Engineering Science Series ENGINES AND BOILERS ENGINEERING SCIENCE SERIES EDITED BY DUGALD C. JACKSON, C.E. PROFESSOR OF ELECTRICAL ENGINEERING MASSACHUSETTS INSTITUTE OF TECHNOLOGY FELLOW AND PAST PRESIDENT A.I.E.E. EARLE R. HEDRICK, Ph.D. PROFESSOR OF MATHEMATICS, UNIVERSITY OF MISSOURI MEMBER A.S.M.E. ENGINES AND BOILERS BY THOMAS T. EYRE DEAN, COLLEGE OF ENGINEERING STATE UNIVERSITY OF NEW MEXICO, FORMERLY ASSISTANT PROFESSOR OF MECHANICAL ENGINEERING, PURDUE UNIVERSITY J?eto gorfe THE MACMILLAN COMPANY 1922 All rights reserved PRINTED IN THE UNITED STATES OF AMERICA COPYRIGHT, 1922, BY THE MACMILLAN COMPANY Set up and electrotyped. Published August, 1922. . - FORESTRY Press of J. J. Little & Ives Co. New York / J ^ , 1 Yo ^es^vv PREFACE This text book on Engines and Boilers is intended for use in engineering schools which offer an elementary course in Heat Engines. No attempt has been made to cover the more advanced work in Thermodynamics, or to give an exhaustive treatment of the subject of Heat Power. This work is the result of the author's experience during the several years that he taught classes in Engines and Boilers and in allied subjects at Purdue University. Much of the material was given to the students first in lectures, and later in the form of mimeographed notes. It is now presented in book form with the hope that it may be of value in other engineering schools. At the end of the book a list of representative problems is given. It has been the author's experience that the student obtains a better understanding of the subject if he is required to work problems related to the matter in the text. The author wishes to thank Professor C. H. Lawrence for valu- able suggestions made in regard to the form of presentation of some of the work. THOMAS T. EYRE. University of New Mexico. M512399 CONTENTS CHAPTER I PAGE PRESSURE, TEMPERATURE AND HEAT UNITS ... 1 CHAPTER II FUEL . ... 4 Anthracite Coal Bituminous Coal Lignite and Peat Natural Gas Oil Coal Fields of the U. S. Coal Storage - Determination of the Heating Value of Fuel Combustion Composition of Flue Gas Flue Gas Analysis Heat Lost in Flue Gas Value of CO 2 for Best Efficiency CO 2 Recorders. CHAPTER III STEAM ..... .... 16 Use of Steam Tables Throttling Calorimeter. CHAPTER IV BOILERS ........... 23 Requirements Rated Horsepower Heating Surface Rules for Finding the Heating Surface Superheating Surface Size of Boiler Tubes The B. & W. Boiler The Sterling Boiler The Wickes Boiler The Return Tubular Boiler The Internally- fired Return Tubular Boiler The Scotch Marine Boiler The Vertical Fire Tube Boiler The Locomotive Boiler Superheaters Horsepower of boilers Factor of Evaporation Efficiency of Boilers A. S. M. E. Boiler Test Code. CHAPTER V BOILER ACCESSORIES AND AUXILIARIES ...... 45 Grates The Plain Grate The Rocking Grate Mechanical Stokers The Chain Grate The Roney Stoker The Under- feed Furnace Smoke Prevention Settings Draft Dam- pers Safety Devices The Pressure Gage The Safety-valve Safety-valve Capacity Napier's Formula Safety-valve For- mula The Water Glass or Gage Glass High-water and Low- water Alarm The Fusible Plug Boiler Feedwater Treatment Scale Prevention and Removal Oil Separators Boiler Feed Pumps The Injector Boiler Feed by Returning Trap The Steam Line The Steam Trap Expansion Joints Steam vii Vlll CONTENTS PAGE Separators Steam-pipe Covering Feedwater Heaters Econ- omizers Condensers The Surface Condenser The Jet Con- denser Cooling of Circulating Water. CHAPTER VI THE STEAM ENGINE 76 History The Plain Slide-valve Engine Parts of the Steam Engine Piston Displacement Clearance Steam Back of Piston During Stroke The Indicator and its Purposes Events of Stroke Location of Events on Diagram Equation of Expan- sion and Compression Curves Hypothetical Indicator Diagram Determination of Clearance from Card Determination of the Mean Effective Pressure Indicated Horsepower Brake Horse- power Mechanical Efficiency Thermal Efficiency Cylinder Condensation Steam accounted for by the Indicator Diagram Valve Setting from the Indicator Diagram. CHAPTER VII COMMON TYPES OF STEAM ENGINES 100 Slide-valve Engine The Corliss Engine The Four-valve Engine The Compound Engine The Tandem-compound The Cross-compound Cylinder Ratio The Combined Indi- cator Diagram The Diagram Factor Ratio of Expansion The Unaflow Engine. CHAPTER VIII VALVES 112 The D Slide-valve Relative Motion of Crank and Piston Valve Diagrams The Valve Ellipse The Bilgram Diagram The Zeuner Diagram Types of Slide-valves Valve with Pres- sure Plate The Piston Valve Double-ported Valves The ' Gridiron Valve The Riding Cut-off Valve Effect of Rocker Arm on the Location of Eccentric Oscillating Valves Poppet Valves Reversing The Stephenson Link The Walschaert Valve Gear The Joy Valve Gear Setting the Slide-valve. CHAPTER IX GOVERNORS ........... 137 General Classification of Governors The Gravity-balanced Spindle Governor The Spring-balanced Governor Governing by Changing Position of Eccentric Governing by Changing Angle of Advance Governing by Changing Both Angle of Advance and Valve Travel Centrifugal and Inertia Governors. CHAPTER X STEAM TURBINES 150 General History Fundamental Principles Available En- ergy in Steam Velocity Due to Expansion Impulse and Reac- CONTENTS lx PAGE tion Bucket Shapes Single-stage Turbine Staging Multi- stage Impulse, Velocity-staging Multi-stage Impulse, Pressure- staging Multi-stage Impulse, Combination Pressure-staging and Velocity-staging Multi-stage Reaction Change of Area of Steam Passage Leakage Loss due to Running Partial Capacity Summary of Losses in the Steam Turbine Common Commer- cial Types DeLaval Single-stage Multi-pressure-stage Im- pulse Turbine Curtis Turbine Parsons Turbine Westing- house Turbine Other Types Low-pressure Turbine Mixed- pressure Turbine Use of Superheated Steam The Marine Turbine. CHAPTER XI GAS ENGINES .......... 193 General Historical Cycles of Operation The Four-stroke Cycle The Two-stroke Cycle Classification from Fuel Used Efficiency Fuels The Gasoline Carburetor The Gas Pro- ducer Cooling of Cylinders Ignition Valves Governing Determination of Horsepower Multi-cylinder Engines. PROBLEMS 213 ENGINES AND BOILERS CHAPTER I PRESSURE, TEMPERATURE AND HEAT UNITS 1. Pressure Units. In steam engineering, pressure is meas- ured in the following units: (1) In pounds per square inch. (2) In inches of mercury. (3) In inches of water. In this country boiler pressures are ordinarily measured in pounds per square inch above atmospheric pressure. Condenser pressures are commonly measured from atmospheric pressure in inches of mercury, i.e. the difference between the pressure in the condenser and atmospheric pressure is read on a mercury column. Draft pressures are usually measured in inches of water. Pressure gages and vacuum gages are so constructed that they read zero at atmospheric pressure. The atmosphere exerts a vari- able pressure, which is about 14.5 pounds per square inch at ordi- nary altitudes and under ordinary conditions. At sea-level, the standard is taken as 14.7 pounds per square inch, which is equiv- alent to 29.92 inches of mercury. As the atmospheric pressure is slightly variable, it is necessary in accurate work that pressures be reduced to an absolute basis. Since the zero reading of a boiler pressure gage means atmospheric pressure, the absolute pressure will be the sum of the gage pressure and atmospheric pressure. A partial vacuum usually exists in a condenser, since the abso- lute condenser pressure is usually less than atmospheric pressure. Since the vacuum gage as well as the boiler pressure gage reads zero at atmospheric pressure, the absolute pressure in the con- denser is the difference between the atmospheric pressure and the vacuum-gage pressure. Figure 1 shows diagrammatically the relation between these pressures. In this figure, a is the boiler-gage pressure, b the 1 2 ENGINES AND BOILERS atmospheric pressure, and c the absolute boiler pressure. Like- wise, d represents the vacuum-gage pressure, which is measured downward from the atmospheric pressure; and e, the difference between 6 and d, is the absolute condenser pressure. I 1 a 1 : r t i 6 d \ l_L_ Condenser pressure V f Z^/ pressure FIG. 1 Barometers used in engineering work in this country are usu- ally graduated in inches. The mercury barometer is used be- cause mercury is the most convenient fluid for this use. A cubic inch of mercury weighs very nearly 0.49 of a pound. Hence the pressure in pounds per square inch is the barometric reading in inches multiplied by 0.49 Vacuum gages also are commonly graduated to read in inches of mercury. Therefore the absolute condenser pressure in pounds per square inch is 0.49 times the difference between the barometer and the vacuum-gage readings. EXAMPLE 1. Find the absolute boiler pressure when the pressure gage reads 110 pounds and the barometer reads 29.4 inches. SOLUTION. When the barometer reads 29.4 inches, the atmospheric pres- sure is 29.4 X0.49 = 14.4 pounds per square inch. The absolute boiler pressure is then 14.4 + 110 = 124.4 pounds per square inch. EXAMPLE 2. Find the absolute pressure in a condenser when the barometer reads 29.8 inches and the vacuum gage reads 27.3 inches. SOLUTION. The absolute condenser pressure is the difference between the barometric and the vacuum-gage pressures, which in inches of mercury is 29.8-27.3 = 2.5. This reduced to pounds per square inch is 2.5. X49 = 1.22. 2. Temperature Units. Ordinary temperatures are measured by means of the mercury thermometer. For higher tempera- tures, such as those that occur in furnaces, special thermometric devices called pyrometers are used. Three thermometer scales are in use, the Fahrenheit, the Cen- tigrade, and the Reaumur. In the Fahrenheit scale, the differ- ence between the temperatures of melting ice and boiling water at sea level is divided into 180 divisions or degrees; the freezing point is 32 and the boiling point 212. This makes the zero PRESSURE, TEMPERATURE, AND HEAT UNITS Fahrenheit Centigrade Reaumur point come 32 below the freezing point of water. In the Centi- grade scale, the freezing point is and the boiling point is 100. In the Reaumur scale, the freezing point is and the boiling point is 80. Figure 2 shows graphically the relation between these scales. It is readily seen how the reading on one scale may be reduced to that of either of the others. It is serviceable --!. - | 67/.S- 1 80' i Boili/tf temperature femperofare \\ '' -J \ A Absolute zero FIG. 2 to remember that a difference of temperature equal to 5 on the Cen- tigrade scale is equal to 9 on the Fahrenheit scale. Experiment shows that a perfect gas under constant pressure at 32 Fahrenheit expands 1/491.5 part of its own volume for each degree (F.) that its temperature is in- creased. Because of this we call the point 491. 5 below the freez- ing point, or 459.5 on the Fahrenheit scale, the absolute zero. This corresponds to a Centigrade temperature of 273. 3. Heat Units. In engineering work in this country, it is customary to use English units. Our common heat unit is the British thermal unit (B.t.u.), which is practically defined as the amount of heat necessary to raise the temperature of one pound of pure water from 62 to 63 Fahrenheit. In the metric system the engineers' heat unit is the large calorie, which is the amount of heat necessary to raise the temperature of a kilogram of water from 15 to 16 Centigrade. As one kilogram =2.2046 pounds, and as 1 Centigrade =1.8 degrees Fahrenheit, one calorie =3. 968 B.t.u., or one B.t.u. =0.252 calorie. 4. Mechanical Equivalent of Heat. Heat experiments have been made to determine the relation between heat-energy and mechanical work. The latest and most refined experiments show that one B.t.u. as above defined is substantially equivalent to 778 foot-pounds of work. This relation is called the mechanical equiva- lent of heat. 1 *For definition of the "mean B. t. u." and corresponding Mechanical Equivalent of Heat see A. S. M. E. POWER TEST CODE, Edition of 1915, p. 28. CHAPTER II FUEL 5. Introduction. The source of the world's supply of energy is the sun. In the use of water power we are drawing, in point of time, almost directly from the sun. On the other hand, in the use of such fuels as coal, gas, and oil, we make use of a store of energy that has been accumulating for ages. While the world's supply of coal, oil, and gas is limited, we have used but a very small part of the known deposits. The past century has seen a marvelous change in our manner of living and in our ways of thinking. Our vast commercial system with its perplexing problems has arisen during the past few generations. One of the chief causes of this great change is our ability to util- ize the vast stores of energy to be found in nature. The steam- engine has been the chief means by which the energy stored in our coal deposits has been tapped and forced to do the work of man. What will become of this modern civilization of ours when the fuel supply upon which it is based is exhausted, is an inter- esting problem of the future. Already conservationists are call- ing to us to stop the great waste of our natural resources. Of all our fuels, coal is the most important. Coal is the re- mains of vegetable matter deposited in remote geological ages. It is well known that wood rots but little when kept under water. If the water be fresh, -the wood is not eaten by the teredo worm or other forms of aquatic life, and will be kept in a fair state of preservation for thousands of years. If tree trunks and other vegetable matter fall into a fresh-water swamp and are sub- merged before they rot, and if this continues for many centuries, there will be a great accumulation of it. Oar coal deposits are the result of such an accumulation of vegetable matter. Under tropical conditions accompanied by a large supply of carbon dioxide in the atmosphere the growth was very rapid and a deep bed would collect in a comparatively short time. Geologists tell us that in the past, parts of the surface of the earth have gradually risen while others have fallen. At a remote time, the tops of our highest mountains may have been the bot- tom of the sea. Suppose a former swamp with its accumulated vegetable matter is now sunk, and that great quantities of silt or 4 FUEL 5 other material are deposited upon it. The weight of the material above will compress the vegetable matter into a compact and dense mass. It is also possible that it will be subject to a high temperature, which will change its chemical composition. Vary- ing conditions of pressure and heat are thought to be responsible largely for the differences between the various kinds of coal. The principal constituents of coal are carbon, hydrogen, oxygen, nitrogen, sulphur, and refractory earths called ash. The wood- fiber of the original vegetable matter was composed chiefly of hydrocarbons. While under the influence of great pressure, it has at some period of its history been subjected to considerable heat and therefore undergone a process of destructive distiilization. This has driven off much of the volatile matter of the original vegetable material and left a considerable portion of uncombined carbon, which is called fixed carbon. The remainder of the car- bon exists in combination with hydrogen. These carbon and hydrogen compounds are called hydrocarbons. They are easily volatilized, and so comprise a part of the volatile matter of the coal. Oxygen and hydrogen are always present in coal in the form of water. This water is volatilized, of course, when the coal is burned. Since heat is required to evaporate and to superheat it, water is a detriment if present in too large a quantity. A small amount of water, however, seems to aid in the combustion of some coals. All coal then contains fixed carbon, volatile matter (hydrocar- bons and water), and ash; and it may contain other substances (e.g., sulphur). The combustible is the fixed carbon, the hydro- carbons, and part of the sulphur that may be present. Excessive sulphur is undesirable because it is harmful to the metal of the boiler and the stack if moisture is present, since it may form sulphurous or sulphuric acid; it combines with the ash to form a fusible slag or clinker, which is commonly objectionable; and it makes the fuel more liable to spontaneous combustion when stored in deep piles. Coals that have been subjected to the greatest pressure and heat are composed mostly of fixed carbon, and contain only a small amount of volatile hydrocarbons. Such coals are called anthracite. Coals containing larger quantities of volatile hydro- carbons are called bituminous. Since there is no definite divid- 6 ENGINES AND BOILERS ing line between these two classes, but the two seemingly overlap, the terms semi-anthracite and semi-bituminous are commonly used to designate coals to which are ascribed certain properties of each class. 6. Anthracite Coal. Anthracite coal contains but a small amount of combustible volatile matter. While it is considered better for some uses, its heating value is less than that of good grades of bituminous coal. It burns slowly, with but a small flame and practically no smoke. Due to its slow burning quali- ties, a relatively large grate area is needed on which to burn it. The supply of this coal that is easily obtained has been diminish- ing in this country, and its demand for domestic use has greatly increased during the past few years. This has led to a rapid in- crease in price and a great diminution of its use for power purposes. Anthracite is considered much superior to bituminous coal for the production of producer gas. This is due to the fact that it is so free from the hydrocarbons that produce tars. The formation of tar has been the great objection to the use of bituminous coals in the producer plant. Average anthracite coal contains about 85% fixed carbon, 4% volatile matter, 9% ash, and 2% water. The heat value averages about 13000 B.t.u. per pound. 7. Bituminous Coal. Most of the coal used for power pur- poses is bituminous. This coal contains a larger percentage of volatile combustible matter. It burns at a lower temperature than does anthracite, and with a much longer flame. The length of the flame varies with the composition, some kinds being called long-flaming and others short-flaming. The ordinary furnace is not usually so constructed as to give the volatile hydrocarbons a chance to be completely burned. This results in the formation of smoke. Engineers have spent much time and study on the prevention of smoke. Upon yielding up their volatile matter some coals fuse and form a solid mass or cake of the nature of coke. These are called caking coals. This action hinders the draft. If rapid combustion is desired, the mass must be broken up. The composition of bituminous coal varies greatly, but the average of the better grades may be taken as 65% fixed carbon, 28% volatile combustible, 5% ash, and 2% moisture, with a heat value of 14000 B.t.u. per pound. FUEL 7 8. Lignite and Peat. Lignite or brown coal is high in vola- tile combustible and also contains much moisture. The evapora- tion of this moisture after mining causes the lignite to crumble or slack. It is usually inferior to anthracite and bituminous coals, but it is used where it is easily obtained and where better coal is expensive. Abroad, lignite is often formed into briquettes. Peat is the partly decayed remains of vegetation that accumu- lates in bogs. While it is an inferior fuel, it is used to a consid- erable extent abroad. It is sometimes pressed into briquettes. 9. Natural Gas. Natural gas is used to some extent for power purposes in sections of the country within reach of the gas fields. It contains about 90% marsh gas (CH 4 ) and has a heat value of nearly 1000 B.t.u. per cubic foot. 10. Oil. Crude petroleum and fuel oil are used to a consid- erable extent in parts of this country. The petroleum produced in the eastern and middle states is of a paraffin base, while that from Texas and California is of an asphalt base. Gasoline and other light oils are distilled from petroleum, and the residue is sold as fuel oil. Petroleum is composed principally of hydrogen and carbon in the form of hydrocarbons and has a heat value of about 20000 B.t.u. per pound. 11. Coal Fields of the United States. Our principal deposits of anthracite coal are in eastern Pennsylvania. The deposits are not of large extent and the best are rapidly becoming exhausted. It is claimed that there is anthracite in Alaska. Only a small proportion of the anthracite mined is being used for power pur- poses, the rest going for domestic heating and like purposes. The best of our bituminous and semi-bituminous coal is taken from the field that includes western Pennsylvania, eastern Ohio, a large part of West Virginia, and eastern Kentucky. Southwest- ern Indiana and most of the state of Illinois are underlaid with coal of a fair quality. There is also a field running north from Okla- homa through eastern Kansas and western Missouri into Iowa. The coal from the latter is generally of poor quality, and is used only locally. Immense coal fields exist in southern Utah and Colorado, and in New Mexico and Arizona. No complete survey of these fields has been made as yet, and they have not been de- veloped up to the present. There is also a coal and lignite field in eastern Montana and western North Dakota. 8 ENGINES AND BOILERS The peat beds of the country are principally in Minnesota, Wisconsin, Michigan, New York, and the New England states. Oil is produced in the territory occupied by the eastern coal fields, in Kansas, Oklahoma, Texas, California, and, to a smaller extent, elsewhere. 12. Coal Storage. In the operation of most steam-power plants, it is essential that a constant supply of coal be available. Owing to unsettled labor conditions at the mines and to uncer- tain transportation facilities, it is necessary that there be some storage capacity. With anthracite coal this is a simple problem, but it is not so with certain grades of bituminous coal. Upon exposure of bituminous coal to the air, there is a considerable oxidation of the hydrocarbons with attendant heat production. If the coal pile is large, this heat may start a fire which is costly and hard to extinguish. The origin of such a fire is called spon- taneous combustion. Even if fire does not start, there is a loss of heat-value up to as high as 10% in some grades of coal. If the moisture content is large, the weathering is accompanied by a crumbling or slacking. Storage piles are often ventilated in order to keep them cool. In some large plants, the storage is so ar- ranged that it may be submerged in water. This obviates the fire risk, and reduces the other losses to a minimum. 13. Determination of Heating Values of Fuel. In plants where large amounts of fuel are used, it is quite common to buy coal on the basis of its heating value. In accurate tests of power plants it is also necessary to know the heating value of the coal used. Care must be exercised in order to get a fair sample of the fuel. The heating value may be determined in two ways, by combustion in a calorimeter, or by chemical analysis. In the calorimeter method a sample of the fuel is placed in a steel bomb along with compressed oxygen. The bomb is placed in a calorimeter containing water, and the fuel is ignited by means of an electrically heated wire. Upon the firing of the fuel, heat is given to the water. It is possible to determine from the rise of the temperature of the water the amount of heat generated. The value thus determined is known as the higher calorific value. It must be remembered that it is seldom possible actually to get this amount of heat by burning in a furnace. This is due to the fact that the hydrogen in the fuel combines with the oxygen of FUEL 9 the air to form water, which ordinarily passes from the furnace in the form of steam, carrying with it the heat of vaporization of the steam. The lower heat-value does not contain this heat of vaporization. The higher value is the accepted standard. There are two methods of chemical analysis, the ultimate, and the proximate. The ultimate analysis may be made on the basis either of moist or dry fuel. The latter is commonly accepted. If the analysis is made on the basis of moist fuel, it may be con- verted to the dry basis by dividing the percentage of the various constituents by one minus the percentage of the moisture. The ultimate analysis gives the percentage by weight of carbon, hydro- gen, oxygen, nitrogen, sulphur, and ash. Knowing the composi- tion of the fuel, the heat-value may be determined by a formula. An accepted formula is a modification of that of DULONG : B.t.u. per pound of dry fuel = 14600 C+62000 (H-O/8)-f 4000S, in which C, H, O, and S represent the proportionate parts by weight of carbon, hydrogen, oxygen, and sulphur. The heat- value of pure sulphur is 4000 B.t.u. per pound, but the sulphur in coal is mostly in a form that is noncombustible. The results of the ultimate analysis agree closely with those obtained from the calorimeter. The proximate analysis gives the proportion of fixed carbon, volatile combustible, moisture, and ash. Since the heating value of the volatile combustible is not determined, the results are not as reliable as those of the previous method. It is very often used, however, because it is easily made and affords a rough comparison of various fuels. In making the proximate analysis, a weighed sample of coal is placed in a crucible of known weight and is kept at a tem- perature a few degrees above the boiling point of water for an hour. From the loss of weight, the moisture in the coal is cal- culated. The sample is next heated in a hot flame a few minutes with the lid on the crucible. This drives off the volatile matter, and the difference in the new and previous weight is the amount driven off. Now the sample is heated for two hours with the lid off, all fixed carbon is burned out, and the weight is determined by difference as before. The weight left is that of the ash. After having found the percent by weight of the moisture, volatile matter, fixed carbon, and ash, the approximate calorific value may be found by means of the chart, shown in Fig. 3, which 10 ENGINES AND BOILERS has been constructed from the values determined in a large num- ber of accurate analyses. On this chart the heat-value in B.t.u. per pound of combustible is plotted against the percent of fixed carbon in the total combustible. It is assumed that the volatile matter and the fixed carbon constitute the combustible, the ash and the moisture being noncombustible. The method of getting the heating value may be shown by examples. EXAMPLE 1. Determine the calorific value in B.t.u. per pound of dry coal having the following ultimate analysis: carbon = 75.46%, hydrogen = 3.34%, oxygen = 2.70%, nitrogen = .53%, sulphur = 2.54%, and ash = 15.43%. SOLUTION. The B.t.u. per pound = 14600 X .7546 +62000 (.0334 - .0270/8) +400t< .0254 = 12880 B.t.u. The oxygen calorimeter gave a value of 13000 B.t.u. per pound for this same sample. Chart for determining Heat Value of Combustible with Different Percentages of Fixed Carbon from Proximate Analysis ^ ^ ^ ^ "X S x s \ 3+O0 s \ s \! S \ | / / ^ ^ / <*" / s / / \ ^ 14000 s J ^ ^ t+fVQ f \ / V * r 1 1 * 1 I tot ^ -14,00 1 14400 JO SS 6S 70 7S &0 Percent of ftxcd C arbor? m Tbta/ FIG. 3 EXAMPLE 2. Determine the calorific value of coal which has the follow- ing proximate analysis: moisture =4. 7%, volatile matter = 24.6%, fixed carbon =62.5%, and ash = 8.2%. SOLUTION. The combustible being composed of the volatile matter and fixed carbon comprises 24.6+62.5 = 87.1% of the weight of the coal. Of this combustible the fixed carbon is 62.5/87.1 = 71.7%. From the chart it is seen that for 71.7% the B.t.u. per pound of combustible is 15550 B.t.u. As the combustible comprises only 87.1% of the total weight the heating value will be .871X15550 = 13550 B.t.u. per pound of wet coal, or if reduced to the dry basis, 13550/(1.00-. 047) = 14200 B.t.u. per pound. FUEL 11 14. Combustion. By combustion is meant the rapid chem- ical combination of oxygen with the carbon, hydrogen, and sul- phur in fuel. Combustion is complete when the maximum amount of oxygen is used in the combination. One atom of carbon will combine with one atom of oxygen to form carbon monoxide (CO). This is not complete combustion, however, for one atom of car- bon will combine with two atoms of oxygen, forming carbon di- oxide (CO2), if sufficient oxygen is present. Since the atomic weight of oxygen is 16, and that of carbon is 12, it takes 32 pounds of oxygen to 12 pounds of carbon to form carbon dioxide, i.e., one pound of carbon requires for its complete oxidation 2.667 pounds of oxygen. Air, by volume, is composed of about 21% oxygen and 79% nitrogen, and by weight, of 23.15% oxygen and 76.85% nitrogen. So one pound of oxygen is con- tained in 4.32 pounds of air. It therefore takes 2.667X4.32= 11.55 pounds of air for every pound of carbon burned. At ordi- nary room temperatures one pound of air occupies about 13.4 cubic feet, so that it requires theoretically 13.4X11.55= 155 cubic feet of air for the complete combustion of one pound of carbon. The hydrogen in the hydrocarbons of the coal is also com- bustible. A part of the sulphur present may be combustible, but it is usually present in such small amounts that it will be omitted in our present computation. As explained previously, not all of the hydrogen content of the coal is combustible, since part of it is already combined with oxygen in the form of water. Therefore the available hydrogen may be expressed as (H O/8). Since hydrogen combines with oxygen to form water, in the ratio by weight of 1 to 8, it will require 8 pounds of oxygen to burn each pound of hydrogen. Since one pound of oxygen is contained in 4.32 pounds of air, it will take 8X4.32=34.6 pounds of air to burn a pound of hydrogen. Hence the total weight of air re- quired to burn a pound of coal to CO 2 and H 2 O is theoretically 11.55 C+34.6 (H-O/8), where C, H and O have the same meaning as in 13. Since the nitrogen of the air is inert, it is of no value to the combustion. Since it passes up the stack at a higher tempera- ture than that at which it entered the furnace, it carries away heat. Any less air than the theoretically correct amount would result in the formation of a mixture of carbon monoxide and 12 ENGINES AND BOILERS carbon dioxide. The heat liberated by the formation of the carbon monoxide is only 4450 B.t.u. per pound of carbon, while it is nearly 14600 B.t.u. for the formation of carbon dioxide. Hence the production of carbon monoxide in a furnace means a large loss of heat. The presence of carbon monoxide in flue gas nearly always indicates a large amount of unburned hydrocarbons and hence an even greater loss of heat. If it were possible to so dis- tribute the air that it all came in close contact with the fuel, and also to give it time enough to combine thoroughly with the fuel, the theoretical amount of air would be sufficient. Under actual furnace conditions, however, it is found that 50% or more excess of air is needed to give complete combustion of coal. A somewhat smaller excess is needed when oil is used as a fuel, because there is better distribution of the air. The greater the amount of air passing through the furnace, the greater the amount of heat it will carry along to the stack. Hence an unnecessary excess of air is not desirable, and leads to lessened efficiency. The neces- sary excess depends upon the conditions of draft and fire as well as upon the kind of fuel and the type of furnace. It can only be determined by actual test. 15. Composition of Flue Gas. As just explained, an excess of air is needed in order to get complete combustion of the fuel. If it were possible to get complete combustion without this excess, our flue gas would be composed chiefly of nitrogen, carbon dioxide, and water vapor. Due to the excess of air, there will be free oxygen present in the flue gas. If there is an insufficient excess of air there will also be carbon monoxide and probably some hydrocarbons present. We have seen that the presence of carbon monoxide indicates incomplete combustion and therefore low fur- nace efficiency. On the other hand, a large excess of air, while it may give complete combustion, gives poor furnace efficiency because the air will carry a large amount of heat up the stack. It is a matter of great importance that just the right excess of air be admitted to the furnace. Since it is difficult to measure directly the amount of air entering the furnace, an easier method is used. This method consists in analyzing the flue gas to deter- mine the amount of each of its constituents. From this analysis we can easily compute the amount of air entering the furnace. Knowing the composition of flue gas, we can regulate the amount FUEL 13 of air entering the furnace so as to give the proper excess to insure the best economy of operation. 16. Flue Gas Analysis. There are various types of apparatus on the market for making the analysis of flue gas, most of which are modifications of the apparatus designed by ORSAT. A com- plete description of the Orsat apparatus will not be given here, but the principle of its operation is as follows. A sample of gas is taken from the rear of the furnace or between the furnace and the stack. After being cooled to the room tem- perature, it is carefully measured by volume at atmospheric pressure. All measurements are taken at room temperature and at atmospheric pressure. This known volume of our sample is first passed a few times through a solution of caustic potash, which absorbs the carbon dioxide. The volume is measured again, and the difference between the new volume and the original volume is the volume of the carbon dioxide absorbed. The same sample is next passed several times through a solution of potassium pyro- gallate, which absorbs the oxygen. The amount of oxygen is determined by the loss in volume, as before. Next the sample is passed several times through a solution of acid cuprous chloride and the carbon monoxide removed, and its amount determined as before. The amount of carbon monoxide is usually quite small. The remainder of the sample is usually assumed to be nitrogen^ 17. Heat Lost in Flue Gas. The weight of flue gas per pound of fuel burned (assumed carbon and ash) may be computed from the formula, where W = weight of flue gas per pound of fuel burned. C = decimal part by weight of total carbon in fuel. N = percentage by volume of nitrogen in flue gas. CC>2 = percentage by volume of carbon dioxide in flue gas. CO = percentage by volume of carbon monoxide in flue gas. A = decimal part by weight of ash in fuel as fired. Unless the ultimate analysis of the fuel is known, the weight of carbon in the volatile matter will have to be estimated and added to the weight of fixed carbon to give C in the preceding formula. Marks has published a chart showing the approximate 14 ENGINES AND BOILERS relation between the volatile carbon in the combustible and the total volatile matter in it. With the aid of this chart (Fig. 4), the value for C in the preceding formula can be approximated from the proximate analysis. The specific heat of the flue gas is usually taken as .24; and the heat lost per pound of fuel burned is equal to the product of the specific heat of flue gas, the weight of gas per pound of fuel, and the difference in temperature between the leaving flue gas and the entering air. Chart for determining the Carbon in the Volatile Matter Marks ts 20 10 20 30 -9-0 SO t Percent of Yo/otifa Afatt-rr jr? ffye Combust/hie FIG. 4 EXAMPLE. How much heat is carried up the stack by the dry flue gases, when the furnace is burning coal of the following proximate analysis? Mois- ture = 3%, fixed carbon = 65%, volatile matter = 26%, and ash=6%. The analysis of flue gases gives: CO 2 = 10%, O = 8%, and CO = .5%. The stack temperature is 500 F. and the temperature of the air entering the furnace is 80 F. SOLUTION. From the chart of Figure 4, we find that the percent ofjvoh tile carbon in the combustible is about 13.5, which corresponds fuel. The total carbon is then 12.1+65 = 77.1%, and the weighiToT flue gases per pound of coal is 3.032 X. 771 --6) = 19 - 1 Pounds. The heat carried up the stack by the dry flue gases is then .24X19.1 (500 80) = 1925 B.t.u. for each pound of coal fired. In this problem, the heat- value of a pound of coal found from the proximate analysis is 13860 B.t.u. Hence the loss is 1925/13860 = 13.9% of the heat available. FUEL 15 Both the fuel and the air contain moisture. This moisture also carries heat up the stack, since it is a superheated vapor upon leaving the furnace. 18. Value of CO 2 for Best Efficiency. As has been stated, the efficiency of the furnace will vary with the excess of air ad- mitted. Since the percentage of CO 2 also varies with the excess of air, we see that an indication of the efficiency .is given by the CO 2 reading. Just what percentage of CO 2 corresponds to the highest operating efficiency depends upon such factors as kind and state of fuel, stack temperature, etc. After the proper amount of C0 2 for best efficiency has been determined under these conditions, the CO 2 reading will indicate whether or not high efficiency is being obtained. Since the determination of the CO 2 is a comparatively simple operation, it is an excellent way to keep check on operating conditions. Some plants go so far as to pay their firemen on the basis of the CO 2 record. In general, high CO 2 means high efficiency, unless there is some abnormal condition such as too much CO. The CO should be kept as near zero as possible. At the same temperature and pressure, CO 2 occupies the same volume as the oxygen from which it was formed. The volume of the oxygen in the air is 21%. Hence, if the products of combus- tion are cooled down to the temperature of the entering air, the CO 2 reading would be 21% for perfect combustion with no excess of air, assuming the fuel to be carbon and ash. In practice, the CO 2 runs 17% or lower. Even 15% is usually considered an in- dication of very good efficiency. 19. CO 2 Recorders. Automatic devices are on the market that will analyze and record the amount of CO 2 on a chart. An analysis is made every few minutes, so that a complete record is kept of the operating efficiency. CHAPTER III STEAM 20. Introduction. Definitions. A perfect gas may be at any temperature under any pressure. For instance, air may be placed under a certain pressure and have its temperature raised or low- ered by the addition or subtraction of heat. On the other hand, a saturated vapor, such as steam, can exist only at a certain definite temperature for each particular pressure. Under ordinary atmospheric pressure, saturated steam can exist only at a temper- ature of about 212 F. Under an absolute pressure of 100 pounds per square inch, saturated steam will be at a temperature of 327.86 F. Let us consider a case in which a pound of water at 32 F. is placed under a pressure of, say, 100 pounds per square inch. The containing vessel is supposed to be so constructed that the pressure remains constant, no matter what change of volume takes place. Now suppose that the water is heated. There will be little change in volume, but there will be a rise of temperature of approximately one degree F. for each B.t.u. given to the water. This will continue until we have added 298.5 B.t.u. The tem- perature will then be 327.86 F. A further addition of heat up to a limit will not cause any change of temperature, but will effect a change in the physical condition of the water, turning it to steam. We shall need to apply 887.6 B.t.u. to effect this change completely. We have now added a total of 298.5+887.6 = 1186.1 B.t.u., and have converted the pound of water, originally at 32 F., into dry saturated steam at a temperature of 327.86 F., and under an absolute pressure of 100 pounds per square inch. Now that the water is all evaporated, if more heat be added to this steam, the temperature will rise at the rate of nearly two degrees per B.t.u. added. We now have superheated steam. As stated previously, the volume of the water will change but little until the boiling point is reached. The space occupied by the saturated steam will be 4.432 cubic feet. This is many times greater than the space formerly occupied by the water. Part of the 887.6 B.t.u. that was used to evaporate the water was evidently used to cause this change in volume under the pressure 16 STEAM 17 of 100 pounds per square inch. The remainder was used to make the physical change in the water, to increase the kinetic energy of its atoms. For pressures other than 100 pounds we would have values different from those given above. Figure 5 represents graphically the relation between the tem- perature and the heat added to a pound of ice, starting at zero degrees F. (with the assumption that the pressure is constant). Upon the first addition of heat, the temperature of the ice will rise until the melting point is reached. Further addition of heat 1 I O* 32 Temperature FIG. 5 causes the ice to melt. This occurs without a change of tempera- ture. The part of the line representing the melting of the ice is therefore vertical. When the ice is all melted, addition of more heat causes the temperature of the water to rise. This will con- tinue until the boiling point is reached. That part of the line representing evaporation will be vertical since there is no change of temperature during that period. When evaporation is com- plete, the addition cf heat again causes a rise in temperature. The amount of heat necessary to raise the temperature of the water, the amount of heat required to change it to steam, and also the volumes of steam formed under different pressures, have been determined by numerous experiments and are published 18 ENGINES AND BOILERS under the name of steam tables. We shall use in our work the tables prepared by C. H. Peabody. 1 These are arranged in two ways. In Table I, the various absolute pressures at which water boils are given for each degree F. from 32 to 428. In Table II, the various temperatures at which water boils are given for each pound per square inch from 1 to 336. The values are arranged in two tables not because they are different, but simply as a con- venience in their use. In Table I, the first column, headed t, gives the temperature at which water boils. The second column, headed p, gives the absolute pressure under which it must be in order that it boil at the temperature given in the first column. The third column, headed q, gives the heat of the liquid, which is the number of B.t.u. necessary to change the temperature of one pound of water from 32 F. to the temperature given in the first column. This does not mean that there is no heat in the water at 32. 'The heat in the water below the freezing point is of no moment to the steam engineer; hence it is chosen as the arbitrary starting point. Column four, headed r, gives the heat of vaporization, which is the B.t.u. necessary to evaporate completely a pound of water at the temperature and pressure given in the first and second columns. This is sometimes called the latent heat of evaporation. The sum of the heat of the liquid and the heat of vaporization is called the total heat. The fifth column, headed p, is that part of the heat of vaporiza- tion that is used in energizing the atoms of the water to turn it to a vapor; it is called the heat equivalent of internal work. The sixth column, headed Apu, is the rest of the heat of vaporization, or that port that is needed to do the work of increasing the volume, under the pressure of column two; it is called the heat equivalent of external work. Columns seven and eight will not be discussed here. Column nine, headed s, gives in cubic feet the specific volume, which is the volume of one pound of dry saturated steam under the pressure of column two. Column ten gives the reciprocals of the values found in column nine. It is the weight of one cubic foot of dry saturated steam under the pressure of column two. i C. H. PEABODY, Steam Tables. STEAM 19 Steam generated in most boilers not equipped with a super- heater is likely to carry with it, when leaving through the outlet pipe, a small amount of water in a finely divided state or mist. Steam containing this moisture is said to be wet steam. The Chart Showing Specific Heat of Superheated Steam Values from Knoblauch and Jakob 700 zoo SO /OO /JO ZOO 250 Pressure In pounds per square /nch ffhso/ute FIG. 6. quality of wet steam is expressed in percent. If in a hundred parts by weight of a mixture of steam and water, five parts by weight are moisture, the quality of the mixture is said to be 95% and the priming 5%. As long as steam is in contact with water it will remain satu- rated, and its temperature cannot be raised under constant pres- 20 ENGINES AND BOILERS sure. If it is conducted away from the water and led to a super- heater, its temperature will be raised by the addition of heat. It is then superheated steam. The amount of heat necessary to superheat depends upon the pressure and upon the degree of superheat. The chart of Fig. 6 gives the specific heat of super- heated steam for the ranges of pressure and temperature com- monly found in practice. The specific heat of steam varies with both temperature and pressure. The chart gives the average values of specific heat as the steam is raised from the temperature of saturation to the temperature of superheat. EXAMPLE 1. How much heat is required to change a pound of water at 70 F. into dry saturated steam at a pressure of 120 pounds per square inch absolute? SOLUTION. On page 48 of Peabody's Steam Tables, we find that the heat of the liquid, q, for 120 pounds pressure is 312.3 B.t.u. This amount of heat would bring the temperature of the water from 32 F. to the boiling point. As the temperature of the water to start with is 70 (p. 36), it already con- tains 38.1 B.t.u. It is then necessary to add to it 312.3-38.1=274.2 B.t.u. in order to bring it to the boiling point. To evaporate the water requires the heat of vaporization, r, at 120 pounds (p. 48), which is 876.9. Hence the total heat required to bring the water up to boiling and to evaporate it is 274.2+876.9 = 1151.1 B.t.u. EXAMPLE 2. If, in Example 1, the quality of the steam formed had been 97%, how much heat would it have required? SOLUTION. The water must all be brought to the boiling point, which takes the same amount of heat as in Example 1, 274.2 B.t.u. As the quality is 97%, only .97X876.9 = 850.6 B.t.u. are needed to evaporate the water. Hence the total amount of heat required is 274.2+850.6 = 1124.8 B.t.u. EXAMPLE 3. Find the amount of heat necessary to generate the pound of steam in Example 1, if it is superheated to a temperature of 475 F. SOLUTION. To generate a pound of dry saturated steam under the con- ditions of Example 1, requires 1151.1 B.t.u. The temperature correspond- ing to 120 pounds pressure is 341.3 F. The superheat is then 475 -341.3 = 133.7. From the chart of Fig. 6, the specific heat of superheated steam is .537. It will take .537X133.7 = 71.8 B.t.u. to superheat the steam. Hence the total heat required is 1151.1+71.8 = 1222.9 B.t.u. EXAMPLE 4. Find the volume of the pound of steam in Example 2. SOLUTION. A pound of dry saturated steam at 120 pounds pressure occu- pies 3.723 cubic feet. The quality being 97%, the volume occupied by the steam is .97X3.723 = 3.611 cubic feet. A pound of water occupies .016 cubic feet, and, as 3% of the pound of wet steam is water, the volume of the water is .03 X .016 = .0005 cubic feet. Hence the total volume is 3.61 1 + .0005 = 3.611 cubic feet. STEAM 21 21. The Steam Calorimeter. In making tests of boilers or engines it is necessary to know the quality of steam leaving the one or entering the other. A steam calorimeter is used in mak- ing this determination. Several types of calorimeters are in use. If the quality of steam is high (between 94% and 100%), the throttling type is usually used. When properly constructed this calorimeter is sufficiently accurate for ordinary purposes. Figure 7 shows a throttling calorimeter attached to a steam- pipe A. If the steam is saturated, and the pressure is known from a pressure gage H, its temperature may be determined from the steam tables. If the steam in A is superheated, it is also necessary to take its temperature by means of a thermometer, and the heat contents may be calculated from the steam tables. If it is wet, the quality must be known to find its heat contents. The tube B in Fig. 7 is a sampling tube through which a sample of steam is taken from the pipe A. This sample is throttled down 22 ENGINES AND BOILERS in pressure at C from the pressure in A, say pi, to the pressure in the calorimeter G, say p%. If the calorimeter is well covered, but little heat is lost by radiation and the heat contents in one pound of steam in A is the same as in the chamber G. The pres- sure and temperature in G are measured by the gage F and the thermometer E, and the heat contents are computed by means of the steam tables. Since the heat content is the same per pound in A as in G, the quality in the former may be computed as follows. The total heat in A equals q\-\-xri where x is the quality and qi and r\ are the heat of liquid and the heat of vaporization in A, respectively. If the calorimeter is working properly, the steam will be super- heated in G, and its heat contents will be equal to where q% and r% are the heat of the liquid and the heat of vaporiza- tion in G, respectively, 3 is the temperature of steam in G as measured by the thermometer E, and k is the temperature of saturated steam at the pressure p%. The term .48 (3 2) is the heat of superheat in G, since .48 is the specific heat of super- heated steam at low pressure and temperature. Then we have from which x may be found. EXAMPLE. Find the quality of steam leaving a boiler when the pressure is 165 pounds gage. The gage pressure in the throttling calorimeter is 3 founds, the temperature is 265 F., and the barometer reading is 29.6 inches. SOLUTION. From the steam tables, q\ =345, n =851, FIG. 13b tion flues, each of which has its own set of tubes. In marine practice the combustion chamber at the rear of the flue and the tubes is internal, there being a water-leg between the combustion chamber and the rear head. There is a combustion chamber for each flue and its set of tubes. Since the surfaces of the combus- tion chamber are flat, they must be stayed. These boilers are large in diameter; the outer shell must there- fore be very thick. The longitudinal seams are usually triple riveted, with two-strap butt joints. The Scotch marine boiler is used to some extent on land, but as the space occupied is usually an element of less importance here, the type is not common. BOILERS 35 34. Fire-tube Boiler. Vertical Type. A boiler often used for small or portable plants is the vertical fire-tube type (Fig. 14). These boilers are internally fired, the fire being enclosed in the lower part of the shell, and surrounded by an annular water-ring or water-leg. The lower tube- sheet is placed but a small distance above the grate. Therefore the space for com- bustion is very limited. The tubes are vertical and are ex- panded into the lower and the upper tube-sheets. ,/ connecfioas 35. Fire-tube Boiler. Loco- motive Type. The type of boiler used on locomotives, and also often used on portable plants, is shown in Fig. 15. In this boiler, the fire-box is at the rear end of the shell, and its top and sides are water- heating surfaces. Since the sheets that form the water-legs at the sides and rear of the fire-box are flat, it is necessary to stay them to prevent distortion. A screw-stay is used for this purpose; it con- sists of a threaded bolt screwed through the parallel plates. The threads on the center part of these stay-bolts are removed so that cracks will not start in the bolt at the root of the thread. On what are called safety stays a small hole is drilled in from the end so that a cracked bolt will leak steam and give warning. The flat or arched sheet at the top of the fire-box is called the crown-sheet. The crown-sheet is stayed in various ways, some- times by radial stay-bolts which run between it and the outer shell, or by sling stays, which are girders slung from the outer shell. Fire-tubes extend from the tube-sheet at the front of the fire-box to the tube-sheet at the front end of the boiler. The tubes in locomotive boilers are smaller than in the types previously described, and are placed as close together as good FIG. 14 36 ENGINES AND BOILERS BOILERS 37 circulation of the water will permit. By making the tubes small and numerous, a large heating surface is obtained. Where a superheater is used, as shown in Fig. 15, some of the tubes are made larger and contain the superheating surface. This super- heating surface is formed by tubes that extend into the fire-tubes from the front end of the boiler and run to within a short dis- tance of the fire-box. The outer shell extends beyond the front tube-plate to form a smoke-box. In this smoke-box vertical nozzles are located, through which the exhaust steam from the engines escapes. This induces a strong draft that allows a very rapid rate of combus- tion. The rate of combustion often exceeds 100 pounds of coal per square foot of grate surface per hour. Since the steaming of this type of boiler is very rapid, the steam is taken from a steam-dome located on the top of the shell. This allows the steam to be taken at a distance from the water surface, thereby insuring fairly dry steam. The throttle valve for the engine is located in the dome. In smaller locomotive boilers, the fire-box is set between the rear drivers. In the larger sizes, this arrangement will not allow a large enough grate area, and so the fire-box is extended later- ally over low trailer wheels. Manholes and hand-holes are em- ployed to give access for cleaning, as in other types of fire-tube boilers. 36. Superheaters. During the past few years superheated steam has come into quite general use, especially if it is to be used in steam turbines. The amount of superheat used is gener- ally not large, usually between 100 and 200 Fahrenheit. The advantages gained by the use of superheated over saturated steam will be considered later. There are two types of superheater. One type is independ- ently fired. The other is formed by the addition of some super- heating surface to the main boiler. The latter is the more com- mon form. In this type the steam is taken from the steam space and led through superheating coils. Often provision is made for the flooding of these coils during the period in which steam is being raised, in order to prevent the coils from burning out. The boiler shown in Fig. 8 has a common form of superheater at- tached. 38 ENGINES AND BOILERS 37. Horsepower of Boilers. As explained previously, boilers are generally rated by the manufacturer on the amount of their heating surface. The rate at which a boiler is working, how- ever, must be determined from a consideration of the amount of steam that is being generated. The amount of heat necessary to evaporate a given quantity of water varies with the temperature of the feedwater, with the pressure at which the steam is formed, and with the quality of the steam produced. Hence it is desirable that there be a stand- ard of temperature and pressure at which we can find the equiv- alent amount of water evaporated, using the same amount of heat as is used under the actual conditions of temperature and pressure. The conditions of temperature and pressure set by the A. S. M. E. 1 as a standard are 212 F. and 14.7 pounds per square inch absolute. The equivalent evaporation is then the amount of water that would be evaporated from and at 212 F. if the same amount of heat were used in its evaporation as is used in the evaporation under the actual working conditions. From the steam tables, the B.t.u. required to evaporate one pound of water from and at 212 F. is seen to be 969.9 B.t.u. This is the unit of evaporation. At the time when it first became necessary to rate boilers, a good engine used about 30 pounds of steam per horsepower per hour at a pressure of 70 pounds gage. The judges at the Cen- tennial Exposition in 1876 awarded prizes using the following unit as a standard. A one-horsepower boiler is one that will evapo- rate 30 pounds of water per hour from feedwater at 100 F. into steam at 70 pounds pressure by gage. The A. S. M. E. has since adopted an equivalent standard and defines a boiler horsepower to be the evaporation of 34.5 pounds of water per hour at 212 F. into steam at 212 F. and at a pressure of 14.7 pounds absolute pressure. As the heat of vaporization at 212 F. or 14.7 pounds absolute is 969.7 B.t.u., it therefore takes 969.7X34.5 B.t.u. per hour for one boiler horsepower. To determine the horsepower at which a boiler is working, we must therefore first find how many B.t.u. are used to generate one pound of steam under the given conditions of steam pressure and temperature of feedwater. The number of pounds of water evaporated per hour multiplied by the number of B.t.u. to gen- iA. S. M. E. POWER TEST CODE, Edition of 1915, Table 4, 21. BOILERS 39 erate one pound of steam under the conditions will give the num- ber of B.t.u. used by the boiler per hour. This product divided by 969.7X34.5 will give us the horsepower of the boiler. EXAMPLE. It is required to find the horsepower of a boiler working under the following conditions: Steam pressure = 115 pounds gage. Temperature of feedwater = 65 F. Quality of steam = 98% (i.e., 2% priming). Water fed to boiler per hour = 3640 pounds. SOLUTION. From the steam tables it is seen that The heat of the liquid at 115 Ib. gage (129.7 Ib. abs.) = 318.4 B.t.u. The heat of vaporization at 115 pounds gage =872.3 The heat of the liquid at 65 F. =33.1 B.t.u. The heat required to evapo- rate one pound of water under the above conditions is 318. 4 + . 98X872. 3 33.1 = 1140 B.t.u., and the B.t.u. used per hour is 1140X3640 = 4150000. Hence the horsepower of the boiler is 4150000/(34.5X969.7) = 124 h.p. 38. Factor of Evaporation. Since it takes 969.7 B.t.u. to evaporate a pound of water from and at 212 F., and since it takes more (1140 B.t.u. in the previous example) to evaporate a pound of water under the actual conditions that exist in the boiler, a certain ratio exists between these amounts. The factor of evaporation is the ratio of the amount of heat required to evap- orate a pound of water under actual conditions to the amount re- quired to evaporate a pound from and at 212 F. In the previous example, the factor of evaporation was 1140/969.7 = 1.176. If the factor of evaporation is known, the equivalent evaporation is found by multiplying the actual evaporation by this factor. 1 By this method, the heat used to raise the temperature of the moisture in the steam from the temperature of the feed water to that of the steam is not considered in computing the factor of evaporation. In most cases the difference in results due to this omission is very small. 39. Efficiency of Boilers. Usually speaking, the efficiency of anything is the ratio of what is got out to what is put in ; output and input being measured in like units. For boilers, the term efficiency means the ratio of the number of B.t.u. in the steam generated to the number of B.t.u. available in the coal fired. Boiler efficiency is usually expressed in percent. 1 In the A. S. M. E. POWER TEST CODE the "Mean B.t.u." and steam tables by MARKS and DAVIS are used, thereby giving 970.4 B.t.u. instead of 969.7 B.t.u. as used above for heat required to evaporate a pound of water from and at 212 F. See POWER TEST CODE, Edition of 1915, pp. 28 and 47. 40 ENGINES AND BOILERS The fact that the combined efficiency of a boiler, furnace, and grate is not 100% is due to several losses. These losses are due to the following causes. (1) A part of the fuel may drop through the grate and be lost in the ash. (2) Heat is lost up the stack. There are several sources of this loss, and to them is due the greatest loss in efficiency. First, unburned particles of solid fuel are often carried from the fur- nace. The amount depends upon the draft and the kind of fuel. In locomotives, with their high draft and with a fine fuel, this loss may be considerable. Second, there is loss due to the un- burned or partially burned hydrocarbons. Black smoke is caused by the incomplete burning of some of the hydrocarbons. Third, heat is carried away by the excess air and the inert nitrogen which have been heated, and by the hot products of combustion. Fourth, heat is required to evaporate and to superheat the moisture in the fuel and in the air. Fifth, there may be loss due to the burning of the carbon to carbon monoxide instead of to carbon dioxide. (3) Heat is lost by radiation from the furnace and from the boiler. It is very difficult to separate all these losses and the attempt is seldom made. It must be remembered that what is often called boiler efficiency is really the combined efficiency of grate, furnace, and boiler. Under the most favorable conditions, using coal as a fuel, efficiencies of over 80% have been attained. Under ordinary conditions of operation, efficiencies vary from 80% to less than 50%. The efficiency may sometimes exceed 80% when underfeed stokers, described later, are used. When oil is used as a fuel, higher efficiencies may be attained, due in part to the better mix- ing of the air and the fuel. EXAMPLE. It is required to find the combined efficiency of a boiler, fur- nace, and grate, working under the following conditions: Steam pressure = 127 pounds gage. Superheat = 190 F. Temperature of feedwater = 180 F. Water fed to boiler per hour = 8750 pounds. Coal fired per hour = 1 160 pounds. B.t.u. per pound of coal as fired = 11540 B.t.u. BOILERS 41 SOLUTION. The B.t.u. required to generate one pound of steam under the above conditions is seen to be 325.4+866.8 - 148+.55 X 190 = 1148.7 B.t.u. The total B.t.u. used in the generation of steam per hour then is 8750 X 1148.7 = 10051000 B.t.u. The total B.t.u. in coal fired per hour is 1160X11540 = 13386000 B.t.u. Hence the efficiency is 10051000/13386000 = .752 or 75.2%. 40. A. S. M. E. Boiler Test Code. 1 In reporting the results of a steam-boiler test it is well to put them in the form prescribed by the A. S. M. E. This form is as follows. DATA AND RESULTS OF EVAPORATIVE TEST CODE OF 1915 (1) Test of boiler located. To determine Test conducted by DIMENSIONS (2) Number and kind of boilers (3) Kind of furnace (4) Grate surface (width length ) sq. ft. (a) Approximate width of air openings in grate in. (6) Percentage of area of air openings to grate surface per cent (5) Water heating surface sq. ft. (6) Superheating surface sq. f t. (7) Total heating surface sq. f t. (a) Ratio of water heating surface to grate surface (6) Ratio of total heating surface to grate surface (c) Ratio of minimum draft area to grate surface (d) Volume of combustion space between grate and heating surface cu. ft. (e) Distance from center of grate to nearest heating surface ft. DATE, DURATION, ETC. (8) Date (9) Duration hr. (10) Kind and size of coal AVERAGE PRESSURES, TEMPERATURES, ETC. (11) Steam pressure by gage Ib. per sq. in. (a) Barometric pressure in. of mercury (12) Temperature of steam, if superheated deg. (a) Normal temperature of saturated steam deg. (13) Temperature of feedwater entering boiler deg. (a) Temperature of feedwater entering economizer deg. (6) Increase of temperature of water due to economizer deg. i A. S. M. E. POWER TEST CODE, Edition of 1915, p. 51. 42 ENGINES AND BOILERS (14) Temperature of escaping gases leaving boiler deg. (a) Temperature of gases leaving economizer deg. (6) Decrease of temperature of gases due to economizer deg. (c) Temperature of furnace deg. (15) Force of draft between damper and boiler in. of water (a) Draft in main flue near boiler in. of water (6) Draft in flue between economizer and chimney in. of water (c) Draft in furnace in. of water (d) Draft or blast in ash-pit in. of water (16) State of weather (a) Temperature of external air deg. (6) Temperature of air entering ash-pit deg. (c) Relative humidity of air entering ash-pit per cent QUALITY OF STEAM (17) Percentage of moisture in steam or number of degrees of superheating per cent or deg. (18) Factor of correction for quality of steam TOTAL QUANTITIES (19) Total weight of coal as fired Ib. C 20) Percentage of moisture in coal as fired per cent (21) Total weight of dry coal (Item 19X (22) Ash, clinkers, and refuse (dry) (a) Withdrawn from furnace and ash-pit Ib. (6) Withdrawn from tubes, flues and combustion chamber Ib. (c) Blown away with gases Ib. (d} Total Ib. (e) Weight of clinkers contained in total ash Ib. (23) Total combustible burned (Item 21 Item 22d) Ib. (24) Percentage of ash and refuse based on dry coal per cent (25) Total weight of water fed to boiler Ib. (26) Total water evaporated, corrected for quality of steam (Item 25 X Item 18) Ib. (27) Factor of evaporation based on temperature of water entering boiler. . . . (28) Total equivalent evaporation from and at 212 degrees (Item 26Xltem 27) Ib. HOURLY QUANTITIES AND RATES (29) Dry coal per hour Ib. (30) Dry coal per square foot of grate surface per hour Ib. (31) Water evaporated per hour, corrected for quality of steam Ib. (32) Equivalent evaporation per hour from and at 212 Ib. (33) Equivalent evaporation per hour from and at 212 and per square foot of water heating surface Ib. CAPACITY (34) Evaporation per hour from and at 212 (Same as Item 32) Ib. (a) Boiler horsepower developed (Item 34-5-34^) bl.-h.p. BOILERS 43 (35) Rated capacity per hour, from and at 212 ........................ lb. (a) Rated boiler horsepower .............................. bl.-h.p. (36) Percentage of rated capacity developed ...................... per cent ECONOMY (37) Water fed per pound of coal as fired (Item 25-r-Item 19) ........... lb. (38) Water evaporated per pound of dry coal (Item 26 -^ Item 21) ........ lb. (39) Equivalent evaporation from and at 212 per pound of coal as fired (Item 28 4- Item 19) .......................................... lb. (40) Equivalent evaporation from and at 212 per pound of dry coal (Item 28^- Item 21) ............................................... lb. (41) Equivalent evaporation from and at 212 per pound of combustible (Item 28 -f- Item 23) .......................................... lb. EFFICIENCY (42) Calorific value of 1 pound of dry coal by calorimeter ............ B.t.u. (a) Calorific value of 1 pound of dry coal by analysis ......... B.t.u. (43) Calorific value of 1 pound of combustible by calorimeter ......... B.t.u. (a) Calorific value of 1 pound of combustible by analysis ...... B.t.u. (44) Efficiency of boiler, furnace and grate / Item40X970.4\ ( 100X -TtoZlar-) ................ percent (45) Efficiency based on combustible / in Item4lX970.4\ ( 100X - Item 43 J ............... P6F C6nt COST OF EVAPORATION (46) Cost of coal per ton of ...... pounds delivered in boiler room ......... ................ dollars (47) Cost of coal required for evaporating 1000 pounds of water under ob- served conditions ......................................... dollars (48) Cost of coal required for evaporating 1000 pounds of water from and at 212 .................................................... dollars SMOKE DATA (49) Percentage of smoke as observed ............................ per cent (a) Weight of soot per hour obtained from smoke meter ..... per cent FIRING DATA (50) Kind of firing, whether spreading, alternate or coking ................ (a) Average thickness of fire .................................. in. (6) Average intervals between firings for each furnace during time when fires are in normal condition ...................... min. (c) Average interval between times of leveling or breaking up ....... . . min. 44 ENGINES AND BOILERS (51) Analysis of dry gases by volume (a) Carbon dioxide (CCh) per cent (b) Oxygen (O) per cent (c) Carbon monoxide (CO) per cent (d) Hydrogen and hydrocarbons per cent (e) Nitrogen, by difference (N) per cent (52) Proximate analysis of coal As fired Dry coal Combustible (a) Moisture (6) Volatile matter (c) Fixed carbon. ... (d) Ash 100 per cent 100 per cent 100 per cent (e) Sulphur, separately determined referred to dry coal per cent (53) Ultimate analysis of dry coal (a) Carbon (C) per cent (b) Hydrogen (H) per cent (c) Oxygen (O) per cent (d) Nitrogen (N) per cent (e) Sulphur (S) per cent (/) Ash per cent 100 per cent (54) Analysis of ash and refuse, etc. (a) Volatile matter per cent (6) Carbon per cent (c) Earthy matter per cent 100 per cent (d} Sulphur, separately determined per cent (d) Fusing temperature of ash deg. (55) Heat balance based on dry coal . Dry Coal B.t.u. Percent (a) Heat absorbed by the boiler (Item 40X970.4) . . . (6) Loss due to evaporation of moisture in coal (c) Loss due to heat carried away by steam formed by the burning of hydrogen (d) Loss due to heat carried away in the dry flue gases (e) Loss due to carbon monoxide OLoss due to combustible in ash and refuse i Loss due to heating moisture in air (h) Loss due to unconsumed hydrogen and hydrocar- bons, to radiation and unaccounted for . . (i) Total calorific value of 1 pound of dry coal (Item 42) 100 CHAPTER V BOILER ACCESSORIES AND AUXILIARIES 41. Grates. Grates are used to support the fuel in a furnace. Most grates are made of cast iron, which is cheap and less liable than other convenient materials to be distorted or twisted under the high temperatures to which it is subjected. The grate must be strong enough to support the load placed upon it, and it must be of such a form that sections can easily be replaced when broken or burned out. It must have sufficient opening for the admis- sion of air to the fuel. The openings or air spaces depend upon the kind of fuel used. The combined area of the openings will usually be from 30 to 50 percent of the total area. The area of the grate depends upon the amount of coal to be burned and the rate of combustion. Under natural or chimney draft, from 10 to 25 pounds of coal can be burned per square foot of grate surface per hour. Under forced draft, from 40 to 130 pounds of coal may be burned per hour. If hand firing is employed, the grate must not be longer than the distance the fireman can throw the coal accurately (six or seven feet) . Depend- ing upon the fuel, draft, and economy of the boiler, the equivalent evaporation per pound of coal will vary from 5 to 12. Various forms of grates are used. For hand firing, plain grates and shaking or dumping grates are used. The plain grate is harder to keep clean than a dumping grate. Moreover, it is necessary to keep the fire-doors open while the cleaning is in process. The grates used in mechanical stokers are of various types; some are stationary, others traveling and rocking. Occa- sionally grates are water-cooled, to prevent their burning out. Since this water is led to the boiler after becoming heated in the grate, the boiler capacity is increased, but in most cases the extra care and cost are prohibitive. 42. The Plain Grate. The grate bars shown in Figs. 16 and 16a are of the stationary type. These grates are cast in small sections so that a section may be easily and quickly replaced when it is burned out. The size of the openings in the bars is governed by the size and kind of coal that is to be burned. If 45 46 ENGINES AND BOILERS anthracite coal is used, the openings are small. If the coal is bituminous, and if it cakes, the openings should be made large. 43. The Rocking Grate. A form of rocking grate is shown in Fig. 17. The bars are supported on pivots, and are dumped or rocked by means of a lever from the front of the furnace. Only the largest clinker need be removed from the top, since the rocking action of the bars breaks up most of the clinker that is formed. In case a strong draft is used, as in the locomotive, this type of grate is usually used in order to keep a clean fire, such as is required with a high rate of combustion. 44. Mechanical Stokers. The first cost of a mechanical stoker is greater than the equipment for hand firing, but it Herat/on \ g cfioi? at 4-3 ' FIG. 16 FIG. 16a tr rr requires less labor and attendance in its operation. A cheaper grade of fuel can be used, a higher efficiency attained, and less smoke is formed than is usual with hand firing. In a fair-sized or large plant, it is usually better economy to use some form of mechanical stoker. There are many forms of stokers in use. 45. The Chain Grate Stoker. Where a low-grade fuel is used, as is often the case in the middle west and in the central states, the chain grate (Fig. 18) is extensively used. The grate is composed of a large number of short links, forming an endless chain. This chain runs over front and rear sprockets. Power is used to drive one of these sprockets, causing the whole chain to revolve slowly at a speech which is regulated by a suitable mechanism. The whole grate is mounted on wheels so that it can be run out in the open for repairing and cleaning. BOILER ACCESSORIES AND AUXILIARIES 47 The coal is fed to the front of the grate from a hopper which extends across the entire width of the grate. At the rear of the hopper there is a plate lined with firebrick that may be raised or lowered, thus regulating the depth of fuel-bed. The volatile matter in the fuel is distilled off as the coal first enters the fur- nace. These volatile products pass back over the part of the fire where the fixed carbon is burning, and are given a chance to burn there. By the time the fuel-bed has reached the rear of the furnace the combustion should be complete. The ash and clinker are dropped off to the ash-pit at the rear. STirry. ca- t "' o. : 'eg' .-ara>o o ^-^^L '6'/l'6 ^0 A d^ "tffiuLllAJiH h a Lir in which D is the diameter of the valve in inches, H is the heat- ing surface of the boiler in square feet, L is the lift of the valve in inches, and P is the absolute boiler pressure in pounds per square inch. It is noticed that smaller valves are required for locomotives, because the maximum draft can be secured only when the steam is being drawn from the boiler by the engine. 57. Other Safety-valve Formulas. Various cities and states have their own rules governing the sizes of safety valves, a few of which are given below. CITY OF CHICAGO. One square inch of pop-valve area (7rD 2 /4) for every three square feet of boiler grate area. CITY OF PHILADELPHIA. For pop valves, A = 22.5XCr/(p 8.62), in which A is the area of the valve in square inches (not the effec- tive opening for the escape of steam), G, the grate area in square feet, and p the gage boiler pressure. U. S. SUPERVISING INSPECTORS. A = .2Q74:XWH/P, in which A is the area of the valve as in the previous formula, WH is the number of pounds of water evaporated by the boiler per hour, and P is the absolute boiler pressure. Safety valves are not made in sizes over 5 or 6 inches in diameter. In large boilers it is therefore necessary to use more than one. EXAMPLE. What should be the size of pop safety valve on a boiler with 1500 square feet heating surface if the pressure carried is 130 pounds gage? SOLUTION. Assuming a maximum rate of evaporation of 8 pounds of water per square foot of heating surface per hour, we get 8X1500 = 12000 pounds of steam to be discharged per hour through the valve. The weight discharged per second is 12000/3600 = 3.33 pounds. From Napier's formula, W = AP/70, we see that 3.33 =AX(130+15)/70, whence A, the area of opening, in square inches, is 1.61 Assuming a lift of valve equal to 1/30 of the diameter, the area of the opening will be approximately .7077rD 2 /30. Then .7077rZ) 2 /30 = 1.61, or D = 4.67 inches. Hence a 5-inch valve should be used. 58. The Water Glass or Gage Glass. In order that the amount of water in the boiler may be known, a water glass is attached. The lower end of the water glass is attached to the 60 ENGINES AND BOILERS water space, and the upper to the steam space. Since there is danger of the glass becoming stopped in these connections, and the water level thereby being falsely indicated, or of the glass being broken, three gage-cocks or try-cocks are placed on the boiler or water column. The top cock is placed above, and the bottom one below, the normal water level. By open- ing these cocks in succession, one may determine whether or not the gage glass is giving the correct level. 59. High and Low Water Alarm. High- water and low- water alarms are sometimes used to attract the attention of the fireman when the water falls below or rises above the safe level. The alarm is operated by means of a float in the water column. When this float rises too high or falls too low it will open a valve, and allow the escaping steam to blow a whistle. (See Fig. 23.) 60. Fusible Plug. Another safety device used to detect low water is the fusible plug. This plug (Fig. 24) has a tin core that will melt when the water level falls below it. These plugs are placed in the crown sheet of an internally fired boiler in the rear head a little above the top tubes in the return-tubular boiler and in the bottom of the steam drum of a water-tube boiler. They should be kept free from scale on the inside and from soot on the outside. None of the above safety devices are absolutely certain in their action. Under conditions of very rapid steaming or with feedwater that foams, the water level in the boiler may be below that in the water column. FIG. 23 //7S/t/f or arfsjure iat/Se ry/*e Outft'efe ft/pes 0ttts/e/e or // FIG. 24 BOILER ACCESSORIES AND AUXILIARIES 61 61. Boiler Feedwater Treatment. The impurities in water that are responsible for most of the scale formation are the car- bonates and sulphates of calcium and magnesium. If muddy water is used, the mud may be deposited on the heating surface and aid in scale formation. The carbonates of lime and magnesium are soluble in water containing carbon dioxide. These carbonates cause what is known as temporary hardness. . Upon heating to 212 F. the carbon dioxide is driven off, and the carbonates are precipitated. If these are the only impurities in the feedwater, an open feedwater heater will remove most of the scale-forming material. Where a heater is not used the carbonates may be precipitated by the addition of a solution of slacked lime. The lime combines with the carbon dioxide to form the insoluble monocarbonate of lime. The sulphates of lime and magnesium are not precipitated at a temperature of 212, but are precipitated at a temperature such as exists in the boiler. They cause what is known as permanent hardness. The addition of carbonate or hydrate of soda (or a mixture of the two) will cause precipitation. The carbonate of soda decomposes the sulphates and forms insoluble carbonates of lime and magnesium, which precipitate, leaving neutral soda and sodium sulphate in solution. If carbon dioxide is present, the soluble bicarbonate of lime is formed, which may be precipitated by heating or by the addition of lime as explained previously. In most purification processes both the lime and soda are used. If organic matter, from sewage or from some other source, is present in the water, it may be removed by filtration. Before passing the filter a coagulant, such as alum, is often used. Organic matter in feedwater is often the cause of foaming. 62. Scale Prevention and Removal. Many substances have been used to prevent the formation of scale. Some of these probably do as much damage to the boiler as would the scale. Aside from the treatment to remove the scale-forming material, the best substance seems to be graphite. When it is injected into the boiler, it is said to help in the prevention of scale formation. Where no precaution is taken to prevent the scale from form- ing, it is necessary to clean it from the tubes periodically. This is usually done by means of a cutter or hammer that is driven by a small air, steam, or water turbine. 62 ENGINES AND BOILERS Figure 25 shows one make of cleaner that is applied to a fire tube. Figure 26 shows the form that is applied to a water tube. FIG. 25 Jca/a tfrr/zeafaf jecf/o/J t/rfer fi/bc FIG. 26 63. Oil Separators. In plants where all or part of the steam is condensed and used again as boiler feed, the oil that was used to lubricate the engine will find its way to the boiler. This does not apply to steam turbines, as oil is not usually used internally with them. This oil forms a very hard scale that it is almost im- possible to remove. To pre- vent this, oil separators are used to remove the oil, either from the exhaust steam or from the water after it is con- densed. The removal before condensation is preferable, since the oil does not have to be contended with in the con- denser or feedwater heater. Figure 27 shows an Austin oil separator. Since the separator is quite large, the steam passes through it with a small veloc- ity and deposits the oil on the surface of the corrugated ver- tical baffle plate shown in plan and section in the figure. With high vacua, a spray of water keeps the surface moist, which aids in the separation FIG. 27 of the oil. BOILER ACCESSORIES AND AUXILIARIES 63 64. Boiler Feed-pumps. The feedwater is usually forced into a boiler by means of a pump. Figure 28 shows a common type of boiler feed-pump. This pump is direct acting, the steam piston and water piston or plunger being fastened to the same piston rod. Since the steam and the water in a boiler are both under the same pressure and since pipes and fittings offer a resistance to the flow of both water and steam, it is seen that it is necessary to FIG. 28 make the steam piston of the feed-pump considerably larger in diameter than the water plunger. Steam is admitted by the steam valve alternately to the two ends of the steam cylinder. At the same time that the valve is admitting steam to one end of the cylinder it is allowing it to exhaust from the other, thus giving the piston a reciprocating motion. The water piston sucks up water from the suction pipe in one end of the water cylinder while forcing it into the delivery pipe on the other. The suction pipe leads either to the hot well or to the cold well from which the feedwater is taken. The lower end of the suction pipe should be provided with a strainer to prevent any large pieces of solid matter from getting into and clogging the valves. A foot valve should be provided also to keep the suction pipe 64 ENGINES AND BOILERS full of water when the pump is not running, thus eliminating the priming of the pump every time it is started. There are two sets of valves at each end of the water cylinder. On the suction stroke, the suction valves are lifted and the water is sucked in back of the piston. During the forcing stroke, these valves are closed and the water is forced out through the upper set and into the delivery pipe. A light spring is employed to help seat the valve. Most valves are faced with a composition disc which may be replaced when it becomes worn. The flow from a reciprocating pump is not steady; to insure a more uniform rate of flow of the water an air chamber is placed on the delivery line close to the pump. This is kept partly filled with air, which acts as a cushion. A check valve is placed in the feedline between the pump and the boiler. This prevents the water from the boiler escaping back through leaky valves when the pump is not in full operation. There should also be a stop valve in the feedline. There are two types of reciprocating steam pump, one in which there is only a single steam cylinder and a single water cylinder, and the other in which there are two steam cylinders and two water cylinders placed side by side. The latter type is called a duplex pump. In this type the valve for one steam cylinder is operated by the movement of the piston of the other steam cylinder. These boiler feed-pumps take steam the full length of the stroke, not allowing it to expand in the cylinder, and they are not eco- nomical in the use of steam. (See Chapter VI on the steam en- gine.) However, only a small proportion of the steam generated by the boiler is needed to run the pump. To secure better economy, feed-pumps are occasionally driven by power taken from the main engine. The supply of water from these pumps is not easily regulated. They are made to pump more water than is normally required, the excess being passed back to the suction through a relief valve. In large electric power plants, triplex pumps, driven by electric motors are often used for boiler feeding. Of late years centrifugal and turbine pumps have been employed for boiler feeding. An automatic regulator is sometimes installed with the pump so that the pump will furnish just the proper amount of water to keep the boiler water-level constant. BOILER ACCESSORIES AND AUXILIARIES 65 65. The Injector. On portable boilers and in small plants, the water is often forced into the boiler by means of an injector or inspirator. This is usual also on locomotives. The principle upon which the injector works is illustrated by Fig. 29. Steam from the boiler enters the injector through a steam nozzle, a, in which it expands and some of its heat energy is transformed into kinetic energy. The steam leaves the nozzle with a high velocity and enters a small combining tube, b. The water inlet leads to a chamber which is located between the nozzle and the combining tube. As the steam flows from the nozzle to the combining tube it tends to form a par- tial vacuum in the water chamber and thus sucks up and car- ries the water along with it. The steam mixes with the water in the combining tube and is condensed. This mix- ture of condensed steam and water has a high velocity and therefore a considerable amount of kinetic energy; its pressure, however, is " atmospheric or less. This mixture passes from the combining tube to the delivery FIG. 29 tube, c, which has an increasing diameter. The mixture therefore loses a large amount of its velocity and kinetic energy. What it loses in kinetic energy it gains in pressure energy, so that by the time it leaves the in- jector it has gained enough pressure to force open the check valve leading to the boiler. An overflow is located at the end of the combining tube, so that when steam is first turned on, it escapes through the overflow. The overflow is fitted with a valve which automatically closes when the pressure inside the combining tube falls below the pressure of the atmosphere, thus pre- venting air from coming into the injector and impeding its action. 66 ENGINES AND BOILERS Unless specially constructed, an injector cannot lift water a very great height. Moreover, since the injector must condense the steam in order to work at all, it is necessary that the water be cold. Considered as a pump, the efficiency of the injector is very low, because the greater part of the energy of the steam goes to heat the water. If it is used to feed a boiler, the heat spent in raising the temperature of the feedwater is not lost, as it goes back into the boiler. Hence it is efficient for this purpose. The injector is light, occupies but little space, and is cheaper than a pump, but it is not so dependable. 66. Boiler Feeding by Return Trap. The condensation from various parts of the plant is sometimes returned to the boiler by what is known as a return trap. This trap is located above the level of the boiler and the water runs into the boiler under the influence of gravity and the pressure of live steam. These traps are quite economical in the use of steam and they may be used to supply all the feedwater. They are not nearly so reliable as the pump or injector, however, and are therefore but little used to furnish the entire feedwater supply. 67. The Steam Line. Steam pipe is made of wrought iron or of steel. The nominal diameter corresponds approximately with the inside diameter. Sizes of standard pipe vary, by the y%" from y 8 " to y 2 ", by the M" from %* to IJ^", by the %' from IY 2 " to 5", and by the 1" from 5" to 15". It has been customary to allow an average velocity of steam in the line of from 4000 to 6000 feet per minute. In modern turbine plants, however, where the flow is uniform, and especially where superheated steam is used, velocities much in excess of these values are used. If a velocity of wet steam much greater than that just mentioned is used, the drop in pressure due to skin friction will be excessive. On the other hand, if a velocity much less is allowed, too large and expensive a pipe will be required. If the volume and velocity of steam to be carried by the pipe line are known, the diameter is easily determined. The volume carried per unit of time equals the product of the area of the cross-section of the pipe and the velocity. If the size and speed of the engine to be supplied are known, we may compute the vol- ume of steam needed. At maximum it may be assumed that the engine takes steam during the full length of the stroke. When BOILER ACCESSORIES AND AUXILIARIES 67 more than one boiler is used it is customary to discharge the steam into a common pipe called a header. In such a case each boiler should be provided with a non-return stop-valve be- tween the boiler and the header. This non-return stop-valve acts as a check valve in case the direction of steam flow should be reversed, which would happen in case a tube blew out or some other similar accident occurred. The piping must be provided with a sufficient number of hangers to prevent breaking due to its own weight. The line should slope downward in the direction the steam is to flow in order that the condensation may be carried along with the steam. If this precaution is not followed condensed steam will collect in the pipe and may be carried in slugs by the steam in amounts large enough to injure and cause leakage or even breakage of the fittings. Provision should be made at the low points of the line to remove condensation. A pipe should be run down from the low point and the Va/re seaf fa/re water collecting in this may be blown out from time to time by open- ing a valve by hand. A trap may be installed that will remove it automatically. 68. The Steam Trap. - Several types of traps are in use. In the more common kinds, the valve is operated by means of either a * T float, the unequal ex- pansion of two differ- ent metals with chang- ing temperature, press- ure of collected water on a flexible diaphragm, or the weight of a bucket as it fills with water. The latter kind is illustrated in Fig. 30. In this type the buoyancy of the bucket keeps the valve closed until enough water flows over the FIG. 30 68 ENGINES AND BOILERS edge and collects in the bucket to sink it. The sinking of the bucket opens the valve and the water collected in the bucket is forced out through the valve by the steam pressure inside the trap. The bucket now being lightened, it again rises, closing the valve. In many traps, the valve is operated by a float. The water collects in a float chamber and raises the buoyant float until the valve is opened. The water then escapes until the float is lowered enough to allow the valve to seat. An air valve is located at the top of the trap to allow the air to escape if enough should be caught there to interfere with the operation of the trap. Another form of trap is one in which the valve is operated by the unequal expansion of two metals. When the trap is cold the valve is open and the water is allowed to escape. As soon as the steam flows through, however, the parts are heated and ex- FIG. 31 FIG. 32 pand unequally, closing the valve. Water then collects again, and as the parts cool, the valve will again open and the opera- tion will be repeated. When large amounts of water are to be handled, dumping traps may be used. The discharged water from the trap is led to the drain or is piped back to the hot well. 69. Expansion Joints. Since the pipe is laid cold, it will ex- pand when steam is turned into it and its temperature becomes that of the steam. The expansion amounts to 2.5 inches per hundred feet of pipe with ordinary steam temperatures, and may be greater when the steam is of very high pressure and is super- heated. The piping must be so arranged that this expansion may take place without injury to the pipe. If the pipe is not laid straight but contains elbows, it may bend enough so that no dan- gerous stresses will be induced. If there is a considerable run of straight pipe, however, expansion joints must be provided. There are several types of expansion joints in use. A very com- mon kind for use with low-pressure steam is the slip- joint. In BOILER ACCESSORIES AND AUXILIARIES 69 this, provision is made for the slippage of one part of the joint on the other. The joint is kept steam tight by means of a stuf- fing box. Figure 31 shows this type. Goosenecks and expan- sion loops (Fig. 32) are used when the steam pressure is high. 70. Steam Separators. Unless superheat is used, steam leav- ing the boiler will always contain some moisture. If the steam- pipe is very long, some condensation also takes place. Due to these causes, the steam is liable to reach the engine quite wet. It is desirable both for safety and for economy to have the steam as dry as possible when it enters the engine. To remove the moisture from the steam, a separator is placed in the line just before it reaches the engine. The steam is given a sudden change in direction upon enter- ing the separator. The moisture resists this change to a greater extent than does the steam. In the type shown in Fig. 33 the steam is first deflected downward and then upward, and as the moisture cannot change its direction of motion as rapidly as the steam, it is caught and collected in the bottom of the separator. In some makes the steam is given a whirl- ing motion and the water, being denser than the steam, is forced to the outside of the separator, where it is collected. Another type, which is similar to the oil separator of Fig. 27, is that in which a corrugated baffle plate is interposed in the path of the steam. The steam passes around the baffle while the moisture is caught by it and runs down the corrugations to the bottom of the separator, where it is collected. A separator should remove most of the moisture, but it should not offer too great a resistance to the passage of the steam, since this would cause a drop in pressure. The moisture, after being collected, is trapped off and discharged to the drain or returned to the hot well. Often the separator is made large and acts as a steam receiver. This reduces the pulsation in the steam line when the steam is used by a reciprocating engine. 71. Steam-pipe Covering. To prevent radiation of heat from the steam-pipe and the consequent condensation, a covering is applied to the pipe. The covering is made from materials that FIG. 33 70 ENGINES AND BOILERS are poor conductors of heat. A finely-divided, dead air space is one of the best non-conductors of heat. In most coverings the object is to get as much finely-divided dead air space as possible. 7 r nX B.t.u. per pound of dry steam where n is the number of pounds of dry steam used per hour per horsepower. The thermal efficiency of a steam engine will seldom, if ever, exceed 25 per cent. This may seem to be a very low value, but it is impossible for the engine to use a very large part of the heat supplied due to the fact that the exhaust steam carries with it its heat of vaporization. On account of condensation of steam in the cylinder and other causes of heat loss, the efficiency of a reciprocating engine seldom approaches the efficiency of an ideally perfect engine working under the same range of pressure, but a well-designed steam turbine of large size may do so. 95. Cylinder Condensation. The largest single loss in the average engine is due to what is known as initial condensation. Since the cylinder walls are made of iron, which is a good con- ductor of heat, they naturally absorb heat from any hotter body or substance placed in contact with them and they give up heat to a cooler body. The steam comes into the cyl- inder at a relatively high pressure and temperature. Both the pressure and the temperature drop in the cylinder, and the steam leaves at a relatively low pressure and temperature. Since the cylinder walls are exposed first to hot, and then to cool steam, their temperature will never be as great as that of the incoming steam, nor as low, during operation, as that of the outgoing steam. When the steam first enters the cylinder and strikes the cooler walls, a part of its heat will be absorbed by the walls. This THE STEAM ENGINE 97 causes a partial condensation of the steam. Since the engine operates by virtue of the steam pressure and volume, it is readily seen that a shrinkage in volume causes a loss of work, and a lower- ing of efficiency. By the time the steam leaves the cylinder, it is cooler than the cylinder, and it takes back some of the heat it gave to the walls, but at too late a time to avoid the loss in efficiency. Depending upon the type of engine and the condi- tions of operation, the condensation may continue until release occurs, or re-evaporation may start during the expansion of the steam between cut-off and release. By computing from the in- dicator diagram the weights of steam at cut-off and at release, we find a net condensation during expansion if the weight at release is less than at cut-off, and a net re-evaporation if the weight at release is greater than at cut-off. The computation of condensation or re-evaporation during expansion is of little value since most of the re-evaporation occurs after release and before the steam leaves the exhaust ports. 96. Steam Accounted for by the Indicator Diagram. The A. S. M. E. code for testing steam engines calls for the computa- tion of the steam accounted for by the indicator diagram at points near the cut-off and release. Mark the points of cut-off and release and a point on the compression curve where we are sure the exhaust valve is closed, as in Fig. 47. Find the ratio of stroke at these points. The volume back of the piston at cut-off is the ratio of stroke at cut-off plus the ratio of p IG 47 clearance, shown by a, times the piston displacement. Scaling the pressure at cut-off from the diagram, we may compute the weight of dry steam back of the piston at cut-off by means of steam tables. Not all of the steam back of the piston at cut-off entered on that one stroke from admission to cut-off, since some of it was in the cylinder during compression. The amount that was ad- mitted is the weight at cut-off minus the weight caught at com- pression. The weight of steam compressed may be computed in a manner similar to that at cut-off. The weight of steam per hour accounted for by the indicator diagram is then equal to 98 ENGINES AND BOILERS FIG. 48 FIG. 49 FIG. 50 FIG. 51 FIG. 52 FIG. 53 FIG. 54 FIG. 55 FIG. 56 FIG. 57 THE STEAM ENGINE 99 (Wc.o.-TF C omp.)XWx60/i. hp., where W c . . and TF comp . are the weights of steam back of the piston at cut-off and compression, respectively. The weight is calculated separately for the head-end and crank-end of the cylinder, and the two values added to give the total for the engine. The weight accounted for at release is computed in the same way, but the two results usually differ slightly on account of the net condensation or the re-evaporation during ex- pansion from cut-off to release. The weight of steam accounted for by the indicator diagram will be considerably less than the actual amount used by the engine, because of the initial condensation of steam when it first enters the cylinder. 97. Valve-setting from the Indicator Diagram. It has been mentioned previously that one of the uses of the indicator is to assist in the setting of the valves While the subject of valve- setting will not be discussed thoroughly here, faulty setting may be recognized from the appearance of the diagram. Figure 48 shows the effect when the admission valve opens too soon. Figure 49 shows the results of late admission. Due to the tardiness of the valve opening, the steam is throttled, and the pressure for a large part of the stroke is lowered considerably. Figure 50 shows the effect of early compression. The steam caught when the exhaust valve closes is compressed to a pressure above that in the steam chest, the admission valve is lifted off its seat, and some of the steam escapes into the steam chest. Figure 51 shows almost no compression. This would cause no harm in a very slow-speed engine, but with higher speeds the steam caught at compression acts as a cushion and makes for smooth running. Figure 52 shows too early a release, and Fig. 53, too late a release. Both cause a loss in the area of the diagram. Figures 54 and 55 show unequal cut-off in the two ends of the cylinder. The crank-end is doing a much larger proportion of the work. The work done by the two ends should be about equal. Figure 56 shows improper lubrication of the indicator piston or the binding of some part. A wavelike motion of the curve is sometimes noticed when the diagram is taken from a high-speed engine, due to the vibration of the indicator spring, but it differs materially from Fig. 56. In Fig. 57, the indicator drum is strik- ing the stop on account of improper adjustment of the length of the cord connecting the indicator and the reducing mechanism. CHAPTER VII COMMON TYPES OF STEAM ENGINES 98. Slide-valve Engine. Where simplicity and reliability are of more importance than high efficiency, the slide-valve en- gine is used. The simplest type of slide-valve was shown in Fig. 39, and the principles of its operation were explained to some extent in the previous chapter. There are many varieties of slide-valves, the more common of which will be described later. Most of the smaller stationary engines in use are equipped with the slide-valve, and all American locomotive engines are of this type. While the plain slide-valve is very simple, it has certain defects. One of these is the impossibility of obtaining the proper steam distribution at all loads, i.e. of making the events of stroke occur at the proper place to give the highest efficiency at light load and also at heavy load. Various modifications and improvements have been made on the slide-valve to remedy this defect, the chief one of which is to place a second slide-valve on top of the main valve, and to control cut-off by a rider. Another defect of the plain slide-valve is the slowness with which the steam ports open up and close at some loads, which cause what is known as wire-drawing. This is simply a throttling of the steam by the valve as it enters the cylinder. This throt- tling usually causes a lowering of the efficiency of the engine. 99. The Corliss Engine. By far the most common type of high-grade reciprocating stationary steam engine in this country is the Corliss engine. The name comes from its inventor and first producer, GEORGE CORLISS, an engineer and engine builder of Providence, R. I. There are two distinguishing features of this engine. The first of these is the oscillating cylindrical valve. The second is the means for disengaging the valve from the mech- anism that drives it, and the quick closing of the valve after its disengagement. To understand the Corliss valve mechanism thoroughly, it is necessary to make a rather thorough analysis of its motion. We shall not do this in this chapter. The general principle of operation of the gear is fairly simple, however. Figure 58 shows a typical Corliss engine cylinder. The right end is cut in section to show the construction of the valves and 100 COMMON TYPES OF STEAM ENGINES 101 their locations relative to the cylinder. The left end shows an ordinary form of the mechanism that moves the steam valves and the exhaust valves. An eccentric on the shaft is connected to the hook rod which operates the valves through an eccentric rod and rocker arm. This gives the hook rod a horizontal recipro- cating motion that is nearly harmonic. The hook rod is attached - 3 fa am fff>e FIG. 58 to a wrist plate which is pivoted to the cylinder at C. The wrist plate is thereby given an oscillating motion. Four rods are attached to the wrist plate. The two upper rods are the steam rods, which transmit the motion to the two steam valves, and the lower or exhaust rods drive the two exhaust valves. The steam rods are attached to bell cranks or double arms that are pivoted on the valve spindle but are not attached to it. It is thus possible for the wrist plate and the bell crank 102 ENGINES AND BOILERS to move without affecting the valve in any way. The cylindrical steam valve has a spindle which extends out of the steam chest. To the outer end of this spindle there is keyed a steam arm. Any motion of this arm causes the valve to move. A block is attached to the back side of the steam arm, and a hook which is carried by the upper arm of the bell crank catches over it. As the steam rod moves to the right, the steam arm is picked up and the valve is turned. After having lifted the steam arm a certain distance, the hook is made to disengage -with the block and the steam arm is released. The steam arm is connected to the piston of a dash-pot by a dash-rod. As the steam arm is raised, a partial vacuum is formed in the dash-pot. When the steam arm is released from the hook, it is suddenly pulled downward by the vacuum in the dash-pot. As the steam arm is lifted, the valve opens and admits steam to the cylinder. When it is pulled down, the valve is closed suddenly, giving a quick cut-off. The time at which the hook is made to release the steam arm is controlled by a cam whose position is regulated by the governor. This cam engages with the tail of the hook and causes the disengagement. At light loads, the trip occurs soon and an early cut-off is given, and the cut-off is retarded as the load of the engine is increased. Figure 59 shows the trip mechanism on a larger scale. DAC is the bell crank. As the point D moves backward and forward in a nearly horizontal direction, the point C moves up and down in a nearly vertical direction. The pin C carries the steam hook. The tail of the hook engages with the knock-off cam, and its jaw engages with the block attached to the steam arm at B. For any one load the knock-off cam is stationary, and as C goes up, the tail of the hook is pushed away from A by the cam, which causes the latch to disengage with the block B. When there is a heavy load on the engine, the governor rod moves to the left, which raises the knock-off cam and makes the trip come later, giving a long cut-off. There is also a safety cam, shown in Fig. 59. If the governor fails to rotate, the safety cam comes into contact with the tail of the hook and prevents the picking up of the steam arm and therefore causes a failure to admit steam to the cylinder. There is no disengagement between the exhaust arm and the exhaust valve, so that the events of release and compression occur COMMON TYPES OF STEAM ENGINES 103 at the same ratio of the stroke for all loads. Both the steam valve and the exhaust valve of Fig. 58 are double-ported, which gives twice the opening for the passage of steam with the same valve movement as with the single-ported type. From the description just given it is seen that cut-off is inde- To forer/ior FIG. 59 pendent of the other events, and that the steam valve closes quickly, thereby preventing wire-drawing at closing. If a care- ful analysis of the motion is made, it will be seen that the steam valve opens the port nearly as widely at light loads as at full loads. The force necessary to operate the valve mechanism is not large and the work done in moving the valves is a very small part of the total output of the engine. 100. The Four-valve Engine. With a single slide-valve, the changing of one event necessitates the changing of all the others. To avoid this difficulty, engines are often made that 104 ENGINES AND BOILERS have four valves, a steam valve for each end of the cylinder, and an exhaust valve on each end. The exhaust valves are driven by a fixed eccentric, so that the release and compression are the same at all loads. The steam valves are controlled by the gov- ernor; hence cut-off and admission will vary for different loads. f FIG. 60 Figures 60 and 61 show one style of four- valve engine. This par- ticular engine has oscillating cylindrical valves similar to those shown in Fig. 58 for the Corliss engine. Some makers, probably to make use of the enviable reputation of the Corliss engine, call this type a non-releasing Corliss. This engine lacks, however, the distinct advantage of the trip found in the true Corliss type. FIG. 61 101. The Compound Engine. When all the expansion of steam takes place in one cylinder, we have what is known as a simple engine. If the steam passes through two successive cyl- inders, the engine is said to be a compound engine. If there are three successive cylinders, it is called a triple-expansion engine. COMMON TYPES OF STEAM ENGINES 105 If there are four successive cylinders, it is called a quadruple- expansion engine, etc. An engine may have two cylinders and not be compound, i.e. it may be a twin-cylinder engine, in which half the steam passes through one cylinder and half through the other. Likewise a compound engine may have three cylinders, all the steam passing through one high-pressure cylinder, and then dividing, half passing through each of the two low-pressure cylinders. The purpose of compounding is to reduce the initial conden- sation. It does not necessarily follow that there is a greater ratio of expansion of steam in a compound engine than in a simple engine, since that depends upon the point of cut-off. We have seen that initial condensation is caused by the range in temperature within the cylinder. The temperature range is less in each cylinder of a compound engine than in the single cylinder of a simple engine of the same capacity. Since the amount of condensation does not vary directly as the total temperature range, there may be considerably less total condensation if the steam is passed through two successive cylinders than if all the expansion occurred in one cylinder. Several years ago the idea of compounding was very popular and was carried to the extreme. Many triple-expansion engines, and some quadruple-expansion engines, were built. Experience proved, however, that there was a practical limit to which the idea might be carried. Now stationary engines are seldom built with more than two pressure-stages, except in direct -acting pumps. In the marine service, the triple-expansion engine is still popular, partly for the reason that it is desirable to have three cranks on the same shaft to give a greater uniformity of torque on the pro- peller shaft, and partly on account of the uniformity of load on marine engines. Of the many types of compound engines that have been built, only two are in common use in land service at present. We shall proceed to consider these. 102. The Tandem-compound Engine. In the tandem-com- pound engine, the pistons of the two cylinders are placed on the same piston rod, as shown in Fig. 62. The cylinder to the left is the high-pressure cylinder, and the one to the right is the low- pressure cylinder. The steam ports of the high-pressure cylinder are at a and 6; the exhaust ports of the low-pressure cylinder are 106 ENGINES AND BOILERS at c and d. The pistons, in Fig. 62, are shown moving to the right. Steam is entering the high-pressure cylinder through a, and leaving it through /. The exhaust steam from the crank end of the high-pressure cylinder passes to the head end of the low- pressure cylinder either directly or through a stationary vessel called a receiver, i.e. the back pressure on the piston A is the forward pressure on the piston B. On the return stroke, steam FIG. 62 enters the port b and the exhaust from the high-pressure cylinder leaves through e and enters the low-pressure cylinder at h either directly or through the receiver. With the tandem arrangement only one cross-head, connecting rod, crank, and frame are needed. In locomotive work, the Baldwin or Vauclain compound engine is sometimes seen. In this engine, the cylinders are placed side by side, and both piston rods attach to the same cross-head. The method of steam distribution is similar to that of the tan- dem type. Figures 63 and 64 show the high-pressure and low- FIG. 63 FIG. 64 FIG. 65 pressure indicator diagrams, as taken from a Baldwin compound engine. The high-pressure card comes from the crank end of the high-pressure cylinder and the low-pressure card from the head end of the low-pressure cylinder. These cards were taken with springs of different scales. Figure 65 shows the same diagrams when drawn to the same scale of pressure. It is noticed that the back- pressure line of the high-pressure card parallels the admission line of the low-pressure card from the left end up to the point of cut-off in the low-pressure cylinder. The reason for this is COMMON TYPES OF STEAM ENGINES 107 obvious, since the exhaust from the high-pressure cylinder passes directly to the low-pressure cylinder. When the admission valve of the low-pressure cylinder closes, compression must necessarily start in the high-pressure cylinder. Since it is neces- sary, with such a high pressure at exhaust as exists in the high- pressure cylinder, to have the compression occur late, it follows that cut-off must come very late in the low-pressure cylinder. This is not an ideal condition, but it is necessary if no receiver is placed between the two cylinders. If a receiver were placed between the two cylinders so that it could act as a reservoir into which to discharge, and from which to draw steam, it would not be necessary to have the preceding relation between compression and cut-off. 103. The Cross-compound Engine. In the cross-compound engine (Fig. 66), each cylinder has its own cross-head, connect- " FIG. 66 ing rod, crank, and frame. The cranks are usually spaced 90 apart. A by-pass is arranged so that the engine can be started even if the high-pressure cylinder stops on dead center, by admitting steam directly to the low-pressure cylinder. Each cylinder has its own valve mechanism, and the exhaust from the high-pressure cylinder passes into a receiver from which the low-pressure cylinder takes its steam. This arrangement per- mits a better steam distribution than that used in the tandem 108 ENGINES AND BOILERS type without a receiver. If the receiver is quite large, the back pressure in the high-pressure cylinder during exhaust will be nearly constant. Figures 67 and 68 show the indicator diagrams from a cross-compound engine. It should be noticed that the engine exhausts into a condenser. 104. Cylinder Ratio. The cylinder ratio of a compound engine is the ratio between the piston displacements of the low-pressure and the high-pressure cylinders. While it is not essential that the length of stroke be the same for both high-pressure and low-pressure cylinders of a cross- compound engine, they are made so. The cyl- inder ratio is then the ratio of the squares of the diameters of the low-pressure and high- pressure cylinders. 105. The Combined Indicator Diagram. The combined diagram is constructed by plot- ting both cards to the same scale of pressure and volume. Usually we do not change the low-pressure diagram but change the scale of the high-pressure card conform to it. Figure 69 shows the combina- tion of diagrams of Figs. 67 and 68. The low- pressure diagram is identical with Fig. 68, while the length of the high-pressure dia- gram equals the length of the low-pressure diagram divided by the cylinder ratio. The high-pressure diagram is placed to the right of the pressure axis its clearance distance, i.e. its distance from the axis equals the ratio of the high-pressure clearance times the new length of the high-pressure diagram. Condenser Pressure. FIG. 69 COMMON TYPES OF STEAM ENGINES 109 106. Diagram Factor. The definition of the diagram factor as given in the 1915 edition of the A. S. M. E. Power Test Code is as follows : The diagram factor is the proportion borne by the mean effective pressure measured from the actual diagram to that of a hypothetical diagram which represents the maximum power obtainable from the steam accounted for by the actual diagram at the point of cut-off; assuming first, that the engine has no clearance; second, that there are no losses through wire- drawing the steam either during admission or release; third, that the expansion line is a hyperbolic curve; and, fourth, that the initial pressure is that of the boiler, and the back pressure that of the atmosphere for a non-condensing engine, and of the condenser for a condensing engine. To determine the steam accounted for by the actual diagram at the point of cut-off, draw hyperbolic curves through the point of compression P and the point of cut-off (Fig. 70) until they FIG. 70 cut the boiler-pressure line at R and S. The length of RS is the length of the admission line for the hypothetical diagram, FA in Fig. 42, drawn to proper scale. The hypothetical diagram is drawn as in Fig. 42 except that the boiler pressure is taken as the initial pressure, release comes at the end of the stroke, the back pressure is the atmospheric pressure (condensing pressure in a con- densing engine), there is no compression, and there is no clear- ance. The hypothetical diagram for the combined diagrams of Fig. 69 is shown dotted. Since we assume there is no clearance, the length of the hypothetical diagram is equal to that of the low-pressure card. The distance RS at boiler pressure is deter- mined from the high-pressure diagram, as in Fig. 70. From S to T construct a hyperbolic curve, using the origin 0', and not 0. Release is at the end of the stroke and the back-pressure line is at the condenser pressure. *-*'*' /}b i<$ CX^Trar 110 ENGINES AND BOILERS In Fig. 69, the mean effective pressure of the combined dia- grams and of the hypothetical diagram are in the same ratio as the areas of the combined and hypothetical diagrams, because they are of the same length. To find the diagram factor of the combined cards, divide their area by the area of the hypothetical diagram. 107. Ratio of Expansion. The A. S. M. E. Power Test Code gives the following rule: To find the percentage of cut-off, or what may best be termed the commercial cut-off, the following rule should be observed: Through the point of maximum pressure during admission draw a line parallel to the atmospheric line. Through a point on the expansion line where the cut-off is complete, draw a hyperbolic curve. The intersection of these two lines is the point of commercial cut-off, and the proportion of cut-off is found by dividing the length measured up to this point by e total length> To find the ratio of expansion, divide the volume correspond- bg to the piston displacement, including clearance, by the olume of the steam at the commercial cut-off, including clear- ance. In a multiple-expansion engine the ratio of expansion is found by dividing the volume of a low-pf essure cylinder, including clearance, by the volume of the high-pressure cylinder at the commercial cut-off, including clearance. 108. The Unaflow Engine. The unaflow, or uniflow, engine is shown diagrammatically in Fig. 71. There is an admission valve at each end of the cylinder. The exhaust steam escapes through a port located around the circumference of the cylinder midway between the two ends. The piston, which is longer than in most engines, itself uncovers the exhaust port at about 90 per cent of the stroke. Compression must start when the piston is at the same place on the return stroke. Under non-condensing conditions this would give a very excessive compression pressure; hence the engine normally is run condensing, under which con- ditions the compression pressure is moderate. The thermal efficiency is about the same as that of a compound engine. The gain in efficiency over the ordinary double flow engine is due to the reduction of initial condensation. The condensation COMMON TYPES OF STEAM ENGINES 111 is reduced with the unaflow principle because the ends of the cylinder are kept hotter than the central portion. High- pressure steam never comes in contact with the central part of the cylinder and the flow of steam is from the ends toward the middle. The exhaust steam passing out through the central port does not cool the walls as much as it would if it flowed back to the ends of the cylinder upon leaving. In actual engines, pro- vision must be made for relieving the excessive compression pres- sure, should the vacuum break. This is done by a relief valve that adds to the clearance or allows the compressed steam to re-enter the steam chest, or by adding an auxiliary exhaust port nearer the end of the cylinder, which is opened automatically when the vacuum fails. CHAPTER VIII VALVES 109. Introduction. From our previous study of the steam engine we have learned that the purpose of the valve is to admit steam to the cylinder, and to release steam from it. The time at which the events occur must be such that the engine is capable of doing the work required, and that it may have as high an efficiency as possible under the conditions of operation. An en- gine may run with the valves improperly set or designed, but more steam will be used than if the valves functioned properly. 110. The D Slide-valve. The engine of Fig. 39 has what is commonly known as a D slide-valve. The valve slides back and forth on its seat, alternately opening and closing the ports. An eccentric on the shaft drives the valve. The eccentric rod is usually quite long in comparison with the throw of the eccen- tric, so that the valve may be considered to have the same motion as the horizontal component of the eccentric. In Fig. 72a, the valve is shown in mid-position; consequently the eccentric will either be directly above or directly below the center of the shaft. In mid-position, the valve laps over the edges of the port; the amount it extends over on the steam side is called the steam lap, and on the exhaust side, the exhaust lap. At the right side of Fig. 72a is shown the relative position of the eccentric and the crank. The eccentric leads the crank by angle 8. This angle 8 will be the same at all times during the revolution. In Fig. 726, the crank is on head-end dead center, and the valve is uncovering the head-end port a small amount. The amount the port is open when the crank is on dead center is called the lead. It is measured in inches. The valve in Fig. 726 is to the right of its mid-position by an amount equal to the steam lap plus the lead. In moving from the position shown in Fig. 72a to that in Fig. 726, the eccentric has moved a hori- zontal distance equal to the steam lap plus the lead, and has turned through a certain angle which is called the angle of advance. It is evident that the eccentric is ahead of the crank by an angle of 90 plus the angle of advance. It is customary to speak of the angle of advance and -not of the whole angle 6. Figure 72c shows the valve at head-end admission. The valve 112 VALVES 113 FIG. 72 114 ENGINES AND BOILERS is on the point of opening the head-end port for the admission of steam, and is traveling to the right. In the admission posi- tion it is to the right of its mid-position by a distance equal to the steam lap. Likewise the eccentric will be to the right of its mid-position by a horizontal distance equal to the steam lap. The crank is back of the eccentric by the angle 6 and is seen to be approaching its head-end dead-center position. If there is any lead, the crank will never be quite up to its dead-center position at admission. At head-end cut-off, Fig. 72d, the valve is to the right of its mid-position by a distance equal to the steam lap and its direc- tion of motion is to the left. The position is the same as for admission, but it is going in the opposite direction. The eccen- tric is now below the center line of the shaft. Figures 72e and 72/ show the relative positions at head-end release and compression. At both of these events, the valve is at the left of mid-position by an amount equal to the exhaust- lap distance. It is moving to the left at release and to the right at compression. In all the six diagrams of Fig. 72, the positions of the piston and its direction of motion are shown. The cylinder section in each diagram is in a horizontal plane and therefore is at an angle of 90 from the diagram showing crank and eccentric positions For an understanding of the slide-valve and the analyses to follow, it is essential that the student have a precise conception of the rel- ative positions of the valve on the seat, the eccentric and the crank relative to the center positions, and the position of the piston in the cylinder. 111. Relative Motion of Crank and Piston. Since the con- necting rod is not very long compared with the crank arm, we cannot assume that the horizontal movement of the crank is the same as the piston movement. This is clearly seen from the diagram in Fig. 73. As the crank moves from A to C, the cross- head moves from E to D. That part of the stroke completed by the cross-head is a'. It is evident that a' is considerably larger than the horizontal movement of the crank in going from A to C. In our analysis of valve motions, it is not customary to draw VALVES 115 in the cross-head D to find the proportion of the stroke at dif- ferent crank positions, but the following scheme is used. The horizontal diameter of the crank circle AB is extended to the left. With the length of connecting rod DC as a radius at the desired scale, and with D as a center, strike the arc shown by the dotted line CG. This gives the distance AG = a, on the diame- ter of the crank circle, that is equal to ED = a', the movement of FlG. 73 the piston or cross-head from E to D. In the valve analysis it is not necessary to draw the crank circle to any particular scale if we keep the proper ratio between the lengths of the crank and the connecting rod. This ratio usually is expressed as R/L. No matter what the scale of the crank circle may be, the ratio of a to the length of stroke will be constant. 112. Valve Diagrams. Many diagrams have been used to show graphically the relation of the movement of the valve to the movement of the piston, or of the relative movements of eccentric and crank. Only those most commonly used in this country will be explained here, i.e. the valve ellipse, the Bil- gram diagram, and the Zeuner diagram. 113. The Valve Ellipse. In this diagram the system of rec- tangular coordinates is used. The valve displacement is plotted vertically and the piston displacements are plotted horizontally. On the left of Fig. 74 is shown a crank and eccentric. The eccen- tric is ahead of the crank by 90 plus a. With the crank at C, the piston is at a distance x from the cen- ter of the stroke. At the same time, the valve is at a distance y from its mid-position. If we plot x against y, we get a point G. 116 ENGINES AND BOILERS The coordinate axes are the horizontal and vertical diameters of the crank circle. On the right of Fig. 74 this same operation is carried out for twelve crank positions with their corresponding eccentric posi- tions. The crank positions are denoted by Ci, 2, C%, etc., and the corresponding eccentric positions by E\, E^ E%, etc. Plotting the displacements, we get the points 1, 2, 3, etc. Connecting the points thus found by a smooth curve, we get what is known as a valve ellipse. It is evident from Fig. 74 that this is not a true fro re/ ' ~ - -^ FIG. 74 ellipse. It would have been but for the distortion due to the short length of the connecting rod. The upper half of the ellipse ABD represents the valve move- ment to the right of its mid-position. The lower half, DFA, represents the valve movement to the left of its mid-position. From A to B it gives the movement from mid-position to ex- treme right; from B to D, from the extreme right to mid-posi- tion; from D to F from mid-position to extreme left; and F to A, from the extreme left to mid-position. The valve ellipse of Fig. 74 is reproduced in Fig. 75. Four horizontal lines are drawn through the ellipse. The head-end steam lap is the distance from the top line to the horizontal axis. The crank-end steam lap is the distance from the bottom line to the axis. The head-end exhaust-lap line is drawn below the axis and the crank-end exhaust-lap line above. When the valve has moved to the right a distance equal to the head-end steam lap, head-end admission takes place. Admission is shown by the VALVES 117 point H on the ellipse, and the crank position corresponding to H is determined by projecting vertically from H to the axis. As the valve moves from its extreme right position back to mid-position, head-end cut-off takes place. This is shown by the point / on the ellipse. The crank position corresponding to I is found by projecting down from I to the axis and striking an arc upward from this point to the crank circle. The radius of the arc is the length of the connecting rod. In like manner, the A fxg/w Ce.mof.ft.0. FIG. 75 crank position at head-end release is found from the intersection of the head-end exhaust-lap line with the ellipse at J. The head-end compression point is at M , where the exhaust- lap line cuts the ellipse. The crank-end events are determined from the points K, L, N, and P. The vertical distance from the top point of the ellipse to the head-end steam-lap line is the head- end maximum port-opening, and the head-end lead is given by the distance from the extreme left point of the ellipse to the head-end steam-lap line. The crank-end maximum port-opening and lead are found in a similar manner. In actual use, the ellipse is rather burdensome because it takes considerable time to construct it. It is evident also that the crank position at admission is not easily determined with accuracy. The ellipse is little used except in locomotive work. 114. The Bilgram Diagram. On the left of Fig. 76, the crank and the eccentric are shown by C and E, respectively. The displacement of the valve from mid-position is y. The dis- tance y is laid off perpendicular to the crank, and a line is drawn parallel to the crank 'at a distance y from it. 118 ENGINES AND BOILERS At the right of Fig. 76, this has been done for twelve crank positions. It is seen that these lines all pass through two points P and P', and that a line drawn from P to P' passes through the center of the crank circle and makes an angle a with the hori- zontal. Hence the perpendicular distance from the points P or P' to the crank at any position is the distance that the valve is from mid-position. The points P and P' are called the con- struction points in the Bilgram diagram. Figure 77 shows the application of the Bilgram diagram. About the point P draw two circles, one whose radius is equal to the head-end steam lap, and the other with a radius equal FIG. 76 to the head-end exhaust lap. Draw the crank-end lap circles with center at P'. The crank positions tangent to these lap circles give the positions at the different events. Head-end ad- mission occurs when the valve is at a distance from its mid- position equal to the head-end steam lap, and when the valve is going away from its mid-position. The head-end admission position shown in Fig. 77 fulfills these conditions. At that time the crank is at a distance from the point P equal to the head-end steam lap, and further motion moves it farther from P. The crank position at head-end cut-off is tangent to the head-end steam-lap circle on the other side, i.e., the valve is then at a dis- tance equal to the steam lap from mid-position and further mo- tion brings the valve nearer mid-position. The crank positions for the other events are shown in Fig. 77. Reasoning similar to the preceding will show them to be correct. It is customary to draw the head-end lap circles about P, and the VALVES 119 crank-end lap circles about P f , although there is no inherent reason for so doing. Half of the valve-travel minus the steam lap equals the maxi- mum port opening if the valve has no over-travel, i.e. if it does not move beyond the far edge of the port. Therefore the maxi- mum port-opening is as shown in the figure; It is remembered that the port is open a distance equal to the lead when the crank is on dead center. In other words the valve is then at a distance equal to the steam lap plus the lead from its mid-position. Therefore the perpendicular distance of P from the horizontal c,e. t a FIG. 77 axis is equal to the steam lap plus the lead. The distance of the steam-lap circle from the axis is then s the lead. 115. The Zeuner Diagram. On the left of Fig. 78, the valve displacement y is laid off radially on the crank from the center outward. This gives a point G on the crank. This has been done for twelve crank positions on the right of Fig. 78 and the points connected by a smooth curve. The points fall on the circum- ferences of two equal circles, the diameter of each of which is one-half the valve-travel. The line which forms the diameters of these two circles makes an angle a with the vertical. The 120 ENGINES AND BOILERS circle whose center is at A shows the movement of the valve to the right of its mid-position, and is called the right valve-circle. The other circle is called the left valve circle; it shows the move- ment of the valve to the left of the mid-position. When the crank is drawn in any position, the displacement of the valve is given by the distance from the center of the crank circle to the intersection of the crank with the valve circle. The application of this diagram is shown in Fig. 79. With the crank on head-end dead center, the eccentric is at E, at an angle a to the right of the vertical. The diameters of valve circles, P'P', are at an angle a on the other side of the vertical. FIG. 78 The extremity P of the diameter of the valve circle is called the construction point in the Zeuner diagram. The lap circles are drawn as shown. The head-end steam-lap circle intersects the right valve-circle at the point T. The crank position through T is then head-end admission, because with the crank in this position the valve is at a distance equal to the steam lap to the right of mid-position. The crank position at cut-off is drawn through the point K, where the steam-lap circle intersects the right valve-circle. Head-end release and compression occur when the head-end exhaust-lap circle intersects the left valve-circle. It may be proved by means of the similar triangles OWE and PRO or by actual construction that a line drawn from A to B is tangent to the steam-lap circle at W . This line is perpendicular to the diameter of the valve circles. In like manner a line drawn from I to F is tangent to the head-end exhaust-lap circle, and is perpendicular to the diameter of the valve circles. The same thing is true of the lines HG and JD for the crank end. It is VALVES 121 often better to draw the lines A B, IF, HG, and JD than to determine the crank positions at the events by the intersections of the valve circle and the lap circle. ON is equal to the steam lap plus the lead, because the valve circle cuts the crank on dead center at N. But ONP is a right triangle, since it is inscribed in a semicircle. If QS is drawn parallel to AW, the triangle OSQ is a right triangle, and it is similar and equal to the triangle, ONP. Therefore OS is equal Right w/re ctrck be. c.o. fivnfr circle Eccentric c/n,/e 4t. comp t.rel. ce. co.- c/rc/e FIG. 79 to the steam lap plus the lead, and WS is equal to the lead. If we then draw a circle about Q as a center, tangent to the line AW, its radius is the lead. This is called the head-end lead circle. The crank-end lead circle is drawn about X tangent to HG. The head-end maximum port opening equals PW, which is one-half the valve-travel minus the steam-lap. The line PK is perpendicular to the crank at cut-off, because PKO is a right angle since it is inscribed in a semicircle. The application of these valve diagrams to practical problems will show their value. Space will not be taken here to give the solutions of the various common problems in which these dia- grams are used. The Bilgram and Zeuner diagrams are both 122 ENGINES AND BOILERS adapted to problems of valve setting, but the Bilgram diagram is the more convenient for use in designing valve gears. 116. Types of Slide-valves. The simple D slide-valve has been discussed and its action explained. This type of valve is much used, but it has certain defects which have been overcome in other types. One of the defects of the simple D valve is the large force necessary to move it when high steam pressure is used. The steam pressure on the back of the valve presses it against the seat. This pressure times the coefficient of friction between the valve and the seat is the force that must be exerted to move the valve. The work done in operating the valve is the force times the distance the valve is moved. If either the force or the distance is decreased, the work necessary to operate the valve will be lessened. 117. Valve with Pressure Plate. The pressure on the back of the valve may be removed by putting a pressure plate above it, somewhat in the manner shown in Fig. 80. A steam-tight fit between the valve and the plate is made by strips set into s | ots j n t h e va i v e. These are pressed up against the plate by springs from beneath, and they act in much the same way as rings on a piston. If any steam leaks by the strips, it may escape to the exhaust through a vent in the valve. This scheme enables us to remove as much of the pressure from the back of the valve as we desire, but some pressure downward is desirable in order to keep the valve firmly seated. Many flat slide-valves have pressure plates. Aside from removing the pres- sure, the plate does not affect the valve in any way. 118. The Piston Valve. Instead of a flat valve such as we have considered, a piston valve is used extensively. Figure 81 shows a form of this type where the valve is cylindrical and slides in a cylindrical chamber. It is readily seen that it is perfectly balanced, since the steam causes no thrust either endwise or on the seat. Piston valves are very easy to operate, but are liable to leak steam when they become worn. Many of them FIG. 80 VALVES 123 have rings similar to piston rings to keep this leakage of steam to a minimum. If rings are used it is necessary to bridge the ports. A broken ring is liable to cause severe damage and care must be exercised to keep them in good condition. In Fig. 81, the steam is led to the inside of the valve, so that the steam lap is on the other side of the port from the valves previously considered. A valve so constructed is said to be an FIG. 81 indirect valve, or to have inside admission. Most piston valves have inside admission, but this is not a necessity. It is easier to keep the stuffing-box tight against exhaust steam than against high-pressure steam. With the inside admission arrangement, also, the live steam has less surface exposed to radiation. With an inside admission valve, the eccentric follows the crank by an angle of 90 a. The valve diagrams studied will have the same form as before, but what was right-hand is now left-hand, i.e. in the Zeuner diagram, for instance, what was formerly the right valve-circle is now the left, but otherwise there is no change in the diagram and no difference is made in the solution of a problem. 119. Double-ported Valves. As has just been mentioned, the work required to move the valve is the product of the force required to move it and the distance it is moved. The distance may be cut in half by making the valve double-ported. Figure 82 shows a so-called double-ported valve. It is seen that for a certain valve movement twice the area for the passage of steam is given in this type compared with a simple slide-valve. There are many different forms of double-ported valves, but they are much the same in principle as that shown in Fig. 82. FIG. 82 124 ENGINES AND BOILERS 120. The Gridiron Valve. If the idea of a double-ported valve is carried a step farther, we may get a very large aggregate opening for the passage of steam with but a small movement of the valve. Figure 83 shows a gridiron valve. In many valves of this type there are a large number of openings, whereas Fig. 83 shows only three. A valve of this character can have no exhaust functions, and separate exhaust valves must be provided. If one exhaust valve takes care of both ends of the cylinder, the engine is called a two-valve engine; if there is a separate exhaust valve FIG. 83 for each end, it is called a three-valve engine. When gridiron valves are used, it is more common to have a steam valve and an exhaust valve for each end of the cylinder. The engine then has four valves. If a governor is so constructed that it regulates the amount of steam admitted to the cylinder by changing cut-off, it is evi- dent from the valve diagrams that the events of release, com- pression, and admission are changed when cut-off is changed if a single valve is used. Under some conditions this is a serious defect, and it makes the use of separate steam and exhaust valves desirable. 121. The Riding Cut-off Valve. To utilize the expansive force of the steam in an engine, it is necessary to have an early cut-off. With a single valve, early cut-off will necessitate either early release or early admission. With an early cut-off and with release near the end of the stroke, compression is bound to occur too soon for satisfactory operation under non-condensing con- ditions. The student only needs to draw a valve diagram to con- vince himself of this fact. To use an early cut-off, and to have at the same time reasonable percents of release and compression, VALVES 125 a riding cut-off valve is often used. There are several forms of riding cut-off valves, but we shall describe only one of them. Figure 84 shows the Myers riding cut-off valve. A main valve slides on the seat in the same manner as an ordinary D valve. The steam lap is made small so that the proper relation exists between the events of admission, release, and compression. If the main valve were acting alone, cut-off would occur very late. To give an early cut-off, a rider valve which controls only the event of cut-off is placed on the back of the main valve. The working edge of the rider valve effects cut-off when it matches with the edges of the main valve at B and at D. In Fig. 84 both valves are shown in their mid-position. This would not occur normallv unless one of the valves were discon- FIG. 84 nected from its eccentric, but the figure is drawn in this manner to give a clearer idea of the laps. Each valve is driven by its own eccentric. The rider valve is made in two parts and the relative position of the two parts may be changed by revolving the valve stem. One part of the rider valve is secured to the valve stem by a right-hand thread and the other part by a left- hand thread. The hand wheel to the left is arranged so that by turning it the valve stem is rotated and the parts of the valve are brought nearer together or moved farther apart, thereby effecting a change in cut-off. With the parts farther apart, cut- off occurs earlier. When the valves are both in mid-position, as shown in Fig. 84, it is easily seen that the rider valve has negative steam lap or steam clearance. To determine the crank position at cut-off from the valve dia- grams, it is necessary to consider the relative motion of the two valves. Figure 85 is the Zeuner analysis for the rider valve. The crank circle and the two eccentric circles are shown by the light lines. The point PI is the extremity of the diameter of the 126 ENGINES AND BOILERS right valve-circle for the main valve, and a\ is the angle of ad- vance for the main eccentric. The point P% is the extremity of the diameter of the right valve-circle, and a 2 is the angle of advance for the rider-valve eccentric. If a number of crank positions be chosen and the displacement of the rider valve relative to the main valve be laid off on the crank from the center radially outward, a number of points will be established. Connecting Relative ffiqht l/e/ve Ctrt/e C.e. f?/cfer fa/re Steam top (neyafiri) ff/JfrVafoe c.e a/t-o/f. FIG. 85 these points by a smooth curve, we find that it is composed of the two circles shown by heavy lines in Fig. 85. The point PS is the extremity of the diameter of the right relative valve-circle and 3 is the angle which that diameter makes with the vertical. It is seen that the diameter of the right relative valve-circle, OPs, is equal in amount and parallel to the dotted line PiP2. Knowing this fact, the solution of a rider-valve problem is quite simple, for it is not necessary to plot the points to determine the relative valve-circle. Since the steam lap for the rider valve is negative, the intersec- tion of that lap circle with the left valve-circle gives the crank position at head-end cut-off. The cut-off positions of the crank are shown by the heavy lines in Fig. 85. The crank positions for admission, release, and compression are determined from a valve diagram for the main valve in exactly the same manner as previously explained for the D valve. 122. Effect of Rocker Arm on Location of Eccentric. In the previous discussion of slide-valves, it has been supposed that VALVES 127 the eccentric rod is attached directly to the valve rod, and that the movement of the valve is the same as the horizontal move- ment of the eccentric. Quite often a rocker arm is interposed between the eccentric and the valve rods, in which case it may be necessary to modify our previous assumption. In Fig. 86, three arrangements of rocker arms are shown. In I, the eccentric rod and the valve rod are both connected to the same pin at B, and our assumption is not changed. In II, the arm B AC reverses the motion of the eccentric. In this case, if a direct valve is used, the eccentric must be placed on the shaft at 180 from the position it would have had without the rocker arm, i.e. if a direct valve is used, the eccentric must follow the crank by an angle of 90 a. If an indirect valve is used with a reversing rocker, the eccen- Eccentr/e /foe/ Eccentric flotf n FIG. 86 m trie is placed 90+ a ahead of the crank. This modification does not effect the work in making the valve analysis. In case III, the two arms of the rocker AB and AC are not of the same length. The travel of the valve is the diameter of the eccentric circle times AB/AC. 123. Oscillating Valves. One of the most common valves is cylindrical and oscillates or rocks on a cylindrical seat. A spindle fastened to the valve extends out of the steam chest and carries an arm that is moved back and forth by the eccentric. The Corliss valve shown in Fig. 58 and the valves of Figs. 60 and 61 are of this type. It would be possible to make an oscillating valve control both the admission and exhaust events, but this is seldom done. Where the oscillating valve is used, four valves usually are employed. Because of the distortion of motion due to the valve arm, it is not possible to show by the valve diagrams 128 ENGINES AND BOILERS the exact motion of the valve. However, the relation of horizontal motion of the eccentric still holds, and so the valve diagrams are of value in the analysis of these valves. 124. Poppet Valves. While they are not common for steam engines in this country, most gasoline engines are equipped with poppet valves. This is a lifting valve and there is no sliding of the valve on the seat. Figure 87 shows this type of valve. These valves do not have to be lubricated and would seem to be well adapted to conditions where highly superheated steam is used, since one of the troubles met with in the use of superheated steam is the difficulty of proper lubrication of the valve. Where poppet valves are used on steam engines, they usually are operated either by a cam or by an eccentric from a lay shaft that is parallel to the axis of the cylinder. The lay shaft usually is driven by mitre gears from the main shaft. 125. Reversing. The direction of rotation of an engine may be changed FIG. 87 by shifting the eccentric to the proper position. Occasionally an ordinary engine must be reversed. Unless the eccentric is keyed to the shaft, this is not usually a difficult task. With a direct valve, the eccentric leads the crank by an angle of 90+ a. This is true irrespective of the direction of rotation. To reverse, then, move the eccentric in the direction the engine has been running through an angle of 180 2 a. The same rule applies with an indirect or inside-admission valve. With certain classes of engines, such as are used for locomotives, in marine work, etc., reversing is a common occurrence. Some handier and quicker means must then be provided than that men- tioned above. The devices used for this purpose are called re- versing gears. There are a very great many types of reversing gears in use, but space will not permit the discussion of more than the most common types. 126. The Stephenson Link. One of the most widely used reversing gears is the Stephenson link gear, which is much used for small locomotives. In this arrangement, there are two ec- VALVES 129 Gentries placed on the shaft at an angle of 180 2 a apart. The forward eccentric controls the forward motion, and the back- ward eccentric controls the reverse motion. In the diagram of Fig. 88, the crank is shown on head-end dead center, and the valve is indirect, so that the eccentrics will be at an angle of 90 a FIG. 88 from the crank. In Fig. 88, FE is the forward eccentric, and BE the backward eccentric. The two eccentric rods connect to the eyes of a link at H and I. This link may be raised or low- ered by the bell-crank BAD and the link DJ. When the link is down, as shown in Fig. 89, the forward eccentric entirely con- trols the motion of the valve. With the link all the way up, the backward eccentric controls the valve. In Fig. 88, the link is FIG. 89 shown in mid-position with both eccentrics controlling an equal amount. With the crank on head-end dead center as in Fig. 88, the valve is as far to the left as it can get, the port is open a dis- tance equal to the lead, and the valve is at a distance equal to the steam lap plus the lead from its mid-position. With the 130 ENGINES AND BOILERS valve opening only lead distance, the engine will not get enough steam to run. At mid-gear, half the travel of the valve is equal to the steam lap plus the lead. With the link in some position between those shown in Figs. 88 and 89, both eccentrics will control the motion of the valve, but the forward one will predominate, and the engine will run for- ward. The cut-off will now be earlier than it would be if the link was all the way down in full gear. It is evident that the Stephenson gear may be used to change the cut-off as well as to reverse. It is possible to give the same motion to the valve with the link in intermediate position, by a simple equivalent eccentric. The method of determining this equivalent eccentric is not difficult but it will not be discussed here. 127. The Walschaert Valve Gear. Most large locomotives in this country are equipped with the Walschaert gear, or some FIG. 90 similar reversing gear. In this type of gear, shown in Fig. 90, the motion of the valve is derived partly from the cross-head, and partly from a return crank or eccentric. That part of the motion coming from the cross-head is constant, while that de- rived from the eccentric is varied for different conditions of load and direction of rotation. Since the eccentric is placed outside the driver, it is commonly called a return crank. The bar CE is fastened to the outer end of the crank pin C, and the eccentric pin E moves in the dotted circle (Fig. 90), about as a center. VALVES 131 The angle between the crank and the eccentric is 90. The hori- zontal motion of the eccentric is transmitted through the eccen- tric rod EM to the lower point of a link. The link is pivoted at its center G to the frame of the engine, so that the point M oscillates about the point G. In this link is fitted a block which can be raised or lowered in the link by the bell crank DAB. As shown in the dotted position, the block is at its lowest posi- tion in the link, and the engine is running forward, taking steam during the largest possible part of stroke. In the full line posi- tion, the block is at the center of the link, and the engine will not get enough steam to drive it. With the block in mid-position in the link, the eccentric will give no motion whatever to the valve. Under this condition the motion of the valve comes entirely from the cross-head. The lever HIJ is called the combination lever, because it com- bines the motion from the eccentric with the motion from the cross-head. The ratio I H/JH is fixed by the condition that I H __ 2(steam lap plus lead) J H length of stroke With the block in the center of the link at G, H has no horizontal motion, but the horizontal motion of / is equal to length of stroke X I H n , , , , , ,. T-PT - = 2 (steam lap plus lead). J ti As the motion of the valve at mid-gear is 2 (steam lap plus lead), it is seen that the valve will open only a distance equal to the lead on each end, and the engine will not get enough steam to run. When the block is dropped to the dotted position, the point H does have a horizontal motion which comes from the eccen- tric, and with the engine running in the direction shown, the horizontal motion of H will add to the port opening. To reverse, the block is raised above the center of the link. By changing the position of the block in the link we may get not only a reversal of direction of rotation but also a change in cut-off. As in the Stephenson gear, it is possible to find an equiv- alent eccentric which would give the motion actually obtained from the mechanism. 132 ENGINES AND BOILERS 128. The Joy Valve Gear. The Joy gear is a so-called radial gear. Unlike those previously described, it has no eccentric. Figure 91 shows diagrammatically the principle of its operation. A point such as D on the connecting rod FC will move in the path of an ellipse, which is shown dotted. A bar BED is pinned to the connecting rod at D. The other end of the bar is con- nected to the frame of the engine by the link AB, A being a point on the engine frame. It is evident that a point E on this bar will have a combination of the elliptic motion of D and the nearly vertical motion of B. The path of E is shown dotted. A bar EGH is connected to BED by the pin E. The point G is in a block which is at liberty to slide along a curved FIG. 91 link. At any particular cut-off this link is held stationary, and the block G slides up and down in it. It is thus seen that the point H gets a combination of the motions of the point E and of the point G. The horizontal component of the motion of H is transmitted through the link I H to the valve. By rotating the link about its center J, a reversal in the direction of rotation of the shaft may be obtained. As in the other gears, cut-off may be varied as well as the engine reversed. Under conditions of light load an early cut-off may be used and sufficient power obtained if the steam is not throttled between the boiler and the engine. It is the practice on locomotive engines to vary the cut-off to suit the load under normal running condi- tions. The power generated by the engine might also be regu- lated by throttling the steam, but it has been found that a higher efficiency is obtained at light loads by using an early cut-off VALVES 133 than by using a late cut-off and a low steam pressure, especially where valve gear is used which increases compression when cut- off is shortened, as in the Stephenson gear. In the reversing gears commonly used, an early cut-off is accompanied by an early compression. The increased efficiency at light loads is due, how- ever, more to the early cut-off than to the high compression, although the high compression aids by heating the clearance space, piston, and cylinder head, thereby keeping initial conden- sation within more reasonable limits. 129. Setting the Slide-valve. On a small engine that can be turned over easily by hand, the setting of a slide-valve is a simple matter. With proper valve setting an engine should run smoothly, should be easy to start, and each end of the cylinder should furnish about half the power. If an indicator is at hand, it should be employed in the setting. If no indicator is available, the valve may be set by linear measurement. In the setting of a slide-valve there are but two things to do, shift the eccentric on the shaft, and lengthen or shorten the valve stem or rod. VALVE SETTING BY INDICATOR. In order to get approximately the same amount of work from each end of the cylinder, cut-off for the two ends should be about the same. A rough adjust- ment may be made easily by placing the eccentric somewhere near its proper position, and adjusting the position of the valve on the rod so that the engine may be started. After the engine is started, take cards and then adjust the length of the valve stem until the cut-off is the same percent on both ends. Next, shift the eccentric until the desired percentage of cut-off is attained. Shifting the eccentric ahead makes cut-off come earlier. By the use of the indicator it is easy to get the exact setting desired. Knowing the valve-travel and the dimensions of the valve, we may compute by means of the Zeuner valve diagram the exact amount the valve stem must be lengthened or shortened and the angle the eccentric must be shifted to give a desired setting. In the upper part of Fig. 92 are shown two cards taken with a valve as now set. It is desired to change the setting so that cut-off will be 50 per cent for each end. The cards are shown in length equal to the valve-travel, but they need not have been, as the percentages of stroke could have been scaled and the corresponding crank positions found. 134 ENGINES AND BOILERS From the cards, locate on the crank circle (assumed for con- venience with its diameter the same as that of the valve-travel circle), the crank positions at the different events. Draw lines between the crank positions at admission and cut-off, and from release to compression, for each end. The distance between the admission-cut-off lines should be the sum of the steam laps as measured on the valve itself. A radial line drawn perpendicular to the line joining admission and cut-off establishes the angle of advance a\. The laps may be measured from the diagram as shown. Now construct a Zeuner diagram (Fig. 93) for the desired c.e.ad/n FIG. 92 FIG. 93 cut-off (50 per cent in the diagram) keeping the sum of the steam laps the same as before. This amount, and also the sum of the steam lap and the exhaust lap for each end, will be the same no matter what adjustment is made. The angle of advance for the new condition is 0:2. Then <* 2 a.\ is the angle that the eccentric must be shifted forward. The difference X between the new and the old head-end steam laps is the distance the valve must be moved on the rod. Since the new head-end steam lap is larger than the old, the valve rod must be lengthened if the valve is direct, and shortened if the valve is indirect. The upper part of Fig. 93 shows the cards that may be expected after the setting has been changed. VALVES 135 VALVE SETTING BY MEASUREMENT. If an indicator is not avail- able, the valve may be set by linear measurement. The steam chest cover must be removed so that the measurements may be taken. It is customary to set either for equal cut-offs or for equal leads. We cannot have equal leads and equal cut-offs at the same set- ting because of the angularity of the connecting rod.* In either case adjust the valve on the stem so that the valve travels about as far beyond the head-end port as beyond the crank-end port. SETTING FOR EQUAL CUT-OFF. By means of marks on the guides and on the cross-head the stroke may be determined, and that proportion from each end at which cut-off is to take place may be laid off on the guides. The procedure is as follows: (1) Place the cross-head at the position for head-end cut-off. (2) Loosen the eccentric on the shaft and turn it on the shaft in the direction the engine is to run until the valve is just on the point of cutting off. Fasten the eccentric to the shaft in this position. (3) Turn the engine to the position for cut-off at the crank-end and measure the distance from the valve to its correct position to give cut-off on this end. Divide the error by two and take up half the error by shifting the valve on the stem and the other half by turning the eccentric on the shaft. (4) Turn the engine back to the position for head-end cut-off and check the setting. If there is an error left, repeat as ex- plained above. Be sure the eccentric is fastened firmly to the shaft and the valve to the stem. Replace the cover of the steam chest. SETTING FOR EQUAL LEAD. (1) Place the engine accurately on head-end dead center. (2) Loosen the eccentric on the shaft and turn it in the direc- tion the engine is to run until the port is open a distance equal to the desired lead. Be sure the valve will open the port if the eccen- tric is turned ahead more. Fasten the eccentric to the shaft. (3) Turn the engine to crank-end dead center and measure the error. Divide the error by two and take up half by turning the eccentric on the shaft and the other half by adjusting the length of the valve stem. (4) Turn the engine back to head-end dead center and check. If there is an error, correct it by repeating as previously explained. * This expression means the deviation from parallelism with the axis of the cylinder, of a connecting rod of finite practical length, except when the crank is at one of the dead centers. 136 ENGINES AND BOILERS The reason that half the error is taken up by moving the valve on the stem and half by turning the eccentric on the shaft may be explained as follows. Suppose the valve has been set to give the correct cut-off on the crank-end, but that when turned over to head-end position for cut-off there is an error as shown in Fig. 94. The eccentric is now at E\. By turning the eccentric from Ei back to E 2) this error would be adjusted, but nothing would have been gained as will be seen when the engine is turned back to the position for crank-end cut-off. The same error now exists on crank-end as shown in Fig. 95. That is, the eccentric is at E* and should be at E%. If we try to take up the error by lengthening the valve stem (Fig. 96), nothing is gained because the valve will be moved to the left a distance equal to the error, and when it is turned back to the position for crank-end cut- off, the valve will be open a distance equal to the error. If now we divide the error by two, and move the edge of the valve to A, FIG. 94 FIG. 95 FIG. 96 in Fig. 94, by turning the eccentric back from EI to D, we will find that the other edge of the valve will be at B in Fig. 95, when turned over to the crank-end cut-off position. If the other half of the error be taken up in Fig. 94 by lengthening the valve stem i.e. by moving the valve to the left on its stem, the valve will be in the correct position for head-end cut-off. Moreover, moving the valve to the left moves the crank-end edge (Fig. 95), from B to the edge of the port, where it should be for crank-end cut-off. Therefore, by taking up half the error in each way at the posi- tion of head-end cut-off, we have made no net change in the position of the valve for the crank-end cut-off position. CHAPTER IX GOVERNORS 130. General. The function of the governor is to keep the engine running at nearly the same speed at all loads. It does this by controlling the amount of steam admitted to the cylinder. It is impracticable for the governor to keep the speed exactly constant at all loads. This may be seen when we understand how a governor must work. Under conditions of changing load, the governor must change the amount of steam admitted to do the work. At any instant we have a certain load on the engine, the governor is admitting enough steam to carry that load. If an extra load is thrown on, the engine will momentarily slow down. This slower speed affects the position of the parts of the governor, and this in turn allows more steam to enter to do the extra work required. This change cannot be affected instantane- ously, although some governors respond in a very short time. Governors that are quick in making the change are said to be sensitive, and those that are slow, sluggish. Under full-load conditions, the engine usually runs slower than that at no load, by an amount that depends upon the construction of the governor. Governors are made that give an engine speed which is nearly as great at full load as at no load. These are said to give doss regulation. It is possible to design and construct a governor that gives the same average engine speed at all loads, in which case the governor is said to be isochronous. Such a governor would tend to hunt, i.e. there would be a constant fluctuation in speed as the governor attempted to regulate the steam supply to balance the load conditions. Practical considerations limit the nearness to which isochronism may be approached. If the no-load speed is greater than the full-load speed the governor is said to be stable. If the governor is isochronous or gives a full-load speed greater than no-load speed, it is unstable. An unstable governor is clearly undesir- able. Various schemes are used to express the relation of speeds at various loads. A common way is to express the variation of speed from no load to full load and from no load or full load to normal load in percent of the speed at normal load. We may 137 138 ENGINES AND BOILERS then say, the percent of variation in speed from no load to full load is equal to 100(ni n^/n, where HI and n% are the speed at no load and the speed at full load, respectively, and n is the speed at normal load. 131. Classification of Governors. Governors may be classi- fied according to the following characteristics. (a) As to the manner of regulating the steam supply: Under this head we have (1) throttling governors, which regulate the amount of steam admitted to the cylinder by controlling a throttle valve, and (2) cut-off governors, which control the steam supplied by changing the point of cut-off. (b) As to the predominant controlling force in the mechanism: We speak of centrifugal governors, inertia governors (although inertia is not a force), and resistance governors. All mass has inertia. If the mass of the moving parts is small and the inertia effect is not used in governing, we call the governor a centrifugal governor. If the inertia is large and its effect is used to aid in governing, we have what we call inertia governors. Even in inertia governors, the centrifugal force is a very important factor. (c) As to the force used to balance the centrifugal force of the rotating parts*: We have gravity-balanced governors and spring- balanced governors. (d) As to the arrangement of the mechanism: There are spindle governors and shaft governors. 132. The Gravity-balanced Spindle Governor. The diagram of Fig. 97 represents a simple gravity-balanced spindle governor. This is sometimes called a conical pendulum, and is also often called the Watt governor, because James Watt first used it on his engines. Two flyballs, at the ends of arms, rotate about a vertical spindle. The arms are pivoted to the spindle at 0. The height of the balls is caused to control the steam supply. In Fig. 97 this is done by raising or lowering the point A with the balls, which, by a suitable mechanism, causes either the movement of a throttle valve or a change in the point of cut-off. A definite relation exists between the height hi of the cone of revolution, and the speed of the spindle. To determine this rela- tion consider one of the balls as a free body. At any certain speed it may be considered to be in equilibrium under the action of the following forces: the tension in the arm T, the weight W, GOVERNORS 139 and the centrifugal force acting radially outward (W/g)X(v z /R). Taking moments of these forces about the point 0, we have X^Xh-WXR = 0. g ti But v = 2irRn, where n is the number of revolutions per unit time. Therefore we may write 1 =WR gR or If we wish to express hi in inches and the speed of the governor in r. p. m., our equation becomes hi 32.2X12X3600 35200 - . 9 9 zj 4?r 2 n 2 n 2 approximately. From this equation it is seen that the height hi of the cone of rev- olution does not depend upon the length of the arm. Figure 97 FIG. 97 shows the r. p. m. corresponding to the height /ii for two-inch increments of /i, up to 20 inches. From these values it will be noticed that for a certain vertical movement of the ball there will be less speed variation with the ball in the lower positions. In other words, to get a reasonable speed variation it will be neces- sary to run the governor very slowly. At low speeds the gov- ernor will not have much power unless the balls are made exces- sively heavy. This practical limitation precludes the use of this governor on modern engines. 140 ENGINES AND BOILERS FIG. 98 If the governor arms are crossed, as shown in Fig. 98, it will be noticed that much less variation in speed exists for the same vertical movement of the balls than for the form shown in Fig. 97. Moreover, this governor is nearly isochronous at 50 r. p. m. By the proper selec- tion of the pivots B and C, we may get quite satisfactory speed regulation for a certain limited range of vertical movement at any desired speed. The pendulum gov- ernor may be made exactly isochronous by making the balls swing up in the arc of a parabola, as in Fig. 99. The subnormal of a parabola is constant, and it is seen in Fig. 99 that hi is the sub-normal of the parabola which is the path of the balls as they swing upward. The balls may be made to take the parabolic path by having the arms made flexible and to unwind from the evolute of the parabola or by having them guided by an arrangement of cams. Of course it is understood that, in practice, the governor never would be made exactly isochronous, but it is seen that iso- chronism may be ap- proached as nearly as practical conditions will permit. In order to run a spindle governor at fairly high speed, and FlG> 99 still have a reasonably small speed variation in the engine, it is customary to load it as shown in Fig. 100. The load L tends to pull the balls down; hence they must rotate faster, to get to the same height as before. To find the relation between the height of the cone of GOVERNORS 141 revolution and the speed, consider the load and the ball each as a free body. With the forces acting on them as shown, we can express the conditions of equilibrium as follows. Expressing the fact that the sum of the vertical forces is zero for the load L, (1) or 2T 2 sin j8 = L, T -- 2sin/3 Considering the ball as a free body, and taking moments about as a center, we have, since the sum of these moments must be zero, W v 2 (2) -p Q rt -WXR- !F 2 sin By (1) this may be written in the form (3) ~^ Xhi ~ WR ~ f X# - | ctn or, since v = 2irnR, W T T 0. (4) j WRtfh, -WR-^R-^ctu whence, since ctn /3 = R/h 2 , TT n l - and finally, solving for n 2 , r^+K^+A.)] , 1 + (6) * = i- If /Z-i = ^2 o, = 0, = 0, L1 )J 7 or fe /Tf +ZA V IT / If hi be expressed in inches and n in r. p. m., W + L\ 35200 142 ENGINES AND BOILERS The usual form of loaded governor is shown in Fig. 101. This is seen to vary slightly from Fig. 100. Taking the load as a free body, T 2 = L/(2 sin /3), as in (1) above. FIG. 101 Taking the upper arm as a free body, and expressing the fact that the sum of the moments about equal to 0, (9) Xj^XH -WXR- (10) -X^XH- g R in PX^R- T 2 cos ftXhi = 0. Substituting the value of T 2 from (1), we have 6 2 ^b 2 But we have (ID hence (12) ~X^H - Since we have also v = 2irRn, hi = H T and o _ w __ _ 2 b 2/i whence ' . GOVERNORS 143 Taking as our units inches and r. p. m., this becomes (15) n 2 = H In the solution of a problem for this type of governor it is best to make a drawing and to scale from it the values of H, h 1} and hz at the different positions of the weights. With these values substituted in the formula (15), the r. p. m. of the governor is readily determined. The governor of Fig. 101 is the one commonly used on Corliss engines. 133. The Spring-balanced Governor. In most high-speed engines, the centrifugal force of the revolving weight is balanced FIG. 102 by the force of a spring. Figure 102 shows a weight W, revolv- ing about the center of a spindle or shaft. The centrifugal force C of the weight is balanced by the spring tension S. The weight is at a distance R from the center of rotation. Hence we have For a certain value of n, C varies directly as R. In Fig. 103, this variation is shown graphically. At speed n and with a radius R, the centrifugal force is C. If R is doubled, C is doubled. For other values of n, as HI and WQ, C will have different values with the same radius R. 144 ENGINES AND BOILERS In any kind of uniform spring the elongation, shortening, or deflection is proportional to the force producing the deformation. Figure 104 shows graphically the relation between the pull of the spring and the elongation. If Figs. 103 and 104 be superimposed, as in Fig. 105, we easily can see the relations that must exist if of p0/h of FIG. 103 o FIG. 104 the spring pull equals the centrifugal force. For the three speeds HI, n and n 2 , the spring pull equals the centrifugal force, as seen at the points b, d, and /. That is, the spring pull will balance the centrifugal force at the speed %when the elongation of the spring is ei. The weight will then be revolving at a radius Ri=a+e, where a is the distance the weight would be from the center of rotation if there were zero tension in the spring. At a speed n, the forces will balance when the elongation is e and the radius is R=a-\-e. In like manner, the forces balance at the speed n 2 with an elongation 62 and radius R%. If the origin A should be moved over to the origin 0, i.e. the distance a be made zero, it is seen that the spring pull could balance the centrifugal force at only one speed. This means that the gover- nor would then be isochronous. If the origin A were moved to the left of the origin 0, the governor would be unstable, because the speed HI at no load would be less than the speed 712 at full load. As stated before, an unstable or isochronous governor could not be used in practice. - Q . FIG. 105 GOVERNORS 145 What is known as scale of spring is the force necessary to pro- duce an elongation of one inch in the spring. If the spring pull is Si at elongation e\, and $2 at elongation e 2 , the scale of the spring is equal to Suppose that we desire to find the scale of spring necessary for a governor such as that shown in Fig. 102. Let us suppose that the no-load speed HI and the full-load speed n 2 , and the corre- sponding radii RI and R 2 , are known. First compute Ci and C 2 for the two speeds. Then, since the spring pull must equal the centrifugal force at all loads, the scale of spring is seen to be because R\ R^=ei e%. In actual governors it is very seldom that the spring pull acts in the same line as the centrifugal force. Figure 106 represents FIG. 106 a more usual case. In the solution of this problem, moments will be taken about the pivot point of the governor arm B. In the full-line position, the moment of the centrifugal force equals the moment of the spring pull about the point B. That is, CXh = SXd. In like manner, CiXi=SiXf. The scale of spring equals (S Si) divided by the elongation of the spring as the weight goes from R to RI. 146 ENGINES AND BOILERS 134. Governing by Changing Position of Eccentric. Most shaft governors regulate the steam supply by changing the per- cent of cut-off. This is accomplished by changing the position of the eccentric relative to the crank. In Fig. 107, a Zeuner valve diagram is shown on which appear two positions of the crank at cut-off. In the full-line construction, cut-off comes late and the angle of advance is i, i.e. when the crank is on head- end dead center, the eccentric will be at E\ (direct valve). By shifting the eccentric forward to E^ the angle of advance is FIG. 107 FIG. 108 changed to 0$. The dotted construction gives the position of the crank with the new angle of advance. It is seen that the cut-off comes earlier with the larger angle of advance. By turning the eccentric on the shaft, the time of cut-off can be changed but the value of the steam lap will be the same for all values of a, because the only way to change the laps is to move the valve on the stem. Figure 108 shows the effect on cut-off of changing the valve travel while keeping the angle of advance constant. The Zeuner construction for a large valve-travel is shown in full line. Cut- off is seen to come fairly late. With a smaller valve-travel, as shown by the dotted construction, cut-off is seen to come earlier. Hence it appears that late cut-off is obtained with large valve- travel and earlier cut-off with small valve-travel. A governor can regulate cut-off by changing either a or the valve-travel. It may be so constructed that it will control by making only the one change or it may change the valve-travel and the angle of advance at the same time. GOVERNORS 147 Changing the angle of advance affects the other events as well as cut-off. When a is increased all events occur sooner. Thus on one-valve engines that control the steam supply by shifting the eccentric, it is found that a high compression accompanies early cut-off, such as is characteristic of the Stephenson valve-gear. 135. Governing by Changing a. In Fig. 109, a governor is shown that controls cut-off by turning the eccentric around the shaft. The two weight arms are pivoted at the points G and F. FIG. 109 Any rotation of these arms about their pivots causes the eccentric to turn on the shaft. The weight arms are connected by the links AC and BD to AB, which carries the eccentric. At light loads the speed of the engine is greater than at heavy loads, and the weights will be farther from the center of rotation. This movement of the weights away from the center of rotation causes the eccentric to turn in a clockwise direction relative to the crank. It is evident that this increases a arid therefore makes cut-off come earlier, as it should at light loads. 136. Governing by Changing both a and the Valve-travel. - In Fig. 110 a governor is shown that changes both a and the valve-travel at the same time. The pivot point on the flywheel carries the governor arm. The eccentric is shown as a pin. When the governor arm moves about the pivot point in a clock- wise direction, it carries the center of the eccentric with it and 148 ENGINES AND BOILERS makes a smaller and the valve-travel larger. It is known that small a and large valve-travel both give late cut-off. Conversely, a counter-clockwise movement of the arm about the pivot point gives early cut-off because it makes a larger and the valve-travel smaller. The centrifugal force acts through the center of rota- tion and the center of gravity. Hence this force tends to give the arm counter-clockwise movement about the pivot. At light loads, and therefore higher speeds, the centrifugal force will be greater than for heavy loads. This tends to give the arm counter- clockwise rotation about the pivot and this makes cut-off come FIG. 110 earlier for light loads, as it should. A great many other governors besides the one shown in Fig. 110, change a and the valve-travel in the same way. 137. Centrifugal and Inertia Governors. As has been previ- ously stated, all governor weights have inertia. If the tendency of the weight to keep moving at the same speed helps to effect the change of position that causes the governing, the governor will respond more quickly than it would if the inertia opposed the change. In the gravity-balanced spindle governors that were considered in 132, inertia acts against the rapid change of posi- tion of the balls and so tends to make the governor sluggish. In the governor of Fig. 110, the inertia of the arm assists in the governing. If the load is suddenly thrown off the engine, GOVERNORS 149 it will momentarily speed up. This means that the flywheel will go ahead of the governor arm or rotate in a clockwise direction relative to the arm. This swings the eccentric nearer the center, which makes a large and the valve-travel small. Hence cut-off occurs sooner, which tends to re-establish the conditions of equi- librium. In Fig. 110 the arm is made very heavy, so that it will have considerable inertia. It must not be assumed that the inertia of the arm is the only factor in the governing. The centrifugal force also plays its part, as has been explained previously. The governor of Fig. 110, while it is very simple in construction, is at the same time sensitive and quick in its action. It is widely used and is known as the Rites inertia governor. CHAPTER X STEAM TURBINES 138. Introduction. The steam turbine of to-day is of as much importance in the world of engineering as is the reciprocating steam engine. Practically all large steam power plants which produce electric current employ the turbine engine. The devel- opment of the turbine has been remarkably rapid. However, it would not be true to say that the turbine has crowded the recip- rocating engine from the power-plant field. The fact is rather that it has developed along new and different lines of use, and now occupies a field that was never held by the reciprocating engine, i.e. as the direct connected prime mover for high-speed electric generating units. The turbine and the alternating-current generator have devel- oped together. Both the turbine and the alternating-current gen- erator are well adapted to high-speed rotation. The cost of a slow- speed electric generator is much higher than that of a high-speed generator of the same capacity. Before the days of the turbine, generators were nearly all of the direct-current type, which could not run at very high speeds. The electric system of power transmission is more economical than the old belt system. Hence the turbine has replaced the reciprocating engine in some manufacturing plants on account of the development of systems of electric power transmission. On land, large turbines are seldom used to drive anything but electric generators. 139. History. The steam turbine is not a modern invention. Hundreds of years ago people knew, as every child knows today, that a pin-wheel would rotate when blown upon. There are rec- ords of turbines built in the quite distant past. They were little but toys, however, like pin-wheels, and of no practical importance. The modern turbine dates from the years between 1880 and 1890. During this period two types of turbines that have become of great practical importance were developed. DE LAVAL, the inventor of the cream separator, sought to drive his separator by means of a turbine. After several experiments, he perfected a type that was satisfactory for that purpose. The 150 STEAM TURBINES 151 same turbine, with improvements, has been used in large numbers for driving centrifugal pumps, fans, and even small generators. During practically the same time, C. A. PARSONS developed the type of turbine that now bears his name. These two pioneers were soon followed by other experimenters. Various forms of tur- bines were developed. Some of these are still used. Others are obsolete and are of interest only from an historical standpoint. In the development of the turbine, there were two obstacles that had to be overcome. The first of these was the lack of knowledge of principles. The second was the need of better mechanical means of manufacture. As will be shown later, the velocity of the rotor in a turbine must be very high. This causes large stresses, and makes necessary a very perfect balance. The clearances between the rotors and the stationary parts must be small to prevent undue leakage. This calls for an excellence of design and construction that did not commonly exist for heavy machines in the past. As in the development of any new machine, satisfactory solutions of the problems grew out of the necessities, so that the modern turbine is as reliable and dependable as any piece of machinery in the power plant. 140. Fundamental Principles. Before making a study of the common types of turbines now in use, we shall discuss the fun- damental principles of the steam turbine. It is not our purpose to give an exhaustive discussion, but only to present the principles in their simplest form. The sketches of blades and nozzles are not exactly correct in shape for the conditions assumed. They are to be considered as only diagrammatic. Steam under pressure contains a certain amount of usable heat. The available amount depends upon the initial pressure, the degree of superheat, and the pressure to which the steam may be dropped. There is the same amount of available heat if the initial and final conditions of the steam are the same, no matter whether we are considering the reciprocating steam engine or the steam turbine. The turbine, or the reciprocating engine, is effi- cient, or is not, according as it uses a large or a small amount of this available energy. Since the turbine and the reciprocating engine both use the same medium, it is not to be expected that one will be much more efficient than the other. Both reciprocating engines and 152 ENGINES AND BOILERS turbines may be made of about the same thermal efficiency. The choice of type of engine depends upon other considerations than efficiency. The greatest loss in the reciprocating engine is due to the initial condensation in the cylinder. Since the cylinder walls are made of a heat-conducting material, they will never be as hot as the incoming high-pressure steam, and they will be hotter than the low-pressure steam leaving the cylinder. The relatively hot steam coming to the cylinder strikes the cooler cylinder walls and some condensation takes place, with a consequent shrinkage in the volume. The condensed steam is mostly re-evaporated before the steam leaves the cylinder, owing to an absorption of heat from the then hotter cylinder walls. The loss in the turbine is due to other causes, such as leakage, friction, etc. The leakage occurs around the ends of the blades or from stage to stage. The friction exists between the steam and the parts of the turbine in the passage of steam through both the stationary and the moving parts. There is also a windage loss between the moving parts and the steam. This friction does not cause a complete loss, because a part of the heat generated may be used in later stages of the turbine. 141. Available Energy in Steam. In order to make clear the nature of the available energy in steam, a concrete example will be taken. (1) Let us assume that the steam is dry saturated steam at a pressure of 150 pounds gage (165 pounds absolute), and that it is allowed to expand adiabatically to a pressure of 15 pounds absolute.* The heat contents of a pound of dry saturated steam at 165 pounds absolute is 1194 B.t.u. The heat contents of a pound of steam at 15 pounds absolute, after expanding adiabati- cally from 165 pounds, is 1019 B.t.u. The difference between these values, which is the amount of heat available for doing work, is 11941019 = 175 B.t.u. At 165 pounds pressure, dry saturated steam occupies a volume of 2.75 cubic feet per pound. At 15 pounds pressure, after the expansion just mentioned, the quality is 87 per cent, and the volume is about 23 cubic feet. In the reciprocating steam engine, this change in volume, work- ing by its pressure, does work on the piston in forcing it forward. * Adiabatic expansion is that in which no heat is added to the steam and none is extracted except by the conversion of heat into work. STEAM TURBINES 153 The velocity of the piston is immaterial. In the turbine, the same change in volume takes place, but the steam is allowed to acquire velocity in expanding. The energy of the steam due to its velocity is imparted to the rotor of the turbine. The efficiency of a perfect engine working on the Rankine cycle,* between the pressures of 165 and 15 pounds absolute, is 175/(1194 181) =17 per cent. Since neither the reciprocating engine nor the turbine is perfect, neither would have an efficiency as great as 17 per cent when worked between the pressure limits named. (2) Assume that the steam is allowed to expand adiabatically from 165 pounds absolute to 1 pound absolute. The heat-drop is 1194 871 = 323 B.t.u. and the efficiency on the Rankine cycle is 323/(1194-70) = 28.6 per cent. 142. Velocity Due to Expansion. Let us next compute the velocity of the steam if all the heat-drop goes to giving the steam velocity. (1) Suppose that the drop in pressure is from 165 to 15 pounds absolute. Since one B.t.u. = 778 foot-pounds, the energy is 175 X 778 = 136,100 foot-pounds per pound of steam. The energy of motion, or kinetic energy, is mv 2 /2 = (1/32) Xv 2 /2. This must be equal to the value 136,100 foot-pounds just calculated. Hence v 2 = 64X136,100 = 8,710,400, v = 2950 ft./sec. (2) If the drop in pressure is from 165 to 1 pound, we find, in a similar manner, K.E. = 778X323 = (1/32) Xv*/2, whence v = 4010 ft./sec. In the steam turbine, the steam must be expanded, and the velocity due to this expansion must be used by imparting its kinetic energy to the rotor of the turbine. If this is done by allowing the steam to expand in the stationary parts of the tur- bine and imparting the velocity thus produced to the moving parts, the turbine is said to be of the impulse type. If it is done * To compare steam engines, the efficiencies based on the Rankine cycle are often used. The efficiency of a steam engine operating on the Rankine cycle is given by the expression (Qi Qi/Q\, where Q\ is the amount of heat required to make dry steam at boiler pressure from water at the temperature of the exhaust, and Qz is the amount of heat rejected from the engine minus the heat of the liquid at the temperature of the exhaust. 154 ENGINES AND BOILERS by allowing the steam to expand in the moving parts, the unbal- anced steam pressure reacting on the rotor, the turbine is called a reaction turbine. In some turbines the steam expands both in the stationary parts and in moving parts, and the turbine is said to be of the impure reaction type. 143. Impulse and Reaction. In order to understand the prin- ciples involved, consider the simplest cases involving the principles of impulse and reaction in which the velocity is created and used. It may be easier to think of the jet as a jet of water, for in that case the fluid does not expand when the pressure on it is reduced. Otherwise, the steam jet and the """"*'* . e '^ * : i water jet follow the same laws of impulse and reaction. FlG jjj Suppose that water issues from a nozzle, as in Fig. 111. The nozzle is stationary, and the issuing jet has a velocity of v feet per second. The unit mass m of water that will be considered is that issuing from the nozzle in one second. A particle of water in the jet will move v feet in one second. The kinetic energy of this unit mass will be mv 2 /2. (1) Consider the case in which the jet strikes a stationary flat surface (Fig. 112). After striking the flat surface, the water flows or splatters out to the sides at right angles to its former direction of motion, that is it loses all its velocity in the direction of the jet. The force exerted by the | ^"," "' * ^ * jet on the flat surface may be measured by the force F neces- sary to hold the flat surface stationary. Since force = mass X change in velocity per second, and since the time in which mass m emerges from the nozzle and strikes the plate is one second, we have F = mv. The force exists in the case of the stationary plate, but no work is done because the plate does not move. (2) Suppose that the flat surface moves with a velocity V (Fig. 112). Then the force F=mX(v-V). The quantity (v-V) is the velocity of the jet relative to the flat surface. It is seen that F is less than before, and will be zero if the velocity of the sur- face is the same as the velocity of the jet. The work done in one Force F * t/t/. & STEAM TURBINES 155 second by the jet on the plate equals F times the distance the plate moves, or W=FxV=m(v-V)V. The velocity of the plate at which the work is a maximum may be found by equating to zero the first derivative of the work with respect to V. This gives Hence the maximum work occurs when V =v/2, that is, when the velocity of the plate is half that of the jet. (3) If instead of striking a flat surface, the jet strikes a station- ary curved surface, such that the jet is turned completely back on itself, or through an angle of 180 (Fig. 113), the force F exerted on the surface is mX2v, which is twice that "*'' ^~* f ,, a , f PIG. 113 exerted on the nat surtace. Since the curved surface is stationary, there is no work done. (4) Suppose the curved surface (Fig. 113) is moving with a velocity V. The velocity of the jet relative to the curved surface is (vV). It follows that the absolute velocity of the jet leaving the surface is (v V) V = (v 2 V ) . Consequently the change in the velocity of the jet is v-\-(v 2V) = 2(v V), because the direction of motion is completely reversed. As before, we have F=mX2(v-V), and the work done per second is W=FV = 2m(v-V)V = 2m(vV-V*), whence For maximum work we must have ^=0 dV Hence v=2V, or V=v/2. That is to say, the curved surface should move at half the velocity of the jet for the production of maximum work. If the latter condi- 156 ENGINES AND BOILERS tion exists, the absolute velocity of the jet as it leaves the sur- face is zero. That is, all of the velocity of the jet has been used. (5) In Fig. 114 we have a tank that is free to move horizontally upon a track. In one side of this tank is placed an orifice or nozzle. The water issues from this nozzle due to the pressure of the water from above. If the tank is stationary, the water leaves the tank with an absolute velocity v. The force F, due to the unbalanced pressure of the water in the tank, tends to force the tank to the left, but since the tank | is held stationary, the force F ^: fe/.v-^ does no work. If the tank is ' allowed to move to the left with .....,, ,,,JJj$,,,,,,^), f ,,,,,,,,,,, r a ve l c ity V, however, the work p iG 114 done will be FV. The absolute velocity of water leaving the tank is (vV). The maximum work will be done when V =v, that is when the escaping water has no absolute velocity. In the first four cases considered, the jet impinged on a sur- face. Work was done by the jet striking and moving the surface. A turbine in which the pressure-drop occurs in a stationary nozzle or part is said to be an impulse turbine (142) because the energy is given to the moving parts by the impulse of the jet. In the fifth case the drop in pressure occurred in the moving nozzle. When this occurs in a turbine, it is said to be a reaction turbine (142). A comparison of the above simple examples shows that the velocity of the moving parts of a reaction turbine must be nearly twice as great as that of the impulse type, other factors being equal. 144. Bucket Shapes. In the common types of steam tur- bines, buckets or blades are mounted on the periphery of a wheel or rotor. The shape of these blades is something like that of the curved surface considered in (3), 143. Of necessity the jet cannot be completely turned through an angle of 180 as in (3), 143, because the steam must have a velocity in the direc- tion of the axis of rotation of the rotor in order to get to the bucket and to leave it. In Fig. 115, let Vi denote the velocity of the jet relative to the bucket or blade at the point where the jet first strikes it, and let a STEAM TURBINES 157 denote the angle it makes with the tangent to the rim of the rotor. Let v% and 0, respectively, denote the velocity and the angle upon leaving the bucket. We see that the component of the velocity of the jet relative to the bucket in the direction of the tangent is vi cos a and the component in the direction of the axis of the rotor is v\ sin a. In like man- ner, the relative velocity v^ has similar components v% cos /3 and t>2 sin /3. These compo- nents Vi sin a and v% sin /3 must be large enough to get the jet through the row of buckets on the rotor, in order that the following buckets shall not interfere with the flow. If the relative velocity of FIG. 115 the jet is Vi and the angle that it makes with the tangent is a, the absolute velocity v of the jet makes a different angle 6 with the tangent (Fig. 116). If V denotes the tangential velocity of the bucket, v is the resultant of the two velocities Vi and V, and 6 is the angle that this resultant v makes with the tangent. In like manner, the resultant of v% and V at the exit is the abso- lute velocity v' at the exit, and it makes an angle with the tangent. If we assume that the jet strikes the bucket in the direc- tion of the tangent to the rim of the rotor, the preceding par- agraph shows that an error will be introduced. In order to make the calculations as simple as possible, however, we shall assume that the jet does strike tangentially, and we shall bear in mind that some error has been introduced. The results previously derived for steam veloci- ties for certain heat drops will now be applied to the problem of the turbine. FIG. 116 158 ENGINES AND BOILERS 145. The Single-stage Turbine. In a single-stage impulse turbine, the steam is expanded in a stationary nozzle, and is directed against the moving buckets or blades, which are mounted on the rim of the rotor. In the single-stage reaction turbine, the rotor itself carries the nozzles, and the steam expands in passing through them. If the expansion of the steam is from 165 to 15 pounds absolute we have seen in (1), 142, that the steam or jet velocity is 2950 feet per second. For maximum work done, the bucket velocity of the impulse turbine is half that of the jet velocity ( 143). Hence the peripheral velocity of the rotor should be 2950/2 = 1475 feet per second. With a reaction turbine, the peripheral velocity of the rotor should be the same as the jet velocity, or 2950 feet per second. If the rotor speed is assumed to be 3000 revolutions per minute, or 50 revolutions per second, the diameter of the rotor should be 1475 -^r- =9.4 feet 50-7T for an impulse wheel. This is obviously very much too large. For a reaction wheel, the diameter should be 18.7 feet, which is absurd. If a speed of 24,000 r. p. m. is assumed, the diameter of the rotor for an impulse turbine should be 1475 or 14 inches. These values for the speed and the diameter of the rotor are not far from those which are used in the DeLaval single-stage turbine. The preceding examples show what a very high peripheral veloc- ity is necessary for a fair efficiency in a single-stage turbine. With a vacuum, a much larger velocity should be used. Since immense stresses are induced in the wheel by these high velocities, it is readily seen that a single-stage reaction turbine is almost out of the question. If such turbines were operated, their efficiency would necessarily be very low. Hence they are not used. In Fig. 117, a diagram of the single-stage impulse turbine is shown. Steam enters the nozzle from the left, expanding as it passes through. As the pressure drops, a high velocity is imparted to the steam. The steam leaves the nozzle at low pres- sure and at a high velocity. The steam now impinges upon the STEAM TURBINES 159 buckets or blades of the rotor, imparting to the rotor its velocity, and therefore its kinetic energy. Upon leaving the rotor, the absolute velocity of the steam is quite low. The graphs at the lower part of the diagram show the changes in the pressure and in the velocity. The steam pressure is shown by the full line, and the steam velocity by the dotted line. While single-stage impulse turbines are widely used, they are never made in large sizes. The diagram of Fig. 118 represents a single-stage reaction tur- bine. The steam passes from the left directly to the rotor. The rotor carries blades so shaped that the spaces between them act as nozzles. The steam expands in these spaces or nozzles. As it expands, its pressure drops, and it reacts upon the blades. This Single-Stage Impulse FIG. 117 Single - Sfagre Reac tioh FIG. 118 force of the steam on the blades causes the rotor to move and to absorb the energy liberated by the expansion. It will be noticed from the graphs that the velocity does not change much in pass- ing through the rotor, but the pressure drops during the passage. In order to decrease the peripheral velocity of the rotor, and at the same time to expand and use all the velocity of the steam, more than one set of rotor blades or buckets are employed. This is called staging. The steam passes successively through the sets of blades in each stage, giving up part of the energy to each set. 160 ENGINES AND BOILERS Staging. In multi-stage impulse turbines, two methods are in use. The first method is to expand the steam in one set of stationary nozzles, and to take out part of the velocity in each stage. This is known as velocity-staging. The second method is to expand the steam partially in one set of stationary nozzles, using up the velocity caused by this expan- sion in one stage, then to expand the steam again in another set of stationary nozzles, using the velocity thus generated in the second stage, and so on. This scheme is called pressure-staging. A combination of pressure-staging and velocity-staging is also used, in which there are two or more velocity stages in each pres- sure stage. 147. Multi-stage Impulse Type with Velocity-staging. The diagram in Fig. 119 shows the velocity-stage impulse turbine. - Velocit y - Stage /mpulse . FIG. 119 The steam enters from the left and passes through the stationary expanding nozzle, where the pressure drops and the velocity is acquired in exactly the same manner as in the single-stage im- pulse turbine of Fig. 117. The rotor in this case, however, has much less velocity than the rotor shown in Fig. 117. Hence the steam loses only a part of its velocity in passing through the first STEAM TURBINES 161 set of buckets. Emerging from the first set of buckets, it passes through a set of stationary blades or vanes which change the direction of flow of the steam, but not its velocity. These sta- tionary blades are necessary, because the steam has a large up- ward component of velocity after leaving the rotating buckets of the first stage; and since the velocity of the rotor is downward, the direction of flow must be reversed so that the steam may impinge on the second set of rotating buckets. In passing through the second set of moving buckets, more of the steam velocity is taken up by the rotor. The direction of flow is again changed by the stator, and so on, till the steam finally emerges from the turbine with its velocity practically all expended. Suppose, for example, that the downward velocity of the steam as it leaves the nozzle is 4000 feet per second, and that the bucket velocity downward is 500 feet per second. As it leaves the first set of buckets on the rotor, the steam will have an upward velocity of 4000-2X500=3000 feet per second, the effect of friction being neglected. In going through the first set of stator blades, the direction of flow is reversed, but is unchanged in magnitude. Upon leaving the second set of rotor buckets its velocity will be upward, and its magnitude will be 3000-2X500=2000 feet per second, and so on for the other two stages. Each set of moving buckets takes out 1000 feet per second of its velocity, and it emerges with no vertical component of velocity. It was shown in 145 that the rotor of a single-stage turbine has to have an absurdly large diameter unless it has a very high speed or the efficiency is very low. The objection to such high speed is that the turbine must have a reducing gear in order that the power may be used. With a multi-stage turbine we can choose the diameter and also the speed, and make the number of stages such that all the velocity can be used. Let us assume that the speed is 3000 r. p. m., and that the diameter of the rotor is 3 feet. Then the peripheral velocity must be (3000/60) X37r =471 feet per second. Neglecting the effect of friction, each stage will absorb a steam velocity of twice the bucket velocity, or 942 feet per second. If the steam expands from 165 to 15 pounds absolute, the steam velocity is 2950 feet per second. Hence the number of stages necessary will be 2950/942 =3.1, and three stages should be used. If the turbine is condensing, and the pressure drops from 165 to 1 pound absolute, 162 ENGINES AND BOILERS the steam velocity will be 4010 feet per second, the number of stages 4010/942 =4.2, and four stages should be used. In a velocity-stage turbine, the efficiency is very low after the first two stages, principally because the jet is broken up by its passage through the blades. As a result, more than two velocity stages are seldom used. It must be remembered also that a very high steam velocity produces a very great friction between the steam and the surfaces of the blades, thereby causing a consid- erable loss. Impulse Type with Pressure-staging. If, instead of expanding the steam completely in one nozzle, we ex- pand it only a little in the first nozzle, then use its velocity, ex- pand it some more in a second nozzle, and again use the velocity generated, and so on, the process is called pressure- staging (146). Figure 120 shows the method diagrammatically. In this diagram there are five sets of nozzles and five pressure stages. The steam enters from the left and passes through the stationary nozzle. The pressure and velocity lines below show that the drop in pres- sure is accompanied by an increase in velocity. The steam with its acquired velocity impinges on the blades of the rotor. The velocity is absorbed in the rotor. As the steam leaves this rotor with low velocity, it is collected and led to a second stationary nozzle in which the pressure is again dropped, and velocity is acquired. The second rotor absorbs this velocity and the steam passes on through the following stages, until its pressure and velocity are practically all used up at its exit. Making the same assumptions for speed and diameter of the rotor as in the preceding type, let us compute the number of stages necessary with the pressure-stage type. If the diameter of the rotor is 3 feet and the speed is 3000 r. p. m., there is a steam velocity of 942 feet per second to be absorbed per stage (147). If a pound of steam loses a velocity of 942 feet per second, it gives up ~XrX942 = 13900 foot-pounds Z oZ of kinetic energy. This is equivalent to 13,900/778 = 17.9 B.t.u. For the non-condensing condition assumed in 147, there was a ^ . STEAM TURBINES 163 heat-drop of 175 B.t.u. available in the whole turbine. The num- ber of stages will then be For the condensing conditions assumed in 147, the number of stages will be 17.9" In making a comparison of this type with that of 147, it might seem at first sight that the velocity-staging were the better, as the number of stages is so much smaller. While this is an advantage, it is overbalanced by the fact that the pressure-stage type is more efficient. This type is used very extensively in Multi- Pressure -Stage Impulse. FIG. 120 medium-sized turbines. It is often called the multi-cellular type because each pressure stage is composed of a cell that is steam- tight except for the openings through the inlet and outlet nozzles. It is evident that it is necessary to keep each cell steam-tight in order to prevent a leakage of steam from one stage to the next. Where no difference in pressure exists, there is no tendency to leak. In the velocity-stage type, the pressure was the same throughout the whole turbine beyond the expanding nozzles, and so there could be no leakage. 164 '; ENGINES AND BOILERS 149. Multi-stage Impulse Type with Combined Pressure- staging and Velocity-staging. Very often a combination of pressure-staging and velocity-staging is used, with the result that some of the advantages of both types are utilized. The diagram of Fig. 121 shows this arrangement. In the sketch three pressure stages are shown, and each pressure stage has two velocity stages. The drop in pressure occurs in the three stationary nozzles. After expanding in the nozzle, the steam passes through the buckets /*" Pressure g & Pressure 5fe?e 3 & Pressure 5hj2 is formed. If not enough air is supplied, the burning is only partial and CO is formed. This CO may later be burned to CO2 by the addition of more air, which is done in the engine cylinder in the case of the producer and engine plant. If steam is passed through a hot carbon bed, a decomposition of the steam takes place. The hydrogen is liberated as H 2 and the oxygen combines with the carbon to form CO. Both of these gases are valuable as fuel, and the mixture is often called water gas. When air is passed through the hot carbon bed, and CO is formed, heat is generated, so that the bed gets hotter and hotter. On the other hand, when steam is passed through, heat is ab- sorbed and the bed gets cooler and cooler. By the proper pro- portioning of air and steam passed through, it is possible to keep the fuel bed at the proper temperature. This is what is done in the gas producer. It is evident that most of the gases in producer gas will be CO, H and N, the nitrogen coming from the air and being inert. The fuel used in the producer may be coke or coal. Better results are obtained by using anthracite coal than by using bitu- minous coal. This is partly due to the fact that bituminous coal 206 ENGINES AND BOILERS tends to cake and needs constant working or poking to keep holes from burning through the cake, thereby letting excess air get through. Moreover, bituminous coal gives off various tars when it is heated. If these are not removed, they clog up the pipes and the engine. The removal of the tar is not easy. It is some- times done by throwing the tarry material out of the gas by centrifugal force by means of a kind of fan arrangement, or by passing the gas through scrubbers. Devices have been tried FIG. 146 whereby the distilled products are passed through the hot fuel bed and their composition thereby changed. This last method resembles the underfeed furnace used in steam plants. Figure 146 shows a form of gas producer. Coal is fed into the hopper A. From this, it is dropped into the chamber B. Pass- ing out of the bottom of B, it is scattered in a uniform layer over the fuel bed by means of the spiral spreader C, which is rotated by the bevel gear Q. The fuel bed may be divided roughly into three zones. The top zone is the green-coal zone, where distillation takes place. The volatile products pass on out with the other gas. As the volatile products are driven off, and the bed settles, the fuel reaches the coke zone. It is here that the burning and de- composition mentioned above take place. As the carbon is burnt GAS ENGINES 207 out, the ash settles to the bottom of the producer, where it is raked out through the water seal R. Air from the duct E is led to the bottom of the fuel bed through D, and passes up through the coke zone and the green-coal zone. Steam is admitted to the air supply through the pipe G. Holes are provided in the furnace walls at F for the working of the fuel bed. The gas leaves the producer through the opening at 7, and passes down the pipe shown to K. From K it passes through the pipe L to the wet scrubber. The wet scrubber is filled with coke, Pj which is continuously sprinkled with water from the nozzle N. The gas passing up through the wet coke is cooled and deposits dust, tar and other impurities. The gas leaves the wet scrubber at 0, and may go either directly to the engine or else to a dry scrubber. The dry scrubber is filled with wood shavings or excelsior, which takes out the remaining tar. Upon starting the producer, the gas is vented to the roof through the valve J. As soon as the quality of the gas becomes good enough, the valve J is closed and the engine is started. The coal in B is kept cool enough to prevent it from burning, by a water- jacket S. The producer is lined with firebrick. Air either may be blown into the producer, or it may be drawn through by the suction of the engine. In the former case a storage tank for the gas is necessary. With the suction type the storage is unnecessary, as the engine draws through only what it needs. When a water seal is used at the bottom, the producer is called a wet bottom producer. The poking of the fuel bed may be done by hand or mechanically. 178. Cooling of Cylinders. The cylinder walls of the internal- combustion engine must be kept cool enough to insure proper lubrication. This cooling is commonly done by circulating water through a jacket around the cylinder, as shown at W, in Figs. 137, 139, 140, 141 and 144. As far as the efficiency of the engine is concerned, the hotter the cylinder walls the better, so that it is evident that they should not be cooled any more than is neces- sary to insure lubrication. In small engines, the cylinder is sometimes placed in the lower part of a hopper filled with water. Fresh water is added as it boils away in the hopper. Another method that is used occasionally to cool the cylinder 208 ENGINES AND BOILERS Met FIG. 147 is by having the outer walls of the cylinder and head covered with fins. These present a large surface to the air and the heat is radiated from them. To increase this transfer of heat, a cur- rent of air is kept moving over the surface of the fins by means of a fan. This latter method is called air-cooling. Figure 147 shows an air-cooled cylinder. 179. Ignition. The charge of combustible mixture in an internal- combustion engine is fired in vari- ous ways. A method formerly used quite extensively but now not very common is the hot-tube method. This method is illustrated by the sketch in Fig. 148. The tube which connects with the cylinder is heated by a gas jet. Arrangement is made so that the flame can be shifted along the tube, heating it closer or farther away from the cylinder. To explain how the scheme works, suppose the charge has been fired. The tube will then be filled with burnt gas. After the fresh charge has been drawn in and compression started, the fresh gas is forced into the tube. When it is compressed into the tube far enough to strike the heated part, ignition occurs. Naturally, this scheme can be used only when the load is constant. As has been explained previ- ously, ignition may be had by using a high compression, so that the temperature of compression will be high enough to fire the charge. This method is used mostly in oil engines and is assisted in those using the lower pressures by a hot bulb or plate. The hot bulb acts somewhat in the same manner as the hot tube just mentioned. With gas FIG. 148 GAS ENGINES 209 or the more volatile liquid fuels this method is not satisfactory, since early ignition is apt to occur. The most common method is electric ignition. There are two general types of electric ignitors, the jump-spark and the make- and-break. In the former, a spark plug is used (Fig. 149), which has two fixed terminals exposed to the gases of the cylinder. At the proper time for ignition a current with voltage high enough to jump the gap is introduced in the circuit. The heat of this spark ignites the charge. The details of timing and of producing the current will not be dis- cussed here, except to say that the current may be furnished either by dry-cell or storage batteries, or by a magneto. In the make-and-break system, two electrodes are brought into contact within the cylinder and are sepa- rated at the proper time for ignition. As the circuit is broken a spark is formed between them. The details of the many schemes used will not be discussed here. This make-and-break system does not require as high an e.m.f. as the jump-spark system, but it is limited to the slower engine speeds. 180. Valves. The earlier types of gas engines were equipped with slide-valves. These required lubrication, which is some- what difficult at high temperatures. Except for a few engines that use sleeve-valves, gas engines of today have the so-called lifting or poppet valves , which are commonly lifted .by cams on a cam shaft. In the four-stroke cycle engines the cam shaft is geared to run at half the speed of the crank shaft, so that the cam shaft makes one revolution per cycle. The cams are placed on the cam shaft so that the valves open and close at the proper time. Each valve is kept seated by means of a spring around the valve stem. Figure 147 shows the cams, and how they are made to lift the valves. In some of the slower speed engines, the inlet valve is not opened by means of a cam, but by the suction inside the cylinder. With this arrangement a strong spring cannot be used to seat the valve. 181. Governing. As far as the governor itself is concerned, the gas-engine governor does not differ materially from the steam- engine governor. Both centrifugal and inertia governors are used. 210 ENGINES AND BOILERS Depending upon how the governor regulates the amount of work done in the engine cylinder, we have three general types of gas engine governors: (1) hit-and-miss governors, (2) quantity gov- ernors, and (3) quality governors. HIT-AND-MISS GOVERNOR. With this type of governor there is a working stroke for every cycle under conditions of maximum load. At lighter loads the governor mechanism fails to admit a charge occasionally, giving what might be termed a blank cycle in which no work is done by the cylinder. This drops the speed of the engine and the governor acts so that fuel is again taken in as before. With this scheme the engine either operates under conditions of maximum efficiency, or it does not fire at all. This method of governing gives better economy at light loads than the other methods, but it does not give close speed regulation. When the engine misses, the exhaust valve is commonly held open so that there is no work done in useless compression. QUANTITY GOVERNOR. For every engine there is a ratio of gas to air, which is nearly constant, with which the engine gives the best efficiency. With the quantity governor, this ratio is kept (theoretically) constant. Regulation is accomplished in two ways : (1) by the cut-off governor, and (2) by the throttling governor. With the cut-off or throttling governor, the normal mixture is allowed to enter the cylinder only during a part of the suction stroke at light loads. The length of time the mixture is ad- mitted is controlled by the governor. With the throttling gover- nor, the normal mixture is taken in during the whole of the suc- tion stroke, but the opening is throttled so that not as much enters at light loads as at full load. QUALITY GOVERNOR. This governor changes the ratio of the fuel to the air at different loads. At full load, a rich mixture is used, and at light load a lean mixture. Mechanically, this scheme is quite simple, but it has the disadvantage of giving low effi- ciency at light loads. If the mixture gets too rich or too lean, it may be impossible to secure ignition. Oil engine governors commonly control the amount of oil admitted per working stroke. This is seen to be the same as the quality-governing scheme. It should be mentioned that the speed of an engine can be regulated by changing the time of ignition. With either too early or too late ignition the full power is not developed in the GAS ENGINES 211 cylinder. The speed of motor-boat engines is often controlled in this way. With high speeds, the spark should occur earlier than in low-speed engines. With a variable-speed engine, such as exists in an automobile or truck, the time of ignition should be adjusted to the speed in order to get the best results. 182. Determination of Horsepower. The indicated horse- power of a gas engine is determined in the same way as for the steam engine, with the exception that for four-stroke cycle en- gines only half the r.p.m. is used in the computation. If a hit-and-miss governor is used on the engine, the number of hits per minute must be counted rather than the r.p.m. of the shaft. 183. Multi-cylinder Engines. The single cylinder, single-act- ing four-stroke-cycle gas engine has one impulse stroke in two revo- lutions. The double-acting steam engine has two impulse strokes per revolution. Thus it is seen that the single cylinder gas en- gine has a much greater variation in angular acceleration of crank shaft than does the steam engine. For some kinds of service this variation in angular velocity is immaterial; in other cases it is a serious disadvantage. For instance, in the generation of elec- tric current to be used for lighting, a single-cylinder engine is impracticable, unless an exceedingly heavy flywheel is used. To approximate the uniformity of torque that exists in a single- cylinder steam engine, it is necessary to use four cylinders on the gas engine. In automotive service, a fairly uniform torque is desirable, and therefore four or more cylinders are used. If more than four cylinders are used, there will be less variation in the angular torque, and the engine speed may be controlled more easily by the throttle. Too large a number of cylinders may cause a de- crease in the efficiency of the engine. This may be explained by the fact that with a multi-cylinder engine, there is more area of cylinder wall exposed to the burned gas for the same volume of gas than there is in the single-cylinder engine. Principle I of 169, as set forth by Beau de Rochas, is a statement of this same fact. While a lowered efficiency may result from the use of a large number of cylinders, this loss may be more than com- pensated by the added smoothness of running. Many of the higher grade automobiles made at the present time are equipped with six, eight, or even twelve cylinders. 212 ENGINES AND BOILERS PROBLEMS 213 PROBLEMS 1. (a) What is the pressure in pounds per square inch that corresponds to a mercury column 16 inches high? (6) What is the atmospheric pressure when the barometer reads 27.4 inches? 2. A steam gage is used to show the pressure in a steam line and is at- tached as shown in Fig. A. If the small pipe leading to the gage is full of water and the gage reads 183 pounds, what is the pressure in the steam line? 3. If the pressure gage on a boiler reads 150 pounds and the barometer reads 29.3 inched, what is the absolute pressure in the boiler in pounds per square inch? 4. The vacuum gage on a con- denser reads 27.2 inches and at the same time the barometer reads 29.1 inches. (a) What is the absolute pres- sure in the condenser in pounds per square inch? (6) What is the vac- uum-gage reading reduced to a JT IG 30-inch basis? 6. A condenser with its air pump is guaranteed by the manufacturer to produce a vacuum of 28.5 inches (on the basis of a 30-inch barometer). Dur- ing the acceptance test the barometer read 28.73 inches. What should be thevacuum-gage reading to maintain the guaranteed vacuum? Q) A boiler-feed pump is located 14 feet below the water line of the boiler. The pump draws water from a tank located 7 feet below the pump cylinder. If the pressure in the boiler is 40 pounds gage, neglecting friction losses due to the flow of water, etc., what is the least total head the pump must act against: (a) In feet? (6) In pounds per square inch? (c) What is the least height of water level in a standpipe above the boiler in order that the water will flow into the boiler by gravity? (Y) Reduce: (a) A temperature reading of 50 Centigrade to the corre- sponding Fahrenheit reading. (6) A temperature of 320 Fahrenheit to Centigrade. (c) 75 (great) calories to B.t.u. (d) 46 B. t. u. to calories. (e] 33000 foot-pounds to B.t.u. 8. If 1,576,000 B.t.u. are given to an engine in an hour and if the engine can convert 6 per cent of this heat into work, what is the horsepower of the engine? (One horsepower is 33,000 foot-pounds per minute.) 9. A sample of Indiana coal gave the following proximate analysis: Moisture =3.81%, Fixed carbon = 76. 16%, Volatile combustible = 13.62%, Ash = 6.41%. The same sample when dried gave the following ultimate analysis: Carbon =84.26%, Oxygen =1.73%, Sulphur = 1.22%, Hydrogen= 4.38%, Nitrogen = 1.75%, Ash =6.66%. The oxygen calorimeter gave a calorific value of 14682 B.t.u. per pound of dry fuel. Find the calorific value of the preceding sample per pound of dry fuel, from (a) the proximate analysis, (6) the ultimate analysis. 214 ENGINES AND BOILERS A sample of West Virginia coal gave the following proximate analysis : Moisture =4.85%, Fixed carbon = 68.36%, Volatile combustible = 16.31 %, Ash = 10.84%. The same sample when dried gave the following ultimate analysis: Carbon =80.34%, Oxygen = 3.11%, Sulphur = .49%, Hydrogen = 4.00%, Nitrogen = 1.05% Ash =11.01%. The oxygen calorimeter gave for a similar dried sample a calorific value of 14,180 B.t.u. per pound. Find the calorific value per pound of dry coal from ^ (a) the proximate analysis (b) The ultimate analysis. (llj Find the theoretical weight of air required to burn completely a pound of the dry coal of Problem 10. 12. A boiler test was run using the coal from which the sample of Problem 10 was taken. During the test the analysis of dry flue-gas was as follows: Carbon dioxide = 8.58%, Carbon monoxide = .05%, Oxygen =11.32%, Nitrogen = 80.05%. Find the approximate weight of air used to burn one pound of dry coal. (13? In the test of Problem 12, the temperature of air entering the furnace was'64 F., and the stack temperature was 559 F. Find the percentage of available heat carried up the stack by the dry flue-gas. 14. Water is fed to a boiler at a temperature of 170 F. The pressure gage reads 140 pounds, and the barometer 29.6 inches. How many B.t.u. are needed to generate a pound of dry steam? What is the temperature of thesteam generated? The volume per pound? Q$ Steam at a gage pressure of 135 pounds is generated from water at 120 F. The temperature of the steam is 490 F. The barometer reads 29.3 inches. Find the B.t.u. required to generate one pound of steam. 16. If the temperature of steam in a condenser is 115 F., what is the great- estjaossible vacuum-gage reading, if the barometer reads 29.16 inches? Qj) If water boiling under a pressure of 185 pounds gage is allowed to escape to the atmosphere (as in a boiler explosion), what percentage of its weight turns to steam? What is the ratio of its new volume to the old? Assume that the barometer reading is 29.6 inches. 18. Dry steam leaves a boiler at a pressure of 180 pounds gage and reaches the engine with a quality of 98 per cent, and a pressure of 177 pounds gage. What percentage of its heat contents has it lost in its passage through the pipe? What percentage of its volume? Assume that the barometer reading is 29.43 inches. Q9^ If one pound of steam of 95 per cent quality at atmospheric pressure is mixed with 8 pounds of water at 70 F., what will be the resultant tempera- ture? Assume that the barometer reading is 29.00 inches. 20. Dry steam enters a turbine at a pressure of 180 pounds gage; leaving the turbine it passes into a condenser in which the vacuum is 27.6 inches (30- inch basis). The quality of steam as it leaves the turbine is 87%. Neglect- ing all losses, find how many foot-pounds of work may be obtained from each pound of steam that passes through the turbine. /2L) A frictionless piston weighing 7000 pounds is placed in a vertical cyimcler 10 inches in diameter. Two pounds of water at 70 F. are placed PROBLEMS 215 under the piston. If 800 B.t.u. are added to the water, how far will the pis- ton move? The barometer reads 29.6 inches. 22. If, in Problem 21, 2500 B.t.u. are added to the water, what will be the weight of steam formed? What will be its temperature? How far will thejMston move? 2iP A sample of steam is taken from a steam line in which the pressure is 150 pounds gage and is led to a throttling calorimeter in which the tem- perature is 230 F. and the gage pressure is 3 pounds. The barometer reads 29.4 inches. What is the quality of steam in the line? 24. A horizontal water-tube boiler (B. and W. type) has 10 vertical rows of four-inch tubes, 9 tubes to the row. The tubes are 18 feet long, and the steam drum is 24 feet long and 42 inches in diameter. Find the heating surface and the rated horsepower (a) by using as heating surface the out- side surface of the tubes and one-half the surface of the drum; (6) by RuJle3, p. 26. j2J5c A horizontal return tubular boiler 60 inches in diameter and 18 feet long has 44 four-inch tubes. Find the heating surface and rated horsepower (a) by Rule 1, p. 26, (6) by Rule 3, p. 26. 26. A Scotch marine boiler-shell is 16 feet 3 inches in diameter and 12 feet long. There are three furnaces, each 43 inches in diameter. The boiler contains three sections of tubes, each section consisting of 110 three-inch tubes 10 feet long. Find the approximate heating surface and the horse- power. 27. A vertical fire-tube boiler (exposed-tube type) has a diameter of 30 inches and a height of 6 feet. The furnace is 25 inches in diameter and 27 inches high. There are 55 two-inch tubes 45 inches long. The normal water level is 10 inches from the top of the tubes. Find the heating surface and rated horsepower by Rule 2, p. 26. ^5P In a test of a B. and W. boiler with a hand-fired furnace at the Sew- age Pumping Station, Cleveland, Ohio, the following data were taken: Rated horsepower of boiler 150 Grate surface 27 square feet Duration of test 24 hours Steam pressure 156.3 pounds gage Temperature of feed water 58 F. Quality of steam formed 99 per cent. Total weight of coal fired (wet) 15078 pounds. Moisture in coal 7.5 per cent. Total weight of water fed to boiler 105100 pounds. Find: (a) Factor of evaporation. (6) Dry coal per square foot of grate surface per hour. (c) Equivalent evaporation per hour (from and at 212 F.). (d) Equivalent evaporation per hour per square foot of water-heating surface. (e) Boiler horsepower developed. (L Percentage of rated capacity developed. (297 In the test of Problem 28, the dry coal had a calorific value of 12292 216 ENGINES AND BOILERS B.t.u. per pound, and the cost delivered at the boiler room was $3.50 per ton of 2000 pounds. Find: (a) Equivalent evaporation from and at 212 per pound of dry coal. Combined efficiency of boiler, furnace and grate. Coal cost per 1000 pounds of equivalent evaporation. In a test of a B. and W. boiler the following data were taken: Rated horsepower of boiler 508 Grate surface 90 square feet Duration of test 16.25 hours Steam pressure 199 pounds gage Temperature of feedwater 48.4 F. Superheat 136.5 F. Total weight of coal fired (wet) 39670 pounds Moisture in coal 4.22 per cent Total weight of water fed to boiler 336200 pounds Find: (a) Factor of evaporation. (6) Dry coal per square foot of grate surface per hour. (c) Equivalent evaporation from and at 212 per hour. (d) Equivalent evaporation from and at 212 per hour per square foot of water-heating surface. (e) Boiler horsepower developed. Percentage of rated capacity developed. The coal in the test of Problem 30 gave the following proximate anal- ysis when dry: volatile combustible, 19 66 per cent; fixed carbon, 75.41 per cent; ash, 4.93 per cent. The cost delivered to the boiler room was S3. 75 per ton of 2000 pounds. Find: (a) Equivalent evaporation per pound of dry coal. (6) Combined efficiency of boiler, furnace and grate. Coal cost per 1000 pounds of equivalent evaporation. Is the boiler of Problems 28 and 29 working harder than that of Problems 30 and 31, or conversely? Give the reason for your answer. 33. Find the size of a pop safety-valve with a 45 seat for a 60-horsepower return-tubular boiler which is to carry a gage pressure of 75 pounds. Assume that the maximum evaporation is 5 pounds of water per hour per square foot of wetter-heating surface, and that the lift of the valve is 1/30 of the diameter. (3p How many 2.5-inch pop safety-valves would one 4.5-inch valve re- place, assuming that the lift is proportional to the diameter? 35. How many 3.5-inch pop safety-valves are required for the boiler of Problem 30? Assume the rate of maximum evaporation as 6 pounds of water per square foot of water-heating surface per hour, and that the lift is 1/30 of the diameter. 36. What should be the size of the pop safety-valve for the boiler of Problem 28 (a) Computed as in Problem 35? (6) Computed from the P. G. Darling formula? See p. 59. (c) Computed from the city of Chicago formula? See p. 59. (d) Computed from the city of Philadelphia formula? See p. 59. PROBLEMS 217 (e) Computed from the U. S. Supervising Inspectors' formula? See p. 59. (/) Computed from the A. S. M. E. Boiler Code Committee's require- menjt^? See Report of Boiler Code Committee of A. S. M. E. /3T,/ What should be the size of a steam pipe leading from a 250-horse- power boiler if the pressure carried is 160 pounds gage? Assume a velocity of florin the pipe of 5000 feet per minute. (ZSy/A 5000-kw. steam turbine requires 16 pounds of dry steam per hour per kw. at 160 pounds gage pressure. The vacuum in the exhaust of the turbine is 27.5 inches of mercury (30-inch barometer). The quality of steam in the exhaust is 85%. If the velocity of flow of steam to and away from the turbine is to be 7500 feet per minute, what should be the size _ol steam and exhaust pipes? /39.) If a steel steam pipe is to carry steam at a pressure of 200 pounds gageand may be as cold as 30 F. when the steam is cut off, how far apart should expansion joints be placed if each joint gives a 3-inch movement? 40. If 9536 pounds of water at a temperature of 60 F. are mixed with 1160 pounds of steam at 3 pounds gage pressure, the steam being of 90 per cent quality, what will be the resultant temperature of the mixture? ^41. The exhaust from a 65-horsepower steam engine is led to an open feedwater heater. The engine uses 30 pounds of steam per hour per horse- power, and the quality of the exhaust steam is 80%. The heater is at atmos- pheric pressure; water enters at 50 F. and is heated to 200 F. fa) What horsepower of boilers will the heater supply? (6) What should be the size of steam and water pipes leading to the heater? Assume a steam velocity of 5000 feet per minute and a water veloc- ity of 150 feet per minute. 42. A 4000-kw. steam turbine is equipped with a surface condenser The turbine uses 16 pounds of steam per kw. per hour, which enters the condenser at a quality of 85 per cent. The vacuum to be maintained is 28 inches (30-inch basis). The circulating water enters the condenser at a temperature of 60 F., and leaves at a temperature 10 cooler than that of the incoming steam. (a) How much circulating water is needed per hour? (6) If the same amount of water is circulated as in part (a), but if it enters at 90 instead of 60, and leaves at 10 cooler than the incoming steam, what vacuum can be maintained? 43. An 18"X24" steam engine has a piston rod 2.75 inches in diameter. Find the head-end and the crank-end piston displacements in cubic feet. 44. If it takes 10.6 pounds of water to fill the head-end clearance space and 11.2 pounds to fill the crank-end clearance space of the engine in Problem 43, what is the percentage of clearance for each end of the engine? 45. Find the volume of steam back of the piston of the engine of Problem 43: when the piston is at 12.4 per cent of the head-end stroke; when it is at 14.0 per cent of the crank-end stroke. ^ 46. Find the weight of dry steam back of the piston of a 24" X 36" engine when it is at 30 per cent of the head-end stroke. The head-end clearance is 4 per cent and the steam pressure back of the piston at the above position is 105 pounds gage. If we know that at this time there is actually 1.06 pounds of wet steam back of the piston, what must be its quality? 218 * ENGINES AND BOILERS Construct a hypothetical indicator diagram, using the following data. Length of diagram =4 inches (this does not include clearance). Initial pressure = 150 pounds per square inch (gage). Back pressure =5 pounds per square inch (gag). Cut-off = 25 per cent, Release =95 per cent. Compression = 15 per cent, Admission = 2 per cent. Clearance = 7 per cent. Atmospheric pressure = 15 pounds per square inch, as a scale of pressure 60 pounds per inch. (Construct a hypothetical indicator diagram for a uniflow engine 108), using the following data. Length of diagram =4 inches. Initial pressure = 170 pounds. Back pressure = 12 pounds (engine is running condensing). Cut-off = 20 per cent. Release and compression each =90%. Admission = 2 per cent. Clearance = 3 p r cent. Also show, by a dotted line on the same diagram, the compression curve when the engine runs non-condensing (back pressure = 0). State in what ways this excessive compression may be relieved. 49. Compute approximately the percentage of head-end and of crank-end clearance of the engine from which the cards of Fig. B were taken. Use two methods. Cards were taken with an 80-pound spring. 50. Compute the engine con- stant, or the horsepower constant (LA/33000), for the head end and for the crank end for a 10"X14" engine with a 2" piston rod. Your answer must be correct to within onephalf of one per cent. -p IG -g (51/ Find the indicated horse- power (i. hp.) of the steam engine of Problem 50 when the head-end mean effective pressure is 34.2, and the crank- endgjke.p. is 35.4 pounds per square inch. The engine is running at 260 r.p.m. (52/ A test was run on a 14" X 18" steam engine with a 2" rod. The head-end m. e. p. was found to be 35.2 pounds per square inch and the crank- end m.e.p. 34.6 pounds per square inch. The speed of the engine was 250 r.p.m. The power was absorbed by a Prony brake whose arm is 6' 5" long. The effective weight of the brake arm on the scales was 45 pounds. During the test the pressure on the scales was 382 pounds. Find (a) the indi- cate4\h rse P wei S (&) the brake horsepower; (c) the mechanical efficiency. H>3/ The test of Problem 52 was run for 45 minutes, during which time thVengine used 2750 pounds of steam at a pressure of 120 pounds gage, and at a quality of 97 per cent. Find: (a) Dry steam used per indicated horsepower per hour. (6) B.t.u. per indicated horsepower per minute, (c) Thermal efficiency based on i. hp. (d) Thermal efficiency based on b. hp. PROBLEMS 219 The indicator diagram in Fig. C and the following data were taken during cut-off re/ease compression FIG. C a test of a Buckeye engine. Size of engine, 7.75" X 15", \\\" rod. Radius of Prony brake arm =6.02 feet. Room temperature = 73.5 F. Temperature in throttling calo- rimeter =221.5 F. Steam pressure at throttle = 128.7 pounds per square inch gage. Steam pressure in calorimeter = 1.125 pounds gage. Barometer = 28.5 inch. R.p.m.= 222.5. Net brake load = 140 pounds. Scale of indicator spring = 80 pounds. Steam used per hour = 1161 pounds. 54. Find the m.e.p. of the cards by the mean-ordinate method. 65. Find the indicated horsepower, head-end, crank-end, and total. Find the brake horsepower. Find: Mechanical efficiency. Pounds of steam per i. hp. per hour. B.t.u. per i. hp. per minute. Thermal efficiency based on b. hp. Determine from each card the percentage of stroke and the steam pres- sure for each of the following events : (a) Cut-off. (6) Release (c) Compression. 59. Determine the weight of dry steam back of the piston for each end at the events of cut-off, release, and compression. 60. Find the amount of re-evapo- ration or condensation per hour dur- ing expansion. 61. Find the weight of dry steam per hour per indicated horsepower accounted for by the cards. 56. 57. (a) (6) (c) (d) 58. ZO pound s/orinq. FIG. D 62. Combine the indicator dia- grams shown in Fig. D, and deter- mine the diagram factor. The cards of Fig. D were taken from an 8.02" X15"X24" cross-compound Corliss steam engine, running at 85 r.p.m. The head-end clearance of the high-pres- sure cylinder is 7.4 per cent and the head-end clearance of the low-pressure cylinder is 6.01 per cent. 220 ENGINES AND BOILERS Determine the size of cylinders for a compound, two-cylinder, double- acting steam engine (receiver type), assuming the following data: i. hp. = 120, r. p.m. = 100, cylinder ratio = 1/3, piston speed = 600 feet per minute, initial steam pressure = 140 pounds absolute, termin^ljgressure of hypothetical dia- gram =14 pounds absolute, vacuum = 24 inches (30-inch basis), and diagram f actor = .85 QJ4,/ In a certain two-cylinder compound steam engine the number of ex- pansions is 10, the initial steam pressure is 120 pounds absolute and the back pressure is 5 pounds absolute. The receiver pressure is 30 pounds abso- lute. The cylinder ratio is 1 to 3. Neglecting clearance and piston rods, compare the work done in the two cylinders and the stresses on the two piston rods. 65. Given a cross-compound steam engine, show by means of a graph the l/j/ve shown in mid-position f- >- "4* >l \ head end /ohfon -*r cnwft \ end FIG. E variation in power distribution when the governor varies the cut-off equally in each cylinder (choose at least three cut-offs). 66. Proceed as for Problem 65, but assume that the governor varies the point of cut-off in the high pressure cylinder only. 67. Consider a 12"X18" steam engine (section of cylinder and valve shown in Fig. E), with the following given data. Connecting rods 6 feet long. Val ve- travel = 6 inches. Head-end lead = crank-end lead = .25 inch. Head-end steam lap = 1.25 inches. Head-end exhaust lap = .5 inch. Width of port = 1.75 inches. (a) Draw the valve on its seat, the crank position, the eccentric position, and the position of the piston in the cylinder when the crank is on head-end dead center. (Make your drawing y actual size.) (6) Draw the same parts for head-end cut-off. (c) Draw the same parts for head-end admission. (d) Draw the same parts for head-end release. (e) Draw the same parts for head-end compression. (/) Determine the percentage of the stroke for each of the above events. PROBLEMS 221 'Consider a 14"X16 L aruTwith the following data. engine, running over, with direct slide-valve valve-travel = 4 inches, lead = 1 A inch,' steam lap = 1 inch, v exhaust lap = K inch/ / (a) Valve ellipse. Draw the crank circle K actual size, and about the same center draw the eccentric circle full size. Choose 12 equidistant crank positions and find the corresponding eccentric positions. For any crank position (as C, Fig. F 2 ), the piston is at a distance x from its mid-position, and at the same time the eccentric is at a distance y from its Jfeam /ap+/ead Crank on head-end dead center FIG. F 2 FIG. F 4 mid-position. Plot y vertically and x horizontally, for all 12 positions of the crank. Connect the points thus found by a smooth curve. Label on this diagram the following details: the crank position at each event of the stroke, the lead, the steam lap, the exhaust lap, the maximum port-openings, and the angle of advance. (6) Bilgram diagram. Draw the crank and eccentric circles and choose 12 equidistant crank positions as in (a) . For each crank position (as C in Fig. F 3 ), draw a dotted line parallel at a distance y from the crank. The inter- section of these dotted lines is the Bilgram construction point P. About this point P, draw in the steam-lap and exhaust-lap circles. Show on this diagram the crank position at each event, the lead, the steam lap, the exhaust lap, the maximum port-openings, and the angle of advance. (c) Zeuner diagram. Draw the crank and eccentric circles as before, and choose 12 equidistant crank positions. Lay off radially on the crank from the center of the crank circle the eccentric displacement y (Fig. F 4 ) ; connect all points thus found by a smooth curve. Show on this diagram the crank position at each event, the lead, the steam lap, the exhaust lap, the maximum port-openings, and the angle of advance. 222 ENGINES AND BOILERS 69. Consider an engine with the following given data. Direct slide-valve. Head-end steam lap = 1| ". Engine running over. Crank-end steam lap = 1 inch. Valve-travel = 5 inches. Head-end exhaust lap = J4 inch. Head-end lead = 1/8 inch. R/L = 1/5. Find the head-end and crank-end crank positions, and the percent of stroke at each event by means of (a) The valve ellipse, (6) the Bilgram diagram, fc) the Zeuner diagram. 70. Consider an engine with a direct slide-valve and with the following given data: Engine running over. Crank-end cut-off =50 per cent. Valve-travel = 3 inches. Head-end compression = 25 per cent. Head-end admission = 1 per cent. Crank-end compression = 25 per cent. Head-end cut-off = 50 per cent. R/L = 1/6. Find the percentage of stroke at all events, the angle of advance in degrees, the steam laps, the exhaust laps, the maximum port-openings, and the leads, by means of (a) The Bilgram diagram, (6) the Zeuner diagram. Draw the eccentric circle full size and the crank circle to such a scale that it is the same size as the eccentric circle. Label all of the dimensions asked for directly on the diagrams, also label the head end of the diagram and in- dicate by an arrow the direction of rotation of the crank. 71. Consider an engine with an indirect slide-valve and with the follow- ing given data. Engine running over. Valve-travel = 4 inches. Head-end lead = | inch. Crank-end lead = Y inch. Head-end cut-off = 35 per cent. Head-end compression = 15 per cent. Sum of steam lap and exhaust lap the same for both ends. R/L = %. Find the percentage of stroke at all events, the angle of advance in degrees, the steam laps, the exhaust laps, and the maximum port-openings, by means of (a) The Bilgram diagram, (6) the Zeuner diagram. 72. Consider an engine with a direct slide-valve and with the following data. Engine running over. Head-end admission = 2 per cent. Head-end cut-off = 60 per cent. Head-end maximum port-opening = 1 .25 inches. Crank-end maximum port-opening = 1.25 inches. Head-end compression = 20 per cent. Crank-end compression = 20 per cent. Find the valve-travel, the angle of advance, and each of the laps, by means of (a) The Bilgram diagram, (6) the Zeuner diagram. Also draw to scale the valve on the seat in its mid-position. PROBLEMS 223 73. Consider an engine with a direct slide-valve and with the following data. Engine running under. Head-end lead = J^ inch. Crank-end lead= f inch. Head-end cut-off = 55 per cent. Head-end compression = 20 per cent. Crank-end compression = 20 per cent. Head-end maximum port-opening = 1.25 inch. Find the valve-travel, the angle of advance, and each of the laps by (a) The Bilgram diagram, (6) the Zeuner diagram. Draw the valve to scale in mid-position. FIG. G 74. Consider an engine with a valve whose dimensions and seat are as shown in Fig. G. The valve is not shown in mid -position. The valve-travel is 4 inches; R/L = l/5; the engine runs over. The cards of Fig. H are taken with the valve as now set. Find the angle through which the eccentric must be shifted (state whether backward or for- Cut-otf Cuf-off Crank-end card FIG. H ward), and the amount the valve stem must be lengthened or shortened (state which), in order to give a cut-off of 25 per cent on each end. Draw the ap- proximate cards for the new setting. 224 ENGINES AND BOILERS 75. Consider an automatic shaft governor with the following given data (The Rites inertia governor is shown in Figs. I and J.) : Head-end steam lap = 1.75 inches. Head-end exhaust lap = 0. Lead at normal position of eccentric = 5/32 inch. Distance of eccentric center from pivot (R) = 12". Distance from center of shaft to pivot point (x) = 13f ". FIG. I Location of point DE = 25", /=18". Location of point A: c = 30", 6 = 52". I. Cut-off at no load = 10 per cent. Cut-off at full load = 65 per cent. Direct slide-valve, engine running over. Draw the governor analysis (full size) and find the valve-travel and angle of advance (a) At 10% cut-off, (6) at 65% cut-off, (c) at normal cut-off. (d} Also find percent of normal cut-off. II. Draw the head-end Zeuner diagram, or the Bilgram diagram, for (a) 10% cut-off, (6) normal cut- off, (c) 65% cut-off. From the events thus deter- mined, draw the theoretical indicator diagrams, using 6 per cent clearance, 150 pounds initial steam pressure, 5 pounds back pressure, and 80 pounds per inch as the scale of spring. Find the elongation of governor spring (drawing J/ size) . (a) From no load to normal, (6) from normal to full load. FlG PROBLEMS 225 76. Consider a four-valve engine, such as is shown in Figs. 60 and 61, p. 104, with head-end valves as shown in Fig. KI and K 2 , and with the follow- ing data. Radius of steam valve arms = 5". Maximum diameter of steam or cut-off eccentric circle = 4". Diameter of exhaust eccentric circle = 4". With the crank on head-end dead center, the pivot point of the governor arm for the cut-off eccentric is on the horizontal center line 8^" beyond the center of the shaft. Radius of locus of eccentric centers = 7 5/16". Cut-off at maximum load = 65 per cent. Compression = 15 per cent. Cut-off at normal load = 25 per cent. Release = 95 per cent va/ve in extreme open position (Fu// /oad) 'exhaust rtt/ye !/r extreme c/osed joosift'on FIG. K At normal load the steam-valve arm is vertical when the valve is in extreme closed position. The exhaust-valve arm is vertical when that valve is in ex- treme open position. R/L = 1/Q. Engine runs over. Find the angle of advance for each eccentric at normal load. Find the location of the steam-valve arm when in mid-position, at cut- off, at admission, and when it is in extreme open position at normal load. Find the maximum port-opening at normal load, and the lead at normal load. Draw the head-end steam valve in extreme position, and in open position at normal load. Draw the head-end exhaust valve in extreme open position. 226 ENGINES AND BOILERS 77. The necessary dimensions of a Corliss engine are given in Figs. L and M. Consider such an engine with the following data. A 12"X24" Corliss engine running at 150 r.p.m. All valves operated by one eccentric. Engine runs over. FIG. L Diameter of all valves =3" (d, Fig. L). m, Fig. L=sy 2 "; k, Fig. L = 15". Length of valve arms = 4' / . Center of wrist plate is equidistant from all valves. Radius AO and BO on wrist plate = 5". Radius HO on wrist plate (0 =6". Angles EC' A' and FD"B" = 90. Release = 98 per cent. PROBLEMS 227 Compression = 4 per cent. Crank angle at admission = 3. Throw of eccentric = 6f." Radii of rocker arms are equal for eccentric and hook rods. Normal cut-off = 20% . Width of admission port = %". Width of exhaust port = l". Single-ported valves. In Fig. M, Radius of arm EG =3%' Radius of arm EH =4^". Radius of arm El =4". Radius of cam EJ =2". Radius of latch EK = %1". Center G is V/i" above horizontal center line at trip position for normal cut-off. Make the general layout % actual size, and that of the trip mechanism full size. Find the lengths of the steam rod AC and the exhaust rod BD, the angle of advance, the steam lap, the lead, the exhaust lap, the maximum FIG. M cut-off with trip working, the maximum cut-off when beyond control of trip, the maximum port-opening for maximum cut-off, the maximum port-open- ing for normal cut-off, the maximum port-opening for 10% cut-off, the move- ment of the governor rod from normal to 10% cut-off, and the movement of governor rod from normal to maximum trip cut-off. 228 ENGINES AND BOILERS PROBLEMS 229 78. The necessary dimensions of a Stephenson link are as follows. (Fig. N.) Valve-travel at full gear = 5>". Steam lap = I". Exhaust lap = T y. Lead at full gear=0". Steam port = W. Exhaust port = 2Y 2 ". Bridge = 1". = l/7.5 a = b = 45". m = 3". e=U". h=U". .7 = 1713/32". k = l8". #==48.8". Make the drawing one-half actual size and proceed as follows. (1) Make a template of the link as shown in Fig. N. (2) Locate the center of the link-block with the crank at head-end dead center at full gear. (3) Find the center of travel of the link-block, neglecting for the time being the angularity of the eccentric rods. (4) Place the center of the rocker shaft above the point found. (5) Place the crank and eccentrics in their positions at 40% head-end cut- off, running forward. (6) Find by trial the position of the link (template) for the preceding position of the crank. (Remember the center of the link-block is now at a distance equal to the steam-lap from its mid-position.) Locate the saddle- block pin and the position of the bell crank. (7) Assume twelve equidistant crank positions and the corresponding ec- centric positions for the preceding cut-off. Then draw in the center of link for each crank position, by trial by means of the template. (8) Draw a valve ellipse from the valve displacement found above. Locate all events. (9) Check as to the assumed cut-off. (10) Find the amount of slip between the link-block and link when running at the assumed cut-off. 230 ENGINES AND BOILERS PROBLEMS 231 79. The arrangement of parts and the necessary dimensions of a Wal- schaert gear are shown in Fig. O. The valve is of the piston type and has inside admission. This is the common type used on modern locomotives. Fig. O shows the piston in its mid-position and the link-block set at mid-gear. As the link is pivoted to the frame of the engine at the point L, there will be no motion of the link-block when set at mid-gear. Hence all the motion the valve gets at this position of the block comes from the cross-head. There- fore, as the cross-head is in mid-position, the valve will also be in mid-position. Suppose that such a gear is used on an engine with the following data. 26J/TX30" engine. Engine runs forward (under). Diameter of valve = 14." Maximum valve-travel = 6>". Steam lap = l 1/16". Exhaust lap = 0. Lead =3/16". Dimensions as in Fig. O. Make the drawing one-fourth actual size, and proceed as follows: (1) Lay out the gear in the position shown in Fig. O, with the piston in mid-position and the link-block set at mid-gear. Indicate each of the parts by its center line only. (2) Make a template of the link. (3) Set the valve and the point H for head-end cut-off. Then place the crank at 40% of forward stroke, assuming that the engine is running forward. (4) Locate the eccentric E at 90 back of the crank, and the point F, and draw in the center line of the link. (5) With the cross-head K in position for 40% cut-off, the location of the point / will be determined. Since H was located in (3), the point G is found by connecting / and H. The distance that G is to the right of the mid- position gives the distance that the link-block center is to the right of its mid-position. This locates the link-block for 40% cut-off. Now locate the points D, M and B. (6) With the link-block set for 40% head- end cut-off, take twelve equidis- tant crank positions and find the valve displacement for each position. Plot these valve displacements against the corresponding piston displacements either as in a Zeuner diagram or as in a valve-ellipse diagram, and connect the points thus found by a smooth curve. (7) Draw in the laps on the diagram just constructed, and check the cut-off with the assumed value of 40%. (8) Find the amount of slip between the link-block and the link when the posi- tion is that of 40% cut-off, with the engine running forward. 80. The Russell Engine Co. makes a four- valve engine. In this engine, the exhaust is taken care of by oscillating or Corliss valves, and the admis- sion by a direct slide-valve. This slide-valve admits steam and carries on its back a rider-valve that cuts off the steam. The main valve is driven by an eccentric keyed to the shaft, while the rider-valve is driven by an eccentric whose angle of advance is controlled by the governor. The gov- ernor simply rotates the eccentric about the shaft, changing the angle of 232 ENGINES AND BOILERS advance, but affecting in no way the absolute travel of the valve. Hence, it is necessary to consider the relative motion of the rider-valve and the main valve in making an analysis of the cut-off valve. The necessary dimensions of the valves and the seat are given in Fig. P. The throw of both eccentrics is 5K inches, and the angle of advance of the main eccentric is 32.5 degrees. Proceed with the analysis in the following manner. (1) Draw the two eccentric circles about the same center and locate the extremity of the diameter of the valve circle for the main valve. Trgyf/ grf both IM/HCS S4? Seaf FIG. P VALVES AND SEA^T OF RUSSELL FOUR-VALVE ENGINE (2) Place the crank for cut-off at 25% of the head 7 end stroke, and locate the main eccentric. Then determine from the dimensions in Fig. P how far from the mid-position the cut-off eccentric must be to give the proper posi- tion of rider-valve for cut-off. This determines the angle of advance of the cut-off eccentric at this particular cut-off. (3) Take twelve equidistant crank positions and determine the relative position of the rider-valve to the main valve for each position. Plot these distances as in a Zeuner diagram. It will be noticed that the diameter of the relative valve-circle is equal in amount and parallel in directio'n to a line connecting the extremities of the diameters of the valve circles in the Zeuner diagrams for the main valve and the rider- valve. Also determine the relative steam lap, which is negative. It will be found that this is equal to the distance between the working edges of the rider-valve and the main valve when both are in mid-position. PROBLEMS 233 (4) Now determine the diameters of the relative valve-circles in amount and in direction by repeating the process of (3) for eight cut-offs (0% to 70%), and draw the locus of the extremities of the diameters of the relative valve- circles. It is seen that this locus is the arc of a circle whose radius is equal to the eccentricity of the rider-valve eccentric and whose center is the ex- tremity of the diameter of the valve circle for the main valve. (5) Make the Zeuner analysis for cut-offs of 10%, 30% and 60%, finding the relative angle of advance and the relative valve- travel by drawing a per- pendicular to the crank position at the point where the crank cuts the rela- tive steam lap. The point where this perpendicular intersects the locus of the extremities of the diameters of the relative valve-circles determines the construction point for the relative Zeuner diagram. 81. A gravity-balanced spindle governor built as shown in Fig. Q has arms 20 inches long. At normal speed the arms are at an angle of 45 with the horizontal. ' (a) Find the normal speed of the governor. (6) Find the percentage of variation in speed from no load to full load. (c) If the normal speed is increased 30 per cent and the range of vertical movement of the point A is the same as before, what is the percentage of varia- tion in speed from no load to full load? o - S/o/oaJ - formal - - FU///04J FIG. Q FIG. R 82. In the cross-armed gravity-balanced spindle governor shown in Fig. R,. the upper arms are 30 inches long and the lower arms are 20 inches long. Find the percentage of variation in speed from no load to full load, and compare the result with that of Problem 81. 83. The governor of Problem 81 is now loaded with a weight of 60 pounds. If the normal speed is now 100 r.p.m., and the vertical movement of the point A is the same as before, what is the percentage of variation of speed from normal? 84. A Corliss engine is governed by a loaded gravity-balanced spindle governor. The pulley on the governor and pulley on the engine are both 10 inches in diameter. It is desired to change the speed of the engine from 100 to 120 r.p.m. In what three ways may this be done without affecting the speed regulation? Give your calculations. 234 ENGINES AND BOILERS 86. Find the percentage of variation in speed from no load to full load for the governor shown in Fig. S. 86. Suppose that the spring of a spring-balanced centrifugal governor is fastened directly to a weight of 40 pounds, as shown in Fig. T. If the speed FIG. S FIG. T at no load is 206 r.p.m., and at full load 200 r.p.m., what must be the scale of spring? 87. If the spring in Problem 86 is replaced by one of 40 pound scale, and the speed at full load is 200 r.p.m., what is the speed at no load? Is the governor stable or unstable? 88. What scale of spring would make the governor of Problem 86 iso- chronous at a speed of 200 r.p.m.? 89. The no-load speed of a governor is 300 r.p.m. and the full load speed is 290 r.p.m. At no load, the tension in the governor spring is 500 pounds. At full load, the spring is 2 inches shorter than at no load. The scale of spring is 100 pounds. If the spring is tightened by shorten- ing it up half an inch, what will be the effect on the no-load and full-load speeds? 90. With the governor of Problem 89, how much would the spring have to be taken up to make the governor isochro- nous? What would the speed then be? 91. In the steam-turbine governor shown in Fig. U, the weights of 8 pounds each are 4 inches from the center of the spindle at full load when the speed is 600 r.p.m. At no load, the weights are 5 inches FIG. U from the spindle and the speed is 610 r.p.m. The weights are pivoted at points A and A . (a) Compute the scale of spring. (6) Design the spring, i.e. find the size of spring wire, the diameter of the coil, and the number of turns that will give the correct force and scale. M512399 TJ255 B9 Forestry