Direct-Acting Rod Mine Pump, 
 
MINE PUMPS 671 
 
 This is the force of the steam acting on the system. The 
 unbalanced weight of the moving parts of the whole system 
 per square inch of low pressure piston area is next found as 
 
 The value of this force which reaches the piston rod will 
 depend on the inclination of the bell-crank levers and by con- 
 structing the perpendiculars from the pivot, p h and p v to the 
 line of pump-rod pull and piston-rod pull, respectively, the 
 pull in the direction of the piston rod may be computed. 
 
 This gives the curve c, Ci, c 2 . 
 
 The weights of all the parts of the rods, balance weights, 
 levers and pump pistons are now found. These must have 
 an acceleration equal to the acceleration of the piston, multi- 
 plied by the ratios of the various lever arms up to the part 
 considered. Thus the acceleration of the main pump rod is 
 
 If T is the ratio of the arms of a balance lever the acceler- 
 b 
 
 ation of this part is 
 
 p v a 
 
 If now each weight is examined and its acceleration found 
 in terms of a p , the force which must be exerted on the piston 
 to accelerate all of these parts is 
 
672 PUMPING MACHINERY 
 
 This means that if the sum of the terms 7 ^~ T W x is found, 
 
 *~*pk9 
 
 this will give a weight which may be assumed to have the same 
 motion as the piston in computing the force on the piston 
 
 required to give these parts their motion. These ratios, ~, 
 
 Ph 
 
 J-, etc., vary slightly, so that this sum will change as the piston 
 
 moves. The value of this is laid off for different positions 
 of the piston along the stroke, as in C, Fig. 502. 
 
 The pressure per square foot on the pistons from the water 
 is equal to 
 
 if the air chambers are of the proper size to prevent the 
 fluctuation of pressure, and the connections to the air chambers 
 are so short that the iriertia of the water in this connection 
 can be neglected. If the connection is not short, then the weight 
 of the water between the chamber and the pump must be added 
 as a weight of one of the reciprocating parts. In general, how- 
 ever, the terms involving X and a of Eq. (21), Chapter V, are 
 so small when compared with hi that they may be neglected. 
 
 The sum of the products whA~j- for the various pump pistons 
 
 when divided by AI will give the pressure per square inch 
 which must be exerted per square inch of L.P. piston required 
 to lift the water. This is shown by the curve w, w it w^ etc., 
 for the different piston positions. 
 
 If there is no air chamber on the pump, the weight of the 
 whole column of discharging water must be added to the 
 weight of the reciprocating parts in the determination of the 
 weight to be accelerated. 
 
 The friction of the various balancing and supporting levers 
 should next be found and with this the friction of the rods 
 against their supports when not vertical as well as the friction 
 of the stuffing boxes. These forces are overcome by a force 
 
MINE PUMPS 673 
 
 in the direction of the piston travel which may be found by 
 the use of the moments of the various forces about different 
 pivots. In this way the curve / 1? / 2 , fs . . . is found, showing 
 the force per square inch of low-pressure piston area which 
 must be exerted on the piston in order to overcome the friction. 
 These curves are combined, giving the resultant curve shown 
 at D. The point i is found as 
 
 i = a ll -a l w l -a 
 Ef =a\bi a\Wi a 
 
 Since the term a\c\ is added in one case and subtracted in the 
 other, it is seen that the work on the steam end is less on the 
 one stroke than on the other unless the weights are balanced. 
 If under balanced, the work of the steam end is greater on 
 the up stroke, while if overbalanced, the work of the steam 
 end is greater on the down stroke. 
 
 The area of the curve /'/i, which is the resultant of the 
 pressures which act during the stroke, represents unbalanced 
 work and hence the positive and negative parts must be equal, 
 or the moving parts of the pump will not come to rest at the 
 end of the stroke. In the actual engine of the direct-acting 
 form it may be said that the stroke ends, the moving parts 
 coming to rest, when the negative area equals the .positive. 
 On the return stroke the same must be true. 
 
 If now the net unbalanced pressure on each square inch 
 of piston area at some point of the stroke be divided by the 
 equivalent weight per square inch of piston area at that point 
 and multiplied by g, the result will be the acceleration in the 
 parts caused by the force, or 
 
 ap= %7 
 
 If now a curve be plotted with the values of a p for different 
 points, the result will show how the acceleration varies. 
 This is shown in Fig. 503. Now 
 
 dv dv ds dv 
 
674 
 
 PUMPING MACHINERY 
 
 Hence 
 
 and 
 
 -y 
 ds 
 
 ;r V 2 
 Vdv=\ads=. 
 
 TIP b 
 
 Down 
 
 FIG. 503. Acceleration and Velocity Curves. 
 
 The area of the curve just drawn from the end to any point 
 
 V 2 
 is . If then the integral curve be drawn in Fig. 503 the 
 
 V 2 
 ordinates of this curve will be and from these V may be 
 
 found, giving the dotted curve. From this curve the velocity 
 time curve may be constructed as shown in Fig. 193 and from 
 it, the time taken to complete one stroke. The curve is con- 
 structed for the two strokes. The time found will determine 
 the number of strokes per m'nute and consequently the capacity 
 of the pump. 
 
MINE PUMPS 675 
 
 If it is desired to cut down the speed of the pump more, 
 a balanced mass could be added to the system, causing the 
 acceleration curve to be lower, although the force diagram, 
 Fig. 502, would not change. This may also be accomplished 
 by decreasing the maximum steam pressure and carrying the 
 cut off later. The total area of the indicator card cannot be 
 changed, as that must be equal to the work done on the water 
 and in overcoming the friction. 
 
 If it is desired to increase the speed of the pump, the 
 balanced mass must be made smaller if possible and if this 
 cannot be done the steam pressure is increased and the cut off 
 is made earlier. The first of these changes the acceleration 
 curve, increasing its height without any change in the curve 
 of unbalanced force, while the second method increases the 
 height of the acceleration curve by increasing the height of 
 the curve of unbalanced force, since there is no change in the 
 weight of the moving parts in this case. In the latter case 
 also, it must be remembered that the area of the indicator 
 cards cannot be changed. 
 
 The mass in some cases may be so great that even with high 
 steam pressure the speed must be slow and the capacity small. 
 In such cases balance weights may be omitted; then it 
 becomes necessary to cut down the amount of steam used on 
 the down stroke. A better method is that shown in Fig. 495, 
 in which two rods and pump sets are so used that as one is 
 descending the other is ascending. In this case the effect is 
 a balanced one without the inertia effect of balance weights, 
 for each side of the apparatus is designed to do pumping, and 
 the mass is that of an unbalanced pump. Such an arrange- 
 ment will give the fastest moving pump of this type. 
 
 In using these pumps it is well to note that should the steam 
 pressure be increased without the reduction of the cut off, 
 the positive area of unbalanced force would be greater than 
 the negative area, and consequently the piston would strike 
 the head with considerable force. For this reason a large 
 cushion space should be used or springs should be applied 
 to take up the shock. In case the pressure is reduced the 
 
676 
 
 PUMPING MACHINERY 
 
 piston will come to rest as soon as the negative area is equal 
 to the positive area. In some cases the steam may be admitted 
 to the opposite end, at which time the full boiler pressure will 
 act to stop the motion and reverse the pump. This would 
 mean a large increase in the negative residual pressure, which 
 would mean a rapid negative acceleration, bringing the parts 
 to rest. 
 
 The use of direct-acting pumps with steam cylinders at the 
 pump involves the same principles of design as the pumps 
 just considered. In this case, however, the total mass recip- 
 rocated is quite small and for that reason the residual steam 
 
 Combined H.P. and L.P. Steam 
 
 Water Resistance 
 
 FIG. 504. Compound Direct-acting Pump Diagram. 
 
 pressure cannot be as large as used with the other type of 
 pump, since the acceleration 
 
 would be very great. This would cause pounding and there 
 would be some danger of wrecking the pump if the valve gear 
 did not work properly. For this reason the method is to 
 carry the pressure practically the full length of the stroke, get- 
 ting the advantages of expansion in the use of several cylinders 
 in which the cut off is at or near the end of the stroke. For 
 such the combined card is shown in Fig. 504. The steam 
 pressure is slightly above the resistance for the major part of 
 the stroke and the piston is brought to rest by the cushion 
 steam or the valve reversal. 
 
MINE PUMPS 677 
 
 The principles of Chapter V may be applied in this case 
 for the resistance of the various parts, the terms which depend 
 on the acceleration being so small as to be neglected in com- 
 parison with the height through which the water is lifted. 
 Should this be such that these terms are not small, they would 
 have to be neglected in the first approximation, when the 
 motion is found; then computed for the motion thus deter- 
 mined, and a second approximation made in which they are 
 used. If these terms are changed much by the new curve 
 of time and velocity a third approximation should be made. 
 
 In this manner the probable capacity of the pump may be 
 found. Of course in this type of pump with a proper cushion 
 chamber the speed may be increased by increasing the steam 
 pressure. After the speed reaches its limit, pounding begins, 
 as the cushion cannot care for the energy stored up in the 
 moving parts. 
 
BIBLIOGRAPHY 
 
 THE list of references given below have been used in the preparation of 
 this treatise. The references are given by volume number or date with the 
 page on which the reference is found. The following abbreviations are used: 
 
 Am. Mch. American Machinist. 
 
 Engng. London Engineering. 
 
 Eng. Mag. Engineering Magazine. 
 
 Eng. News Engineering News. 
 
 Eng. Rec. Engineering Record. 
 
 Engr. The Engineer (American). 
 
 A.S.C.E. Transactions of the American Society of Civil Engineers. 
 
 A.S.M.E. Transactions of the American Society of Mechanical Engi- 
 neers. 
 
 J.A.S.M.E. Journal of the American Society of Mechanical Engineers. 
 
 N.E.W.W. Journal of the New England Water Works Association. 
 
 A.W.W.A. Transactions of the American Water Works Association. 
 
 Z. f. d. g. Turb. Zeitschrift fiir das gesampte Turbinenwesen. 
 
 Z. d. V. d. Ing. Zeitschrift des Vereines deutscher Ingenieure. 
 
 Elec. World Electrical World. 
 
 Power Power and the Engineer. 
 
 A.E.S. Journal of Associated Engineering Societies. 
 
 Cassiers Cassiers Magazine. 
 
 These are arranged under the following heads: 
 
 Texts. 
 
 Air Pump and Screw Pump Tests. 
 Centrifugal Pumps, Theory and De- 
 sign. 
 
 Centrifugal Pumps. 
 Centrifugal Pump Stations. 
 Centrifugal Pump Tests. 
 Costs. 
 
 Fire Pumps and Stations. 
 Historical. 
 Injectors and Direct Steam Pumps. 
 
 Irrigation, Drainage and Sewage Pumps 
 and Stations. 
 
 Mine Pumps. 
 
 Reciprocating Pumps, Tests and Re- 
 sults. 
 
 Reciprocating Pumps, Design and 
 Theory. 
 
 Rotary and Disc Pumps. 
 
 Simplex Pumps. 
 
 Special Pumps. 
 
 Water Works and Water Works Pumps. 
 
 679 
 
680 BIBLIOGRAPHY 
 
 TEXTS 
 
 Centrifugal Pumps and Turbines. Chas. H. Innes. 
 
 A Descriptive and Historical Account of Hydrau- 
 lic and Other Machines for Raising Water. Thomas Ewbank. 
 
 Encyclopedia Britannica. Article "Steam Engine." 
 
 Growth of Steam Engine. R. H. Thurston. 
 
 Hydraulic Motors. G. R. Bodmer. 
 
 Hydraulic Power Engineering. G. Croydon Marks. 
 
 Hydraulics. Mansfield Merriman. 
 
 Hydraulics. F. C. Lea. 
 
 Hydraulics. A. H. Gibson. 
 
 Hydraulics. W. C. Unwin. 
 
 Irrigation Works in India. R. B. Buckley. 
 
 Lives of Boulton and Watt. Smiles. 
 
 Life of Newcomen. Smiles. 
 
 Lives of George and Robert Stevenson. Smiles. 
 
 Machine Design. W. C. Unwin. 
 
 Machine Drawing and Design. Low and Bevis. 
 
 Mechanical Engineers' Pocket Book. Kent. 
 
 Die Pumpen. Hartmann-Knoke-Berg. 
 
 Practice and Theory of the Injector. Strickland Kneass. 
 
 Pumping Machinery. Wm. M. Barr. 
 
 Water Supply and Irrigation Papers of the U. S. 
 
 Geological Survey No. i. Herbert M. Wilson. 
 
 Water Works Pumps. Chas. A. Hague. 
 
 Die Zentrifugal Pumpen. Fritz Neumann. 
 
 AIR PUMP AND SCREW PUMP TESTS 
 
 Experiments on Air Pumps. Engng., 86: 703. 
 
 Screw Pumps. Eng. Rec., 58: 426; Eng. 
 
 News, 60: 269. 
 
 Air Lift Pumps at Redlands, Cal. Eng. Rec., 51: 8. 
 
 Air Lift Pump. Eng. News, 32: 27; Engr., 
 
 Aug. 15, 1904. 
 
 Merrill Compressed Air Pump. Engr., July ! 5, 1904. 
 
 Air Lift Pump. Engr., June i, 1906. 
 
 CENTRIFUGAL PUMPS, THEORY AND DESIGN 
 
 On the Construction of the Impellers of High- 
 Speed Low-head Centrifugal Pumps, by P. Zeit. f. d. g. Turb., Nov. 10, 
 Riebensahm. 1909. 
 
 Investigation of Centrifugal Pumps, Part II. C. 
 
 B. Stewart. Bui. U. of Wis., No. 318. 
 
 Design of High Lift Centrifugal Pumps, by F. z. 
 
 Nedden. Eng. Mag., Jan.-May, 1910. 
 
BIBLIOGRAPHY 681 
 
 Le Calcul de la hauteur de refoulement des Le Genie Civil, Nov. 6, 
 
 pompes centrifuges. 1909. 
 
 Notes on Centrifugal Pump Design. J. B. 
 
 Sperry. Am. Mach., Nov. 19, 1908. 
 
 Die Grundlagen der Lorenzschen Theorie des Zeit. f. d. g. Turb., May 10, 
 
 Kreiselrader. R. Lowy. 1909. 
 
 Zur Theorie der Zentrifugal pumpen. E. Busse. Zeit. f. d. g. Turb., Jan. 9- 
 
 Feb. 10, 1909. 
 Ausfiihrungen und Versuchsresultate von Hoch- Zeit. f. d. g. Turb., Feb. 20, 
 
 druckzentrifugalpumpen, by Griessman. 1909. 
 
 Die Wirkungsweise der Kreiselpumpen und Mit. u. Forsh., Oct. 12, 
 Ventilatoren. Dr. R. Biel. 1907, Eng. Rec., Oct. 12, 
 
 1907. 
 The Design of Centrifugal Pumps, by J. Richards. Eng. News, July 29 and 
 
 Aug. 5, 1897. 
 Design of Centrifugal Pumps, E. U. Percy. Jour. Elec. Power and Gas, 
 
 April 27, 1907. 
 Theory of Centrifugal Pumps and Fans. A.S.C.E., 50, No. 963; 51, 
 
 No. 936. 
 
 Design of a -Centrifugal Pump. Am. Mach., Sept. 22, 1910. 
 
 Ueber der hydraulischen Wirkungsgrad von Tur- 
 binen bei ihrer Verwendung als Kraftma- 
 schinen und Pumpen. R. Proell (Springer, 1904). 
 
 CENTRIFUGAL PUMPS 
 
 Allis-Chalmers Electrically Driven Fire Service, Power, Nov. 2, 1909, 760; 
 
 New York, 79% Eff. ' J.A.S.M.K, Sept., 1909. 
 
 Worthington, Electrically Driven, 80% Eff., 94.2 Engng., Feb. 5, 1909; Zeit. 
 
 M. head. f. d. g. Turb., Sept. 20, 
 
 1909. 
 
 A New Turbine Pump. Engr., June 14, 1907. 
 
 Some Practical Experiences with Centrifugal 
 
 Pumps for Water Works Service, C. A. Hague Eng. News, Aug. i, 1907. 
 Tan -gyro Centrifugal Pump. Engng., Jan. 6, 1911. 
 
 The Worthington Turbine Pump. Am. Mach., 27: 514. 
 
 Fire Boats of Chicago. Fire & Water Eng., April i, 
 
 1908. 
 
 Large Centrifugals. Engng., 42: 233; 38: 12. 
 
 Appold Pumps at Portsmouth Docks. Engng., 2: 382. 
 
 Stadil Fjord Pumps. Engng., i: 13. 
 
 Andrews Pumps. Engng., 3: 495. 
 
 Gwynne Pumps. Engng., i: 82; 6: 185; 50: 
 
 755- 
 High Lift Centrifugals. Engng., 6: 519; Eng. Rec., 
 
 58: 112. 
 Bernays Pump. Engng., 15: 456. 
 
682 BIBLIOGRAPHY 
 
 New York Dredging Pumps. A.S.C.E., 25: 599. 
 
 Irrigation Pumps. A.S.C.E., 54: 163. 
 
 Dredging Pumps. A.S.C.E., 54: 285; 54: 
 
 391; A.E.S., 19: 140; 
 
 Eng. News, 27: 290. 
 
 Vertical Shaft Centrifugals. Engng., 37: 138. 
 
 Malta Dock Pumps. Engng., 44: 224. 
 
 Southwark Foundry Pumps. Eng. News, 14: 349. 
 
 Compound Centrifugals. Eng. News, 26: 246. 
 
 Balancing Centrifugals. Eng. News, 25: 114. 
 
 Lea-Degen Pumps. Eng. Rec., 54: 352. 
 
 Well Pumps. Eng. Rec., 50: 177. 
 
 High Lift Centrifugals. Eng. Rec., 58: 112. 
 
 Helicoidal Pumps. Engng., 42: 570. 
 
 Twin Pumps for Dock Service. Engng., 86: 719. 
 
 CENTRIFUGAL PUMPS' STATIONS 
 
 New York Fire Station. Power, Nov. 2, 1909; A.S. 
 
 M.E., 31: 437; Eng. 
 
 Rec., 57: 22. 
 
 Auxiliary Pumping Station at Charleston, W. Va. Eng. Rec., Sept. 25, 1909. 
 The Turbine Pumps of Montreal Water Works. Eng. Rec., 54: 488. 
 Chicago Fire Boats. Fire & Water Eng., April i, 
 
 1908. 
 
 Electric Mine Drainage in Europe. Elec. World, Nov. 17, 1906. 
 
 Electric Pumping Equipment for the Mexico Eng. Rec., Dec. 8, 1906, 58: 
 
 Water Works. 128. 
 
 Low Head Pumping Plant at New Orleans. Eng. Rec., April 23, 1910. 
 
 Peoria Station. Eng. Rec., 51: 139. 
 
 Montreal Water Works. Eng. Rec., 54: 488. 
 
 Schenectady Station. Eng. Rec., 51: 640. 
 
 CENTRIFUGAL PUMPS TESTING 
 
 Tests of a New Centrifugal Pump (78% EfL). Eng. Rec., 54: 352. 
 Versuche an einer Zentrifugalpumpe. Z. f. d. g. Turb., March 10- 
 
 20, 1908. 
 
 Twin Centrifugals for Dock Service. Engng., 86: 719. 
 
 Centrifugal Pumps of 1885. Engng., 40: 124, 215. 
 
 Early Centrifugals. Engng., 2: 382. 
 
 Centrifugal Pump Efficiency. A.S.M.E., 7: 598. 
 
 Centrifugal Pump Test. , Engng., 52: 696. 
 
 Allis-Chalmers Pumps at Pittsburg. Eng. Rec., 58: 649; Eng. 
 
 News, 60: 573. 
 Richards Tests. Eng. News, 38: 75. 
 
BIBLIOGRAPHY 
 
 683 
 
 De Laval Centrifugal. 
 Worthington Turbine Pumps. 
 
 Sewage Pumps at Carlisle. 
 Vertical Shaft Pump. 
 Drainage Pumps of the South. 
 Pacific Coast Pumps. 
 Duty Test at Torresdal Station. 
 
 Eng. Rec., 50: 216. 
 Engng., 87: 181; 
 
 News, 58: 112. 
 Engng., 87: 46. 
 Engng., 37: 138. 
 A.S.C.E., 56: 159. 
 A.S.C.E., 56: 148. 
 Power, Nov. 16, 1909. 
 
 Eng. 
 
 COSTS 
 
 Cost of 2o-Million Gallon Pumps. 
 
 Cost of Large Pumps in 1870. 
 
 Cost of Pumps in 1886. 
 
 Cost of Pumps. 
 
 Cost of Pumps in 1886. 
 
 Cost of Piping in 1884. 
 
 Cost of Holly Engines. 
 
 Cost of Large Pumps. 
 
 Cost of Pumps for Chicago. 
 
 Cost of Small Plants. 
 
 Comstock Lode Pumps. 
 
 Cost of Boston Plants. 
 
 Cost of Raising Water. 
 
 Cost of Pumping. 
 
 Cost of Raising Water. 
 
 Cost of Pumping Water. 
 
 Cost of Electric Pumping. 
 
 Cost of Operating Producer Gas Stations. 
 
 Oil Engine Stations. 
 
 Olean Gas Engine Station. 
 
 Commercial Pumping Engine. 
 
 Steam Driven vs. Electric Driven Pumps. 
 
 High vs. Low Duty Pumps. 
 
 Possibilities of Economy in Pumping Engines. 
 
 Small Pumping Plants. 
 
 Duty and Costs of Operating Pumps. 
 
 Water Waste in New York. 
 
 Receipts from W. W. of the Eight Largest Cities in 
 
 America, for 1884. 
 
 Financial Management of Water Works. 
 Depreciation. 
 
 American Water Works Statistics. 
 Comparative Merits of Various Pumps. 
 
 Eng. News, 23: 578. 
 Eng. News, 15: 93. 
 Eng. News, 16: 93, 95. 
 Eng. News, 13: 340. 
 A.E.S., 14: 24. 
 Eng. News, 12: 47. 
 Eng. News, 30: 67. 
 Eng. News, 30: 342, 
 
 435, 484- 
 
 Eng. News, 32: 225. 
 N.E.W.W., 14: 163. 
 Eng. Rec., 51: 360. 
 N.E.W.W., 15: 299. 
 Van Nos. i : 648. 
 A.S.C.E., 4: 369. 
 Eng. News, 16: 231. 
 Eng. News, 30: 181. 
 N.E.W.W., 10: 184. 
 Eng. Rec., 59: 786. 
 Eng. Rec., 58: 230. 
 Eng. Rec., 58: 112. 
 Eng. Rec., 51: 537. 
 A.W.W.A., 1907, 189. 
 A.W.W.A., 1908, 725. 
 N.E.W.W., 13: 163. 
 A.W.W.A, 1907, 157. 
 A.W.W.A., 1907, 210. 
 Eng. Rec., 48: 340. 
 
 Eng. News, 12: 47. 
 N.E.W.W., ii : 63. 
 A.W.W.A, 1903, 473. 
 Eng. News, 13: 340. 
 Engng., 37: 394. 
 
 424, 
 
684 
 
 BIBLIOGRAPHY 
 
 FIRE PUMPS AND STATIONS 
 
 Arrangement of Hydrants and Water Pipes. 
 Capacity of Steam Fire Engines, Hydrants and 
 
 Hose. 
 
 Fire Protection, Amount of. 
 High Pressure Stations. 
 
 Boston. 
 
 Chicago. 
 
 New York. 
 
 Philadelphia. 
 
 Providence. 
 
 Hand Pump, Horse-drawn. 
 Gould System. 
 Building a Fire Pump. 
 Merryweather Pump. 
 
 N.E.W.W., 7: 49. 
 
 N.E.W.W., 9: 151. 
 N.E.W.W, 3: 97. 
 N.E.W.W., 13: 304. 
 Eng. Rec., 48: 138. 
 Eng. Rec., 57: 22; J.A.S. 
 
 M.E., Sept., 1909. 
 A.S.M.E., 3 i: 437; Power, 
 
 1909, 760. 
 
 Eng. Rec., 59: 748; 49: 309. 
 N.E.W.W., 13: 85. 
 Engng., 15: 301. 
 Engng., 21 : 432. 
 Am. Mach., Jan. 5, 1911. 
 Engng., 10; 192, 
 
 HISTORICAL 
 
 See Text-Book List. 
 
 Centrifugal Pumps. 
 
 Early History of Centrifugal Pump. 
 
 Development of Centrifugal Pumps. 
 
 Invention of Centrifugal Pump. 
 
 History of Centrifugal Pump. 
 
 Engineering 300 Years Ago. 
 
 Fire Pumps. 
 
 Flash Pumps in Holland. 
 
 Humphrey Pump. 
 
 Hydraulic Rams. 
 
 Newcomen and His Work. 
 
 Pumping Machinery. 
 
 Pumping Machinery. 
 
 Development of Pumping Machinery. 
 Some Heavy Modern Pumping Machinery. 
 Worthington, Henry R. 
 
 Cassiers, 28: 154. 
 Engng., i: 275. 
 A.W.W.A., 1904, 175. 
 Engng., 50: 670. 
 Prac. Mech. Jqur., 1851. 
 Cassiers, 8: 97. 
 Cassiers, 7: 307. 
 Eng. News, May 19, 1910. 
 Engng., 88: 737, 512, 514. 
 Cassiers, 28: 65. 
 Cassiers, Dec., 1891. 
 Engr., Nov. 16, 1903. 
 .Eng. Mag., May, 1891, i: 
 
 141. 
 
 Eng. Mag., 5: 451. 
 Cassiers, Jan., 1895. 
 Eng. News, Sup., March 23, 
 
 1893. 
 
 INJECTORS AND DIRECT STEAM PUMPS 
 
 American Steam Jet Pump. 
 Direct Displacement Pump. 
 Korting Bilge Pump. 
 Pulsometer. 
 
 Engng., 2: 237. 
 Engng., March 15, 1905. 
 Engng., 19: 477. 
 Engng., July 15, 1905. 
 
BIBLIOGRAPHY 
 
 685 
 
 Pulsometer, Hall's. 
 Pulsometer, Test. 
 Steam Ejector. 
 Water Jet Pump. 
 
 Engng., 22: 56. 
 Tech. Quar., Sept., 1901. 
 Engng., ii : 416. 
 Engng., i: 117. 
 
 IRRIGATION, DRAINAGE, AND SEWAGE PUMPS AND 
 STATIONS 
 
 Automatic Sewage Station. 
 Boston Sewage Engine and Plant. 
 Carlisle (Eng.) Plant. 
 Centrifugals for Irrigation. 
 Cheswick Drainage Pump. 
 Chicago Sewer System. 
 
 Dorchester Leavitt Pump. 
 
 Dortdrecht Drainage Station. 
 
 Ferrara Drainage Station. 
 
 Flash Pumps in Holland. 
 
 Hampton Institute. 
 
 Hull Plant. 
 
 Huntly Plant. 
 
 Irrigation on the Pacific Coast. 
 
 Leavitt Pumps. 
 
 Memphis Storm Sewer Station. 
 
 New Orleans System. 
 
 Portsmouth Pump. 
 
 Providence Pumps. 
 
 Scoop Wheels. 
 
 Screw Pump for Milwaukee. 
 Stadil Fjord Pump. 
 Thompson Sewage Valves. 
 Triplex Drainage Pump. 
 Washington, D. C, Plant. 
 McLaren Sewage Pump. 
 
 Eng. Rec., 57: 272. 
 Engng., 40: 555. 
 Engng., 87: 46. 
 A.S.C.E., 54: 163. 
 Engng., 27: 477. 
 Eng. Rec., 52: 578; 58: 
 
 426; Eng. News, 60: 
 
 269. 
 
 Eng. Rec., 51: 676. 
 Engng., 21 : 192. 
 Engng., 21 : 9. 
 Eng. News, May 19, 1910. 
 Eng. Rec., 52: 566. 
 Eng. News, 23: 126. 
 Eng. News, 60: 262. 
 Eng., 44: 479. 
 N.E.W.W., 9: 163. 
 Eng. Rec., 53: 496. 
 Eng. Mag., 25: 342. 
 Engng., 10: 44. 
 Eng. News, 32 : 259. 
 Engng., 8: 174; 9: 183, 
 
 194, 230, 274, 321, 441. 
 Eng. News, 23: 218. 
 Engng., i: 13. 
 Engng., 3: 127. 
 Engng., 29: 23. 
 Eng. Rec., 58: 200. 
 Engng., 19: 317. 
 
 MINE PUMPS 
 
 Allis Pumps at Chapin Mine. 
 Bohemia Pumps. 
 Cameron Pump. 
 Comstock Lode Pumps. 
 
 Cornish Pump at Ontario Mine. 
 
 Deep Mine Pumps. 
 
 Duplex Sinking Pump, Electric Drive. 
 
 Eng. News, 30: 310. 
 Engng., 27: 155. 
 Engng., 14: 180. 
 Eng. Rec., 51: 360; Engr., 
 
 May 15, 1905. 
 Eng. News, 32: 440. 
 5 2 : 40. 
 
 Eng. News, 29: 471. 
 
BIBLIOGRAPHY 
 
 Electric Drive. 
 
 Electric Driven Knowles Pump. 
 Eston, Eng. Pumps. 
 
 Fowey Consols Mines (Recent Practice in Pump- 
 ing). 
 High Speed Pumps. 
 
 Mine Pumps by Howell Green. 
 
 Mine Pumping. 
 
 Mine Working at Great Distances with Rods. 
 
 Rittinger Pump. 
 
 Underground Mine Pump. 
 
 Engng., 44: 534. 
 Eng. Rec., 54: 403. 
 Engng., 16: 294. 
 
 N.E.W.W., 8: 85. 
 Eng. Mag., 24: 772: 
 
 July 15, 1904. 
 A.S.M.E., 4: 217. 
 Cassiers, 31: 125. 
 
 En g n g-, 3 i: 281. 
 Engng., 30: 151. 
 Engng., 43: 285. 
 
 Engr, 
 
 RECIPROCATING PUMPS TESTS AND RESULTS 
 
 Allegheny Allis Pump. 
 
 Belfast Engine. 
 
 Belmont Pump. 
 
 Blake Pump. 
 
 Boston Allis Pump. 
 
 Brooklyn Pumps. 
 
 Brooklyn Pump. 
 
 Chestnut Hill Pump. 
 
 Chicago Allis Pump. 
 
 Cleveland Snow Pump. 
 
 Coal Consumption of Pumps in 1867. 
 
 Cornish Mine Pump. 
 
 Davison Pump at Norwood. 
 
 Davy Pump. 
 
 Duties on Cornish Engines. 
 
 Fall River Davison Pump. 
 
 Farcot Pump. 
 
 Gaskill Pump. 
 
 Hannibal Pump. 
 
 Hathorn, Davey & Co. 
 
 High Duty Tests. 
 
 Improvements in Pumping Engines. 
 
 Lawrence Pump. 
 
 Lowell Pump. 
 
 Memphis Worthington Pump. 
 
 Moreland Pump. 
 
 Newark Pump. 
 Norwood Pump. 
 Odessa English Pump. 
 Pawtucket Pump. 
 
 Engng., 41: 33. 
 Eng. Rec., 51: 517. 
 Engng., 15: 411. 
 Eng. News, 29: 137. 
 N.E.W.W., June, 1901. 
 Eng. News, 28: 561. 
 Engng., 9: 369. 
 Eng. News, 28: 578. 
 Eng. News, 30: 149. 
 Eng. Rec., 48: 341. 
 Engng., 4: 122. 
 Eng. News, 32: 440. 
 Eng. News, 15: 254. 
 Engng., 22: 421. 
 Eng. News, 22: 602; 
 
 Engng., i: 107. 
 Eng. News, 10: 423. 
 Engng., 26: 70. 
 Eng. News, n: 118. . 
 Eng. News, 14: 142. 
 Engng., 86: 37. 
 Eng. Mag., 20: 281. 
 N.E.W.W., March, 1899. 
 Engng., 27: 58. 
 Eng. News, 27: 374. 
 Eng. News, 22: 151; 26: 
 
 233- 
 Engng., 38: 319, 384, 472, 
 
 520. 
 
 Engng., 10: 224. 
 Eng. News, 15: 254. 
 Eng. Rec., 50: 638. 
 Engng., 28: 189. 
 
BIBLIOGRAPHY 
 
 687 
 
 Pontiac Pump. 
 
 Recent Practice. 
 
 Richardson Triple Expansion Pump. 
 
 Superheated Steam Test. 
 
 Trenton Allis Pump. 
 
 Water Works Pumps of 1875. 
 
 Worthington Pumps. 
 
 Eng. Rec., 48: 280. 
 Eng. News, 30: 119. 
 Engng., 50: 158. 
 Eng. Rec., 59: 788. 
 Eng. News, 32: 483. 
 
 A.S.C.E., Vol. 4. 
 
 Engng. News, 16: 289; 30: 
 
 230; 27: 167; Engng., 
 
 42: 340. 
 
 RECIPROCATING PUMP DESIGN AND THEORETICAL 
 ARTICLES 
 
 Pump Parts. 
 
 Steel Forgings. 
 
 Specifications for St. Louis. 
 
 Pump Piping. 
 
 Foundations for Pumps. 
 
 Reducing Water Ram in Direct Acting Engine. 
 
 Steel Pipe. 
 
 Gutermuth Valves. 
 
 Ueber Freigehende Pumpenventile. 
 
 Valves on Hydraulic Pumps. 
 
 Packing for Hydraulic Pressure. 
 
 Experimental Study of the Resistance of Flow of 
 
 Water in Pipes, Saph and Schoder. 
 Superheated Steam with Pumps. 
 Friction in Pumping Mains. 
 Flow of Water through Pipes. 
 Test of Air Lift Valves. 
 Flow of Water in Pipes, by Williams, Hubbell, 
 
 and Fenkell. 
 
 A.W.W.A., 1005, 151. 
 N.E.W.W., 12: 120. 
 N.E.W.W., ii : 172. 
 Engr., April i, 1905. 
 Cassiers, 31: 42. 
 N.E.W.W., 15: 493- 
 N.E.W.W., 13: 314. 
 Engng., 79: 391. 
 Z. d. V. d. Ing., March 25, 
 
 1905. 
 
 Am. Mach., April 14, 1910. 
 Am. Mach., Sept. 22, 1910. 
 
 A.S.C.E., 54: 253. 
 A.S.M.E., 21. 
 N.E.W.W., 10: 234. 
 Cassiers, 29: 22. 
 N.E.W.W., u: 51. 
 
 A.S.C.E., Vol. 47. 
 
 ROTARY AND DISC PUMPS 
 
 Behrens Rotary Pump. 
 Bennison Rotary Pump. 
 Boulton and Imray Helical Pump. 
 Mauley Rotary Pump. 
 McFarland Rotary Pump. 
 Oscillating Pump. 
 Clark Patent Rotary Pump. 
 Phillips Rotary Pump. 
 Portland Rotary Pump. 
 Stannah Pendulum Pump. 
 Von Mottoni Pendulum Pump. 
 
 Engng., 
 Engng. , 
 Engng., 
 Engng., 
 Engng. , 
 Engng. , 
 Engng., 
 Engng., 
 Engng., 
 Engng., 
 Engng., 
 
 10 : 200. 
 19: 69. 
 14: 196. 
 
 25: 45i. 
 
 20: 332. 
 
 18: 262. 
 
 44: 187. 
 
 39: 351- 
 
 33* 59- 
 
 23: 56. 
 
 10 : 485. 
 
688 
 
 BIBLIOGRAPHY 
 
 SIMPLEX PUMPS 
 
 Blake Pump. 
 
 Bradfer Pump. 
 Baummann Pump. 
 Cameron Pump. 
 
 Cherry Pump. 
 
 Clarkson Pump. 
 
 Cope & Maxwell Pump. 
 
 Croyden Water Works Pump. 
 Davey Differential Valve. 
 Davison Pump. 
 
 Deane Sinking Pump. 
 
 Decker Bros. Pump. 
 
 Deep Well Pump. 
 
 Direct Acting Simplex Mine Pump. 
 
 Earle Pump. 
 
 Hall Pump. 
 
 Imperial Pump. 
 
 Parker and Weston Pump. 
 
 Plam & Co. Pump. 
 
 Pickering Pump. 
 
 Ramsbottom Pump. 
 
 Shanks Pump. 
 
 Silver Pump. 
 
 Stone Pump. 
 
 Walker Pump. 
 
 Walker and Holt Pump. 
 
 Engng., 20: 37; Eng. News, 
 
 7: 91. 
 
 Engng., Feb. 19, 1904. 
 Engng., 9: 293; 12: 237. 
 Engng., 5 : 2 53 : 4- 213; 
 
 14: 180. 
 Engng., 22: 77. 
 Engng., 18: 210. 
 Engng., 7: 334; ii : 46; 
 
 22: 57; 35- 327- 
 Engng., 24: 356. 
 Engng., 19: 273; 26: 197. 
 Eng., News, 8: 436; 12: 
 
 220. 
 
 Engng., 36: 125. 
 Engng., 16: 371. 
 
 Engng-, 5 : 5 2 8. 
 Engng., 14: 180. 
 Engng., 3: 625. 
 Eng. News, 14: 247. 
 Engng., 24: 29, 36. 
 Engng., 22: 121: 24: 183. 
 Engng., 39: 587. 
 Engng., 20: 366. 
 Engng., 9: 192. 
 Engng., 32: 35. 
 Engng., 21 : 268. 
 Engng., 30: 138. 
 Engng., 20: 44. 
 Engng., ii : 100. 
 
 SPECIAL PUMPS 
 
 Air Lift Pumps. 
 
 Air Lift Pump at High Bridge. 
 
 Air Lift Pump. 
 
 Bernay's Steam Pump. 
 
 Dock Pump. 
 
 Duplex Pump with Single Valve. 
 
 High Pressure Gas Power Station. 
 
 Humphrey's Gas Operated Pump. 
 
 Eng. News, 29: 541; 32: 27. 
 Eng. Rec., 49: 672; A.S.C. 
 
 E., 54: i. 
 
 Eng. News, 50: 675. 
 Engng., 39^ 525- 
 Engng., Feb. 26, 1904. 
 Engng., 42: 168. 
 Eng., Oct. i, 1906. 
 Engng., 1909, 737; 1909, 
 
 512-14; Am. Mach., 
 
 1911, Jan. 5. 
 
BIBLIOGRAPHY 
 
 Hydraulic Pressure Pump. 
 
 Hydraulic Engine Pump for New London. 
 
 Hydraulic Pump Machines. 
 
 Hydraulic Pumping Plant at Gloucester, Eng. 
 
 Hydraulic Ram, Large. 
 
 Hydraulic Rams. 
 
 Natural Gas Pump. 
 
 Noria at Hannah. 
 
 Oil Pumping Station. 
 
 Oil Pumps, Worthington. 
 
 Oscillating Pump (Undulating). 
 
 Oscillating Pump. 
 
 Producer Gas Pumping Plant. 
 
 Application of Gas, Gasoline, and Oil to Pumps. 
 
 Rubber Pump, Hard. 
 
 Smith's Explosion Pump. 
 
 Steam Turbines for Water Works. 
 
 Valves for High Speed Pumps. 
 
 Well Pump without Suction Valves. 
 
 Theory of Air Lift Pump, E.G. Harris. 
 
 Am. 
 
 Eng. News, 31: 278. 
 Eng. News, 29: 65. 
 N.E.W.W., i: 34. 
 Engng., Feb. 12, 1904. 
 A.S.C.E., 54: 159- 
 Cassiers, 28: 65; 
 
 Mach., April 14, 1910. 
 Eng. Rec., 50: 712. 
 Eng. News, 19: 159. 
 Eng. Rec., 57: 676. 
 Engng., 40: 108. 
 Eng. News, 30: 70. 
 Engng., 18: 262. 
 A.W.W.A., 1908, 61. 
 N.E.W.W., 13: 206: Fov 
 
 er, Dec., 1903. 
 Eng. News, 28: 477. 
 Eng. News, May 19, 1910. 
 A.W.W.A., 1905, 302. 
 Engng., 87: 662. 
 Engng., 40: 55. 
 A.S.C.E., 54: i. 
 
 WATER WORKS AND WATER WORKS PUMPS 
 
 Allegheny Leavitt Engine. 
 
 Allis Pumps. 
 
 Atlantic City Station. 
 
 Beam Engines. 
 
 Beam Engine of 1881. 
 
 Belmont Worthington Pump. 
 
 Berlin Water Works. 
 
 Birmingham Water Works. 
 
 Birmingham, Ala., Worthington Pump. 
 
 Boston-Leavitt Pump. 
 
 Boston Allis Pump. 
 
 Brooklyn Engine. 
 
 Brooklyn Pumps of 1870. 
 
 Brooklyn Pumps. 
 
 Brooklyn, Davison Pump. 
 Brooklyn Water Works. 
 Buda Pesth Water Works. 
 Buffalo Holly Pump. 
 Cambridge, Mass., Water Pipe. 
 Cast Iron Pipe. 
 Chicago Pumps of 1866. 
 
 Engng., 41: 33. 
 N.E.W.W., 13: 172. 
 Eng. Rec., 48: 215. 
 Engng., i: 149. 
 Engng., 31: 64. 
 Engng., 15: 411. 
 Engng., 10 : 260. 
 Engng., 40: 298. 
 Eng. News, 27: 368. 
 Eng. News, 28: 578. 
 N.E.W.W., June, 1901. 
 Engng., i: 250. 
 Engng., 9: 369. 
 Engr., May i, 1904; Engng., 
 
 9: 369. 
 
 Eng. News, 12: 220. 
 Eng. News, 25: 225. 
 Engng., 39: 528; 574, 623. 
 Engng., 28: 365. 
 N.E.W.W., n: 121. 
 N.E.W.W., n: 27. 
 Engng., 7: 242.. 
 
690 
 
 BIBLIOGRAPHY 
 
 Chicago Water Works of 1875. 
 Chicago Pumping Station. 
 
 Chicago Allis Pumps. 
 
 Cincinnati Pump. 
 
 Cleveland Plant. 
 
 Cornish Pump of 1881. 
 
 Columbus Gaskill Pump. 
 
 Compound vs. Triple Expansion 
 
 Compensated Gear. 
 
 D'Auria Pump. 
 
 Davy Differential Pump. 
 
 Dean High Duty. 
 
 Development and Peculiarities of Water Works 
 
 Pumps, by I. H. Reynolds. 
 Deep Well Plant. 
 East Jersey Water Supply. 
 English Beam Fly Wheel Engine. 
 Fall River Davison Pump. 
 Foreign Water Supply. 
 Garden City Station, Air Lift. 
 Gas Engine Plant. 
 Hannibal Allis Pump. 
 Horizontal Compound Pump. 
 Kley, Flywheel Beam Pump. 
 Lambeth Beam Engine. 
 Lardner's Point Station. 
 Lawrence Leavitt Pump. 
 Leavitt Pumps. 
 Memphis Vertical Worthington Pump. 
 
 Milwaukee Allis Engine. 
 Moreland Pump. 
 
 Moreland and Thompson Pump. 
 
 Mystic Pump. - , 
 
 Newark Worthington Pump. 
 
 New Bedford Pumps. 
 
 New London Pump. 
 
 Ottumwa, Iowa. 
 
 Paris Farcot Pump. 
 
 Pawtucket Pump. 
 
 Peoria Plant. 
 
 Philadelphia (see Belmont, Lardner's Point). 
 
 Philadelphia Water Works of 1868. 
 
 Port Washington, N. Y. 
 
 Prague Water Works. 
 
 Present Pumping Engine Practice, Reynolds. 
 
 Engng., 19: 30. 
 
 Eng. News, 51: 485; Eng 
 
 Rec., 48: 120. 
 Eng. News, 23: 506. 
 Engng., 3: 53*- 
 Eng. Rec., 49: 348. 
 
 Engng-, 3 2: 2O 9- 
 Eng. News, n: 118. 
 N.E.W.W., 13: 218. 
 Engng., 44: 349, 409. 
 A.W.W.A, 1905, 304. 
 Engng., 22: 421. 
 Eng. News, 29: 137. 
 
 A.S.C.E., 1904, No. 95. 
 Engng., Feb. 25, 1910. 
 N.E.W.W., 8: 18. 
 Engng., 44: 457. 
 Eng. News, 10: 423. 
 N.E.W.W., 9: 109. 
 Eng. Rec., 59: 535. 
 Eng. Rec., 53: 196. 
 Eng. News, 14: 142. 
 Engng., 24: 10. 
 Engng., 15: 50. 
 Engng., 2: 273. 
 Eng. Rec., 52: 315. 
 Engng., 27: 58. 
 N.E.W.W., 9: 163. 
 Eng. News, 22: 151; 26: 
 
 2 33- 
 
 A.E.S., 14: 24: 
 Engng., 38: 319, 384, 472, 
 
 520. 
 
 Engng., 6: 544. 
 Eng. News, 32: 176. 
 Engng., 10: 224. 
 Eng. Rec., 49: 797. 
 N.E.W.W., 7: 148. 
 Eng. Rec., 53: 430. 
 Engng., 26: 70. 
 Engng., 28: 189. 
 Eng. News, 28: 58. 
 
 Engng., 6: 267. 
 Eng. Rec., 52: 205. 
 Engng., 44: 200, 396. 
 N.E.W.W., 13: 172. 
 
BIBLIOGRAPHY 
 
 691 
 
 Pumping Machinery for Water Works, by F. H. 
 
 Pond. 
 
 Pipes, see Syracuse, Cast Iron, East Jersey, Cam- 
 bridge. 
 
 Richardson Triple Expansion Pump. 
 Sidney Pump. 
 St. Louis Beam Engine. 
 St. Louis Pumps of 1874. 
 St. Louis Water Works of 1894. 
 St. Louis Plants. 
 St. Paul Allis Engine. 
 Small Pump Plants, Farr. 
 Syracuse Steel Pipe. 
 Schenectady Plant. 
 Toledo Water Works. 
 
 Vienna Exposition Pump. 
 
 Vienna Water Works Pump. 
 
 Vertical Pump. 
 
 Washington, D. C, Plant. 
 
 Water Supply for Small Cities. 
 
 Water Works, Ancient and Modem. 
 
 Water Works Pumps of 1875, Sizes, Costs, Duty. 
 
 Windsor, Ont, Plant. 
 
 Worthington High Duty Pump in England. 
 
 Worthington Pump. 
 
 Zurich Plant. 
 
 Eng. News, 13: 340. 
 
 Engng., 50: 158. 
 Engng., 24: 10; 42: 574. 
 Eng. News, 16: 93. 
 
 En gng., 31 : 143, 200. 
 A.E.S., 1894. 
 Eng. Rec., 49: 700. 
 Eng. News, 15: 13. 
 A.W.W.A., 1907. 
 N.E.W.W., 8: 40. 
 Eng. Rec., 51: 640. 
 Eng. Rec., 52: 377; Power, 
 
 Feb. 21, 1911. 
 Engng., 16: 242. 
 Engng., 28: 432. 
 Engng., 39: 485. 
 Eng. Rec., 53: 64. 
 N.E.W.W., 5: 83. 
 Engng., 21 : 502; 22: 7 
 A.S.C.E, 4: 369. 
 Engr., Nov. 16, 1903. 
 Eng. News, 21: 87. 
 N.E.W.W., 13: 229. 
 Eng. News, 32: 34. 
 
INDEX 
 
 Absolute path of water, 628 
 
 Acceleration, 190 
 
 Acceleration and velocity curves, 
 
 674 
 
 Acceleration diagram, 241 
 Acceleration of parts, 333 
 Admiralty pumps, 137 
 Ahrens pump, 277 
 Air chambers, 242 
 
 advantage of, 243 
 
 pressure, 247 
 
 size of, 248 
 
 suction, 305 
 Air compressor, 520 
 Air-lift pump, 122 
 Air-lift pumps and pneumatic 
 
 pumps, 512 
 Air-lift well tops, 514 
 Air valve, 489 
 Air pump, 36 
 
 Alberger, 465 
 
 cards, 459 
 
 dry, 464 
 
 Edwards, 461, 462 
 
 Mullen, 463 
 
 navy, 460 
 
 Air pumps for beer racking, 498 
 Air pumps, size of, 458 
 Aix-la-Chapelle mine pump, 659 
 Alberger Pump Co., 575, 650 
 Alberger centrifugal condenser, 583 
 Alberger multistage turbine pump, 
 
 576 
 Alberger standard volute pump, 
 
 vertical shaft, 581 
 Alberger two-stage volute pump, 
 
 58i 
 Allis-Chalmers, 112 
 
 Allis-Chalmers duplex Riedler ex- 
 press pump, 163 
 Allis-Chalmers pump, 438, 590 
 Allis-Chalmers Riedler valve, 294 
 Allis pumps at the Baden station, 
 
 _ 435 
 
 Allis pumps at Bissel's Point, 436 
 Allis pump of Milwaukee, 286 
 Allis screw pump, 116 
 American Fire Engine Go's, metro- 
 politan engine, 277 
 Andrews, 45, 46 
 Archimedean screw, 14 
 Area curve through bucket, 627 
 Areas at entrance and discharge, 
 
 608 
 
 Area scales, 339 
 Arrangement of pump, 318 
 Arrangement of vanes, 608 
 
 Bach, 208, 210 
 Back vanes, 628 
 Balancing pumps, 400 
 Ballast pump, 141 
 Ball valve, 290 
 Barlow's formula, 309 
 Barr, 3*6 
 
 Bearing power of soil, 402 
 Bearings, 391 
 
 design of, 389 
 Behrens pump, 113 
 Beighton, Henry, 19 
 Belmont water- works, 77 
 Belted volute pump, 567 
 Benjamin, C. H., 562 
 Bibliography, 679 
 Bilge pump, 141 
 Blake, 45, 62 
 
 693 
 
694 
 
 INDEX 
 
 Blake pump, 174, 176, 177 
 
 Boiler-feed pump, 134 
 
 Borsig valve, 301 
 
 Boulton, Mathew, 36, 39 
 
 Box end design, 386 
 
 Braithwaite, 52 
 
 Brass liner, 266 
 
 Brooklyn high pressure station, 
 
 596, 655 
 
 Brooklyn station, 440 
 Brotherhood engine, 119 
 Bucket, 4, 260 
 Bucket, Roman, 10 
 Bucket pump, 129, 130 
 Buffalo balanced two-stage pump, 
 
 587 
 
 Buffalo pump, 587, 588 
 Buffalo vertical underwriter fire 
 
 pump, 589 
 Bull, Wm., 40 
 Bull Cornish engine, 67 
 Burnham pump, 183, 184, 281 
 Burnham pump valve, 300 
 Bushing rings, 565 
 Butterfly valves, 204, 289 
 
 Cage, pressure valve, 300 
 
 Calibrated nozzle, 590 
 
 Galley, John, 30 
 
 Cameron pump, 165, 166, 169 
 
 Cameron valve, 292 
 
 Cards, combined, 334 
 
 Carpenter, 403, 411 
 
 Central outside packing, 132 
 
 Central Park Avenue station, 434 
 
 Centrifugal pressure head, 539 
 
 Centrifugal pumps, 535 
 
 Centrifugal pump dimensions, 570 
 
 Centrifugal pump, efficiency cf, 
 
 594 
 
 Centrifugal pumps for special pur- 
 poses, 646 
 
 Chaplets, 10 
 
 Check valve, 490 
 
 Chesney, Col., 6 
 
 Cincinnati pump, 286, 428 
 
 Circular arc vanes, 618 
 
 Clack valves, 208, 220 
 
 Clack valve, hinged, 288 
 
 leather, 288, 289 
 
 double, 289 
 
 metal, 289 
 
 Clack valves, rectangular, 281 
 Clearance, 235, 257 
 Cloth dryer, 38 
 Cock valves, 204 
 Combined cards, 334 
 Combined entrance and discharge 
 
 diagram, 549 
 
 Compound direct-acting pump dia- 
 gram, 676 
 Compound outside end packed 
 
 pressure pump, 144 
 Compound pumps, 71 
 Compound outside center packed 
 
 boiler feed pump, 135 
 Compressor with clearance, 522 
 Concrete foundations, 402 
 Condensers, 36, 406 
 
 Buckley, 86 
 
 Connections, water pipe, 319 
 Conical valve, 289, 290 
 Connecting rods, 361, 367 
 
 design, 370 
 
 marine end, 369 
 
 strap end, 368 
 Connecting-rod box, 382 
 Connecting-rod ends, design, 386 
 Consolidated Cal. shaft, 66 1 
 Controllable valve arrangement, 489 
 Cooling water, quantity, 457 
 Core hole cap, 310 
 Corliss, 74, 78, 79, 81, 95 
 Corliss valves, 95, 96 
 Cornish, 40, 41, 42, 43 
 Cornish engine, 86, 87 
 Cornish valves, 95 
 Coupling, 590 
 
 Covers, manhole or handhole, 310 
 Crank pin design, 374 
 Critical speed, 636 
 Cross-head design, 370 
 Cross-head, 274, 361, 363, 364 
 two rod, 365 
 pump end, 365 
 cross-head pin design 3 7 1 
 
INDEX 
 
 695 
 
 Cup leathers, 265 
 
 Cup-leather packing, 261 
 
 Ctesibius, 15 
 
 Curves of area, 622 
 
 Curves of quantity, 622 
 
 Curves of variation of coefficient 
 
 of friction, 561 
 Cylinder, design of, 307, 310, 
 
 355 
 
 Cylinder openings, 310 
 Cylinder ratio, 329 
 Cylinders, sizes, 330 
 
 Cylinders, steam, 348 
 
 triple expansion, 346 
 
 Dalby, W. E., 400 
 
 d'Auria, Luigi, 104 
 
 Davey, Henry, 105 
 
 Davey compensator, 106 
 
 Davies, J. D., 98 
 
 Davison, 97 
 
 Davidson pump, 180, 181 
 
 Davidson vertical duplex pump, 
 
 138 
 Davidson horizontal steam cylinder 
 
 for deep-well pump, 156 
 Dean Bros, pump, 186, 187 
 Deane pump, 176, 177, 178 
 Deane triplex vertical single-acting 
 
 power pump, 161 
 Deep- well pump, 153, 278 
 de Lorme, Philibert, 126 
 Desagulier, 33 
 Design of bearings, 389 
 Design of centrifugal pumps, 602 
 Design of cross-head, connecting- 
 rod shaft, 370 
 
 Design of cylinder, 307, 355 
 Design of parts, 260 
 Design of piston rod, 361 
 Design of springs, 316 
 Design of steam piping, 403 
 Details steam end, 344 
 Diagrams of velocity, 237, 239, 241 
 
 acceleration, 241 
 
 space, 238, 239, 241 
 Diagram, scales, 340 
 Diameters of impellers, 610 
 
 Differential bucket pump, 131 
 
 Differential plunger pump, 131 
 
 Diffuser, 536, 574, 631 
 
 Diffuser vanes, 536 
 
 Diffusion chamber, 630 
 
 Direct-acting water works pumps, 
 152 
 
 Discharge chamber, 265 
 
 Discharge cone, 488 
 
 Discharge, rate of, 192 
 
 Discharge, velocity of, 189 
 
 Disc valve, 291 
 
 Doon, ii 
 
 Double-acting plunger pump, 131 
 
 Double-beat valve, 296 
 
 Double-flow pump, 537 
 
 Double-flow volute pump, 568 
 
 Double leather packing, 262 
 
 Double-ported valve, 294, 296 
 
 Drain cocks, 493 
 
 Dredging pumps, 652 
 
 Driving valves, mechanism, 346 
 
 Dry air pumps, 464 
 
 Dry dock units, 647 
 
 Dunkerley's values for shafts, 643 
 
 Duplex pumps, 64, 133, 134, 344 
 
 Duty trial of test of pumping en- 
 gine, 410 
 
 Duties, 43 
 
 Dynamics of steam end, 324 
 
 Dynamics of water end, 189 
 
 Effect of changing quantity with 
 
 fixed speed, 555 
 Efficiency, 418 
 
 Efficiency of centrifugal pump, 594 
 Efforts, tangential, 338 
 Electric sinking pump, 150 
 Emerson pump, 508 
 Engine, Cornish, 86 
 
 horizontal fly-wheel, 87 
 
 rotary 36, 38 
 
 double-acting, 38 
 
 trunk, 38 
 
 Ericsson, Capt. John, 52, 53 
 Euler, 43 
 Eve, J., 48, 49 
 Expansive use of steam, 36, 38 
 
696 
 
 INDEX 
 
 Explosion pump, 128 
 Express pump, 163 
 Express pumps in series, 663 
 
 Fairbanks-Morse deep-well pump, 
 
 I 53. J 54 
 
 Fairbanks-Morse pump, 270, 271 
 Fielding, 105 
 Filling ring, 565 
 Fire pump, Ahrens, 277 
 
 Amer. Fire Engine Co., 277 
 
 double-acting, 275 
 
 Metropolitan, 276 
 Fire pumps, workmanship, 468 
 
 duplex only, 468 
 
 sizes, 468 
 
 capacity, 469 
 
 speed, 469 
 
 capacity plate, 470 
 
 strength of parts, 470 
 Fire engine, Merryweather's, 299 
 Fire engine, shop inspection, 470 
 
 steam cylinders, 471 
 
 steam ports, 471 
 
 steam clearance, 471 
 Fire-pump, French, 51, 52 
 Flanges, 311 
 Flange, shrunk, 404 
 
 F, screwed, 404 
 
 F, welded, 404 
 
 Flash wheels, or scoop wheels, 55 
 Fluctuation factor, 341 
 Fly-wheel design, 393 
 Fly-wheel energy, 340 
 Fly-wheel, sectional, 394 
 
 size, 342 
 
 Follower plate, 266 
 Foot bearing, 536 
 Foot or suction valve, 278 
 Foot valve, 305, 306, 492 
 Forces in direct- acting rod mine 
 
 pump, 670 
 Force pump, 15 
 
 Forms of centrifugal pumps, 565 
 Forms of impeller, 613 
 Foundations, of concrete, of piles, 
 
 402 
 Fountain, 16 
 
 Fowler, Geo., 651 
 
 Frame design, 400 
 
 Frames for vertical engine, 400 
 
 Frames, 391, 400 
 
 Francis, 593 
 
 Friction losses, 197 
 
 Friction in pipes, 200 
 
 Friction at bearings and stuffing- 
 boxes, 559 
 
 Friction of water on back of im- 
 peller, 563 
 
 Frizell, J. B., 122 
 
 Gas engine for deep-well pump, 156 
 
 Gaskell, H. F., 82, 83 
 
 Gate valve, 204 
 
 Gate valve, Ludlow, 405 
 
 Gauge, mercury steam, 38 
 
 water, 38 
 
 Gelpcke-Kugel, 590 
 General service pumps, 140 
 Giffard, H. J., 125, 126 
 Globe valve, 404 
 Gould steam fire pump, 119 
 Governor, engine, 38 
 Graff, Fred, 70, 71, 72 
 Graphical method of centrifugal 
 
 pump design, 611 
 Green, D. M., Prof., 96 
 Guest, J. J. 381 
 Gutermiith valve, 301 
 Gwynne, John, 45 
 Gwynne, J. and H., no 
 
 Hague, C. A., 94, 95 
 Valve method, 219 
 Hammer, steam, 38 
 Hand fire pump, 12 1 
 Hand holes, 270 
 Harris, E. G., 124 
 Harris pneumatic pumps, 425, 531 
 Hartmann and Knoke, 213, 215 
 Head limit, 233 
 Heat determination, 413 
 Helical pump, 118 
 Helical rings, 290 
 Helicoidal pump, 118 
 Heisler pump, 105 
 
INDEX 
 
 697 
 
 Hero, 51 
 
 Hero of Alexandria, 14 
 
 Hesse, F. G., 563 
 
 High-duty pumps, 428 
 
 High-pressure pumping stations: 
 
 Brooklyn, 655 
 
 New York city, 656, 657 
 
 Philadelphia, 654 
 Hinged clack valve, 288 
 Holly, Birdsill, 68 
 Holly Manufacturing Co., 82 
 Holly pump at Boston, 437 
 Holly pump at Washington, D. C., 
 
 439 
 
 Holly triple expansion pump, 108 
 Hornblower, Jonathan, 39, 40 
 Horizontal fly-wheel pumping engine, 
 
 87 
 
 Horse-power, 167 
 Hose valves, 488 
 Humphrey, explosion pump, 128 
 Humphrey, H. A., 126 
 Hydraulic ram, 53, 500 
 Hydraulic presses, 310 
 Hydraulic pressure pumps, 496 
 
 Impeller, 536, 565 
 Indicator, steam-engine, 38 
 Indicator cards, actual, 331 
 
 combined, 335 
 Inertia, effect of, 230 
 Interference of blunt vanes, 554 
 Injectors, 125 
 
 Injector and pulsometer, 503 
 Involute curves, 616, 617 
 
 Jackets, 36, 418 
 Jacobus, J. S., 339 
 Jeanesville mine pump, 668 
 Jeanesville valve pots, 669 
 Jordan, Johann, 43 
 Joseph's well, 7 
 
 Katweh, 10 
 
 Kent, 411 
 
 Kent's pocket-book, 417 
 
 Knsass, Strickland L., 505, 506 
 
 Knowles, 62 
 
 Knowles express pump, 665, 666 
 Knowles pump, 172, 173, 174 
 Knowles simplex pump, 134 
 Knowles steam pump works, 665 
 Knowles underwriters' pump, 142, 
 
 144 
 Knowles vertical duplex electric 
 
 sinking pumps, 149 
 
 Lagging, 36 
 
 Lambeth Water Works, 88 
 
 Lame formula, 309 
 
 Larner pump, 584 
 
 Larner, C. W., 582 
 
 Lat, 4 
 
 Latha, 10 
 
 Lawrence pump, 283 
 
 Leakage, 557 
 
 Leather clack valve, 288, 289 
 
 Leavitt, E. D. jr., 70, 71, 77, 78, 
 
 107, 281, 347 
 Lewecki, 564 
 Liner, independent, 351 
 Lloyd, 45 
 
 London Bridge Water Works, 18 
 Loss in bends, 202 . 
 Loss in passages, 199 
 Loss in pipes, 198 
 Loss in valves, 202 
 Loss due to inertia, 206 
 Loss due to sudden contraction, 205 
 Loss due to velocity changes, 204 
 Losses in centrifugal pumps, 552 
 Loss through valve, 198, 208 
 Ludlow gate valve, 405 
 Luitwieler pump, 155, 157, 158 
 
 MacFarland's rotary pump, 114 
 McCarty, 43, 44, 45 
 Maltby, F. B., 651 
 Manhole or handhole cover, 309 
 Manhole cover, 288 
 Mair, J. I., 101 
 Marine boiler feed pump, 136 
 Marine end, 369 
 
 Marsh pump, 265, 266, 267, 268, 
 269 
 
698 
 
 INDEX 
 
 Marsh steam pump, 169, 170, 187 
 
 Marsh valve, 293 
 
 Massachusetts pump, 44, 45, no 
 
 Mean pressure, 326 
 
 Measuring pressure, 412 
 
 Measuring water pumped, 411 
 
 Mechanical efficiency, 325 
 
 Mental, 10 
 
 Merriman's hydraulics, 593 
 
 Merryweather's fire engine, 299 
 
 Metal clack valve, 289 
 
 Metal disc valve, 292 
 
 Metallic packing, 355 
 
 Method for volute pumps, 613 
 
 Mexican union mine pump, 659 
 
 Milk pump, 266 
 
 Milwaukee pump, 285 
 
 Mine pumps, 146, 659 
 
 Aix-la-Chapelle, 659 
 
 consolidated Cal. shaft, 66 1 
 
 express pumps, 663 
 
 German, 66 1 
 
 Jeanesville, 668 
 
 Mexican union, 659 
 Mine pump, Scran ton pattern, 149 
 Mixed flow pumps, 619 
 Modern fire engine, American, La 
 
 France Co., 122 
 Modern forms of pumps, 129 
 Montgolfier, 54, 55 
 Moody, Prof. L. P., 599 
 Moreland's compound steam end, 
 
 80, 81 
 
 Moreland, Rich., 68 
 Moreland, 96, in 
 
 Sir Saml., 24 
 
 Morrys, Peter, 18, 20, 21 - 
 Mot, n 
 
 Moving parts, weights of, 332 
 Multistage compression, 522 
 Multiple expansion, 326 
 Muntz metal, 314 
 
 Neumann, 547 
 Neumann's curves, 548 
 Newcomen, Thos., 28, 29, 30, 33 
 New York high-pressure station, 
 656, 657 
 
 Non-aligning ring oiling bearing, 
 
 559 
 
 Norberg quadruple-expansion en- 
 gine, 107 
 
 Noria, 2, 4 
 
 Number of revolutions, 255 
 
 Number of stages, 602 
 
 Odell, 564 
 
 Oil cellar, 566 
 
 Oil rings, 566 
 
 Oil thrower, 566 
 
 Ontario pump, 284 
 
 Openings, 310 
 
 Outboard bearing, 566 
 
 Outside packed plunger boiler feed 
 
 pump, 136 
 Outside packed pump, 169 
 
 Packings, 36, 314 
 Packing, U-leather, 263 
 
 plunger, 260, 263 
 
 cup leather, 261 
 
 double leather, 262 
 Paecottah, 4 
 Pallets, 10 
 
 Papin, Denys, 27, 43 
 Pappenheim, 47 
 Parabolic vanes, 615 
 Paris pump, 88 
 Parts, design of 260, 
 Peabody, 409 
 Persian wheel, 7 
 Philadelphia high-pressure station, 
 
 654 
 
 Phillips rotary pump, 115 
 Piles, 402 
 Pipe friction, 200 
 Pipe loss, 198 
 Pipes, size of, 256 
 Piston, 260, 353, 354 
 
 pressure, 197 
 Piston proportions, 358 
 Piston pump, 129, 130 
 Piston ring design, 358 
 Piston rods, 314 
 Piston rod design, 361 
 Pistons, water, 313 
 
INDEX 
 
 699 
 
 Piston with sectional ring, 354 
 Pitot tube, 412, 590 
 Plunger, packed, 260 
 Plunger and ring, 260, 271 
 Plunger and ring packing, 132 
 Plungers, outside packed, 274 
 
 center, oustide packed 274 
 Plunger packing, 260, 263 
 Plunger pump, 129, 130 
 Pohle, Julius, 122 
 Pointed vanes, 554 
 Points of leakage, 558 
 Power pumps, horizontal duplex, 
 
 1 60 
 
 Power head deep- well pump, 162 
 Preparation for test, 419 
 Press, letter copy, 38 
 Prescott duplex outside - packed 
 plunger pot-form boiler-feed 
 pumps, 138 
 
 Prescott steam sinking pump, 150 
 Pressure cylinder, 310 
 Presses, hydraulic, 316 
 Press and pump, 500- ton, 497 
 Pressure on discharge without 
 air chamber, 237 
 
 with air chamber, 243, 247 
 Pressure, measuring, 412 
 Pressure pumps, 145, 279, 280, 
 
 281 
 
 Pressure, resultant, 238, 336 
 Pressures, terminal, 328 
 Pressure valves, 298, 299 
 Priming, 489 
 Priming tank, 492 
 Pulsometer, 126 
 Pump, air-chamber, 303, 304 
 Pump arrangement, 318 
 Pump, dredging, 652 
 Pumping engines, test of, 410 
 Pump, Fairbanks-Morse, 270, 271 
 
 Worthington, 272, 273 
 . Burnham, 281 
 
 Marsh, 265 
 
 milk, 266 
 Pumps, condenser, 456 
 
 combined air circulating, 457 
 Pumps, pressure, 279, 280, 281 
 
 Pump with air chamber, 242 
 Pump rods, 278 
 Pump, sewage, 282 
 
 Cincinnati, 286 
 
 Snow, 287 
 
 Lawrence, 282 
 
 Ontario, 284 
 
 Milwaukee, 285 
 
 Riedler, 451, 452 
 Pump specifications, underwriters', 
 
 468 
 Pump, sewage, 466 
 
 underwriters', 467 
 Pumps, turbine, 649 
 Pumping machinery, special, 456 
 
 Quadruplex pump, 69 
 Quimby screw pump, 127 
 
 Railroad pump, 277 
 
 Ramelli, Agostino, 18, 20, 23, 28, 
 
 47 
 
 Ramseye, David, 18, 23 
 Rankine, 381 
 Rateau, 584 
 pump, 585 
 
 Reciprocating pump, 51 
 Reciprocating parts, weight of, 
 
 333 
 
 Rectangular clack valves, 281 
 Reheaters, 329 
 Relief valve, 488 
 Resultant pressure, 235, 336 
 Return chamber, 588 
 Revillion, 51, 52 . 
 Revolutions, number of, 258 
 Reynolds pump, 92, 95 
 Reynolds, I. H., valve method, 219 
 Riedler pump, 163, 164, 451 
 Riedler valve, 294 
 Results of actual test, 419- 
 Roebuck, Dr., 34 
 Root rotary pump, 116 
 Rotary pumps 
 
 Behrens, 113 
 
 MacFarland's, 114 
 
 Phillips; 114 
 
 Root, 116 
 
700 
 
 INDEX 
 
 Rotary pumps 
 
 Silsby, 115 
 
 "VYilkins, 115 
 
 Rotary pump, Ramelli's sixteenth 
 century, 47 
 
 Trotter's, four-bladed, 50 
 
 Watt's Eve's, 48 
 Rubber check valve, 491 
 Rubber valve, 293 
 Runner, 536 
 
 Safety valve, 488 
 
 Sakias, 7 
 
 Savannah pump, 62, 63 
 
 Savery, Thos., 25, 26, 27, 28, 33 
 
 Scales of diagrams, 340 
 
 Scales, area, 339 
 
 Schwartzkopff, 585 
 
 Louis, in 
 Scoop wheels, 55 
 Screw, Archimedean, 14 
 Sectional fly-wheel, 394 
 Section of impeller, 614 
 Section of mixed flow impeller, 
 
 620 - 
 Self-aligning ring oiling bearing, 
 
 560 
 
 Sellers injector, 503 
 Serviere, 46, 49, 116 
 Sewage pump, 282, 466 
 Shaft design, 370, 633 
 Shaft sketch, 376 
 Shields, Geo., 67 
 Shrunk flange, 404 
 Shadoof, 2 
 
 Silsby rotary pump, 115 
 Simplex boiler feed pump, 135 
 Simplex pumps, 133, 165, 172 
 Simpson's pumping engine, 88 
 Simpson & Co., 649 
 S ; nking pump, 148 
 Size cylinders, 330 
 Size of air chamber, 248 
 Size of air pumps, 458 
 Size of pipes, 254 
 Size of valves, 213 
 Smeaton, John, 31, 33 
 Snow compound pump, 108 
 
 Snow pump end, 287 
 Soil, bearing power, 402 
 Somerset, Edward, 23, 24 
 Space diagram, 238, 239, 241 
 Specific speed, 598 
 Specific speed curves, 603 
 Speed, critical, 636 
 Speed, specific, 598 
 Spindle, valve, 292 
 Split-casing pumps, 572 
 Spring-controller valves, 270 
 Spring design, 316 
 Springs, 316 
 Stations : 
 
 Brooklyn, 440 
 
 Cincinnati, 428 
 
 Ferrara, 648 
 
 High Bridge, 444 
 
 Jersey City, 454 
 
 Kinnickinnic River, 445 
 
 Lardner's Point, 447 
 
 Memphis, 441 
 
 Zurich, 443 
 Steam cylinder, 348 
 Steam end details, 344 
 Steam end dynamics, 324 
 Steam piping, 403 
 Steam piston, 352 
 Steam valve, 345 
 
 Steam valve of duplex pump, 344 
 Stodola, 564 
 Strainer, 305, 306 
 Strap end connecting rod, 368 
 Strap end design, 388 
 Stroke ratio, 255 
 Stuffing-box, 262, 263, 266, 315 
 Suction air chamber, 305 
 Suction bearing, 566 
 Suction chamber, 265 
 Suction head, 565 
 Suction pipe, 536 
 Sulzer, 583 
 Sulzer pump, 584 
 Swape, 4 
 
 Table of lost head, 607 
 
 Taboot, 7 
 
 Tangential effort construction, 337 
 
INDEX 
 
 701 
 
 Tangential efforts, 338 
 
 Tank pumps, 140 
 
 Taper threads, 278 
 
 Terminal pressures, 328 
 
 Test at Lardner's Point, 419 
 
 Test curves, 595,. 596, 597, 607, 
 
 667 
 
 Test data, 417 
 Test for: 
 
 acceptance, 493 
 internal friction, 494 
 strength and tightness, 494 
 internal leakage or slip, 495 
 maximum working pressure, 495 
 maximum delivery, 495 
 Testing centrigufal pumps, 590 
 Test of air lift pump, 515, 518 
 Test of internal friction, 494 
 Test of pumping engines, 411 
 Test of strength and tightness, 
 
 494 
 
 Test precautions, 416 
 Test results, 453 
 Thames-Ditton, 71 
 Thompson, David, 68 
 Thurston, R. H., 107 
 Thrust bearing, 566, 575 
 Tobin bronze, 143, 314 
 Trevithick, Rich., 40 
 Triangular weir, 594 
 Triple expansion pump, 95, 96 
 Triple expansion cylinders, 346 
 Triplex pump, 133,159 
 Trirotor volute pump, 572 
 Trombone frames, 273 
 Trotter, John, 49 
 Turbine pump, 573 
 Two-stage pump, 112 
 Twin cylinder casting, 310 
 Tympanum, 13 
 
 Underwriters' fire turbine type, 
 
 579 
 
 U-leather packing, 263, 265 
 U-leathers, 265 
 Underwriters' pump specifications, 
 
 468 
 Underwriters' pumps, 141 
 
 Unwin, 358, 396, 564 
 
 Valve acceleration, 216 
 Valve, air, 489 
 Valve backing, 292 
 Valve, ball, 290, 291 
 
 Borsig, 301 
 Valve box, 286, 297 
 Valve, Burnham, 299, 300 
 
 butterfly, 204, 289 
 Valve cage, 286 
 Valve, Cameron, 292 
 Valves, check, 490. 
 
 clack, 208, 220 
 
 cock, 204 
 Valve, conical, 289 
 
 dash relief, 74 
 Valve deck c.over, 265 
 Valve deck-plate, 265 
 Valve design, Hague's method, 219 
 
 Reynolds' method, 219 
 Valve dimensions, 220 
 Valve, disc, 291 
 Valve discs, 317 
 Valves, double beat, 296 
 Valve, double-ported, 294, 296 
 
 foot, 305, 492 
 
 foot or suction, 378 
 
 gate, 204 
 Valve gear, 349 
 
 pressure control, 350 
 Valve, globe, 404 
 
 Gutermuth, 301 
 Valve handle, 490 
 Valves, head, 351 
 Valve loss, 202 
 Valve, Ludlow gate, 405 
 
 Marsh, 293 
 Valve movement, 217 
 Valve, multiported, 224 
 
 positive water and steam, 59 
 Valve pot, 300 
 Valve, pressure, .298 
 Valves, rectangular clack, 281 
 Valve, resistance, 229 
 Valves, relief, 488 
 Valve, relief motion, 60 
 Valves, resistance of, 224,. 226 
 
INDEX 
 
 Valve, Riedler, 294 
 Valve rod yoke, 345 
 Valve, rotary steam, 74 
 
 rubber, 293 
 
 safety, 488 
 
 Valves and springs, 316 
 Valves, hose, 488 
 Valve spindle, 292 
 Valves, spring-controlled, 270 
 Valve, spring- thrown, 57, 59 
 Valves, steam, 345 
 Valve weighted, 295 
 
 Witting's, 298 
 Vane curves, 613 
 Velocity diagram, 238, 239, 241 
 Velocities of discharge, 199, 541 
 Velocities at entrance, 541 
 Venturi meter, 412, 590 
 Vertical pump, 101 
 da Vinci, Leonardo, 17, 28 
 Vitrio, 126 
 
 Volute casing^ 536, 565, 632 
 Volute centrifugal pump, 565 
 
 Warren steam pump, 172, 173 
 Warren Steam Pump Co., 499 
 Waterbury Farrel Foundry & Ma- 
 chine Co., 497 
 Water cyliners, 260 
 Water end, dynamics, 189 
 Water-pipe connections, 319 
 Water pistons, 313 
 Water works pump, 151, 153 
 Water works, 428 
 
 Allegheny, 93 
 
 Berlin, 89 
 
 Brockton, 108 
 
 Brooklyn, 63 
 
 Buffalo, 70 
 
 Cambridge, 63 
 
 Charleston, 63, 67 
 
 Cincinnati, 67,^428 
 
 Dunkirk, 68 
 
 Eastbourne, 68, 8 1 
 
 East London, 89 
 
 Fall River, 101 
 
 Hannibal, 96 
 
 Lambeth, 89 
 
 Water works, Lardner's Point, 445 
 
 Lawrence, 78 
 
 Lockport, 68, 94 
 
 Lynn, 70, 71 
 
 Milwaukee, 93 
 
 Newark, 77 
 
 Pawtucket, 78, 79 
 
 Philadelphia, 67, 72, 77 
 
 Pittsburg, 107 
 
 Providence, 88 
 
 Rochester, 70 
 
 St. Maurs, 88 
 
 Saratoga Springs, 82 
 
 Savannah, 63 
 Watt.J., 21,33, 38 
 Weighted valve, 295 
 Weights of moving parts, 332 
 Weight of reciprocating parts, 
 
 333 
 
 Weise & Monski pump, 585 
 Welded flange, 404 
 Wheeler Condenser & Engineering 
 
 Co., 456 
 
 Wheeler system, 517 
 Whirlpool chamber, 537 
 Whitehurst, 53 
 Wilkin, J. T., 115 
 Witting's metallic valves, 298 
 Wood 45-inch centrifugal pump, 
 
 650 
 
 Wood propeller pump, 117 
 Wood, R. D., 649 
 Worcester, Marquis, 23, 24 
 Work in centrifugal pumps, 543 
 Worthington ballast pump, 143 
 Worthington beer-racking pump, 
 
 498 
 Worthington eight-stage turbine 
 
 mine pump, 577 
 Worthington fireboat turbine 
 
 pump, 578 
 Worthington four-stage boiler feed 
 
 pump, 579 
 Worthington high speed pump, 
 
 580 
 
 Worthington mine pump, 147 
 Worthington packed-plunger 
 
 pump, 139, 148 
 
INDEX 
 
 703 
 
 Worthington piston pump, 141 
 
 Worthington, 58, 59, 
 
 pump, 57, 60, 61, 62, 63, 67, 
 73, 75, 76, 77, 85,95, 96, 99, 
 101, 137, 139, 141, 165, 272, 
 
 273 
 Worthington pumps, Fall River, 
 
 Mass., 434 
 Worthington ten-stage pump, 577 
 
 V/orthington three-stage turbine 
 
 pump, 586 
 Worthington trirotor volute pump, 
 
 572 
 Worthington turbine sinking pump, 
 
 580 
 Worthington volute pump, 562 
 
 Zig-zag balance, 1 1 
 
This 
 
 DEC 121 
 MAY 241953 
 
 Jfi 
 
 fl 
 
Enci 
 
 neering 
 
 UNIVERSITY OF CALIFORNIA UBRARY *Jl : -,'