u. . G/'/es. y.o/O. /&3***A o. THE LIBRARY OF THE UNIVERSITY OF CALIFORNIA LOS ANGELES //- STEAM TURBINES Published by the Me G raw- Hill Boot Company Ne^v York Successons to theBookDepartmervts of the McGraw Publishing Company Hill Publishing Company Publishers of Books for Electrical World The Engineering and Mining Journal Engineering Record Power and The Engineer Electric Railway Journal American Machinist Metallurgical and Chemical Engineering STEAM TURBINES A SHORT TREATISE ON THEORY DESIGN, AND FIELD OF OPERATION BY JOSEPH WICKHAM ROE, M. E. MEM. AM. SOC. MECH. ENGRS., MEM. AM. INST. MINING ENGRS. ASSISTANT PROFESSOR OF MECHANICAL ENGINEERING, SHEFFIELD SCIENTIFIC SCHOOL, YALE UNIVERSITY. McGRAW-HILL BOOK COMPANY 239 WEST 39TH STREET, NEW YORK 6 BOUVERIE STREET, LONDON. E. C. 1911 COPYRIGHT, 1911 BY THE MCGRAW-HILL BOOK COMPANY Printed and Electrotyped by The Maple Press York, Pa. Engineering Library TJ PREFACE. The purpose of this book is twofold, to provide, first, a text- book suited to a short course on steam turbines in engineering schools and, second, a book for the engineer on the principles and general design of turbines without going into refined treat- ment of the more difficult problems entering into their design. At the end of each article is given a short list of references where those wishing to pursue the subject will find it treated most thoroughly. Other references are given among the prob- lems to sources available to most students and it is suggested that these be made subjects for short reports, thus extending the reading of the student beyond the limits of the present text. The writer wishes to thank Mr. Geo. A. Orrok of the New York Edison Company, the engineers of the prominent turbine manufacturers, for data and illustrations used, and especially Mr. C. C. Perry of the Sheffield Scientific School for his help in the preparation of the material. JOSEPH W. ROE. SHEFFIELD SCIENTIFIC SCHOOL, YALE UNIVERSITY, April, l. 1911. 713873 TABLE OF CONTENTS. CHAPTER I. EXPANSION OF STEAM. PAGE. Art. 1. The Energy of a Jet 1 " 2. The Temperature-entropy Diagram 4 " 3. The Heat-entropy Diagram 7 " 4. The Velocity of a Jet 9 " 5. Weight of Steam Delivered and Impulsive Force . . 13 CHAPTER II. UTILIZATION OF KINETIC ENERGY IN STEAM. Art. 6. Expanding Nozzles 16 " 7. Design of an Expanding Nozzle 22 " 8. Blade Forms and Blade Velocities 26 9. Conditions of Maximum Efficiency . 29 " 10. The Trajectory of the Steam . . : 35 CHAPTER III. CALCULATIONS OF TURBINE BLADING. Art. 11. Single-stage Impulse Turbines 40 " 12. Multi-stage Impulse Turbines 44 " 13. Single Flow Impulse-reaction Turbines 54 " 14. Double Flow Impulse-reaction Turbines 66 CHAPTER IV. MECHANICAL PROBLEMS. Art. 15. Centrifugal Strains 73 " 16. Bearings 78 " 17. Governing 82 CHAPTER V. COMPARISON OF TYPES. Art. 18. Vertical Turbines 88 " 19. Tangential Flow Turbines 91 " 20. Multicellular Turbines ...... 94 " 21. Special Forms of Impulse-reaction Turbines ... 98 " 22. Low Pressure Turbines and Regenerators . . . . . 101 CHAPTER VI. CONDITIONS OF OPERATION. Art. 23. Effect of Superheat 108 " 24. High Vacuums in Engines and Turbines 110 CHAPTER VII. THE POSITION AND FIELD OF THE STEAM TURBINE. Art. 25. Relative Economy of Engines and Turbines . . . .116 " 26. General Comparison of Engines and Turbines. . 126 " 27. Relative Fields of Engines and Turbines 131 INDEX 137 vii STEAM TURBINES. STEAM TURBINES. CHAPTER I. THE FREE EXPANSION OF STEAM. Art. i. The Energy of a Jet. The power developed in a steam turbine is derived from one or more jets of steam. The power available depends on the kinetic energy of the steam available under the given conditions of operation. The total kinetic energy in a moving mass of weight W and velocity V is WV 2 (1) where the body is brought to rest. Where the velocity is reduced from V to V l the kinetic energy available in the change is Or, from the standpoint of the force of the jet we have Ft mV. Taking the time as unity the impulsive force acting for one second is WV F=mV = -- . (3) 9 If A = area of the jet and w= the weight per cubic foot of the steam at the given pressure and dryness or superheat, the weight delivered per second is W = wAV. Substituting, Eq. (3) becomes ' (4) 2 STEAM TURBINES. For any given pressure and dryness w is known. Both the impulsive force and the kinetic energy, K, available become determinate as soon as V is known. An important part of the theory of turbines has for its object the determination of this V for any given set of conditions. Taking the weight of steam in Eq. (1) as unity we have, since the value of 2g is a known constant, a definite relation between K and 7 by which the velocity V may be calculated when the energy available per pound of steam is known. The energy of a gas such as steam may exist in two forms, heat and kinetic energy. It may consist wholly of heat. In this condition the gas is confined and under pressure from the energy of its particles, due to heat. If allowed to do so it would expand into a medium at lower pressure, doing work; but since it is confined and at rest, this capacity for work is potential only and analogous to that of a weight suspended and doing no work, but capable of doing it if released. Or, theoretically, having been allowed to expand under the influence of its heat tension or pressure until all the heat had been expended, the energy would be wholly kinetic. This condi- tion, however, is unattainable, and is analogous to the energy of the weight if it had fallen freely from its point of suspension to the center of the earth. Practically, therefore, there is a third or intermediate condition where the energy exists in both forms, and the total energy of the steam is the sum of its heat energy and the kinetic energy. As the weight can fall only to the sur- face of the ground, so steam in expanding can not give up all its heat, but can part with it only to a point corresponding to the temperature of the medium into which it expands. In the British system of units heat is measured in British Thermal Units and energy in foot-pounds, the two being inter- changeable at the ratio of 1 B.T.U = 777.5 foot-pounds. 1 From the Law of Conservation of Energy the total energy in a pound of steam during expansion remains a constant. Expressing this in the form of an equation, WV 2 WV^ 777. 5Sl+ -m.5H.+HJ. i The heat values throughout this book are those given in Marks and Davis' "Steam Tables and Diagrams." THE FREE EXPANSION OF STEAM. 3 where H^ \ and H 2 V 2 are the heat energies and steam velocity for any two stages of the expansion. Remembering that W is unity y 2_ V 2 777.5 (H 1 -H 2 )=---^-. (5) V 2 In the expansion of steam from a vessel --- may be neglected, therefore Eq. (5) becomes in which V is the theoretical velocity attained, and H l and H 2 the total heat content in B. T. U. per pound of steam before and after expansion. Problems. 1. What is the impulsive force when five nozzles are each discharging .96 Ib. of steam at a velocity of 3450 ft. per sec.? 2. What velocity must a pound of steam have to deliver 87,000 ft. Ibs. of work per sec.? 3. What energy per Ib. will steam give up when slowed down from 2150 ft. per sec. to 1210 ft. per sec.? 4. What energy per Ib. of steam will be available for a decrease in velocity from 3090 ft. per sec. to 2150? 5. What weight of steam must be delivered per sec. to give 60,000 ft. Ibs. of work per sec. when the velocities before and after ex- pansion are 2500 and 1500 ft. per sec.? 6. What is the theoretical H. P. of a jet delivering 1.3 Ibs. of steam per sec. at 3500 ft. per sec.? 7. What must be the initial velocity of steam which delivers in one jet 40 H. P. per Ib. of steam delivered, the final velocity being 1600ft. per sec.? 8. What is the impulsive force in .092 Ib. of steam moving at 1330 ft. per sec., also at 3680 ft. per. sec.? 9. How many H. P. are available in a heat drop of 42 B. T. U. per sec.? 10. What heat drop is required to generate a theoretical velocity of 3380ft. per sec.? 4 STEAM TURBINES. 11. What is the H. P. per Ib. of steam of a jet where the heat drop is 235 B. T. U.? 12. What is the heat equivalent in B. T. U. per Ib. of the work done when steam is reduced in velocity from 3400 ft. per sec. to 1300 ft. per sec.? 13. How many B. T. U.'s are the equivalent of a H. P. hr.? References. THOMAS: "Steam Turbines." JUDE : " The Theory of the Steam Turbine." Standard works on thermodynamics. Art. 2. The Temperature -entropy Diagram. The determination, then, of the theoretical steam velocities which are to govern the design, depends on the heat drop in the stage considered. This heat drop may be calculated analyt- ically from steam tables, or determined more conveniently, and with sufficient accuracy, from the entropy diagrams published with steam tables or the works on turbine engineering. J FIG. 1. Temperature-entropy diagram. In the temperature-entropy diagram for steam, the absolute temperature of the water or steam is shown by the ordinate of a point in a curve, and the total heat per pound by the projected area underneath the curve. The corresponding abscissa repre- sents a quantity known as the entropy. As engine and turbine THE FREE EXPANSION OF STEAM. 5 problems are concerned with certain changes in heat which never involve the total contents, that portion only of the diagram above the melting point of ice need be drawn. The addition of a small amount of heat, dQ, at some temper- ature T would be represented by the shaded slice, Fig. 1, or TdE = dQ. (7) If S be the mean specific heat during a change the heat added is Q=S(T 1 -T 2 ), or for differentials, dQ=SdT. Substituting, we have (7a) Integrating, the change of entropy for a rise in temperature from T 2 to T v as from A to B, is T e . (8) Equations (7) and (7a) represent the curve OB, the projected area beneath this line showing the heat contents of water. When heat is added as latent heat and goes wholly into changing the condition, as during vaporization, the temperature remains dQ constant and Eq. (7) becomes - 7 -- = a constant, i.e., the entropy dhi and the heat vary directly and the curve is a horizontal straight line represented by As the latent heat varies for different temperatures the corre- sponding horizontal lines, such as BC , will vary in length and a curve, CI', drawn through their ends will be the locus of the saturation points. After the water has been wholly converted into steam, further addition of heat raises its temperature and the equation of the curve resumes the form (8), with S as the variable specific heat of superheated steam. 6 STEAM TURBINES. Expansion through an orifice or short nozzle is so rapid as prac- tically to give the conditions of adiabatic expansion, in which there is no exchange of heat between the steam and the walls surrounding it. Such heat as is given up goes into external work only, which is done at the expense of the internal energy of the steam. For instance, let the steam be at the condition indicated at C, that is, at the temperature T v and just dry and saturated. If allowed to expand adiabatically to a lower tem- perature !T 2 the vertical line CD" would indicate adiabatic expan- sion, since any other line as CD' or CD" would have an area under it indicating heat added to or subtracted from the steam, as heat, during the expansion. The external work done, equal to the heat energy represented by the area A BCD is accounted for by the lowering of the quality of the steam, or the condensa- tion of a portion. If the quality of the steam was originally below the saturation point, as at E, then the quality at the lower pressure will have some other value, as shown by the position of F. If the steam were originally superheated, as indicated at H, adiabatic expansion will reduce the amount of superheat as at /, or if continued bring it to saturation as at /', or if still further continued condense part of the steam as indi- cated by the position of /. Problems. (Use standard steam tables, such as Marks and Davis'.) 1. What is the increase in entropy from water at 65 Fahr. to dry, saturated steam at 320 Fahr.? 2. The entropy of evaporation at 360 Fahr. is 1 .0514. What is the latent heat of evaporation for that point? 3. Referring to Fig. 1, under what conditions will steam expanding adiabatically from Tj to T 2 have its quality raised, when will it remain about the same, when will it be lowered? 4. What will be the final quality of steam expanded adiabatically from the dry, saturated point, C, 140 Ibs. gauge pressure to 28 inches of vacuum?/*, V,A ,177 5- Given steam at 170 Ibs. absolute and superheated 100 Fahr., at what pressure will it be dry and saturated when expanded adiabat- ically, at what pressure will its quality, be 90%? (Take the specific heat of superheated steam at 0.6.) 6. Given steam at 160 Ibs. gauge pressure and 98 1/2% quality, what is THE FREE EXPANSION OF STEAM. 7 the quality or superheat at atmospheric pressure which corresponds to the same total heat? 7. Given 1 Ib. of water at 70 Fahr., what will be its condition when 1148 B. T. U.'s have been added and its absolute pressure is 27 Ibs.? tf, : 8. What is the B. T. U. drop per Ib. for adiabatic expansion from" 150 Ibs. gauge pressure and 80 Fahr. superheat to a pressure cor- responding to 28 1/2 inches of vacuum?- j-ll .^ References. THOMAS: "Steam Turbines." JUDE: "The Theory of the Steam Turbine." NEILSON: "The Steam Turbine." MOYER: " Steam Turbines." FRENCH: " Steam Turbines." Art. 3 The Heat-entropy Diagram. 'Prof. Mollier has developed a heat diagram based upon the temperature-entropy diagram, but more convenient to work with. Copies of this chart, for both British and Continental units, are published with a number of the standard works on steam turbines. An excellent one is published with Marks and Davis' Steam Tables, a part of which, on a reduced scale, is reproduced here. 1 In this chart the heat content of the steam at any given condi- tion of pressure and dryness or superheat, appears as the ordi- nate in a right-angled co-ordinate system, instead of an area as in the temperature-entropy diagram, and the entropy appears as the abscissa. Any condition of the steam may be expressed by the position of a point in this plane. The points of equal pressures being connected, there results the series of curves of equal pressures running upward from left to right. Similarly, by connecting points of constant dryness and constant super- heat we have curves of constant quality and superheat. A vertical line from an initial condition to the pressure line of the final condition shows the heat drop due to adiabatic expansion where there is no change in entropy, and the heat change may be read off directly on the scale of the chart. Horizontal lines indicate changes at constant heat or changes of condition, and therefore give the change in quality for moist 1 The Marks and Davis Diagrams I, Total Heat-entropy Diagram and II, Tota Heat-pressure Diagram, may be purchased at $0.40 net, from the publishers, Longmans, Green, and Co., 4th Ave. and 30th St., New York City. 8 STEAM TURBINES. steam, or in temperature or superheat for superheated steam. As the theoretical velocity developed in a nozzle is a function of certain constants and the heat drop only (Eq. 6), an additional scale may be plotted from which the velocity corresponding to the heat change may be read off at once. A third scale of the foot-pounds available per pound of steam is also added, in which the values are 777.5 times the B. T. U. drops. Problems. 1. What is final condition of dry steam expanded adiabatically from 150 Ibs. gauge to 1 Ib. absolute? J'l 7 2. What will be the quality, of a Ib. of steam at atmospheric pressure which has the same total heat as steam at 50 Ibs. absolute and 95% quality? . > > <^ 24 STEAM TURBINES. pounds pressure and . 985 dryness, drop a perpendicular to B on the line of the given final pressure, 1 pound. The distance AB, laid off on the heat scale gives 319 B. T. U., the heat drop for adiabatic expansion, corresponding with the calculated value in line 6 of the table. But 15% of the heat drop is lost in friction. Laying off Bb' = . 15 of AB, we have Ab' as the actual heat drop. Projecting b' horizontally until it crosses the 1 pound pressure 1" 2" 3' FIG. 11. Curves for expanding nozzle. line as at b", we have, reading from the quality lines, .813 as the final dryness, which agrees with the calculated value line 12. To ascertain the condition at any intermediate pressures, with B as a center and Bb' as a radius, draw the arc b'c. Draw a tangent to the are b'c through A. From the points d, e, etc., where AB cuts the pressure lines in question, draw arcs tangent to Ac. Project the points d' , e' , where the arcs cut AB, hori- zontally over to the points d", e", on their respective pressure UTILIZATION OF KINETIC ENERGY IN STEAM. 25 lines. These points referred to the quality lines give the dryness. A "quality curve/' Ab", may be drawn through d", e", etc., from which the dryness for any intermediate pressure may be read off at once. The qualities thus found will agree with the values as calculated within limits of error justified by the nature of the problem. It may be noted here that this qual- ity curve may be laid off almost as readily for a varying friction Fio. 12. Condition curve for expanding nozzle. loss as for a constant one. This is of value in calculations for multistage machines where the friction losses increase as the steam expands. Problems. I. Design an expanding nozzle for the following conditions: Initial steam pressure = 120 Ibs. gauge; superheat =40 Fahr. ; atmospheric discharge pressure ; energy loss, 16% ; weight of steam delivered = .11 Ib. per sec. ; length of nozzle =3 ins. 26 STEAM TURBINES. 2. Design a nozzle for same conditions as Prob. 1, except that dis- charge pressure is 1 .5 Ibs. absolute and length =4 ins. 3. Calculate the H. P. available in jet at throat of nozzle, Prob. 1, and also after expansion to atmospheric discharge pressure. 4. Calculate the H. P. available at throat and exit of nozzle in Prob. 2. 5. If the expanding portion of the nozzle in Prob. 2 is 4 inches long, what are the pressure and velocity at points 1/2 inch and 2 1/2 inches from the throat? References. STODOLA: "The Steam Turbine." THOMAS: "Steam Turbines." MOYER: "Steam Turbines." Art. 8. BLADE FORMS AND BLADE VELOCITIES. The velocities which enter into the problem of turbine design are: The absolute velocity of the jet as it enters the blades. The velocity of the moving blades. The relative velocity of the jet and the blades. The absolute velocity of the jet as it leaves the blades. The ideal condition in all turbines, whether steam or water driven, is that the driv- ing fluid " enter without shock and leave without velocity." Let V, above, be the abso- lute velocity of the entering steam, and u the blade veloc- ity. W, the relative velocity, is found, both in magnitude and direction, from the ve- FIG. 13. Velocity diagram. , . . , T7 . locity triangle V, u, and W. The angle /? determines the direction of the back of the blade on the entering side. The curved surface of the blade de- flects the jet, discharging it backward with a velocity W 1} equal to W, less a certain loss due to friction. This friction can not be determined theoretically, but may be assumed from results of experiments and experience in turbine design. These indicate that W l varies from . 8 to . 95 of W, according to conditions, . 92 being a fair average value for the single stage turbine. The angle UTILIZATION OF KINETIC ENERGY IN STEAM. 27 /?' is ordinarily made equal to /? for convenience in manufacture and to reduce end thrust. Having W in magnitude and direc- tion, the final absolute velocity 7 t is readily obtained by com- bining it with the bucket velocity u. From equation (2) the energy yz _y 2 per pound of steam developed in buckets is and the theo- 2 # retical efficiency is Fig. 14 shows a convenient method of determining the form of the bucket, where entrance and exit angles are equal. Lay off u, the bucket velocity, horizontally. Draw V to the same scale, making the angle a. In the De Laval wheel this angle is 20. Closing the triangle we have the relative velocity e, f Pitch-^ FIG. 14. Determination of blade sectioas. W in direction and magnitude. Draw ab parallel to u and lay off /?' = /? on the exit side. Take c and e equidistant from ab and the width of the blade apart. Lay off cc' and ee' equal to the pitch of the blades. The face of the bucket may be drawn as the arc of a circle about the intersection, /, of the perpendiculars cf and ef and the center line ab. 1 The back of the blade is formed by two lines parallel to be and be connected by an arc with /as the center. The best width ce and pitch cc' are empirical and determined by experience, as no formulae for them have been developed. Moyer says, "for turbines of less than 100 H. P., the width of the 1 A well known German rule is to make the radius equal to twice the pitch times sine /?, or r =2p sine )9. 28 STEAM TURBINES. blades is often made about I", increasing to about 1.5" in tur- bines of 1000 H. P." "The most efficient blade pitch appears to be between the limits of 1/2" and 1". Between these two values the efficiency of blades made according to conventional designs is practically constant. The usual blade pitches are 5/8, 3/4, and 7/8 inch." It is more important that the backs of thq vanes should be parallel to the entering steam than the faces, for steam striking the backs is deflected outward, retarding the wheel and causing loss. If it strikes the face the energy of deflection goes into useful work in the wheel. In the spacing of the blades there must be enough free area to provide for the unobstructed passage of the steam at the large volume corresponding to the exhaust side of the blade, and it should be as near uniform in section as con- ditions will permit. It must be borne in mind, too, that the opening of a round inclined nozzle is an ellipse, and the greater part of the steam is discharged through these openings near the middle of the nozzle. Problems. 1. If steam issues from a nozzle with a velocity of 3400 ft. per sec. at an angle a. =20 and the blade velocity u =1200 ft. per sec., what are the relative velocity W and blade angle /?? 2. Assume blade angles /? and /? t as equal, what is the final absolute exit velocity neglecting steam friction in the blading? 3. What is the energy per Ib. of steam available for the conditions of Probs. 1 and 2? What is the efficiency? 4. Given 7=3400 ft. per sec. a =20, w=1200 ft. per sec., and an energy loss of 20% in passage through the blading, what is the final velocity of exit F x ? 5. Determine the energy absorbed by the wheel per Ib. of steam and the efficiency for conditions of Prob. 4. 6. Draw a cross section of the blading for conditions of Prob. 4. What length of blade could be used with a nozzle .8 inch exit diameter? 7. How many r. p. m. should a 42-inch diam. wheel run for a nozzle velocity of 2300 ft. and a =22 1/2 and /? =27. 8. With a blade velocity u =400 ft. per sec., a nozzle velocity of 1720 ft. per sec., and a =42, what will be the absolute angle of exit of the steam for a symmetrical blade? 9. Report on article on "Principles of Steam Turbine Buckets." Power, March 17, 1908, p. 391 UTILIZATION OF KINETIC ENERGY IN STEAM. 29 References. STODOLA: "The Steam Turbine." THOMAS: " Steam Turbines." MOYER: " Steam Turbines." JUDE: "Theory of the Steam Turbine." NEILSON: " The Steam Turbine." Art. 9. Conditions of Maximum Efficiency. Let Fig. 15 be the velocity diagram for the entrance and exit velocities, where the relative velocity W is supposed to remain constant. FIG. 15. Velocity diagram. From trigonometry FIG. 16. Velocity diagram. V 1 2 = W 2 + u 2 -2uW cos V 2 - VS = + 2u W (cos p V 1 2 = V 2 -2u W (cos p + cos subtracting, (17) Y 2 V 2 The efficiency of the bucket, - , Eq. (16), will be highest when Vi 2 is least. V l 2 will be least, V 2 being constant, for the maximum value of the second term in Eq. (17). * As in hydraulic problems the ideal condition is given by complete reversal where cos /? and cos /? t are each positive and unity. FIRST CASE. Where p, & and V are fixed. Since the above quantities are fixed it is clear from Eq. (17) that YI is least where uW is greatest. In Fig. 16 the area of abxtrn, the triangle abc = bc.%ac sin p = lab.cm, therefore = 30 STEAM TURBINES. oc.oc = W.u. But ab = F, and sin /? are constant, therefore, comparing with Eq. (17), F t 2 is least when cm is maximum. The length ab and the angle acb are fixed, therefore the locus of the possible positions of the intersection c with reference to ab is an arc acb and cm is maximum when ac = cb, or when u W. We have then F U= ^3 (18) and (19) Taking the relative velocities TF and W \ as equal, the efficiency, F 2 - F^ 2w. TF(cos + cos ft) ^ becomes - ^L_ If, in addition, /? and ft are also equal FIG. 17. Velocity diagram. Since W cos p= V cos a-u, the efficiency in terms of a is (20) SECOND CASE. Where V and a are fixed, and ft and ft are equal but not determined. Eq. (17) above becomes V l 2 = V 2 -4uWcos^, V, 2 ^ V 2 -4u.cd. UTILIZATION OF KINETIC ENERGY IN STEAM. 31 V l is least when u.cd is maximum, bd, the sum of u and cd is constant and = V cos a. Therefore, u.cd is maximum for u = cd, or V cos (21) and tan ad ad ad 1 = = = 2 = 2 tan a. dc bd bd Also, since W = W l} ec = ac and the triangle cef can easily be shown to be similar and equal to the triangle acd. Therefore cf= ad or V l =V sin a. V 2 -V 2 V 2 -V 2 sin 2 a The efficiency 2 becomes then = V 2 (I -sin 2 a) 2 or simply =cos 2 a (22) FIG. 18. Velocity diagram. THIRD CASE. When V, a, and ft are fixed. The triangle abc is fixed. W and u are therefore determinate. Cos /? u then, is the only variable on the right hand side of Eq. (17), and V l is least for ft = 0. In the foregoing, losses in the relative velocity W, due to fric- tion, have been neglected. If these be considered, V 2 = W 2 + u 2 + 2uW cos p as before, but Vf becomes V 1 2 =W 1 2uW l cos ft, where W t =yW. 32 STEAM TURBINES. The energy developed is 72 _ V 2 w*-WS + 2u(W c The general expression for the efficiency V 2 -V 1 2 _W 2 -W 1 2 + 2u (W cos l cos ft) y* F 2 For Case II where /3 = ft it becomes TF 2 - W* + 2w. cos p (W + JFj) (23) It frequently happens that the best value of u must be sacri- ficed for safety or for cheapness of construction, or there may be offsetting gains in mechanical efficiency by adopting a FIG. 19. Velocity diagram. slower speed. Take as an example, a De Laval nozzle, coming under Case II, delivering steam at 3680 feet per second at an angle, a = 20. Then u = y- = - . 94 - 1730 feet per sec- ond. This would give a value for the relative velocity w = X/3680 2 + 1730 2 - 2 co8~26~3680 . 1730 = 2130 feet, and a blade angle, p = tan i 2 tan 20=- tan' 1 . 728 = 36. Experi- ence has shown that the blade velocity should not exceed 1400 feet per second. If the actual speed be 1250 feet per second the value for the relative velocity is UTILIZATION OF KINETIC ENERGY IN STEAM. 33 W = A/3680 2 + 1250 2 - 2 cos 20 .3680 .1250 = 2540 and since sin /3 = ~ -_ - - , the blade angle = 3680 sin 20 FIG. 20. Velocity diagram. In the De Laval turbine a nozzle angle of 20 has been estab- lished for all sizes. The angles /? and /3 X are equal, varying from 30 or 32 up to 36. /? and /? t can not be equal to zero, unless the plane of the rever- sal of the steam makes an angle with the line of motion. This is done in some turbines, as the Stumpf and Terry machines. FIG. 21. Velocity diagram tangential type. The introduction of this new angle partially offsets the advan- tage of complete reversal, as will be seen in considering Fig. 21. If V and u be represented by ab and ac the velocity, u, can be resolved into two components, u cos a = ad, and u sin a = cd. The component cd will be impressed on the entering jet giving it the direction ab' so that W is the relative velocity of entrance, not 34 STEAM TURBINES. cb or ef. The complete reversal of the jet gives a relative exit velocity of W, equal to W, and an absolute exit velocity of V 1} as shown. Neilson shows that V 1 is minimum for u = -- . For this F 2 (l -cos'a.) value 7, 2 = 3 cos 2 a+l The third of the principal forms of buckets used in single- expansion turbines is the Pelton type, which lends itself to use FIG. 22. Pelton type of blade. with steam as well as water. A complete reversal of the steam y jet and a value of u = - would give a theoretical efficiency of 100%. It is necessary, of course, to give the steam some exit velocity to clear the wheel. This velocity, V 1} can be determined graph- ically as in Fig. 22, or analytically by the equation V* = W 2 + u*-'2 W ucos ft. Since W=V-u = (V-u) 2 + u 2 -2u (V-u) cos ft, = V 2 -2u (V-u) (1+cos ft). The theoretical efficiency as before = _2u W (1+cos /9) (25) UTILIZATION OF KINETIC ENERGY IN STEAM. 35 Problems. 1. Given V =2420 ft. per sec., /?. =45 and 0, =40, what is the best value for the blade velocity and for the nozzle angle? Neglect steam friction. 2. What is the blade efficiency for the conditions of Prob. 1. 3. Given a steam velocity of 2250 ft. per sec. and a nozzle angle of 22, what is the best blade velocity, neglecting steam friction? What is the efficiency of the blading? 4. With symmetrical blading what nozzle angle should be used when the steam velocity is 2700 ft. per sec. and the blade velocity 1430 ft. to get an efficiency of 87%? 5. What angles a and /? should be used for a DeLaval nozzle and blade to give 88 1/2% efficiency with an entrance velocity of 3550 ft. per sec.? What blade velocity does this call for? 6. What is the power developed per Ib. of steam for the conditions of Prob. 5? 7. Find the power developed in a DeLaval blade with V =3550 as in Prob. 5., u limited to 1240 ft. per sec., a =20. 8. Find the power developed under the conditions of Prob. 7 if in addition there is a velocity loss of 10% in passage through the blades; what is the blade efficiency under these conditions? 9. Show that for maximum efficiency, V and u are at right angles. 10. Prove that "dilution of the jet" or lowering the velocity of the steam by having mixed with a heavier fluid, as water, is a less efficient way of reducing speeds than by lowering the velocity. 1 References. STODOLA: "The Steam Turbine." NEILSON: "The Steam Turbine." THOMAS: "Steam Turbines." JUDE: "Theory of the Steam Turbine." MOYER: "Steam Turbines." RATEAU: "Method of Calculating Steam Turbines," London Engi- neering, Dec. 10, 1909, p. 804. BRILING: "Energy Losses in Turbine Buckets," Zeitschrift d. v. Deut. Ing., Feb. 12, 1910. Art. 10 The Trajectory Of The Steam. The absolute direction of the steam at any point may be deter- mined by combining the velocity u of the vane with that of the steam as it moves relatively to the vane. Neglecting frictional losses, this would have a magnitude W, and a direction tangent "See French, Steam Turbines, p. 63. 36 STEAM TURBINES. to the vane surface at the point in question. By taking a suc- cession of points the absolute directions chould be found and the envelope of these would give an approximate trajectory. The error in this method is cumulative, however, and a better way is as follows: Let W be the relative velocity at entrance, and u the vane velocity as heretofore. If the inner vane surface be semi- circular the time required to move from A to B will be . The distance forward covered in the time will be u W Lay- ing this off along the path of B we have the end of the trajectory. For any in- termediate point, as C, the time required arc AC will be and the distance forward arc AC covered by C will be CC t = ==- u. The path will depend on the size and shape of the vane and the relative velocities u and W. If, as is usual, the vane surface is only a portion of a semicircle, the method is not altered and the distance CC l for any point C is laid distance AC off along its path equal to u, and so on for any number of points needed to determine the curve. If, however, the relative velocity W be reduced by friction to W l during the passage through the vane it will lengthen out the trajectory. Just at what rate the retardation occurs would be difficult to determine, but it may be assumed to be uniform. If W and W l be the relative velocities at entrance and exit the loss from A to B=WW 1 , and assuming this loss as uniform the AC velocity at C = W- t (W ' WJ and the time required to move Fio. 23. Trajectory of the steam. from A to C AB AC ave.rel. vel. AC AC UTILIZATION OF KINETIC ENERGY IN STEAM. 37 The point corresponding to C on the trajectory would then lie on its path a distance forward AC A succession of points, as before, will determine the complete trajectory. Below are given the trajectories for a blade of the De Laval form, for values of = . 35, . 4, . 45, . 5, . 55, and . 6, and a fric- tional loss of . 08. FIG. 24. Trajectory curves. The cross section of the passage between the vanes should be as nearly uniform as possible. Where the steam fills the passage and the backs of the blades differ in form from the faces, there maybe a marked difference in the trajectories of steam particles, which means eddies and inevitable loss of energy. Fig. 25 shows a difference in the paths of particles mov- ing along the fronts and backs of the blades, which is to be 38 STEAM TURBINES. guarded against as far as possible in designing turbine blades. In impulse turbines the expansion occurs in the nozzles or fixed vanes and there is no drop in pressure as the steam goes through the buckets. This renders the determination 1 of the trajectory easy, but in reaction turbines the moving passages are themselves nozzles, and the steam velocity increases sharply in passing through them. The determination of the trajectory under these conditions becomes very involved. Approximations have been Trajectory for Trajectory for backatoc. face adc. FIG. 25. Trajectory at front and back of blade. made by assuming that the relative velocity increases in some regular manner. Having made some such assumptions trajec- tories may be plotted by the method outlined, but as they are at best guess work and of necessity differ throughout the entire length of the rotor they are of little value. The methods out- lined above of course apply to the Pelton type of bucket as well as to the side entrance type. Problems. 1. With a steam velocity of 2200 and a blade velocity of 800 ft. per sec. and semicircular blade as in Fig. 23, how far will the blade travel during the passage of the steam, assuming no friction loss? 2. Given nozzle and blade velocities of 3300 ft. and 1200 ft. per sec., a nozzle angle of 20, a blade pitch of 5/8 in. and a width of 1 in., how far has the blade moved when the steam crosses the center line, and how far at exit? UTILIZATION OF KINETIC ENERGY IN STEAM. 39 3. Assuming the W and blade face of Prob. 2, plot the trajectory for u = .3, .4, and .5 of W. 4. Given V =3400 ft. and u =1220 ft. per sec., a =20 and a pitch of 3/4 in., draw a section of the blading and plot the trajectories of the steam moving along the face and back of the blade (see Fig. 25). References. JUDE: "Theory of the Steam Turbine." STODOLA: "The Steam Turbine." CHAPTER III. CALCULATIONS OF TURBINE BLADING. Art. ii. Single-stage Impulse Turbines. The simplest form of successful turbine is the single-stage impulse turbine. In this all the expansion occurs in one set of nozzles, where the heat energy availablejs converted into kinetic energy and directed against a single row of moving vanes. As the steam velocity varies with the heat drop and the vane velocity with that of the incoming steam, the speed of the rotat- ing vanes in the case of single expansion is very high, from 1,000 to 1,400 feet per second, or about half the velocity of a modern high powered rifle bullet. De Laval frankly accepted these speeds and developed a machine capable of withstanding the strains produced. Some idea of the problems involved may be had from the fact that in the 300-horse-power turbines the blades, weighing only 1/28 of a pound each, but running with a velocity of 1380 feet per second on a 15-inch radius, develop a centrifugal force of 3/4 of a ton. It called for engineering ability of a high order, and greatest care in the details of design and manufac- ture, for not only was a built up rotor required capable of running from 10,000 to 30,000 R.P.M., according to the size, but the velocities developed had to be reduced through gearing to ones which could be used. An efficient and practical machine has been developed, however, which runs with gear velocities of 100 feet per second with but little wear, no vibration, and high efficiency. The relation between the pressures and velocities during the passage of the steam through an impulse turbine of the De Laval type is shown schematically in Fig. 26. The boiler pressure drops in a single expansion to the exhaust pressure before the steam enters the vanes. The wheel therefore has equal pressure on both sides. The velocity V of the steam is maximum as it enters the vanes and is nearly all absorbed in the passage through 40 CALCULATIONS OF TURBINE BLADING. 41 them. The steam enters the exhaust chamber with the low velocity v, sufficient only to clear the wheel. As an example of calculation for the blading of a turbine of this type let the jsteam conditions be the same as for the example on page 22, namely, Initial steam pressure, Initial dryness Final steam pressure 150 pounds gauge 981/2 % 1 pound absolute Boiler Pressure Velocity in Steam Pipe Exhaust chamber Condenser Pressure" Fia. 26. Scheme of single stage impulse turbine. In addition let the Output Revolutions per minute of wheel = Blade velocity be limited to Number of nozzles Angle of nozzles 110 H. P. 13000 1250 feet per second 6 20 Angle of entrance and exit of blades to be equal. As already pointed out in Art. 8 the design of the blading, and in fact of the whole turbine, is materially affected by the frictional losses in the nozzles and blades. In determining the proportions for a given case assumptions must be made for these, based on experience. Stodola gives the following values for the various losses. 1 1 Steam Turbines, 2nd ed., p. 224. 42 STEAM TURBINES. Losses in the nozzle Losses in the blades Losses at exit Total loss = 15% of the available energy = 21% of the available energy = 4.6% of the- available energy = 40.6% of the available energy From the given conditions of steam and exhaust the theoretical heat drop available is 319 B. T. U. But assuming a nozzle loss of 15%, only 319X- 85 = 271 B. T. U. will be transformed into kinetic energy. From Eq. (6) we have 7 = 223.8^/271 = 3680 feet per second. (See problem in Art. 7.) 3680 X 94 From Eq. (21) u= =1730 feet per second for maximum efficiency. For mechanical reasons u will be limited to 1250 feet per second as given in the initial conditions. The FIG. 27. Velocity diagram. 21 energy loss in the blades is 21% of the total, or '- =24.7% of .85 that delivered to the blades by the nozzles. If 2540 feet per second be the relative velocity at entrance, as shown by the velocity diagram, Fig. 27, the relative velocity at exit will be such that the kinetic energy is 24 . 7 % less, or 2540 2 W* x.753= - 1 -. Hence W l = A/2540 2 X .753 = 2200 feet per second. The diagram shows that the blade angles are 30 and the final velocity F 1 = 1280 feet per second. Comparison of Fig. 27 with CALCULATIONS OF TURBINE BLADING. 43 Figs. 19 and 20 shows how friction and mechanical limitations modify the theoretical conditions. Comparing reported tests of similar turbines, 18 pounds per B. H. P. hour may be assumed for the. steam rate on which to .base 1 8 V 1 1 the nozzle areas. Then = .55 pounds of steam must be oOUU delivered per second by the six nozzles or . 092 pounds for each nozzle. The proportions of a nozzle for these conditions have already been worked out in Art. 7 and are shown in Fig. 11. 13000 R. P. M. have been assumed for the rotor shaft, and 1250 feet per second for u, therefore - - = 5 . 75 feet is the cir- loUUU cumference of the blade circle, which corresponds to a pitch diameter of 22 inches. From Art. 8 the pitch may be assumed as approximately 5/8 inch, which gives us 110 blades. Since the exit diameter of the nozzle, Fig. 11, is .99 inches we may make the blades, say, 11/8 inches high. Problems. -A i. Given the steam conditions of Prob. 1., Art. 7, namely: Steam pressure =120 Ibs. gauge. Superheat = 40 F. Exhaust at atmospheric pressure Output = 100 K. W. R. P. M. of wheel =12000. Blade velocity = 1200 ft. per sec. Number of nozzles =8. Angle of nozzles =20. Blades symmetrical. Design nozzles and blading for a single stage impulse turbine assuming a steam rate of 50 Ibs. of steam per K. W. hr. 2. Design nozzles and blading for a condensing turbine, conditions as in Prob. 1, except exhaust pressure =1.5 Ibs. absolute. What changes of design are involved ? Assume a steam rate of 26 Ibs. per K. W. hr. 3. Design nozzles and blading for a single stage impulse turbine. Steam pressure 15 Ibs. absolute; exhaust pressure 1 Ib. absolute; quality of steam at admission = 88%; blade velocity = 1200 ft. per sec. Take steam rate at 42 Ibs. per K. W. hr. 4. Report on the construction and efficiency of the reduction gearing of the DeLaval steam turbine. 44 STEAM TURBINES. References. STODOLA: " The Steam Turbine." NEILSON: "The Steam Turbine." THOMAS: " Steam Turbines." LEA AND MEDEN: Paper, "DeLa- vol. xxv. Steam Turbines," A. S. M. E. Art. 12. Multi-stage Impulse Turbines. As seen in the last article a single-stage impulse turbine must run at a very high velocity and be geared down. There are two ways by which the velocity of an impulse tur- bine may be reduced, by subdividing the heat drop in a series of FIG. 28. Scheme of impulse turbine with single pressure stage and several velocity stages. expansions, and by having the steam impinge on two or more rows of buckets after expansion. The shaft velocity can thus be reduced to a point where the turbine can be directly connected to a generator, blower, or centrifugal pump. Such turbines are more complex and theoretically less efficient than the single- stage machine. This is offset, however, by the advantages of direct connection and lower speed. The De Laval type of single- CALCULATIONS OF TURBINE BLADING. 45 stage turbine finds its limit at about 500 H. P. Beyond this the strains become too great and the problem of speed reduction too troublesome. In the multi-stage type, however, there seems to be almost no limitation of the power which can be developed by a single machine. Curtis turbines of this type are in use developing 20,000 K. W., and larger units might be used were they commercially desirable. The field of the multi-stage turbine is much wider than that of the single-stage, and it constitutes the second of the main divisions into which turbines may be separated. Fig. 28 illustrates the action of the original Curtis and the Riedler turbines. Complete expansion occurs as before in a Fia. 29. Scheme of Curtis turbine, multi-pressure, multi-velocity stage. single set of nozzles, and the steam enters the first row of blades at high velocity as in the previous case. The steam leaves this row, however, with considerable velocity, is redirected by a set of stationary guide vanes and impinges on the next row, and so on to the exhaust chamber. The drops in steam velocity occur in the passage through the moving blades. The stationary vanes are merely guides, and the velocity during the passage through them is substantially constant, as shown. The increase of length in the successive rows of stationary blades is to allow for the decrease in velocity, not for increase in volume. Theo- retically the exhaust pressure with its corresponding volume is reached at the nozzle, and all the rows of blades are driven by 46 STEAM TURBINES. kinetic energy, revolve in a vacuum with the least possible resist- ance, and the wheels are balanced for pressures. This is a single- pressure, multi-velocity stage type and is still used on the smaller Curtis turbines up to about 35 K.W. As now built the Curtis turbines have the pressure drop divided into two or more " pressure stages." Steam is partially expanded in a set of nozzles to some lower pressure and passes through a num- ber of rows of moving blades as at a, Fig. 29, constituting one stage where its kinetic energy is absorbed. It is then expanded again and goes through another series of blades, or second stage, b. The number of stages and of rows in each stage varies with the FIG. 30. Scheme of multicellular turbine, multi-pressure, single- velocity stage. conditions. As before, the wheels are balanced, and there is no end thrust. The pressure throughout each stage is constant, the drops occurring in the expanding nozzles at the begin- ning of each stage. This may be described as a multi-pressure, multi-velocity stage machine. This idea of breaking up the pressure drop is carried to the limit in the "multi-cellular" turbines, Fig. 30, such as the Rateau, Zoelly, Kerr, and Wilkinson machines. Here a set of ex- pansion nozzles after each row of blades replaces the fixed guide vanes, and accelerates the steam to a certain velocity, which is absorbed in the following running wheel. The multi- cellular turbine is therefore a multi-pressure, single-velocity tage machine. A turbine has been patented 1 which is similar to Fig. 30 except that, while the velocity drop is about as usual, 'U. S. Patent, J. F. M. Patitz, Oct. 4, 1910. CALCULATIONS OF TURBINE BLADING. 47 both velocities V andFj are high i.e., the broken line in Fig. 30 is raised as a whole. The energy available in a single stage varies with V 2 V ^, and is much greater for a given drop when both velocities are high. It is hoped by this means to reduce greatly the number of stages, but this will be partially offset by greater steam friction. See Probs. 3 and 4, p. 3. In all of these forms there will be noticed the characteristic which defines impulse turbines, that there is no fall of pressure in the moving blades. Expansion occurs only in fixed passages or nozzles, and the moving blades are free from end thrust. FIG. 31. Velocity diagram for two stages. The determination of the velocities in the multi- velocity stage turbine is an extension of the method already outlined. The final velocity, V lt at exit from the first row, Fig. 31, is deflected and becomes the initial velocity for the second row of blades, and so on for the whole number of rows. When the rows of buckets have the same mean diameter, as is usual, u has the same value for each row. The diagram may be arranged conveniently and more compactly by extending the line of the vane velocity as in Fig. 32, and laying off a succession of lengths, each equal to u, and connecting each with the point a, as shown. With an initial velocity of 2430 feet per second, a blade velocity of 400 feet per second, and a nozzle angle of 22 1/2, the exit velocity from the second row of moving blades is 1130 feet per second, and the 48 STEAM TURBINES. blade angle 42. Proper allowance for friction, however, materially affects the design and should, as before, be taken into account. Fig. 33 shows how this may be done, and gives the diagram where friction in the entering nozzle has reduced the velocity, V, to 2320 feet per second, and in each succeeding row a friction loss of 10% is assumed. Comparison of the blade sec- tion as determined by the two diagrams shows a marked differ- ence. If the radii of the successive rows on the shaft are not FIG. 32. Velocity diagram for Curtis turbine, no allowance for steam friction. equal, the length of u should be made to increase directly as the diameter. In general, the friction loss is not constant through the successive stages, but increases from 3% or 4% to perhaps 10% with increasing presence of water in the steam. Proper values can be assigned only from experience. The method of finding the velocities in Fig. 33 is applicable for an increasing friction loss as well as for a constant one. As an illustration of calculations for blading in a multi-stage impulse turbine, let the following conditions be given: CALCULATIONS OF TURBINE BLADING. 49 Initial steam pressure Initial superheat Final pressure Assume a blade velocity u Assume nozzle angles = 175 pounds absolute = 140Fahr. = 1 pound absolute = 400 feet per second = 22 1/2 Assume 10% energy loss in the nozzles and 10% velocity loss in each row of vanes. Let the turbine be of the Curtis type with three stages and two rows of moving blades in each stage. FIG. 33. Velocity diagram for Curtis turbine with allowance for steam "friction. For convenience part of the heat chart is reproduced in Fig. 34. The initial condition of the steam is given by the intersection of the 175-pound pressure line and the 140 superheat line. A vertical line dropped to the 1 pound line gives 357 B. T. U. as the total heat drop for pure adiabatic expansion. Dividing this drop as nearly equally as possible at AC, CD, and DB, we have the pressure ranges for the first three stages. First, 175 pounds to 45 pounds absolute. Second, 45 pounds to 8 pounds absolute. Third, 8 pounds to 1 pound absolute. 50 STEAM TURBINES. Deducting CC' = 10% of AC for the assumed heat loss in the ex- panding nozzle, we have AC', which, referred to the velocity scale, gives 2320 feet per second initial velocity. As u and a are given, we may draw the velocity diagram, allowing for the 10% velocity loss in each row. This is given in Fig. 33, where FIG. 34. Condition curve, Curtis turbine. the final velocity is found to be 673 feet per second. The blade sections thus determined are given in the same figure. From Eq. (3.) the force per pound of steam acting in the line W of motion, due to entrance, is V cos a and that due to the . TP, ' exit is - F/ cos /?'. Remembering that W here is unity the total force acting on the first row per pound of steam is CALCULATIONS OF TURBINE BLADING. 51 - and the work done is this expression multi- plied by the blade velocity, u, or the Work done in the first row per second. per pound of steam = U - (Vcosa-Vicosp). (26) 9 A similar expression gives the work done in the second row of moving blades. The velocities required may be scaled from the velocity diagrams. The actual energy per pound of steam which has gone into work during passage through the stage is thus found to be 56540 foot-pounds or in its heat equivalent, 72.8 B.T.U. This is laid off on the heat drop at AE in Fig. 34 and the remainder EC, in this case about 40% of AC, is still in the steam as heat energy. Since the steam is now at a pressure of 45 pounds absolute, and has a total heat corresponding to the ordinate of E, the con- dition of the steam at the beginning of the second stage is found by projecting E horizontally until it meets the 45 pound line as at F. From the position of F it is seen that the steam, instead of being 98% quality, is superheated about GO 9 Fahr. Dropping a vertical from F to the 8 pound line shows an adiabatic heat drop of 126 B. T. U., which, with 10% allowance for losses, gives a velocity of 2380 feet per second. Laying out the diagram as before, we have an exit velocity of 710 feet for the second stage. The work for this stage is found, by the previous method, to be 59,150 foot-pounds per pound of steam, and the corresponding actual heat drop, 76 B. T. U. Applying this to the heat chart as before, we find that the steam enters the. third or last stage with a quality of .986. The adiabatic heat drop for the last stage is 129 B. T. U., and the velocity, allowing far losses, is 2400 feet per second. The energy developed in this stage is 59,600 foot- pounds per pound, and the corresponding heat drop, 76 . 7 B. T. U. From this last we can obtain the final condition of the steam at /, where it is found to have a dryness of .947%. Adding the work of the three stages, we have First stage ................. ..... 56,540 Second stage .................... 59,150 Third stage ..................... 59,600 Total work per pound of steam = 175,290 52 STEAM TURBINES. CALCULATIONS OF TURBINE BLADING. 53 1,980,000 The theoretical steam rate = = 11.3 pounds per 175,290 H. P. hour. Allowing 8% for mechanical friction, windage, etc., 11 3 this becomes ' =12.3 pounds. . 92 If the generator efficiency be taken at 95% we have a steam rate of 13 pounds per E. H. P. hour, or 17.5 pounds per K. W. hour, on the basis of a power factor of unity. As in Fig. 12, a quality curve may be drawn through the points A, F, H, and I, which will give the condition of the steam for any point during the expansion. With this determined, and with the H. P. to be developed, and the steam rate as found above, we have the basis for the calculation of the lengths of the blades. Problems. 1. Make diagrams and calculations for a Curtiss turbine with 5 pres- sure stages, each of 2 velocity stages. Initial steam pressure = 155 Ibs. abs. Superheat =75 F. Final pressure =28 in. vac. Assumed blade velocity =425 ft. per. sec. Nozzle angle =22. Velocity loss in each nozzle 5%. Velocity loss in each row of blades 10%. Calculate the nozzle area and blading for a 500 K. W. turbine, the mechanical and generator friction being taken at the values given on p. 117. 2. Make diagrams and calculations for a L. P. Curtis turbine, the con- ditions being those of Prob. 1, except that the initial pressure is 15 Ibs. absolute and there are two pressure stages, each with two velocity stages. The nozzle angles to be 25. 3. Report on the Small Curtis Turbine. See paper of G. A. Orrok, Trans. A. S. M. E., vol. xxxi, p. 263. 4. Report on a Comparison of the Rateau or multicellular, and the Curtis Principles of Design. Power, Dec. 20, 1910, p. 2218, and Jan. 3 and 10, 1911, pp. 19 and 64. References. STODOLA: " The Steam Turbine." THOMAS: " Steam Turbines." EMMET: "The Steam Turbine in Modern Engineering," Trans. A. S. M. E., vol. xxv. See also references in the problems. 54 STEAM TURBINES. Art. 13. Single Flow Impulse -reaction Turbine. An impulse turbine as indicated in Art. 12 is one in which expansion occurs only in the fixed passages or nozzles. A reaction turbine is one in which the energy is derived from the reaction of steam issuing from openings or nozzles in the mov- ing wheel. No pure reaction turbine is being built at this time. A number were built in this country many years ago by Avery, but they have long since been abandoned. Theoretically, there is no reason why a pure reaction turbine might not be built, but the velocities involved are so high as to render the type impracti- B oiler Press. I Velocity FIG. 36. Scheme of impulse- reaction turbine. cal. A combination of the impulse and reaction principles has, however, led to one of the largest and most important types of turbines yet developed. Figs. 55 and 67 show sectional elevations of two Parsons or impulse-reaction turbines, both single-flow, i.e., the steam enters at the end and flows axially in one direction. Fig. 36, corresponding to Figs. 23 to 26, illustrates their operation. The general form of the buckets is shown in Fig. 37, in which it is seen that the curve at the entering end of the moving vane acts as in impulse turbines, to absorb the kinetic energy discharged from the fixed nozzles. The long straight portion at the exit end of the moving blades forms a nozzle which restores the velocity just absorbed and adds the power of the reaction of exit to that of the impulse first received. In this type there is an almost uninterrupted lowering of the pressure. If there were CALCULATIONS OF TURBINE BLADING. 55 Moving blades no motion in the rotor there would be a corresponding rise in the steam velocity, as shown in the curved line, Fig. 36. In fact, the long passage between the rotor and the casing is essentially an enlarged nozzle, its sectional area increasing steadily, and con- forming roughly to that which would be given an expanding nozzle. The impulse-reaction turbine allows the lowest velocities which have been attained with high economy. Its construction is much more complicated than the simple De Laval wheel, but it has been successful in a very wide range of service. In the impulse-reaction dia- gram, Fig. 37, the fixed vanes beyond the first row of moving blades receive the steam at a velocity, V 1} using the same notation as on page 29. The steam, in passing through the fixed vanes, expands, and the velocity increases to V. With respect to the following row of moving blades, the steam enters at the velocity W. In passage through the moving blades there is further expansion, and the steam leaves them at a high velocity, W v The total work done Fixed Vanes in the moving blades is : First, the energy 29 per pound of steam produced in the guide blades and expended upon the moving vanes; second, the reaction accompanying the increase of steam velocity in the moving blade from W to W lf which is equal to - per pound of steam used. The guide blades and moving vanes usually have the same form and angles, consequently these two expressions are equal, and the total work is divided equally between impulse and reaction. This type of turbine is subject to end thrust which is balanced by pistons on the rotor as shown in Figs. 55 and 67. Before the vanes can be designed in this type, the heat drop in the various stages must be determined, and for this the follow- ing data and assumptions are required: 56 STEAM TURBINES. 1. The initial steam pressure, with dryness or superheat. 2. The exhaust pressure. 3. The absolute velocity, V, of steam as it leaves guide blades. This varies progressively through the turbine. 4. Friction losses of the steam in passing through the turbine. 5. The exit angles of fixed and moving blades. 6. The peripheral velocity of the blades for the various cylin- ders or steps. The exit angles are made the same for both guide vanes and rotating blades, ordinarily between 20 and 30. The smaller angles than these, while apparently increasing the power, lead to contracted passages and increased friction. On the other hand, with too great angles the power for each stage decreases, requiring too many stages, and increasing the size of the turbine. The peripheral velocity u (item 6 above) at the H. P. stage varies from 100 to 135 feet per second and increases as the size of the rotor increases in the later stages. Table I, from a paper of E. M. Speakman, gives some basis for selecting the vane velocities. Table I. PARSONS TURBINE PRACTICE. Electrical Work. Peripheral Vane Speed Rated Capacity Number of Rows Revolutions Per Minute First Last Expansion Expansion 5000 K. W. 135 330 70 750 3500 K. W. 138 280 75 1200 2500 K. W. 125 300 84 1360 1500 K. W. 125 360 72 1500 1000 K. W. 125 250 80 1800 750 K. W. 125 260 77 2000 500 K. W. 120 285 60 3000 250 K. W. 100 210 72 3000 75 K. W. 100 200 48 4000 CALCULATIONS OF TURBINE BLADING. Marine Work. 57 Type of Vessel Peripheral Vane Speed Mean Ratio of ^ No. of Shafts In H. P. In L. P. High speed mail steamer Intermediate mail steamer Channel steamers 70-80 80-90 90-105 85-100 105-120 110-130 110-130 110-135 120-150 115-135 130-160 160-210 .45-. 5 .47-. 5 .37-. 47 .48-. 52 .47-. 5 .47-. 51 4 3 or 4 3 4 3 or 4 3 or 4 Battleships and cruisers. Small cruisers Torpedo craft The steam velocity (item 3) varies from 2 to 3.5 times the blade velocities. When it would rise above this in the pro- gressive expansion along the rotor, the diameter of the rotor is increased, beginning a new step or drum. The heat drop in any stage (i.e., one row of fixed vanes and one row of moving blades) may be found readily by the following .Example of calculation for "Stage Dropj'-First Stage A Aa=Aa below =V= 200 'per sec. Ab=Ab =n=120' " " In diagram bf =170, by scaling FIG. 38.- Velocity curves and heat drops, single flow impulse-reaction turbine. simple graphical construction, Fig. 38. In the velocity tri- angle formed by V, W, and u, produce u and drop a perpendic- ular aa' upon it from a. With a', the base of this perpendicular, as a center, and Aa f as a radius, describe the arc afp. Drop a 58 STEAM TURBINES. perpendicular from b, cutting the arc at /. The work done in foot- pounds per pound of steam in the stage is W , where bfis y measured to the same velocity scale as u and V r The proof of this is as follows: The separate velocity diagrams in Fig. 33 may be combined as shown, and where the work is divided equally between impulse and reaction the figures will be symmetrical with respect to aa'. The velocities of exit from the fixed and moving blades are W and Vi, and from Eq. (26) the work is done u(V cosa+ViCos /3) u(Aa'+a'o) But bf, being the perpendicular from /to the hypothenuse of the right-angled triangle a, f, p, inscribed in a semicircle, is the mean proportional between the segments ab and bp, therefore bf 2 AbXbp and since bp = ao, = Ab(Aa'-\-a'o) = u(V cos a+V^os $ 9 The heat drop producing the work will be -mfa (-m) If the heat drops in the various stages were equal, the number of stages could be determined at once by dividing the total heat drop from inlet to condenser by the stage drop assumed. The stage drops, however, vary, those nearest the condenser being much greater than the earlier ones. A method of determining these, outlined by Prof. Stodola, will be used in working out a definite case. The conditions will be taken as follows: Normal capacity =2000 K. W. Overload capacity, without by-pass, =2400 K. W. Initial pressure = 175 pounds absolute Superheat = 50 Condenser pressure = l pou nd absolute CALCULATIONS OF TURBINE BLADING. 59 R. P. M. =3,600 Number of drums = 4 Loss due to steam friction 28% Loss due to leakage = 6% Loss due to windage, bearings, etc., = 15% Energy in exhaust = 3% Total losses in addition to steam friction = 24%. Efficiency of generator 94% We must first assume blade velocities. From reference to Table I on page 56 these may be taken as it = 120 feet for first stage, increasing to w = 320 feet at exhaust end. Assume also exit angles for all blades and vanes =24, and a clear area through the blades = 1/3 of the cross section of the opening. Lay off, as in the lower part of Fig. 38, a horizontal base AB to represent the length of the rotor. Let the steam velocity in the first stage be 200 feet per second, and in the final stage be 950 feet, these values being based on experience. The curve of steam velocities must be assumed. It follows approximately a hyperbolic curve except toward the exhaust end, where it rises rapidly. The curve here drawn is about as followed in practice. The velocity w=12Q feet per second, for the first cylinder, is laid off at the left, and the velocity 320 feet, for the last cylinder, is laid off at the right. The most advantageous ratio between v the steam and blade velocities is from 2 to 3 and it should vary u between about the same limits on each drum. It is customary also to make the ratio between the successive drum diameters substantially constant. The blade velocities, on this basis, come to 120, 165, 225, and 320 feet per second. Moyer gives an empirical formula for determining the number of stages: (28) 60 STEAM TURBINES. where C is a constant which varies from 1,500,000 for marine tur- bines to 2,600,000 for electric generator service, and u is the mean blade velocity in feet per second of the cylinder considered. The n found is the number of rows required if the blade speed for the whole turbine were u. As the cylinder considered develops a certain fraction only of the total power, the number of rows required for that cylinder will be that fraction of the n found in the formula. It can be seen from the wide range in the constant that reliance on this formula calls for experience on the part of the designer with the working conditions and their relation to the constant used. It may be used to great advantage by any one, however, as giving a good approximation for the number of blades on each drum or cylinder. Let us assume that the power developed by the various cylinders shall be approximately 1/6, 1/5, 1/4, and 3/8 of the total. Then, taking C at 2,500,000, for 2500000 1st cylinder, ^ = -.1/6=30 stages 2500000 2nd cylinder, n 2 = - -- .1/5 = 18 stages loo 2500000 3rd cylinder, n 3 = 00 ^ .1/4 = 12 stages 22o 2500000 4th cylinder, n t = - ~ . 3 /8 = 6 stages Total 66 stages. The stages are spaced off uniformly on AB and the drum lengths located tentatively as at cd, ef, and gh. The heat drop which would occur in a stage located at any point may be determined by the graphical method of Fig. 38. Having taken a sufficient number of these points, and plotted the corresponding heat drops, we get the broken curve marked "stage drops," the breaks being as shown in dotted lines. The average ordinate of this curve gives the average of the heat drops for all the stages, which, in this case, =3.58 B. T. U. This average drop, divided into the total heat drop from inlet to exhaust, will give the total number of stages. The total heat CALCULATIONS OF TURBINE BLADING. 61 drop, determined from the steam-entropy tables, or from the heat-entropy chart (see Fig. 40) is found to be 338 B. T. U., for pure adiabatic expansion, but there is a loss of energy in the steam of y = 28 % ; therefore 338 X ( 1 - 28) = 338 X . 72 = 243 B. T. U. Actual heat drop per pound of steam. Curve of Net Voll per Ib. of Steam, FIG. 39. Pressure and volume curves, single-flow, impulse-reaction turbine. The number of stages then will be the Total heat drop 243 B. T. U. =68 instead of 66. Average stage drop 3 . 58 B. T. U. The 68 stages may be laid off on the base AB and the cylinders, which have up to this point been divided only tentatively, may be readjusted, as shown in full lines. The division made here is as follows: Heat Drop, as Scaled from Curve First cylinder 32 stages 43 Second cylinder . 18 stages 54 Third cylinder Fourth cylinder 11 stages 7 stages 64 82 Total 68 stages 243 62 STEAM TURBINES. The progressive heat drop through the turbine may be shown in a curve (see Fig. 39) in which the ordinates represent the progressive sum of the heat drops from A. The last ordinate should check with the total net heat drop 243 B. T. U. By the initial conditions there are further losses amounting to 24% of the actual heat drop. Therefore, 243x76% = 185 B. T. U. per pound of steam, are available for power. Then 1980000 - =13.8 pounds of steam per B. H. P. 185X777.5 Since the efficiency of the generator = 94%, 13.8X1.36 = 20 pounds = steam per K. W. hour. . i/4 2000 X 20 =11.1 pounds steam per second for normal load. 3600 The steam rate may rise slightly for overloads, but hardly enough to change the rate inside of 20% overload; we have 20 X 2400 therefore =13.4 pounds steam per second for overload 3600 conditions. Where a by-pass valve is used, as in Fig. 55, the steam rate should be further raised for the blading beyond it. There remains the determination of the steam volume under the given conditions, and of the dimensions of the blades and vanes. The condition of the steam may be taken from Fig. 40, which is taken from the heat chart. A, as before, indicates the initial conditions. The vertical drop to B on the 1 pound line gives a quality of 79.2%. But the 28% energy loss raises the quality of the exhaust to 88.5%. The condition curve gives the quality of the steam for any intermediate pressure. The heat drop for any point, read off from the curve of total drops, Fig. 39, and applied to Fig. 40 by measuring down from A and projecting over to the quality curve AD, gives the pres- sure and condition corresponding to the pressure. Curves of pressure and volume of Fig. 39 may thus be derived. The blade lengths may be derived as follows: From Eq. (14) the area of the steam passage required, in square inches, = weight^fjteam per second X vol. per pound X 144 steam velocity in feet per second. CALCULATIONS OF TURBINE BLADING. 63 and the length of blade required to give the necessary area (assuming the blades as occupying 1/3 of the clear passageway, and the blade angle = 24) is Area X 1.5 1.5 A ~ 7rXmean blade dia. sin 24 FIG. 40. Condition curve, single-flow impulse-reaction turbine. Combining these equations and reducing, . 5 _ 1 7 1 . 5 v AxVd ~~ Vd ' For the normal load, 2000 K.W.,TF=11. 1 pounds (p. 62). For the overload, 2400 K. W.,W = 13.4 pounds (p. 62) . 64 STEAM TURBINES. The mean blade diameter, d, for the various cylinders, is determined from the values of u, Fig. 38, and the initial con- dition of 3,600 R. P. M., or 60 revolutions per second. u = 6QXxXd, from which 120 For first cylinder d = - = .635'= 7 5/8 ins. mean dia. .toy For second cylinder d = -= .875' = 10 1/2 ins. mean dia, loc/ 095 For third cylinder d = = 1. 19' = 14 1/4 ins. mean dia. ic?y 320 For fourth cylinder d = = 1.7' =20 1/2 ins. mean dia. lOc/ For a slower R. P. M. or higher blade velocities the mean blade diameters would of course be correspondingly larger. Taking V from the curve V, Fig. 38, and volumes from Fig. 39, and with the values for w and d above, the lengths of the various blades may be determined. For manufacturing reasons the lengths, instead of increasing progressively along the rotor, are stepped into groups as shown in Fig. 44. Problems. 1. Make diagrams and calculations for the blading of a single flow Parsons turbine, having four drums. Capacity, rated, 1000 K. W. Capacity at opening of by-pass valve, 1300 K. W. Initial pressure, 150 gauge. Superheat, 50 F. Exhaust pressure, 28 1/2 in. vac. R. P. M., 3600. Exit blade angles, 25. Steam velocities, 250 ft. to 1000 ft. Blade velocities, 150 ft. to 350 ft. Losses and generator efficiency, as given on p. 62. 2. Make diagrams and calculations for the blading of a single flow L. P. turbine of Parsons type. Initial steam pressure, 16 Ibs. absolute. Quality, 95%. Exhaust, 28 1/2 vacuum. R. P. M., 3600. Exit blade angles, 28. CALCULATIONS OF TURBINE BLADING. 66 STEAM TURBINES. Assume same steam and blade velocities as for the low pressure stages of Prob. 1. 3. Report on the Avery turbine, interesting as being a pure reaction turbine and probably the first commercial one. See American Machinist, Nov. 9, 1905, p. 631. 4. Report on the article on Blading Calculation, Power, Aug. 9, 1910, p. 1412. 5. Report on article on Construction Details of a Reaction Turbine, Power, May 19, 1908, p. 761. 6. Report on article on the Internal Losses in a Parsons Turbine, by Prof. A. G. Christie. Power, Aug. 24, 1909, p. 299. References. STODOLA: "The Steam Turbine." THOMAS: "Steam Turbines." MOYER: "Steam Turbines." HODGKINSON: Paper, A. S. M. E., vol. xxv. PARSONS: Paper, Inst. Naval Arch. (London), May, 1904. See also references in the following problems. Art. 14. Double Flow Impulse -reaction Turbines. On account of the drop in pressure through the moving blades in the Parsons turbine steam must be admitted around the entire circumference, which leads to very short blades in the high- pressure stages, and for the same reason the clearance over the ends of the blades must be very small. Leakage is further minimized by using only small pressure drops, which increases the number of H. P. stages, lengthens the rotor, and introduces trouble from expansion which in the larger sizes becomes serious. While the heat drop per stage is less at the H. P. end of the rotor than at the L. P. end, the pressure drop is greater. This, coupled with the fact that the end clearances over the short H. P. blades must inevitably give a greater percentage of leakage, makes the H. P. end of a Parsons turbine its least efficient portion. The pure reaction type, where the blades are pressure balanced and admission may be had through a portion only of the blade circle, is more efficient in the high-pressure ranges than the reaction type, but in the later stages where water is present in the steam it is less efficient. These considerations have led to the combination of the two types in one machine, utilizing the advantages of both, Steam CALCULATIONS OF TURBINE BLADING. 67 68 STEAM TURBINES. is admitted through nozzles to a multi-stage impulse wheel and expanded to about atmospheric pressure. The rest of the ex- pansion takes place in fixed and moving blades of impulse- reaction type. The reaction portion is divided in the latest form and steam flows axially in both directions from the impulse wheel. Fig. 42 shows a Westinghouse turbine of this design where the central impulse wheel is clearly seen, with the reaction portions on each side. Fig. 43 is a sectional elevation of the same turbine and shows clearly the course of the steam as it leaves the impulse wheel and is divided into two streams flowing in opposite directions. This arrangement automatically balances [Exhaust h- Exhaust FIG. 43. Section of 10,000-K. W. Westinghouse double-flow turbine. the end thrust of the reaction blades and eliminates the balancing pistons necessary in the single-flow type. Besides a symmetrical casing, with smaller exhaust connections it materially shortens the distance between the bearings. Fig. 44 shows two rotors, both for 2000-K.W. Westinghouse machines, the upper one of the double-flow combined type and the lower the old single-flow type. Double-flow turbines are found most desirable in the larger sizes, but for comparison with the previous problem of a single-flow turbine, let the same conditions be assumed for one of this type. Let the steam be expanded in a set of nozzles from inlet pressure to 25 pounds absolute and pass through two velocity stages of an impulse wheel of the Curtis type as shown in stage a, Fig. 29. Let the remainder of the expansion to 1 pound absolute take place in two sets of Parsons blading as in Fig. 42. Let the CALCULATIONS OF TURBINE BLADING. 69 osses in the impulse element, due to disc and blade friction, leakage, etc., be assumed as 40%, the energy so lost reappearing in the increased quality of the steam. The friction losses in the low-pressure blading will be taken at 28% as before. Fig. 45 gives the expansion diagram for these conditions. The net drop available for the impulse stage is 91.2 B. T. U. and for the low-pressure stages 142.8 B. T. U. It will be seen from last article that the number of stages in the low-pressure FIG. 44. Comparison of single- and double-flow Westinghouse rotors 3600 and 1800 R. P. M.^ame H. P. blading depends on the available heat drop, the steam and blade velocities and the exit angles. Dividing the flow does not affect the number of stages on each side. It merely halves the length of the blades, an advantage in the last few stages where in the single-flow machines the length is excessive. Greater blade angles may be used to advantage in low-pressure stages than in high-pressure ones without unduly increasing the number of stages, and 30 will be used instead of 24 as before. Let the blading be carried on two drums or cylinders running at 225 and 320 feet per second and the steam velocity increase from 400 feet per second to 950. This gives the same conditions as for the last two drums in Fig. 38. If 3/8 of the total work be carried by the impulse wheel, 1/4 70 STEAM TURBINES. by the smaller drums, and 3/8 by the larger ones, the application of formula (27) gives 13 stages and 9 stages for the ^ first and second drums, respectively. Laying these off tentatively and proceeding as in Fig. 35 we find the mean stage drop to be 6.8B. T. U. (Fig. 46). Then = 21 = number of stages. 6.8 FIG. 45. Condition curve, double-flow turbine. Re-adjusting we may take 13 for the first cylinders and 8 for the second. The stage drops are laid off in Fig. 45, those from B to C being on the first cylinder, and those from C to D on the second. The pressures and conditions are found directly and the blade lengths determined as in the last article, which will be those for a single-flow turbine of 21 rows. In the present CALCULATIONS OF TURBINE BLADING. 71 one there will be two complete sets of blades of 21 stages each, each set of blading one-half the length required for the single- flow. Curves of pressures and net volumes per pound of steam are shown in Fig. 46. It must be remembered that in this and the foregoing articles certain constants, velocities and friction coefficients have been Vssumed Curve_ofSteamVeloci Curve of StagePv B C FIG. 46. Velocity curves and heat drops, double-flow turbine. more or less arbitrarily assumed. Accurate values for these can be assumed only after wide experience and from intimate knowledge of working conditions. Such information on the choice of values as is available may be found in the standard works on turbines referred to at the end of this article. Problems. \ / 1. Calculate the blading and draw diagrams for a double-flow turbine similar to Fig. 43. Capacity, 2000 K. W. Initial steam pressure, 160 Ibs. gauge. Superheat, 60. Exhaust pressure, 28 ins. vac. Impulse wheel to have two velocity stages and to carry expansion down to 20 Ibs. absolute, with reheating losses of 40%. 72 STEAM TURBINES. 2. Make diagrams and blade calculations for a turbine for conditions similar to Prob. 1. The first pressure stage to be a two-stage Curtis wheel expanding to 25 Ibs. absolute. The remainder of the expansion to take place in a series of wheels of Rateau or multi- cellular type (Fig. 30) . Consult J. A. Moyer, Steam Turbines, p. 86. 3. Report on the Westinghouse double flow turbine described in Power, June 16, 1908, p. 931. 4. Report on the construction and test of a 10,000-K. W. double-flow, impulse-reaction turbine, described in paper by Naphtaly, Trans. A. S. M. E., vol. xxxii, Dec., 1910. (Shown in Fig. 43.) 5. Report on the Melms-Pfenninger turbine (Curtis-Parsons), described in Lond. Engineering, July 9, 1909, p. 39, and Power, Sept., 1907, p. 643. References. MOYER: " Steam Turbines." STODOLA: "The Steam Turbine." NAPHTALY: "Test of 10000-K. W. Turbine," Trans. A. S. M. E., Dec., 1910. CHAPTER IV. MECHANICAL PROBLEMS. The problems already considered have dealt with the using of steam at the high velocities of turbine practice. These involve problems of a purely mechanical nature, among which are provision for centrifugal strains due to speeds of 3,000 to 30,000 R. P. M., the balancing of the shafts, bearings suitable for such speeds, and close and rapid regulation under varying loads. Art. 15. Centrifugal Strains. A thorough consideration of this subject involves extended and complex analysis. Some only of its phases will be touched on here. As the velocities in the De Laval turbine are much higher than in any other, it is in this turbine that the interesting be- havior of a disc rotating at high velocities is most marked. If a flexible shaft carry a round disc, one side of which is slightly heavier than the other, the center of gravity will lie at one side of the geometrical center, as in Fig. 47 "a". If they are rotated rapidly the unbalanced centrifugal force of the heavy side will deflect the shaft toward the heavy side, and cause the geometrical center of the disc to describe a circle as at "b". This action increases with the speed up to a certain point called the "critical speed," when the vibration becomes momentarily excessive. On further increase of speed they settle down and will run quietly, however much the speed be increased. At this critical speed the axis of rotation shifts from the center of the path of the geometric center, o, to the center of gravity of the wheel, as shown in Fig. 47 "c". It can be shown that the deflection is e v= ^ '-fe 73 74 STEAM TURBINES. where e is the eccentricity of the center of gravity and V c the critical velocity referred to. The greater the velocity V, the smaller y becomes, approaching the value y = e, until the wheel bursts. In the De Laval practice the critical speed is from 1/8 to 1/5 of the normal R. P. M. FIG. 47. Action of the flexible shaft. For cylindrical rings where the radial thickness is small, the tensional stress due to the centrifugal force is equal to 12 w (29) where S = stress in pounds per square inch. w = weight per cubic inch of material. v = linear velocity of ring in feet per second MECHANICAL PROBLEMS. 75 Calculation of the stresses in discs rotating at very high speeds is difficult. They vary with the weight of the material and the square of the speed. Fortunately the materials available have been so improved that a factor of safety can be used, sufficient FIG. 48. Large and small De Laval wheels. to cover any uncertainty as to the centrifugal strains. Two facts are shown clearly by both theory and experiment. First. The stresses are greater nearer the center than near the rim. Second. The stress in the nave of a bored disc is greater than in a solid disc. A small hole may even double the strains. 76 STEAM TURBINES. These principles have been recognized in the design of the disc in the De Laval turbine shown in Fig. 48. The profile is a form of logarithmic curve asymptotic to the central axis of the disc. In the smaller turbines the flexible shaft runs through the disc, but in the larger ones the shafts are flanged to the sides, to avoid piercing the disc. The curved lines added near the center of the large wheel show the theoretical form, to which the attachments are tangent. Wheels of this form, when tested to destruction, would burst through the center into large pieces, and it is stated that these pieces have been driven through a steel wheel case 2 inches thick. This menace is obviated, however, by turning grooves, A, A, which reduce the thickness of the wheel close to the periphery, decreasing the strength of the wheel at this point so that the stresses here are about 50% higher than in the rest of the wheel. The factor of safety at this point being about 5, the rim of the wheel will tear off at a little over twice the normal speed, breaking up into small fragments, which are unable to damage the wheel case. When the rim lets go, the blades which are the impelling force of the wheel go with it, the centrifugal forces are at once much reduced, and the wheel is unbalanced. The heavy hub which projects into the casing at B with but small clearance comes into contact with the sides and acts as a brake, bringing the wheel to rest in a few revolutions. Discs of rolled plate are not used for a single-stage machine, as streaks or lines of weak- ness, due to "piping," difficult to detect, may become elements of danger. In connection with the improvement in materials referred to, it may be noted that Krupp and Co. of Essen, Germany, have for this use a special nickel steel of 125,800 pounds tensile strength, 92,300 pounds elastic limit, and 12% elongation. They have recently produced nickel steel of even higher strength but of less elongation. The following figures are taken from their published tables: MECHANICAL PROBLEMS. 77 Ult. Tensile Strength, Pounds per Sq. Inch Elastic Limit, Pounds per Sq. Inch 255,600 7. 136,320 Measured 252,760 5.5 153,360 on .472 inch 251,340 6. 210,160 diam. bar, 258,440 4.1 227,200 3 . 937 between 211,580 6.8 187,440 points. 310,980 (?)' 213,000 Broke in center punch mark. FIG. 49. Section of rotor, five-stage Curtis turbine. These elastic limits are erratic, but the results are remarkable. It is a question whether such hard material is practical but discs 78 STEAM TURBINES. are used which show under test 134,900 pounds tensile strength, 14% elongation, and 106,630 pounds elastic limit.. In multi-stage impulse turbines where the velocity is about half that of the single-stage type, the question of centrifugal strains is not so difficult. In the Curtis turbine the wheels are built up of castings and plates. The blades are milled in a sectional blade ring bolted to the discs at the edge. Fig. 49 shows the detail of this construction. In turbines of the Parsons type the velocities are still further reduced (see table, p. 56). The principle adopted by Mr. Parsons was to have the drum and shaft stiff, balanced as closely as pos- sible, and carried in a bearing which would provide sufficient lateral motion to allow the shaft to find its own center. Problems. i. Report on the conclusions of paper on "Critical Speed Calculations." by S. H. Weaver, Jour. A. S. M. E., vol. xxxii, June, 1910. References. STODOLA: " The Steam Turbine" (especially 4th ed.). JUDE: "Theory of the Steam Turbine." MOYER: "Steam Turbines." J. M. NEWTON: "Design and Construction of High Speed Turbine Rotors," London Engineering, July 8 and 15, 1910. WEAVER: "Critical Speed Calculation," Jour. A. S. M. E., June, 1910. Art. 1 6. Bearings. The earlier form of bearing used by Parsons is shown in Fig. 50, where the space between the shaft and the journal wall was taken up by two sets of rings. Those of one set fitted the shaft closely, and had play between their outer ' diameters and the journal wall. Those of the other set, which alternated with these, fitted the journal box tightly, and were loose on the shaft. They were held in contact by a side spring. A lateral movement of .005 inch or more was thus provided. This arrangement, while satisfactory, has been superseded by one (Fig. 51) in which there is a nest of bronze sleeves concentric with the shaft, each having about . 002 inch play on the diameter. Holes are bored through them to permit the lubricating oil to reach the inner sleeves and shaft. Side play is thus provided as before, but with a much MECHANICAL PROBLEMS. 79 simpler construction. In the large machines, running below 1200 R. P. M., the flexible bearing is dispensed with and solid self -oiling bearings are used. In the vertical Curtis turbine the problem has been to design FIG. 50. Old Parsons shaft packing. a step bearing which would carry the entire weight of the shaft, wheels, and the field. The end of the shaft carries a cast iron block, which runs upon a similar block in the bearing. Oil or water is forced through the pipe ^l,Fig. 52, into the recessed portion between the bearing faces, and circulates out and FIG. 51. Later Parsons shaft packing. upward, through the journal, leaving the bearing at the top 13. The entire weight of the moving parts is thus carried on the film of lubricant between the blocks a and 6. The pressure of the lubricant is from 150 pounds to 900 pounds per square inch, according to the size. 80 STEAM TURBINES. Problems. 1. Report on the principles of labyrinth packings, see Power, Sept. 1908, p. 401. 2. Report on the general principles of the labyrinth packing, i Lond. Engineering, Jan. 10, 1908, p. 35. FIG. 52. Step bearing of a Curtis vertical turbine. References. STODOLA: "The Steam Turbine" (4th ed.). JUDE : " Theory of the Steam Turbine." For theory of the labyrinth packing see London Engineering, Jan. 10, 1908, p. 35, also "Power," Sept. 8, 1908, p. 401. Revue de Mechanique, Nov., 1910. MECHANICAL PROBLEMS. 82 STEAM TURBINES. Art. 17. Governing. There are three types of turbine governors in general use: First, balanced throttling valves. Second, multiple closing or relay valves. Third, intermittent admission. The well known turbines present examples of all three of these classes. A fourth method of control is used in conjunction with the third as supplementary to it. It consists of means for admitting H. P. steam directly to the later stages by means of a by-pass valve, in order to provide for heavy overload conditions. Throttling Type. As applied on the DeLaval turbine, this governor has a spring- actuated double-beat valve, controlled by a spring and weight governor located at one end of one of the gear shafts where the speed has been reduced 10 to 1. Even here the speed is from 1,000 to 3,000 R. P. M. This high velocity alters the weight arms materially from the well known shape, and they take the form of semi-cylindrical leaves, D, turning about the knife edges, and closely enfold the controlling spring. The governing mechanism has a further control in addition to that of the throttle. If the balanced valve fails to govern and the speed continues to rise, the tappit, P, comes into contact with an independent valve, operating a butterfly valve, V, in the ex- haust passage and raising the back pressure. This immediately puts a brake on the wheel and prevents any further accelera- tion of speed. The effectiveness of this action may be shown by the fact that a 150 H. P. turbine, with nozzles designed for 150 pounds gauge pressure, and 26 inches vacuum, if operated on a back pressure equivalent to an atmospheric exhaust, will slow down and can not be brought up to full speed even with the entire load removed. The controlling mechanisms of the De Laval, Kerr, Zoelly, and a large number of other turbines follow these general lines. Multiple or Relay Control. In the smaller sizes of Curtis turbines governing is effected by throttling, as already described. In the larger sizes the speed is controlled by the opening and closing of small valves admit- MECHANICAL PROBLEMS. 83 1. Govenor. 2. Steam Chest. Nozzle Moving Blades Stationary Blades Moving Blades Stationary Blades Movmg Blades 1 I J, Fro. 54. Relay control of Curtis turbine. STEAM TURBINES. MECHANICAL PROBLEMS. 85 ting to the separate nozzle sections, as shown in the lower part of Fig. 54, which is a schematic section of the casing 2 above, the number in action depending upon the load. These nozzle valves are operated from the governor mechanism, 1, at the top of the unit, and are designed to be fully open or shut, like a pop-safety valve, in order to prevent cutting of the seats by wire drawing. The control of the nozzle valves by the governor has been effected by means of positive mechanical connections, electrical connections, and hydraulic cylinders. The last has been found most satisfactory, and is used on the later machines. This method of control by multiple valves gives excellent economy, as the steam is used at an almost constant thermo- dynamic efficiency, whatever the load. Intermittent Control. The general arrangement of the controlling mechanism is shown on the general section of a Westinghouse turbine, Fig. 55. FIG. 56. Governing mechanism of the Westinghouse turbine. Admission is through a single double-beat valve controlled by a steam-operated piston above it, Fig. 56. A heavy spring above the piston tends to hold the valve to its seat. Steam is 86 STEAM TURBINES. admitted to the piston B through the annular clearance X, around the valve steam and exhausts through a controlling valve A, which is actuated by the governor. The governor is of the bell-crank type, the horizontal arms bearing against an ad- justable spiral spring. E and D are fixed fulcrums, F, a floating one, moving up and down with the governor sleeve. C is an oscillating lever driven by worm gearing and an eccentric, from the main turbine shaft. The motion of C is communicated to the pilot valve, A, and thence to the main valve, admitting steam to the turbine in gusts. The distance that the valve A opens as it moves up and down, and consequently the time it & 1BO SI 125 fi 75 5 '0 26 Atmos- pheric^ Absolute Zero 8 2210 O FIG. 57. Pulsation diagram, Westinghouse turbine. remains open, is controlled through F by the position of the governor. At light loads the main valve opens for only short periods and remains closed for the greater part of the stroke of A. As the load increases the valve remains open longer, until finally a practically continuous pressure is maintained in the H. P. end of the turbine. The effect of the governor on the initial pressures is shown in Fig. 57. The fourth method of control, that of providing for overloads by admission of steam to later stages is also shown in Fig. 55 where a secondary admission valve V s is provided, to admit high pressure steam to the second drum of the turbine for overloads. It is not opened and closed regularly, as the main one is, but is controlled by another pilot valve con- nected directly with the governor sleeve. Another device for accomplishing the same thing, used by Brown, Bovari, and Co. in Europe, opens by-pass valves to the later stages as the pres- MECHANICAL PROBLEMS. 87 sure rises in the H. P. chamber. It is obvious that steam admitted directly to the later stages is not used to the greatest advantage, but as overload conditions are intermittent the loss in steam economy is admissible. A comparison of tests of turbines using the three methods of governing 1 shows that from the standpoint of steam consump- tion under light loads the multiple valve control is most economi- cal, intermittent admission next and throttling valve last. The throttling governor is, however, the simplest and is used almost universally on small units. Problems. 1. Look up and report on the various methods of operating the multiple nozzle valve on the Curtis turbine, see Bulletins of the General Electric Co. 2. Report on the regulation of the Rateau low pressure turbine, de- scribed in Power, May 10, 1910, p. 845. References. STODOLA: "The Steam Turbine" (4th ed.). JUDE : " Theory of the Steam Turbine." MOVER: " Steam Turbines." Revue de Mechanique, Oct., 1910. 1 Mechanical Engineer, Jan. 20, 1906. CHAPTER V. COMPARISON OF TYPES. Art. 1 8. Vertical Turbines. Up to the introduction of the Curtis turbines no one seems to have thought of arranging the shaft to run vertically, influenced, perhaps, by the accustomed position of the reciprocating engines. The builders of this turbine adopted this position and still adhere to it for their large units. Its use is limited to self con- tained generating sets. The following reasons are advanced in its favor. 1. Accurate control of the relative position of its parts by means of the adjustable step bearing. (Fig. 52.) 2. Symmetry of design. 3. All bearings relieved from side strain and lateral wear. 4. Friction practically eliminated. 5. A short shaft, free from deflection. 6. Small floor space. 7. Small foundations. The arrangement, shown in Fig. 54, unquestionably has advan- tages. That it reduces friction to a minimum is shown by the fact that unless a brake is applied a 5,000-K. W. unit will run for four or five hours after steam has been shut off. The elaborate step bear- ing necessitated, with its system of high pressure, forced lubrica- tion is one of the chief objections, and though little trouble seems to have developed from this source, it is an object of constant watching. The generator is located above the turbine, ex- posed to the heat arising from it, and to steam leaking past the glands. While the vertical position offers convenient access for inspection and minor repairs, it is inconvenient in case of overhauling. In the horizontal type the turbine and generator may be overhauled independently, and work can be done simul- taneously on both ends. In the vertical, the machine has to be taken apart from the top down, and the discs stripped off one by one. 88 COMPARISON OF TYPES. 89 An interesting example of the vertical type was built in Berlin by the Allgemeine Electricitats Gesellschaft or "A. E.G.," com- bining the features of the Curtis turbine with those of the Reidler- Stumpf . A section of this turbine is shown in Fig. 58. The weight of the moving parts was carried by a thrust bearing (B) between the generator and turbine, and a jet condenser was located at FIG. 58. Section of old A. E. G. turbine. the lower end of the shaft. The steam exhausted into the passage E, where it was condensed by water sprayed in from the side. The condensed water drained into a centrifugal pump on the end of the main turbine shaft. This arrangement is said to have given a high vacuum, and proven very satisfactory. The A. E.G. have, however, discontinued this vertical type and now 90 STEAM TURBINES. FIG. 59. 100 H. P. Terry steam turbine, opened. COMPARISON OF TYPES. 91 build for their larger sizes a horizontal machine illustrated in Art. 20. This has a Curtis single-pressure, multi-velocity stage wheel and a low pressure end with a number of pressure stages, each having a single velocity stage. A section of this turbine is shown in Fig. 66. Problems. i. Report of paper, Emmet, "Steam Turbine in Modern Engineering," Trans. A. S. M. E., vol. xxv., p. 1041. References. STODOLA : ' ' The Steam Turbine. ' ' THOMAS: " Steam Turbines." JUDE : " Theory of the Steam Turbine." FRENCH: " Steam Turbines." EMMET: Paper, "Steam Turbine in Modern Engineering," Trans. A. S. M. E., vol. xxv. Art. 19. Tangential Flow Turbines. The Reidler turbine is one of the few turbines which use tan- gential jets. The buckets for tangential wheels may have two forms, those which curve in one plane only, as the Reidler FIG. 60. Bucket arrangement, Terry turbine. and Terry turbines, with the jet divided or not, or they may be of the Pelton type, which curve in all directions. Prof. Rateau and Zoelly both experimented extensively with the divided bucket, but have abandoned it in favor of multicellular machines with axial flow. In the Reidler-Stumpf machines, STEAM TURBINES. however, it has been developed with success, and these have used it in units of 2,000 K. W. and over, having four or more stages. The Terry turbine, Fig. 59, is manufactured in this country. In this, undivided tangential jets are used. The steam is expanded in one or more nozzles to exhaust pressure and enters at A, Fig. 60, is deflected to one side, and compounded on a single wheel by looping back into buckets B, in the casing, and returned from these into succeeding buckets in the single row on the wheel. FIG. 61. Nozzle arrangement of Reidler-Stumpf Turbine. This reversal is accomplished four or five times, a portion of the steam escaping at each loop, through the exhaust openings, C. Some of the dimensions of one of these turbines are: Diameter of wheel Number of buckets Width of buckets Pitch of buckets R. P. M. Peripheral velocity = 2 feet. = 70. =2 1/2 inches. = 1 inch. = 2600. = 270 feet per second. It developed 30 H. P. with 145 pounds steam pressure on 32 pounds steam per H. P. hour. Remarkable economy is not claimed for this type but it gives a compact and simple turbine COMPARISON OF TYPES. 93 with low cost and running at moderate speeds. In the Reidler turbines the return loops are made in separate sets of buckets, as shown in Fig. 61. FIG. 62. Section of Kerr turbine. The Kerr turbine, developed by Mr. C. V. Kerr, uses buckets of the Pelton divided-flow type, mounted on discs. It is of the multicellular type shown in Fig. 30. The buckets are drop forgings, finished on the inside. The stages are repeating FIG. 63. Rotor of Kerr steam turbine. sections with the nozzles increasing in number, as rendered necessary by the expansion of the steam. The design shows careful consideration of the manufacturing problems involved, and adaptability for commercial production. The speed of 94 STEAM TURBINES. the Terry and Kerr turbines is from 1200 to 3,000 R. P. M., and they are used direct connected for blowers, pumps, and generators, also for belt driving. The governors are of the throttling type, closely resembling the De Laval. Problems. 1. Report on the Terry turbine, see Power, Nov. 1907, p. 801, and paper of Orrok, A. S. M. E., vol. xxxi, p. 263. 2. Report on the Kerr turbine as described in paper of G. A. Orrok, Trans. A. S. M. E., vol. xxxi, p. 263, also Am. Machinist, Mar. 21, 1907, p. 415. 3. Report on Sturtevant steam turbine, see paper of G. A. Orrok, A. S. M. E., vol. xxxi, p. 263. 4. Report on the Dake steam turbine. See Power, December, 1907, p. 886, and paper of Orrok, A. S. M. E., vol. xxxi, p. 263. 5. Report on the Bliss turbine as described in paper of G. A. Orrok, A. S. M. E., vol. xxxi, p. 263, and Power, August 10, 1909, p. 250. 6. Report on the Richards turbine described in the American Machin- ist, November 16, 1905, p. 660. References. STODOLA: "The Steam Turbine." JUDE : " Theory of the Steam Turbine." See also references in the following problems. Art. 20. Multicellular Turbines. As mentioned before, Prof. Rateau's investigations led him into a multicellular type of machine having many pressure stages, each with a single velocity stage, which has been brought to a high degree of perfection in France and is one of the best known of the European turbines. It is also one of the few turbines which have been applied to Marine Service. Fig. 64 shows a section of a L. P. Rateau turbine as manufactured in the U. S. The H. P. machines are practically the same except for the num- ber of stages. The pressure and velocity changes are those shown in Fig. 30. The running joints are on the shaft, where for a given clearance there is a minimum of leakage. So many rotating discs would appear to offer large steam friction loss, but tests of turbines up to 1,000 and 2,000 H. P. show a friction of only from 2 to 4%, or COMPARISON OF TYPES. 95 about the same as drum turbines of the Parsons type. The expansion nozzles are inserted in openings in the fixed diaphragms. In the high-pressure stages they occupy but a small portion of the circumference, but extend farther around as the pressure falls and the volume increases. The nozzles or vanes in each successive diaphragm are set around in angular advance of the preceding one, corresponding to the trajectory of the steam, so that the steam as it leaves the moving blades is opposite the openings of the next expansion set. An advantage in this type is that the clearance around the blades may be very large, FIG. 64. Section of Rateau low-pressure turbine. running from 1/16 to 1/4 of an inch; in fact the builders do not trouble about giving it any precise value. The only small clear- ances are at the hubs, where friction is less to be feared. The diaphragm bushings are of soft material with little attempt at clearances, for the machine when started wears a sufficient play to turn without touching. The Zoelly turbine has been mentioned as having originally had tangential flow instead of axial flow. The old form had a wheel with long spokes milled on their faces for two grooves, which gave them a Pelton form where the jet impinged upon them. 96 STEAM TURBINES. COMPARISON OF TYPES. 97 With the change to side admission the long spokes were aban- doned, and as now built the Zoelly is a multicellular turbine differing from the Rateau only in its construction and details. It has been very successful and is widely known in Europe. FIG. 66. Section of A. E. G. horizontal turbine. As stated in Art. 18, the Algemeine Electricitats Gesellschaft are building a turbine combining a Curtis single-pressure, multi- velocity stage wheel, in which the steam is expanded to about 98 STEAM TURBINES. atmospheric pressure, and a multicellular low-pressure end, making a Curtis-Rateau machine analogous to the Curtis-Parsons described in Art. 14. The Bergmann turbine is also of this type. Both are being built in large sizes. Problems. 1. Report on the A. E. G. turbine described in Lond. Engineering, May 20, 1910, p. 639. Also Power, Feb. 7, 1911. 2. Report on the Rateau turbine as described and illustrated in Lond. Engineering, May 15, 1908, p. 639. 3. Report on the Zoelly turbine as described and illustrated in Lond. Engineering, July 3, 1908, p. 1. 4. Report on multicellular turbine described in U. S. Patent issued to J. F. M. Patitz, October 4, 1910. No. 971,555. References. STODOLA: "The Steam Turbine" (4th ed.). RATEAU: Paper, "Different Applications of SteamTurbines," Trans. A. S. M. E., vol. xxv. JUDE: "Theory of the Steam Turbine." REY: Paper, "Rateau Steam Turbine and its Applications," Journal Am. Soc. Naval Eng., Nov., 1905. See also references in the problems. Art. 21. Special Forms of Impulse -reaction Turbine. The Allis-Chalmers turbine shown in Fig. 67 is a single-flow machine of the Parsons type differing from the Westinghouse only in the method of balancing, blading, and in various details. In the typical single-flow Parsons turbine the end thrust is carried by the pistons P t , P 2 , and P 3 shown in Fig. 55. In the large sizes the large low-pressure piston, P 3 , distorts under pres- sure and repeated heating and cooling, so that it has been necessary to give them clearances which permit leakage of steam. In the Allis turbine this large balancing piston is replaced by a smaller one, a, at the low-pressure end of the rotor where it is practically free from distortion, and nearly the whole area is effective. In the later forms of Westinghouse machines the balancing is accomplished by the double flow, as already described. In the blading, the ends are riveted to a channel-shaped shroud ring, B, Fig. 68, stiffening them against vibration. The openings COMPARISON OF TYPES. 99 100 STEAM TURBINES. for the blade tips are stamped in the ring, accurately spacing the blades. The shape of the ring forms a natural baffle, and the clearance may be smaller than in previous practice. Only the thin edges of the channel can touch the casing, and if they do, they wear free without any material heating. The FIG. 68. BJading of Allis-Chalmera turbine. shroud ring itself is not new, as both the Rateau and Zoelly have used them, but the gain from its use is greater in reaction turbines than in these types. Problems. 1. Look up and report on the early history of steam turbines, see Neilson, French or Gentsch. 2. Look up the Elektra turbine described in Power, April 6, 1909, p. 635. Compare with the new type of small Westinghouse turbine. 3. Report on the Wilkinson turbine and small type of Westinghouse turbine. 4. Report on the Eyermann turbine, a turbine having combine radial and axial flow. See Power, November 24, 1908, p. 855. COMPARISON OF TYPES. 101 References. STODOLA: "The Steam Turbine" (4th ed. illustrates and describes all the turbines of any prominence). JUDE: "Theory of the Steam Turbine." GENTZCH: "Steam Turbines." Translated by A. R. Liddell. SOZNOWZKI: "Turbines a Vapeur," Paris, 1897. (The last two are valuable for the history of steam turbines.) See also references in the problems. Art. 22. Low-pressure Turbines and Regenerators. Reference to the pressure-temperature curve, Fig. 69, or to the heat chart, shows that the heat drop on which the available energy depends is greatest at the low pressures. A reciprocating Absolute Pressures 70 80 90 100 110 120 130 MO 160 180 170 ISO FIG. 69. Pressure-temperature curve. engine is most efficient at the higher pressures, down to about that of the atmosphere. The turbine is most effective in utilizing the lower vacuums, as excessive expansion in an engine entails 102 STEAM TURBINES. prohibitive sizes of L. P. cylinders, and, with the wide range of power available below atmospheric pressure, it can be coupled with a reciprocating engine to great advantage. In fact, the economy of such a combined unit where the engine exhausts into a low-pressure turbine, is higher than when the entire pres- sure range from boiler to condenser is utilized in an engine or a turbine alone, and the installation of combined units of this kind is coming more and more into use. For marine service there is a further advantage of superior maneuvering power and economy at slow speeds over turbines alone, and the new White Star steamers Oceanic and Titanic are being equipped with engines and exhaust turbines. For power station work the total economy of a combined unit of new engines and new turbines, taking into account all running expenses, fixed charges, etc., is not so high as for a H. P. turbine plant alone. It furnishes, however, a valuable means of utilizing existing reciprocating plants and bringing them up to the highest efficiency. Where a good engine plant already exists its capacity may be increased more econom- ically by the addition of exhaust turbines than by replacing the engines with H. P. turbines. The conditions of operation for exhaust turbines fall under two classes: First, where the steam supply is substantially constant or varies with the turbine load. Second, where it is widely fluctuating or intermittent and is independent of the turbine load. In the first class fall marine engines and those for power station service. A conspicuous example of the successful installation of exhaust turbines for the latter service is found in the Fifty-ninth Street Station of the Interborough Rapid Transit Co., N. Y. City. This station was equipped with 7500-K. W. compound Corliss condensing engines direct connected to generators, each unit having two 42-inch H. P. and two 86-inch L. P. cylinders, all of GO-inch stroke, and an average steam rate of 17 to 18 pounds per K. W. hour. After consideration of various means of pro- viding additional power, exhaust turbines were installed on three units to determine their desirability for the remaining sets. They were 3-stage vertical Curtis turbines, of 7500 K. W. maxi- mum rating, each having 6 fixed nozzles, and 6 operated by hand COMPARISON OF TYPES. 103 so as to control the division of load between the engine and its exhaust turbine. An overspeed governor operating a 40-inch butterfly valve on the steam line connecting the separator and the turbine, and an 8-inch vacuum breaker on the condenser were the only forms of governor used. Mr. Stott summarizes the results of the tests as follows: "a. An increase of 100% in maximum capacity of plant. "b. An increase of 146% in economic capacity of plant. "c. A saving of approximately 85% of the condensed steam for return to the boilers. "d. An average improvement in economy of 13% over the best H. P. turbine results. "e. An average improvement in economy of 25% (between limits of 7500 K. W. and 15000 K. W.) over the results obtained by the engine units alone. "f. An average unit thermal efficiency between the limits of 6500 K. W. and 15500 K. W. of 20.6%." The steam rate of the engines alone was from 17 to 18 pounds per K. W. hour, of the exhaust turbines, 26 1/2 to 28 pounds, and of the combined unit ,13 to 14 pounds. In an installation of this character where the load on the turbine is electrically connected with that of the engine, the turbine load rises and falls with the steam supply and no governor is required. When running at very light loads, where a turbine is inefficient, it is found economical to cut it out, and run the engine alone, condensing. Where the supply of steam from the engine is liable to long-continued interruptions, live steam can be admitted through a reducing valve from an independent source, the exhaust turbine thus drawing its supply from two sources. The second class of installations, where there is intermittent steam supply, comprises rolling mills, forging plants, mine hoists, and other non-condensing engines. Such engines use great quantities of steam under widely varying loads, often with little or no expansion. Prof. Rateau has given much attention to the application of low-pressure turbines to such exhausts, and to equalize the supply from these constantly varying sources he has developed the steam regenerator or heat accumulator, shown in Figs. 70 and 71. The engine exhaust at about atmospheric pressure is condensed in the regenerator when it arrives in large 104 STEAM TURBINES. quantities, and revaporized when the excess ceases. The con- densation and re-evaporation of the steam follow the fluctuations of pressure as the supply exceeds or falls short of that required by the turbine. \Perforations In Mixing Tube for distributing Steam into the Water FIG. 70. Side view of Rateau regenerator. The regenerator takes a form like a tubeless horizontal boiler (Fig. 70) where the water is kept in forced circulation by the injection of steam. It is necessary to provide a relief valve to carry off any surplus steam, an automatic expansion valve, as in the previous case, to supply steam directly to Extension of Mixing Tubes above Water Level Leadiug steam in Regeuera Manifold Water Level Baffle Plate Perforations in Narrowed Portion of Mixing Tube for Distribution of Steam in the Water. FIG. 71. Cross-section of Rateau regenerator. the turbine should the main engine be shut down for any length of time, and check valves to protect the regenerator and main engine when the turbine is running direct. The steam flow from the regenerator to the turbine is constant COMPARISON OF TYPES. 105 compared to that from the engine to the regenerator and is, of course, substantially equal to the average rate of engine exhaust. Let T be the maximum time, in minutes, of stoppage of the engine, during which the regenerator must supply the turbines alone. Let R be the total steam consumption of the turbine in pounds per minute, L the mean latent heat at the regenerator pressures (usually to 4 pounds gauge), t the allowed temper- ature drop of the water. Then (30) is the weight of water required to run the turbine T minutes with an allowed fall in temperature, and consequently pressure. This difference in pressure is not over 3 pounds, at the outside. The capacity to absorb the engine exhaust, however, deter- mines the amount of water required, rather than the question of the turbine supply. A regenerator may be called on to absorb fluxes of steam five times as great as the more uniform flow to the turbine and they may be thrown on it almost instantly and must be cared for. If S be the maximum pounds of steam the regenerator will be called on to absorb, L the mean latent heat of the engine exhaust, and 12.4 B. T. U. the difference in temperature between and 4 pounds gauge pressure, then the water required is ' (3D . . which will usually be found greatly in excess of the w above. Proper values of T, S, and the mean engine exhaust are determined only by careful investigation for each installation. The form of regenerator shown in Fig. 71 has shown the greatest capacity for absorbing the steam. Two regenerators of an older type installed at the Bruay Mines in France to utilize the exhaust of a hoisting engine and drive a 300 H. P. Rateau generator set have been in successful service since 1902. With supply pressures ranging from 12 to 14.5 pounds absolute, and condenser pressures from 2 . 13 to 2 . 62 pounds absolute, the plant showed a steam con- sumption of from 37 . 4 to 45 . 2 pounds per E. H. P. hour. A great number of regenerators and L. P. turbines are now in successful 106 STEAM TURBINES. use in Europe and in this country, some of them exceeding 3000 H. P. of turbine capacity. Where there are long periods, as for instance over night, when the exhaust steam supply is not available, a high-pressure turbine has been coupled with the low-pressure one, the low-pressure tur- bine being so proportioned that it can carry the entire load, but if its steam supply fails regulating valves controlled by the speed and the pressure in the regenerator, cut in the high-pressure turbine, and kept up the output. The load may be carried FIG. 72. Rateau mixed-flow turbine. entirely by either turbine, or distributed between them, as the momentary conditions require. Any exhaust steam used in the H. P. turbine passes through the L. P. one, and the group is thus working at the best' economy possible. Running entirely on the H. P. machine, with steam at 110 pounds gauge pressure, one of the combined units has been found to give an economy of about 18 pounds of steam per E. H. P. hour, and running entirely on the L. P., with steam at atmospheric pressure, it gave about 36. The two turbines are now combined in one to form a "mixed- flow" turbine as shown in Fig. 72, a much more compact arrangement. COMPARISON OF TYPES. 107 Problems. 1. Report on the article on compounding piston engines with turbines, by Prof. Rateau, Power, October, 1907, p. 693. 2. Report on paper on Exhaust Turbines and Regenerators, by R. F. Halliwell, given in Lond. Engineering, February 5, 1909, p. 197. 3. Report on low pressure steam turbine as referred to in article by R. M. Neilson, Power, July 6, 1909, p. 1. 4. Report on article by C. H. Smoot, on Regenerators. Power, November 8, 1910., p. 1979. 5. Report on article, L. P. Turbines, by C. H. Smoot, Power, June 22, 1909, p. 1100. 6. Report on investigation into L. P. turbines for the Cambridge Elect. Light Co. by Prof. Hollis, see Power, November 10, 1908, p. 787. 7. Report on the Brush-Parsons double-flow exhaust turbine, de- scribed in Lond. Engineering, July 1, 1910, p. 2. 8. Report on English and German experience with L. P. turbines as described by Gradenwitz in Engineering Magazine, vol. xxxiv, p. 278. 9. Report on the test of engines and exhaust turbines at 59th street Interborough Station, see paper by Stott and Pigott, Trans. A. S. M. E., March, 1910, see also Power, December 14, 1909, p. 985. References. STODOLA: " The Steam Turbine." MOYER: "Steam Turbines." RATEAU: Paper, "Different Applications of Steam Turbines," Trans. A. S. M. E., vol. xxv. RATEAU: Article, Power, Oct., 1907, p. 693. BUELEIGH: Paper on Exhaust Turbines, Power, Oct., 17, 1908, p. 828. HALLIWELL: Paper on Exhaust Turbines, London Engineering, Feb. 5, 1909. STOTT AND PIGOTT: Paper, "Test of a L. P. Turbine," Trans. A. S. M. E., 1910. See also Power, Dec. 14, 1909, p. 985. SMOOT: Articles, Power, June 22, 1909, p. 1100, and Nov. 8, 1910, p. 1979. See also references in the problems. CHAPTER VI. CONDITIONS OF OPERATION. Art. 23. Effect of Superheat. As with reciprocating engines, there is a gain in the economy of turbines by the use of superheated steam, but how much is gained can not be fully determined until the specific heat of superheated steam is definitely known. While this has been the subject of repeated and extensive experiments from the time of Regnault to the present, the proper values have not been agreed upon. Grindley in England; Greissman, Lorenz and Knoblauch, Linde and Klebe in Germany, and Carpenter and Jones in America, have all done important work in this direction, but the results are contradictory and have not definitely settled how the specific heat varies with the pressure and temperature changes. Modern experiments show that the value .4805 of Regnault is too low. Apparently the specific heat increases with increase of pressure, but decreases with increase of superheat. This conclusion is indicated by Lorenz, and is sustained by Knoblauch, Linde, and Klebe. Mr. Emmett has deduced from a series of experiments made in a Curtis turbine at a pressure of 155 pounds that it varies from about . 53 to .77 for temperatures of 360 to 620 Fahr. The usual practice is to take it as constant and at from . 6 to . 65. The bearing of the specific heat on the theoretical economy of the turbine using superheated steam is well shown by calcula- tions made on a De Laval turbine. By taking first the water rate consumption as a basis for calculating, the gain was found to be 8. 1 per cent. Then on the basis of the heat units in the steam the gain was found to be, for Specific Heat Gain .48 4.8% .6 4 % S 2.7% 108 CONDITIONS OP OPERATION. 109 In the diagram following are given the average of a number of steam consumptions for various degrees of superheat, up to 200. These are based on tests of Parsons, Curtis, Zoelly, and Brown- Boveri turbines, and of such high grade engines as had records of tests under same conditions. The lines for the turbines are more reliable than for the engines, being based on a larger I 19 3 60 80 loo' iao Degrees of Superheat FIG. 73. Effect of superheat. 180' SCO 1 number of tests. According to these curves the gain in economy is greater for engines than for turbines. The curves for the turbines would confirm the statement made by Mr. Parsons and by Messrs. Dean and Main that the gain in economy is about 1% for every 10 "of superheat, up to about 150 or 180 Fahr. The gain from superheat in turbine practice comes chiefly from the reduced steam friction in the blades and passages. Refer- ence to Fig. 40 shows that there may be 10% or 15% of water in the steam in the later stages. Superheating reduces the con- 110 STEAM TURBINES. densation and by lessening the water going through the blades, materially decreases the steam friction. It is claimed that a turbine can operate under much higher superheat than an engine. Doubtless this is true, but it remains to be shown that there is any advantage in the high superheats the turbine may be made to stand. The best practice indicates that a moderate superheat, of 50 to 100 is of advantage. Beyond this the results are not wholly favorable, for while the steam rate can be lowered it is done at the expense of extra fuel required for the high superheats. References. STODOLA: "The Steam Turbine." JUDE: '' Theory of the Steam Turbine." MOYER: "Steam Turbines." FRENCH: " Steam Turbines." MARKS AND DAVIS: "Steam Tables and Diagrams." Trans, of A. S. M. E., 1905 to 1910. Art. 24. Relative Effects of High Vacuums in Reciprocating Engines and Turbines. The turbine gains relatively more than the steam engine as the vacuum is pushed toward its limit. In fact, this is one of the chief elements in its success in competition with the older form. The advantage in economy, under equal conditions, is usually in favor of the steam engine for vacuums lower than 26 inches or 2 pounds absolute. For higher vacuums the turbine gains rap- idly. General practice has set the economical vacuum for recip- rocating engines at 26 inches to 27 inches. Most of the turbine practice at the present time is based on 28 inches to 28 1/2 inches of vacuum. The reasons why the turbine shows this greater benefit for the higher vacuums are: First. The low temperatures of the exhaust pressure can not reach back into the hotter parts of the machine, as in a L. P. engine cylinder, which is alternately exposed to wide differences of temperature. Second. If expansion is carried too far in the steam engine, the L. P. cylinder valves and passages become abnormally large, absorb power, and neutralize any thermodynamic gains. A CONDITIONS OF OPERATION. Ill compound engine with a cylinder ratio of 4 to 1, using steam at 150 pounds, might expand say 15 times. A turbine presents no difficulties to an expansion ratio of 100 to 110 or down to a pressure of 1 pound absolute. To carry this expansion in a compound engine calls for a cylinder ratio of 33 to 1, which is obviously prohibitive. Since the expansion of the engine is limited to the L. P. termi- nal, c, Fig. 74, the only gain in a higher vacuum is the lowering of the back pressure, represented by the area e, d, f, g. FIG. 74. Pressure-volume curve for engines and turbines. Since expansion can be carried down within a turbine to the lowest condenser pressures, the turbines can utilize the large triangular expansion area to h, in addition to all that the engine gains. Theoretically the gain increases very rapidly for the rarer vacuums, but with the turbine, as with the engine, there is a practical limit, though not reached so soon, beyond which the cost, maintenance, and power required for the condensing apparatus offset the theoretical gains. As already stated, the limit is 28 to 28 1/2 inches, but its value for any given set of condi- tions is a question of engineering economics. Into it enter the following: 1. The steam consumption of (a) turbine, (b) auxiliaries. 2. The use-factor, or the ratio of the actual output to the rated output. 3. The cost of the coal, as fired. 4. Cost of injection water. 112 STEAM TURBINES. 5. Fixed charges, on the condensing plant, covering (a) Interest on cost. (b) Interest on cost of masonry, space occupied, etc. (c) Depreciation and repairs. (d) Insurance, taxes, etc. The method of determining the economical back pressure may be illustrated by an assumed case. Let Capacity of turbine = 1,500 K. W. Use-factor (2) =15%. Cost of coal (3) = $3 . 50 per ton, fired. Suppose fixed charges (5) = 10%. Average boiler evaporation = 7 pounds per pound of coal. p o Lss x I X 83 53 Vacanma FIG. 75. -Approx. relation of water rate and vacuum. Let the cost of condensing apparatus, based on injection water of 70 Fahr., obtained from estimates of builders for several different back pressures, be as follows: The plant for 28 1/2 inches and 29 inches would be the same as for 28 inches, except for an increase in the size of the dry vacuum pump. These vacuums are obtainable only with injection water of 60 Fahr. for 28 1/2 inches, and 45-50 for 29 inches. CONDITIONS OF OPERATION. 113 Inches of Vacuum Cost of Condensing Plant Increase in Cost Increase in Annual Charge Based on 10% 26 $6000. 26* 6900. $900. $90. 27 7500. 600. 60. 27* 8000. 500. 50. 28 9500. 1500. 150. 28* 10000. 500. 50. Let the guaranteed steam consumption of the turbine be 21 pounds per K. W. hour, at 28 inches vacuum. The steam rates for varying vacuums may be interpolated from Fig. 75, which is based on some Westinghouse tests. The steam consumption of the auxiliaries may be based on their estimated H. P., and a steam consumption of 1 1/2 times that of the turbine. The total steam consumption then would be - Steam Consumption Vacuum Inches Turbine Auxiliaries Total 1 2 3 4 26 23.4 3.4 % 24.2 26* 22.8 4. % 23.7 27 22.2 5.25% 23.3 27* 21.6 6.9 % 23.1 28 20.9 9. % 22.8 28* 20.2 12. % 22.6 The yearly running cost = X const. Where E A =K. W. capacity of the unit. 5 = Total consumption (Coll. 4 last table). (7 = Power factor. D Cost of coal, fired. E = Average boiler evaporation per pound. 114 STEAM TURBINES. hours per year 8760 Constant = = -=3.92. pounds per ton 2240 Calculating the yearly cost with the various consumptions, and tabulating, we have : Inches of Vacuum Cost of Steam per Year Saving 26 440X24.2 =$10648 26J 440X23.7 = 10428 $220. 27 440X23.3 = 10252 176. 27J 440X23.1 = 10164 88. 28 440X22.8 = 10032 132. 28J 440X22.6 = 9944 188. Comparing these savings per year with the increase in the annual charge as the vacuum varies, we see that the gain changes to a loss at just under 28 inches. We see, too, that if an ample supply of cold injection water is available, it would pay to carry 28 1/2 Fia.J76. Parsons Augmentor. inches where the temperatures will permit. This method would call for a vacuum somewhat too low, if anything, for the steam used in the auxiliaries is not all lost, as a large part of the heat in it is regained in heating the feed water. CONDITIONS OF OPERATION. 115 Higher vacuums than 28 inches are carried in high-class power tations. Some condenser plants have even been installed under a guarantee to maintain a working vacuum within 3/4 inch of the barometer. Whether such a high vacuum is profitable remains to be shown by experience. Mr. Parsons has invented a device, Fig. 76, which has been widely used in Europe, but for some reason has been little used in this country. He places the air pump, proportioned for about 26 inches vacuum, below the main condenser, and connects this by an inverted siphon, forming an air seal. Between the air pump and condenser, as a by-pass connection, is a steam ejector connection, A, delivering through a small surface condenser B, about 1/20 the size of the main one, into the air pump suction beyond the siphon. This ejector draws nearly all the residual air and vapor from the main condenser and increases the vacuum to 27 1/2 inches or 28 inches, without the increase in size of main condenser and air pump demanded by the ordinary arrangement. Mr. Parsons states that the live steam required by the ejector is about 1 1/2% of that used by the main turbine at full load, and is materially less than that required to drive the larger air pump, which would regularly be provided for the higher vacuum. Problems. i. Report on paper of G. I. Rockwood on "Condensers for Steam Turbines," Trans. A. S. M. E., vol. xxvi, p. 383. References. STODOLA: " The Steam Turbine." JUDE: "Theory of the Steam Turbine." ROCKWOOD: Paper, "Condensers for Steam Turbines." Trans. A. S. M. E., vol. xxvi. p. 383. CHAPTER VII. POSITION AND FIELD OF THE STEAM TURBINE. Art. 25. Relative Economy of Engines and Turbines. The heat equivalent of a H. P. hour is 33000 + 60 If H t and H 2 are the total heats per pound of steam before and after adiabatic expansion to the exhaust pressure, and W the weight of steam per H. P. hour as tested, then <*.-*.> e-~ or the available heat drop X% utilized = B. T. U. utilized per pound 2545 This "potential efficiency," P, which has been variously named by different authorities, gives a comparison between the heat actually utilized by the motor, and that available under its operating conditions. It applies equally to reciprocating engines, and to turbines, and can be used as a .basis of compari- son, provided the water rate can be brought to the same basis. The almost universal basis used in determining the economy of the reciprocating engine is the steam rate per indicated H. P. hour. No indicator has been devised for turbines, and it would be difficult to show the internal power developed. While some device might show the energy of the jet of steam from the nozzle, it would be practically impossible to register definitely the energy given up to the blades in a multi-stage turbine, where there are losses from eddies, friction, etc. Engi- 116 POSITION AND FIELD OF THE STEAM TURBINE. 117 neers, however, are so familiar with economies based on indicated H. P. that comparisons are often made by reducing the steam consumption of the turbine to the basis of the I. H. P. of a reciprocating engine having the same electrical output or brake H. P., as the case may be. Where it is possible there is a growing practice of basing the steam rate on the K. W. hours or electrical H. P., which requires no assumptions, especially as the bulk of turbine work outside of the marine service is in direct connected electrical units. Comparisons based on I. H. P. must take into account the efficiency of the generators, turbines, and engines. The efficiencies of engine-driven generators for alternating currents on sizes from 500 to 5,000 K. W. will range from 87 1/2% at 1/4 load up to 96% at 1 1/4 load. For direct current generators, 100 K. W. to 3,000 K. W., it will range from 89% at 1/4 load to 94% at full load. The generator efficiency with the De Laval units is somewhat lower, as the electrical work is divided between twin generators. Their efficiency may be taken at from 88% at 1/2 load to 91% at full load, for direct current, and at 86% at 1/2 load to 92% at full load. The following generator efficiencies are reported in various tests. Capacity Load K. W. i i 1 Full H Westinghouse 1 250 86 93 96 Westinghouse Allis-Chalmers Curtis 400 5,500 94.6 97 96 95.7 97.9 96.6 98.3 97 5 98.5 In comparing with engines a value of 95% may fairly be assumed. The mechanical efficiency of reciprocating engines has^been^exhaustively investigated, and the friction losses appear to be nearly constant for all loads. The efficiency con- sequently falls off rapidly for light loads. A summary of various modern engines as given by French is as follows : 118 STEAM TURBINES. Mechanical Efficiency of Engines at or near Rating. Engine Efficiency, % Engine alone Engine and Generator 96.5to97.5 94.5 91 94 87 to 93 90 92 to 94 90 Large horizontal Corliss, compound High speed, simple Small and medium, hor. compound The mechanical efficiency for turbines, as it involves the com- parison of the brake H. P. with the indeterminate internal power developed by the steam, can only be assumed. This factor is taken from 90% to 95%. For large units the higher figure is probably nearer correct. These values give a combined efficiency of 85% to 93% for a turbine generating unit, which may be used in reducing K. W. output to an I. H. P. basis, for comparison with reciprocating plants. The following tables, from a paper by Mr. C. V. Kerr before the A. S. M. E., 1 give means of comparing reciprocating engines and turbines on the basis of their potential efficiencies. 1 Trans. A. S. M. E., vol. xxv., 1904. POSITION AND FIELD OF THE STEAM TURBINE 119 o >o >o co .p . . . *^ SCO i-4 <0 t* t~ CO COlO CO^t-t* II O 50 rH rH -H CO * O CD -< -^ t* CO CO 03 CO w O U5 * 5 ^ 00 CO t^ b- UP fc,' 00 O 00 OOOfflO^ 2 S i I I ; 1 i jil: I ; li s l! : .3 -a 120 STEAM TURBINES. ' "> S o a o a **\ |J 14 5 If ri ill Ol^t^OOCN) >O5 0000 j t~ t~ S 3 111 b <= O CO O O O rH CO t-_ O 00 lOOMrHrH TjOrH t^ IN1O * CO-*COrHr-J 10 f'lN ** CO CO rH IN 8&Sflo*Sfl8oSio 53 IN IN (NOINrHO (NrH Wrt O5O5 n i.g's a 8 > 11 II s I : I > -a ; -s i'a I IPPP1 I _; 122 STEAM TURBINES. Looking at the economy on a basis of steam per H. P. hour, the following table gives the results obtained from some recipro- cating engines of exceptionally high economy. 2 S" fc V M I ** B ' S Engine 1 , 1 S S j! H Authority i I o 1 Pk ji IB W tf CO Westinghouse Vertical 5400 185 27.3 76 11.93 Eng. Record Brooklyn, N. Y. May 28, 1904. Rockwood-WTieelock 595 159 25.4 76.4 13. F. W. Dean Natick, R. I. A. S. M. E., 1895. Mclntosh and Seymour 1076 123 27.1 99.6 20 12.76 F. W. Dean Webster, Mass. A. S. M. E., 1898. Rice and Sargent 627 151 28.6 121 12.1 D. W. Jacobus Brooklyn, N. Y. A. S. M. E., 1903. Rice and Sargent 420 142 25.8 102 297 9.56 D. W. Jacobus Phila., Pa. A. S. M. E., 1904. Horizontal Four- Valve 658 150.4 26.4 80 16.4 12.03 Barrus' Eng. Tests. Leavitt Pumping Engine 575 175.7 27.25 50.6 11.2 E. F. Miller Chestnut Hill, Mass. 1 Tech. Quart'ly, vol. ix. The following is a table of tests of 23 engines in commercial operation, and may be taken as indicating average results. Indicated Horse Power Steam per I. H. P. Hr. Indicated Horse Power Steam per I. H. P. Hr. i 659. 11.89 725. 13.27 658. 12.03 1714. 13.27 670. 12.29 1030. 13.21 798. 13.28 843. 13.53 1017. 13.26 382. 14.05 689. 12.69 873. 14.18 708. 12.45 676. 14.6 636. 13.28 1540. 14.1 280. 13.37 300. 15.78 719. 13.09 606. 16.28 741. 13.23 716. 19.36 739. 13.01 In comparison with these results are the following. 1 In reducing 1 French. " Steam Turbines.' POSITION AND FIELD OF THE STEAM TURBINE. 123 the given steam rates to I. H. P. the combined efficiency factors used are: For units 300 . For units 500. For units 2000 . . 400 K. W oo 7o .1500 K. W 90% .3000 K. W 92% Where 93% is used the test was made on brake H. P. basis. Engine efficiency alone without generator, for units from 400 to 500 H. P., or 300 to 400 K. W., taken at 93%. Engine efficiency corresponding to the 300 H. P. De LaVal turbine, taken at 92%. Typical Steam Consumptions of Turbines. Turbine Nominal Power Steam Rate Estimated Equiv. Consumption per I. H. P. S** g$| ^5 W %ti W l SS^ o> fc I m Per Elect. H. P. W 1 Saturated Steam De Laval 300 H. P. 500 H. P. 500 H. P. 500 H. P. 400 H. P. 1250 H. P. 15.17 13.63 13.96 13.11 14.12 13.28 12.68 12.72 92. 88. 88. 90. 93. 90. Rateau 14.9 16.05 14.76 21.5 19.78 Zoelly Curtis, American Westinghouse-Parsons . Westinghouse-Parsons . 14.13 18.95 Moderate Superheat Up to 150 De Laval 300 H. P. 500 H. P. 500 K. W. 500 K. W. 300 K. W. 1500 K. W. 3000 K. W. 400 K. W. 1250 K. W. 13.94 12.07 14.05 15.29 13.28 14.96 13.44 11.79 13.78 18.82 20.50 17.79 20.06 18. 15.8 18.48 12.82 12.36 13.76 11.95 13.16 12.10 10.85 11.23 12.40 92. 88. 90. 90. 88. 90. 92. 93. 93. Zoellv Curtis, English Curtis, American Parsons Westinghouse-Parsons . Westinghouse-Parsons . 124 STEAM TURBINES. Typical Steam Consumptions of Turbines. Continued. High Superheat 180 to 290 Curtis, American 500 K. W. 11.26 15.1 10.14 90. Curtis, American 2000 K. W. 11.27,15.12 10.36 92. Parsons 3000 K. W. 11. 14.74 10.12 92. Westinghouse-Parsons . 400 K. W. 11.17 10.39 93. I In comparison with the last test in the table, with the high super heat, may be taken the tests of a Van der Kerchove engine, made by Prof. Schroeter. I. H. P. Superheat in Deg. Fahr. 1 Steam per I. H. P. Hr. Lbs. Equivalent Pounds of Saturated Stm. Heat Units per I. H. P. Hr. 222. 0. 12.06 12.08 250. 226. 43.7 11.58 11.77 244. 227. 97.7 11. 11.44 237. 223. 151.7 10.67 11.33 234. 223. 221.2 9.81 10.69 221. 218. 310.9 1 8.86 10.01 207. The Rice and Sargent engine, in table, p. 122, is perhaps a still better comparison. The following list of tests of engines and turbines may be taken as representing the comparative economy obtained by each type to January, 1911. The B. T. U. per K. W. minute are obtained by multiplying the steam rate per K. W. minute by the difference between the B. T. U. per pound of steam at the initial conditions of pressure and superheat and the B. T. U. in 1 pound of water at the vacuum pressure, col. 4. The last column gives the best basis for comparison. POSITION AND FIELD OF THE STEAM TURBINE. 125 1 2 3 4 6 1 Load K. W. Steam Press. Super- heat or Mois- Vacu- um Lbs Steam perK. B.T.U. per K. W. Abs. ture . Hour Min. Turbines N. Y. Edison, Westinghouse 9870 192 97F 27.31 15 294 City Electric, Westinghouse 9173 182 59 27.90 14.57 282.4 Bklyn Rpd. Transit, Westing- 11466 192 106 28.15 14.45 287.5 house. Chgo. Edison, Curtis 10186 191 147 29.47 12.9 270 Carville, Parsons 5059.4 209 95 29.27 13.35 271.7 Frankfort, Brown-Bovari 3521.6 157 130 28.90 13.70 278 Turin, Escher-Wyss 3540 187 90 28.35 15 297.4 Berlin Moabit., A. E. G 3169 185 215 29 12.7 268.4 Rummelsburg, A. E. G 4239 188 285 29 11.9 258.5 Engines N. Y. Edison, Westinghouse 3872 200 dry 27 16.78 312 Interborough, 59th St., Allis 1 . . 5496 190 dry 25.02 17.82 225 Boston Edison, Mclntosh 1500 176.5 92 25.4 16.47 315 Berlin Moabit 1920 202.7 223 28 13.35 276 Berlin Luisenstrasse | 1950 202.7 264 27 14 295 1 1 Without L. P. Exhaust Turbines, see Art. 22. The results brought out in the foregoing indicate 1. That with high vacuums the larger turbines show economies somewhat better than reciprocating engines. 2. That with moderate vacuum, in the medium sizes, the advantage in economy still lies with the refined types of recipro- cating engine. Problems. 1. Look up and report on methods of testing steam turbines. Paper by Dickinson and Robinson, Am. Inst. of Elect. Eng'rs, December, 1910. 2. Report on "Test of a 9000 K. W. Turbo-generator Set." Paper by F. H. Varney, Trans. A. S. M. E., vol. xxxii, December, 1910. 3. Report on chapters on testing of turbines in Moyer's "Steam Turbines." References. STODOLA: "The Steam Turbine." MOYER: "Steam Turbines." ' JUDE: "Theory of the Steam Turbine." FRENCH: "Steam Turbines." 126 STEAM TURBINES. NEILSON: "The Steam Turbine." THOMAS: " Steam Turbines." KERR: Paper, Trans. A. S. M. E., vol. xxv. Art. 26. General Comparison of Engines and Turbines. Other considerations, however, enter such as steadiness and economy under widely fluctuating working loads, the room required, the foundations, cost of installation and operation. The governors of reciprocating engines regulate the effort from stroke to stroke to meet varying load conditions, but fluctu- ations due to the effect of the connecting rod and the weight of the reciprocating parts can be cared for only by the flywheel. The flywheel can at best only reduce these fluctuations to within prescribed limits. In direct connected generating units, espe- cially when running in parallel, the turbine has the advantage of a total absence of reciprocating motion and a uniform turning effort, the high speed of the rotor and heavy armature acting as a powerful regulating force. In fact it has been found possible with turbines to run railway, power, and lighting circuits from the same machine, and where a turbine has been installed to run in parallel with piston engines it has always been found to have a steadying effect on the whole system. But the turbine is a one-speed machine. In this respect it is distinctly inferior to the reciprocating engine, for it cannot be run at speeds lower than that for which it is designed without serious effect on its economy. This is one of the limitations to its use in marine work, and has proved one of the difficult prob- lems in that connection. Even turbine advocates do not recom- mend them for intermittent services and varying speeds. The field available for the marine turbine seems to be deep sea service, where the work consists of long runs at uniform speed. Tests made on a Curtis turbine for variation in steam consumption in relation to the speed resulted as follows: R. P. M. Steam per K. W. hr. 1900 19.75 1600 20.74 1300 22.7 1000 27. POSITION AND FIELD OF THE STEAM TURBINE. 127 In the torpedo-boat destroyers built by the British Navy for comparing turbines and reciprocating engines, the turbine boats were superior in economy to the engine-driven boats at normal speeds, but invariably inferior when running at slow cruising speeds. In good practice reciprocating engines are not installed to run under frequent overloads greater than 50% of the normal capacity. Corliss engines can be made to carry overloads of 100%, but not with good economy. The tendency is to provide an engine large enough to handle the maximum loads easily, and let it operate at an average load considerably below its rated capacity. Prof. Carpenter has reported tests of thirty-five street railway power plants. In these this tendency is clearly marked. The following table gives a summary of seven of these tests made on compound condensing engines of the Corliss or similar types. The remaining ones not given show the same tendency. H. P. of Mean %of Steam per Engine observed H. P. Capacity I. H. P. Hour. 825 482 58.2 22.7 1,000 277 27.7 21.9 1,000 314 31.4 20.0 350 182 52.2 16.64 500 290 58. 16.90 2,000 814 40.7 14.50 200 145 72. 17.30 Average 48 6 18 6 The average load on these engines was less than half their rated capacity, and they were operated at an average steam rate of over 18 pounds instead of 14 and 15 pounds, which the type of engine is capable of. As turbines have been rated they have been capable of operat- ing under heavy overloads and at very fair steam rates. Guar- antees were regularly made on the rates for overloads of 25%, 50%, and 100%. As a result turbines have been installed more with reference 128 STEAM TURBINES. to the average load to be carried, and less with reference to the maximum load than a steam engine, the turbine operating at nearly its best economy most of the time, whereas an engine, rated nearer the maximum load, would be operating at less than its most efficient rating. The following table shows the variations in consumption reported for a Westinghouse turbine on loads varying from 1/2 load to 100% overload. Steam Rate Steam Rate Load Saturated 100 Superheat 1/2 load 3/4 load Full load 15.86 15.05 13.89 14.34 13.45 12.48 1 1/4 load 1 1/2 load 13.85 12.41 12.79 100% overload 15.12 13.55 Here an overload of 100% was carried with an increase of only 10% in the steam rate. Such a load could not be carried indefi- nitely with safety to the generator, but overloads of 50% are often met. It should be added however, that turbine manufacturers are rating their machines higher now and this difference is becoming less marked. In respect to the room occupied, the advantage in favor of the turbine alone is very marked. If the condensing apparatus be in- cluded in the comparison the difference is not so great, but even then it is usually in favor of the turbine. This is brought out clearly in Fig. 77, which shows a 3,500 K. W. engine-driven unit, l and two tur- bine-driven units, one 5000 K. W. and behind one of 7500 K. W. in the Waterside Station of the N. Y. Edison Co. Four of the engine- driven units similar to that shown in the picture are being re- placed by three turbine units, each of 20,000 K. W. capacity. A comparison of these turbine units with the reciprocating engine shows the equally marked difference in the foundations. The selling price of turbines to-day is governed largely by that 1 The engine of this unit is really capable of turning out 6000 H. P., although the genera- tor is rated at 3500 K. W. POSITION AND FIELD OF THE STEAM TURBINE. 129 130 STEAM TURBINES. of the competing engines. The manufacturing cost is probably less, but as there are few turbines in the market and compara- tively little competition as yet between them, the competition has' been rather with the old type engines. Any estimate of cost, however, must include the land, buildings, and foundations. These elements vary so with different conditions that no general results can be given. The general adoption of the turbine shows that for large units at least it is the cheaper, and it has practically driven out its rival for large power station work. Mr. Stott, Supt. of Motive Power of the Interborough R. T. Co., New York City, in a paper before the A. I. E. E., makes the following comparison of the operating cost of large engine-driven and turbine-driven units. He says: "The turbine shows more uniform steam rate for varying loads, practically as good economy with saturated steam, and a thermal economy of 6.6% better for superheated steam." The various items are derived from actual costs. Comparison of Charges per K. W. Hr. Maintenance Comp. Cond. Turbine Engine 5000 H. P. 7500 H. P. Engine room, mechanical 2 57 51 Boiler room Coal and ash handling apparatus Electrical apparatus 4.61 .58 1.12 4.30 .54 1.12 Operation Coal and ash handling labor 2 26 2 11 Removal of ashes Dock rental Boiler room, labor Boiler room, oil, waste, etc Coal 1.06 .74 7.15 .17 61 30 .94 .74 6.68 .17 57 30 Water Engine room, mechanical labor Lubrication. ... 7.14 6.71 1 77 .71 1.35 35 Waste, etc Electrical labor .30 2 52 .30 2 52 Relative cost of maintenance and operation Relative investment in per cent 100. 100 79.64 82 52 POSITION AND FIELD OF THE STEAM TURBINE. 131 The general statement that the turbine requires much less oil is borne out in this table, where the turbine requires but 20% of that for the engine generator. In a turbine oil is used only in the bearings, free from contact with the steam, so that the steam is returnable to the boilers as soon as condensed, a property of great value in places where feed water is expensive. Problems. i. Report on the power of plant of Public Service Corp'n. of N. J., described in Power, November 23, 1909, p. 853. References. STEVENS AND HOBART: "Steam Turbine Engineering." MOVER: " Steam Turbines." FRENCH: " Steam Turbines." MURRAY: " Electric Power Plants." WALDRON: Paper, "The Steam Turbine from an Operating Stand- point." Trans. A. S. M. E., vol. xxiv. Art. 27. Relative Fields of Engines and Turbines. The reciprocating engine is superior to the turbine in starting power, in reversing, in operating at varying speeds on inter- mittent and irregular service. It requires less condensing apparatus and gives a higher economy for saturated steam at average vacuums. It can run at moderate speeds and will therefore have a monopoly of many services where the high speeds of the turbine are prohibitive. These considerations indicate that the steam engine will hold its place in locomotive work, hoisting, rolling mills, lighter marine work, and belt and rope driving. Turbines are being applied to centrifugal air compressers and to pumps, with considerable success, but it will be some time before they will be a serious rival in this field or in blowing engines. If the steam engine encounters serious com- petition in the smaller sizes for miscellaneous work it seems much more likely that it will come from the gas engine than from the turbine. The turbine, on the other hand, has demonstrated that it, too, has its field. For direct connected services such as for generators and marine work, especially in the larger units, it has proven 132 STEAM TURBINES. superior to the reciprocating engine. In central power station service it has practically superseded the old type of engine. Up to February, 1911, the sales of the three foremost manu- facturers of large turbines in the United States were as follows : General Electric Co. 2,150,000 K. W. Westinghouse Machine Co. 1,600,000 K. W. Allis-Chalmers Co. 3 50 ! 000 K W - Total 4,100,000 K. W. When it is realized that scarcely a dozen of these machines are 9 years old, and the vast majority not more than 5, one can see the hold the turbine has obtained on this field. About 70 or 75% of the turbines built are for electric light, power, and traction companies. Here the turbine is at its best. It is practically immune from danger from priming. Its uniform velocity and close regulation, its high economy under widely varying loads, its compactness and inexpensive foundations, and its exhaust free from oil, together with the great saving in the generators due to the high speeds, have made it distinctively the central station machine for the generation of alternating current power. In marine service the turbine has also been successful, although its ascendancy here is not so marked as in power station work. Mr. C. A. Parsons has lead the way into this field, and it is largely to his indomitable will that the successful application of the turbine to marine service is due. When he first began his experiments, propeller shafts ran at from 80 to 120 R. P. M., and the slowest turbines at 3,000 to 5,000 R. P. M. The "Turbinia," the first experimental vessel, was equipped with 2,000 H. P. turbines designed for 3,000 R. P. M. It was found that at the designed speed of rotation only 18 knots could be obtained. Experimental investigation developed the exist- ence of "cavitation," a cylindrical vacuum formed around the propeller, causing great loss of power. After clever and extensive experiments propellers were developed on new lines, the power was divided up between 3 turbines driving 3 shafts, the revolutions reduced about 1/2, and a combination of turbine and propeller obtained which gave remarkable results. A speed of over 34 knots, never before reached by any vessel, was obtained. POSITION AND FIELD OF .THE STEAM TURBINE. 133 From that time, 1894, the development has been steady. From the Turbinia, with its 100 feet of length, 9 feet beam and 44 1/2 tons displacement, the turbine has been successfully applied to destroyers, small passenger steamers on the Clyde, the Channel, and Irish Sea services, to the Allan liners, the largest Cunard liners and to the heaviest battleships. The Mauritania and Lusitania have maintained average speeds of over 25 knots, and developed over 70,000 H. P., 70% more power than has ever before been concentrated in one set of engines. The main advantages of the turbine in the marine service are: 1. Absence of engine vibration. 2. Lower center of gravity with consequent greater stability. 3. Less weight of machinery with consequent greater carrying power. 4. Higher average economy. 5. Smaller propellers and therefore, from this and item three, less draft. 6. Lower cost of operation i.e., attendance, supplies, etc. Mr. Parsons, in a paper before the Institute of Marine Engi- neers, speaking of the future of the marine turbine, says: " I think we are safe in predicting that it will soon supersede entirely the reciprocating engine in vessels of 16 knots sea-speed and upward, and of over 5,000 I. H. P., and probably also in- cluding vessels of speed down to 13 knots, of 20,000 tons and up- ward, and possibly still slower vessels in course of time. At present it may, I think, be said that the above most suitable field comprises about 1/5 of the total steam tonnage of the world; but it must be remembered that the speed of ships tends to increase, and the turbine to improve, and so the class of ships suitable for the turbine will increase. Much interest and attention are being centered on the combina- tion of reciprocating engines and turbines for marine work as in the Olympic. The advantages to this combination have been con- sidered in Art. 22, but it seems especially adapted for ships, giving high economy and a greater flexibility than turbines alone. 1 1 "Combination System of Reciprocating Engines and Steaui Turbines," paper by C. A. Parsons, Institution of Naval Architects., Lond., April 9, 1908. Reprinted in London Engineering, April 17, 1908. See also articles on the Engine of the White Star SS "Olympic," London Engineering, Oct. 11, 1910.