The Una-Flow Steam-Engine By Prof. Dr.-Ing. h. c. J. Stumpf Technische Hochschule, Berlin. Translated by the Stumpf Una-Flow Engine Company, Inc., 401 S. A. & K. Building. Syracuse N.Y. Second Edition 1922 SI Copyright 1922 by Prof. J. Stumpf, Berlin THE BA TTLE OF THE ELEMENTS By J. A. STUM PP. (Chaos 1831, Nr. saw what he had made and found it good, wrote a man of noblest mind and mood. No longer with this doctrine is the world content, The doubter does in bitter words lament: One need but cast a fleeting glance at life, What sees one there? True happiness? No, strife, Death, need and misery far and wide, The elements in constant war abide, And storms of passion breeding endless hate Rob life of peace, till many curse at fate. We ask ourselves why we are so surrounded By raw materials and forces of all kinds, On what the longing in our breasts is founded, United with the curious impulse of our minds? As master of the earth, man shall create! All labor does the builder's hand await. To carry out His plan, God made him wise, That he might force and matter utilize. Such is the man whose work success has crowned, Who truth and light by his research has found, Who tested fire's flame and water's might, And thus their deepest secrets brought to light: Who through their elemental strife conceived, Instead of ruin, mankind's gain achieved. Nature's forces he has sought, And under his control has brought. The foes who storm with rage and hate, He keeps by thin walls separate. Around the boiler roars the flame, The seething waves within to tame, Who, in revenge, their enemy to reach, Strive through the prison's walls to force a breach. A polished rod ascends, by magic trained, Propelled by steam within a pipe contained. But lo! into the angry steam so bold, Now pours a rage-appeasing flood of cold; Down slides the rod, but in an instant back Pursued again by live steam in its track. The shining steel glides to and fro And, driving other parts, all show A striving to one goal. The great machine Obeys the master's mind, it may be seen. How many nature's wondrous course deride, And what they do not grasp, they claim unproved, The man of science does regard with pride How parts and whole in best accord are moved. 520223 ERRATA Page 70-71 124 125 125 206 234 234-235 303 Line or Fig. Fig. 1 Line 17 Heading Curve Last line Line 11 Fig. 13 Title Change from Max. Continuous L.H.P. Omit "& T. Hall" Omit "& Hall" Corliss Lead Fig. 14 German State Rys. Omit "Triple speed" To Max. Cont. I.H.P. Counterflow Load Fig. 13 Russian State Rys. TABLE OF CONVERSION FACTORS 1 mm. 1 cm. 1 m. 1 km. 1 sq. cm. 1 sq. m. 1 kg. I metric ton 1 metric ton-km. 1 m. kg. 1 kg. per sq. cm. 1 atmosphere 1 metric H.P. 1 C. Temp, in F. 1 calorie 1 cal. per kg. 1 cal. per I.H.P. hr. (metric) 1 kg. per I.H.P. hr. (metric) 0.039 in. 0.394 in. 3.28 ft. 1.609 miles 0.155 sq. in. 10.76 sq. ft. 2.205 Ibs. 2204.6 Ibs. 0.685 ton-mile 7.233 ft. Ibs. 14.22 Ibs. per sq. in. 0.9863 H.P. 1.8 F. 1.8 X temp, in C. + 32 3.968 B.T.U. 1.80 B.T.U. per Ib. 4.024 B.T.U. per I.H.P. hr. 2.236 Ibs. per I.H.P. hr. GENERAL INDEX Acceleration, valve, 88, 308 Admission, 14 to 16 A. E. Cr., 69 Area, inlet valve, 50 to 58 Area, exhaust port, 60, 69 Auniliary exhaust valves, 46, 47, 167, 177, 208, 225, 264 B Balancing, 72, 297 Bearings, proportions of, 73 Belt, exhaust, 12, 69 Blast, 112, 248 Bleeding, 187 automatic control for, 190 Bonnet, valve, Lentz, 87 Stumpf, 128 Cage, valve, 2, 80, 170 Cam oscillating, 89, 143, 144, 164, 167, 170, 172, 206, 246, 264, 270 reciprocating, 128, 160, 208, 217, 231, 257, 277, 279, 287, 291, 292 revolving tapered, 208, 218, 297, 300 revolving stepped, 271, 293, 294 Compounding, 2, 5, 190, 194 Compression, best length of, 17 to 34 Compressor, 218 combined steam and air cylinder, 213 Condensation, initial, 1 Condenser, 68 jet, 69 surface, 276, 294 Westinghouse-Leblanc, 69 Condensing engine, 68, 126 Connecting rod, 173, 264 Consumption, steam, lowest for various compressions, 17 to 26 lowest for various clearances, 27 to 29 Convection, losses due to, 105 Corliss una-flow engine, 182 Crank, 74, 75 center, 71, 132 Crank pin, proportions of, 73 Crosshead, 171, 294 pin, proportions of, 73 Cut-off, range of, 126 Cylinder, boring of, 93, 155 material of, 92 D Diagram, indicator, 166, 175, 203, 269, 286 Diffusor, 109, 110, 249 Doerfel, 172 Draft, smoke box, 112, 113, 248 Drop, pressure,. 51 Dual clearance, 177 E Eccentric gear, single, 150 Efficiency, mechanical, 71 Engines Ames Iron Works, 160, 161 automotive, 272 blowing, 213, 223 Borsig, A., 245 Burmeister & Wain, 132, 137, 278 compressor, 213, 218 Corliss una-flow, 182 Dehne, A. L. G., 218 Ehrhardt & Sehmer, 137, 139, 143, 195, 196, 202 Elsaessische Maschinenfabrik, 129 Erste Bruenner Maschinenfabrik, 129, 262 Filer & Stowell Co., 178, 181 Frerichs & Co., J., 273 Goerlitzer Maschinenbauanstalt, 144 Gutehoffnungshuette, 209 Harrisburg Fdy. & Mch. Works, 177, 178 hoisting, 207 Hungarian State Rys., 264 Kingsford Fdy. & Mch. Works, 292 Kolomna Engine Works, 225, 234, 237, 238, 262 Linke Hoffmann Works, 218 List, Gustav, 213 Locomotive, 225 Marine, 272 Maschinenbauanstalt Breslau, 232 Maschinenfabrik Augsburg-Nuern- berg (M. A. N.), 144 Maschinenfabrik Badenia, 252, 255 Maschinenfabrik Esslingen, 155 Mesta Machine Company, 206, 223 Musgrave & Sons, Ltd., 143 Neuruppin-Kremmen-Wittstock Ry., 240 Nordberg Mfg. Co., 172, 177 Northeastern Ry. of England, 240 Northern Ry. of France, 234 portable, 252 pumping, 213, 218 Robey & Co., 262 Rolling mill, 139, 195 Schmid, Karl, 294 Schweizerische Lokomotivfabrik, 232 Skinner Engine Co., 168 to 171 Soumy Machine Works, 159 stationary, 127 Stork & Co., 144 Sulzer Bros., 10, 76, 150, 151, 155 Vulkan Engine Works, 225, 227, 240, 274 Worthington Pump & Mchy. Corp., 218 Ejector effect, 110, 112 to 117, 245, 249, 250, 269 saving due to, 117 Expansion, cylinder, 128 Experiments, Prof. Naegel's, 118 to 124 Friction of driving parts, 77 H.P., 75-77 loss due to, 71 single eccentric, 150 valve, cam (locomotive), 230, 245 double-speed, 295, 303 Gooch, 207 Klug, 275, 278, 280, 290 link, disadvantage of, 16 Marshall, 237 Saeuberlich, 274 Skinner auxiliary exhaust, 171 Stumpf, 128, 129 Walsnhaert, 22, 230, 247, 285 Zvonirek, 143, 195 Governor, flywheel, 252 Gueldner, 110 I Inertia, curves of, 72 Insulation, 105 J Jacketing, 27 effect of, 2, 3, 6, 10, 11, 12, 33 Sulzer's tests on, 11 Jackets, proportions of, 12 Joints, piston ring, 94, 95 Lagging, 105 Lap, 128, 202, 207, 240 Lead, exhaust, 5, 61, 108, 239, 269 Leakage losses due to, 79 valve, effect of, 85 Lentz packing, 100 Liner, 160 Locomobile, 252 Locomotive, 225 Borsig, 245 Kolomna Engine Works, 225, 234, 237, 238 Maschinenbauanstalt Breslau, 232 Neuruppin-Kremmen-Wittstock Ry., 240 Northeastern Ry. of England, 240 Northern 'Ry. of France, 234 Schweizerische Lokomotivfabrik, 232 Vulkan Engine Works, 225, 227, 240 pistons for, 99 , three cylinder, 116, 240 Losses friction, 71 incomplete expansion, 108 leakage, '.^ radiation and convection, 105 surface, 1, 4 throttling, 49 volume, 13 Lubrication cylinder, 99, 155, 227, 248, 273 driving parts, 78 valve gear, 87, 128 M Machining of clearance surfaces, 2 of cylinders, 93, 155 of pistons, 93 Marine engines Burmeister & Wain, 278 Frerichs, & Co, J, 273 Kingsford Fdy. & Mch. Works, 292 Schmid, Karl, 294 Vulkan Engine Works, 274 N Nagel's experiments, 118 to 124 Nozzle, 109, 116, 249, 269 Parts driving, proportions of, 71, 73, 75, 76 reciprocating, 71, 72 Packing, piston rod, 100 to 104 Pin crank, proportions of, 73 crosshead, proportions of, 73 Pipe blast, 111, 112, 115, 248, 249 exhaust, 109, 110 steam, 105 Piston cast steel, 78, 227 expansion of, 93 floating, 77, 92 to 94, 99, 155 machining of, 93, 155, 227 overrunning of, 94 radial clearance of, 93, 155 self-supporting, 77, 92, 155 shoes, 77, 92, 155, 240 three-piece, 227 two-piece, 77, 240, 264 Portable engines Erste Bruenner Maschinenfabrik, 262 Hungarian State Rys., 264 Kolomna Engine Works, 262 Maschinenfabrik Badenia, 252, 255 Ports, exhaust, 5, 69 area of, 60, 113 Pressure, back, 29 to 33, 42 ' critical, 34 to 43 Pump, tube, 223 Pumping engines List, Gustav, 213 Worthington Pump & Mchy. Corp., 218 R Radiation, losses due to, 105 Rating, load, 71 Relief, compression, for starting, 203, 207, 229, 230, 240, 247, 275 Resilient valve, calculation of, 81 Rings piston, 94 hammered, 96 overrunning of, 97, 98 proportions of, 96 wear of, 98 Rod, connecting, 72, 173, 263 tail, 77, 93, 206 Rolling mill engines Ehrhardt & Sehmer, 195, 196, 202 Mesta Machine Co., 206 Schlick, 297 Seats, valve, 80, 85, 86 Separators, oil, 69 Series arrangement of parts, 85 Shaft, crank, proportions of, 73 Shoes, piston, 77, 92, 155, 240 Slap, piston, 94 Speed, piston, 71 Stack, smoke, 112 Stationary engines Ames Iron Works, 160, 161 Burmeister & Wain, 132, 137 . Ehrhardt & Sehmer, 137, 143 Elsassische Mascheninfabrik, 129 Erste Brunner Maschinenfabrik, 129 Filer & Stowell Co., 178, 181 Gorlitzer Maschinenbauanstalt, 144 Harrisburg Fdy. & Mch. Works, 177, 178 Maschinenfabrik Augsburg- Nurn berg, 144 Maschinenfabrik Esslingen, 155 Musgrave & Sons, Ltd., 143 Nordberg Mfg. Co., 172, 177 Skinner Engine Co., 168 to 171 Soumy Machine Works, 159 Stork & Co., 144 Sulzer Bros., 10, 76, 150, 151, 155 Stepped cams, 271, 293, 294 Stodola, 110 Strahl, 112 Superheater, 258 Superheating, effect of, 2, 3, 5, 8, 273 Surface clearance, 1 to 3, 7 Surface clearance, minimum, 1 T Temperature steam, experiments on, 118 to 124 high, 8 Tests, engine, 129, 137, 144, 157, 161, 234, 236, 265 Throttling, losses due to, 49 Tube pump, 203 V Vacuum, high, 68 Valves clearance, 45, 46 Corliss, 79, 183 to 186 double beat, leakage, of, 79 locomotive, 230 resilient, calculation of, 81 tightness of, 79, 80 two-piece, 168 exhaust, auxiliary, 46, 47, 166, 167, 177, 208, 225, 264 automatic, 166, 264 inlet, area of, 50 to 58 leakage of, 85 piston, 79, 177, 178, 196, 210, 239, 240, 275 single beat, 86, 245, 247, 262, 272, 295, 303 slide, 79 locomotive, 237 Valve spring, calculation of, 88, 308 Volume clearance, additional, 44 to 46, 126 clearance, % of, 48, 126, 127, 228 w Walschaert gear, 22, 230, 247, 28.5 Westinghouse-Leblanc, 69 z Zeuner, 111 Zvonicek gear, 143, 195 Preface. The second edition of this book represents a complete revision of the first one, little of which remains. The first edition contained a good many opinions in addition to facts and was intended rather for the purpose of defending the una- flow engine against antiquated theories and attacks. In the meantime the una- Fie. 1. flow principle has been widely tried out and scientifically investigated. It has become an accomplished fact and is in common use. This book therefore contains scientific proof from an objective point of view, as well as a description of the development of the una-flow engine. In the opening chapters of the book the different losses of the steam engine are investigated. The causes and effects are defined, as well as the relations be- tween them, and the manner is pointed out in which the minimum value of each loss is obtainable. After considering all the seven different losses occurring in a steam engine, the question is asked as to how a steam engine must be designed in order to have a minimum total of all the seven losses. In answer to this query two different designs are presented, one being a stationary una-flow engine with single-beat valves for condensing operation, and the other a una-flow locomotive also equipped with single-beat valves. On the former, the use of single-beat valves was made possible by the use of a double speed valve gear (see Chapter VI). Both types of engines were developed during the year 1920. Lower steam consumption figures than those given by the best multi-stage engines are Fig. 2. obtainable with these types for both saturated and superheated steam. The expe- rience gained with una-flow engines in widely different fields and under the most varying conditions was utilized in the design of these engines to the fullest pos- sible extent. A number of Chapters is devoted to the description of this development in all its phases. The novelty in this case is the single-beat valve, which has so far been used only in internal combustion engine practice. All previous attempts to apply it to steam engines have miscarried. The application of this type of valve to the una- flow engine represents figuratively the keystone in the development of the latter. The fundamental conformity between the new una-flow engine and the two-stroke internal combustion engine is surprising. This refers to the uni-directinal flow, the single stage expansion, the piston-controlled exhaust and the single-beat inlet valve. ?! Surprising also is the close agreement in the essential parts of the cylinder between the latest design shown in Fig. 3 (see Chapter VI) and the first original sketch of a una-flow engine made in the year 1900 and reproduced in Fig. 2, which also incorporates single-beat valves. Since, as Descartes says, doubt may be considered the origin of every philo- sophy, the question regarding the doubt which originated the una-flow philosophy may well be asked. This doubt arose in the year 1896 during the starting up of two pumping engines designed by myself for the Pope Mfg. Co., of Hartford, Conn. (Fig. 1). These were vertical triple expansion engines with Corliss valves and a central condensing system, in which everything then considered good practice was carried to the extreme limit. This resulted in a very complicated construction which appeared to me to be a sign of weakness. The doubts which then arose in my mind eventually led to the sketch shown in Fig. 2 during the year 1900. The construction of steam turbines of several stages, which began at that time, Fig. 3. was developed along the lines of pure uni-directional flow, arid this brought up the question whether it would not be possible to raise the reciprocating steam engine to the same thermal plane as the turbine by the use of the una-flow prin- ciple. The application of the una-flow action of the turbine to the steam engine, although in a somewhat imperfect manner, by properly designing the cylinder, valve gear, steam jackets and condenser connection, etc., finally led to the una- flow engine with single-beat valves as shown in Fig. 3. The object in view was the attainment of the minimum total of all the seven different losses of the steam engine, as well as the utmost simplicity and reliability of operation. This goal now seems to have been fully reached. The fact that the una-flow engine pos- sesses the uni-directional flow in common with the steam turbine and has a con- structional basis similar to that of the two-stroke internal combustion engine, may be cited in support of this. The design and adaptation of the una-flow engine to different requirements and conditions of service represents an immense amount of work, in which I received the full support of my assistents as well as that of Mr. Rosier of Miihlhausen, Alsace, Mr. Arendt of Saarbriicken, Prof. Bonin of Aachen, Mr. Dutta of London, and Drs. Mrongovius and Meineke of Berlin. To the splendid support of the last four gentlemen may be attributed the positive developments of the chapters on volume loss, throttling loss, exhaust ejector action, the una-flow locomotive, and the valve gear with double speed lay shaft. I am particularly indebted to those gentlemen who took up my proposals at a time when no one would yet believe in the una-flow engine, namely, Prof. Nol- tein of the Technical Hochschule at Riga, Messrs. Hnevkovsky and Smetana of Brtinn, Mr. Lamey of Miihlhausen, Alsace, Mr. Mtiller of Berlin, and Mr. Schiller of Grevenbroich. Berlin, January 1921. J. Stumpf. Index. Page Preface V I. Steam Engine Losses 1 1 a. Losses due to Cylinder Condensation 1 1 b. The Una-Flow Arrangement as a Means for reducing Surface Losses. Jacketing of the Cylinder 4 2 a. The Influence of the Clearance Volume upon the theoretical Steam Con- sumption (Volume Loss). Critical back Pressure 13 2b. Additional Clearance Space 44 3 a. Losses due to Throttling. Determination of Inlet and Exhaust Port Areas 49 3b. Relation between the Una-Flow Engine and the Condenser * . . . . 68 4. Lossesdueto Friction (mechanical Efficiency). Dimensioning of drivingParts 71 5. Losses due to Leakage. Valves, Pistons, Piston Rod Packings .... 79 6. Losses due to Radiation and Convection 105 7. Losses due to incomplete Expansion. The Exhaust Ejector Effect . . 108 8. Prof. Nagel's Experiments 118 II. 1. The Una-Flow Stationary Engine 126 2. The Una-Flow Corliss Engine 182 3. . The Una-Flow Engine arranged for Bleeding 187 4. The Una-Flow Rolling Mill Engine 195 5. The Un^-Flow Hoisting Engine 207 6. Una-Flow Engines for driving Air Compressors, Pumps, etc 213 III. The Una-Flow Locomotive 226 IV. The Una-Flow Locomobile and Portable Engine 252 V. The Una-Flow Marine Engine 273 VI. The Una-Flow Engine with single-beat Valves and double-speed Lay Shaft 303 Summary 316 I. Steam Engine Losses. The losses in a steam engine may b.e classified as follows: 1. Losses due to cylinder condensation (surface loss). 2. Losses due to the volume of the clearance space (clearance volume loss). 3. Loss due to throttling or wire drawing. 4. Friction loss. 5. Loss due to leakage. 6. Loss due to heat radiation and convection. 7. Loss due to incomplete expansion. > la. Losses due to Cylinder Condensation (Surface Loss). The amount of initial (or cylinder) condensation (termed surface loss in the following) is determined by the size, kind and arrangement of the harmful sur- faces, by the steam jacket, by the quality of steam passing these surfaces, by the temperature gradient and the number of stages used, by the amount and period of the steam flow and the path of the steam through the cylinder (counterflow or una-flow). Initial condensation, which is usually over by the end of admission, is caused by the clearance surfaces and increased by any moisture carried over with the steam. The ability of these surfaces to receive and give off heat forms a kind of heat bypass, with a corresponding loss to the cycle. Part of the steam condenses during admission and re-evaporates during exhaust and the last part of expansion. The harmful surfaces comprise the inner surfaces of the cylinder and* the inwardly exposed surfaces of piston, piston rod and steam distributing parts. Surfaces which are continually exposed even in the dead center position of the piston may be termed harmful surfaces of the first order, and those which are progressively uncovered by the piston during its motion, harmful surfaces of the second order. The former usually cause the essentially greater part of the surface loss. Since the amount of surface loss is determined by the extent of the harmful surfaces, the latter should be kept as small as possible and should also be machined. A good many designers pay attention only to the amount of clearance volume without considering its surface. The minimum harmful surface of the first order comprises an area equal to twice the cylinder cross-section (cylinder head and piston), and it is convenient to express the additional surface of the first order in percent of this minimum surface. In actual engines these additional surfaces, which are mostly not machined, are found to be from 150 to 200% of this minimum harmful Slump/, The una-Flovv steam engine. surface, although it is possible by careful design to reduce tiiis figure to 3 or 5%. Piston valves with snap rings working in separate bushings, as well as slide valves with long curved ports, which latter are in most cases left rough and serve for both steam admission and exhaust, have large surfaces which are especially harmful on account of their very nature and arrangement. Engines with separate admission and exhaust ports are far better in this respect, because the latter are usually very short and the hot admission and cold exhaust steam enter and leave the cylinder through separate passages, thereby avoiding the alternate heating and cooling of these surfaces and the corresponding surface loss which takes place in engines having common inlet and exhaust ports. Slide and piston valve engines with their perpetual reversal of flow are sub- ject to extensive turbulence and heat exchanges, although careful design can generally improve conditions. The Corliss engine may be considered an improve- ment by reason of .the smallness and different arrangement of its additional sur- faces; and, with the valves in the heads, the ports are straight and short with inlet and exhaust separated. Conditions can be improved still further if the exhaust valve, which forms the major part of the additional surfaces, is made to fill its bore completely and has a straight port through its center only. Poppet valve engines in general have rather large additional surfaces, espe- cially where valve cages are used ; and notwithstanding separate inlet and exhaust ports the frequent flow reversal has a deleterious effect. Valve cages considerably increase the additional surfaces. Machining of the latter may be provided for in many cases by clever design, thereby reducing their extent and the corre- sponding turbulence and surface loss. Further means of reducing the surface loss are:' 1. Jacketing, 2. Compounding, 3. Superheating, 4. The una-flow system. Jacketing of the harmful surfaces is a further step in reducing surface losses. The heating medium is usually steam, seldom flue gases. Engines with great sur- face losses will be largely benefited by jacketing, and single cylinder condensing engines working with saturated steam will show the greatest gain since they offer the largest scope for improvement. The effect of the jacket is diminished if the expansion takes place in two, three, or four stages, and if the steam is superheated in addition. These means improve the thermal condition to such an extent that there is little to be gained from jacketing. This applies to superheating, espe- cially if the whole working cycle takes place in the range of superheat. Locomo- tives, in which the steam, when leaving the cylinder is still superheated, will derive no benefit from jacketing. Superheating is such a far reaching remedy that the number of stages in counterflow engines working with superheated steam has been generally reduced from three to two for condensing operation, and from two to one when operating non-condensing. Increased speed, later cut-offs, and larger units tend to reduce the surface losses and hence the effects of jacketing. High speed damps the temperature fluctuations of the walls; the late cut-off raises the mean wall temperature; and the larger size gives a more favorable ratio of volume to surface. For these reasons a large, fast running, and heavily loaded engine will show the least gain from jacketing, especially if working with superheated steam or a small temperature gradient. All this applies to superheated steam locomotives as an example. Jacketing has the greatest effect in low pressure cylinders, since surface losses, temperature gradient, and jacket surfaces are large and the weight ratio of jacket to working steam is the most favorable. The gain from jacketing is accordingly smaller in intermediate and high pressure cylinders, and in many cases there is hardly any in the latter. Similar conditions prevail in two cylinder compound engines. Head jacketing is usually more effective than cylinder jacketing because the cylinder surfaces are temporarily covered by the piston, and the oil film acts as a heat insulator. These surfaces of the second order cause small sur- face losses and consequently show a smaller gain from jacketing. Saturated steam is very bad in this respect because the water particles act as heat conductors and increase the surface losses. Dry steam is better, and best of all is superheated steam. Saturated steam is an excellent, and superheated steam a very poor heat conductor. The action of the cylinder becomes the more adiabatic the more the superheated region extends through the cycle. Superheat, furthermore, by increasing the specific volume, makes the steam lighter and reduces both the weight per cycle and the surface loss. The commonly prevailing amount of superheat allows non-condensing engines to work in the superheated region throughout the cycle; but this is not the case with condensing engines, assuming proper ratios of expansion in both cases. In condensing engines the low pressure part of the cycle always extends beyond the saturation point. Superheating is of such far reaching effect that the reduction of harmful surfaces, their arrangement, and in many cases even jacketing, lose their impor- tance. Generally speaking, among the different ways of reducing surface losses, one or the other may be so effective that there is nothing left for the remaining ones. The amount, extent and kind of steam flow, especially of the exhaust, may have considerable influence in engines working with saturated steam. Wet exhaust steam flowing with high velocity through long unfinished ports having large sur- faces may cause great surface losses. Summing up, it may be stated that the application of the different means for reducing initial condensation resulted in the common use of the two-cylinder compound engine for condensing service, for the reason that the low pressure part of the cycle takes place in the saturated region. The una-flow principle, however, permits the use of the single cylinder single-stage expansion engine for this ser- vice, since the uni-directional flow of steam eliminates initial condensation despite the fact that a part of the cycle takes place in the saturated region. The una- flow principle is also of great advantage for non-condensing and multi-stage engines. 1* Ib. The Una-Flow Arrangement as a Means for Reducing Surface Losses. The una-flow engine, as its name indicates, utilizes the steam energy by a um-directional flow, i. e. the steam passes through the cylinder always in the same direction. As shown in Fig. 1, the steam enters the cylinder head from below, ; heats the surface of the latter, and then enters the cylinder through the inlet valves located in the top portion of the head. Doing useful work, the steam follows the piston and after having expanded, leaves through ports at the opposite end of the stroke, i. e. in the middle of the cylinder; the opening and closing of these ports being accomplished by the piston during its motion. This is in marked contrast to the ordinary or counterflow engine, where the steam enters at the end of the cylinder, follows the piston during the working stroke, and, returning with the piston, leaves at the cylinder end. The result of this kind of flow is an intensive cooling action upon the clearance surface, the exhaust steam being usually wet and thus an excellent heat conductor. The consequence is increased initial or cylinder condensation during the following admission. The una-flow principle avoids the cooling of the clearance surfaces, thereby eliminating initial condensation to such an extent that compounding becomes superfluous. Una- flow engines may, therefore, be built with a single cylinder and single-stage expan- sion, and yet show the economy of compound or triple-expansion engines. The exhaust ports of the una-flow cylinder have an area about three times as large as can be realized with slide or poppet valves, with a consequent complete pressure equalization between cylinder and condenser if long and restricted pas- sages between them are avoided. In other words, if the condenser is placed close to the cylinder and the connection is of ample area, then complete equalization of pressures is assured. In order to get a clear conception of the magnitude of this port area one has to consider that the engine piston acts as a piston valve and the crank as eccentric, while the cylinder constitutes the valve bushing. The exhaust lead is usually taken at 10%, which fixes the compression at 90%. The una-flow exhaust ports do away with separate exhaust valves and their leakage loss, their additional clearance volume and surface, as well as the neces- sary valve gear. The elimination of exhaust valves is therefore an additional advantage of the una-flow construction. Indicator cards of una-flow engines show adiabatic lines for expansion and compression. Adiabatic expansion results in considerable moisture even with highly super- heated steam. The entropy chart shows at a glance that with an initial pressure of 12 at. and a temperature of 300 C an expansion to 0.8 at. abs. produces 7% moisture. During the exhaust period this expansion continues until at a terminal 'pressure of 0.1 at. abs. the moisture amounts to 17%. On account of the un- dvoidable heat losses, the temperature at the end of admission will be somewhat less than the above, with a consequent increase in the final moisture. The extension of the working cycle into the wet region rendered compound engines necessary, the high pressure cylinder working with superheated, and the low pressure cylinder with saturated steam. The una-flow system made a return to the single stage engine possible for both superheated and saturated steam. Superheated steam is a very effective means of combating initial condensa- tion. The combined use of superheat and the una-flow construction will still better conserve the heat during admission. Expansion will therefore start at a higher temperature and terminate with less moisture; or in other words, better economy will result. The head jackets have the effect of a partial regeneration of the steam during expansion and exhaust. On account of the large difference of temperature, the large heating surface, and the great difference in the specific weights of the live steam and the exhaust or compression steam, an intensive heating action takes place during expansion, exhaust, and compression. This affects particularly that part of the steam located close to the cylinder head, while in consequence of the adiabatic expansion that part following the piston will sustain both a drop in tem- perature and an increase in moisture. The greater part of the moisture produced will accordingly be found close to the piston head with a progressive dryness and 6 an increase in temperature towards the cylinder head. This moist steam, being close to the exhaust ports, will escape first when they are uncovered, while that part which received heat from the head jacket is trapped by the returning piston and compressed along adiabatic lines, partly as saturated, and partly as super- heated steam, but mostly the latter, because during the early part of compression heat is still being transmitted to it from the head jacket. This elimination of liquid condensate avoids its deleterious effect of heat exchange as well as damage due to water hammer. Tests conducted on triple-expansion engines in regard to the action of steam jackets have shown that there is no gain in high pressure cylinders, very little in intermediate cylinders, but a large gain in low pressure cylinders, despite the large heat losses prevailing in cylinders of the usual type. Owing to the counter- flow action, a great part of the jacket heat is necessarily carried by the exhaust steam into the condenser. One has only to consider that at the time of exhaust valve opening a considerable amount of pressure energy is available to exhaust the steam with velocities as high as 350 to 400 m. per sec., that the steam with this velocity impinges on the clearance surfaces, depositing water, and that on account of the sudden drop in pressure the heat absorbed by them during admis- sion starts an intensive re-evaporation which extracts considerable quantities of heat from these surfaces. The fact that the latter are in many cases jacketed presents a most unfavorable picture of the very inefficient way in which admission and jacket heat are utilized. From the point of exhaust valve opening up to the beginning of compression the jacket heat is carried into the condenser in the most wasteful fashion. During the remaining part of the cycle, heat exchange occurs under more unfavorable conditions and at lower velocities, but notwithstanding the immense losses of jacket heat the low pressure cylinder derives the greatest benefit from steam jackets. This may be explained by the fact that the low pres- sure cylinder has the greatest temperature differences, the greatest heating sur- faces, the greatest surface losses, and a very favorable density ratio of jacket to working steam. It follows that una-flow cylinders necessarily have a particularly energetic heating action, since, as in the case of low pressure counterflow cylinders, heating takes place under the influence of the full temperature difference, the large surfaces, and the great difference in density of jacket and working steam. The counterflow of steam, with its great losses, is replaced by the uni-directional flow where no jacket heat is lost to the exhaust. As shown in Fig. 2, exhaust steam in a una-flow cylinder never passes heated surfaces. The layer next to the cylinder head, at the very worst, approaches the exhaust ports without leavipg the cylinder, and therefore hardly any jacket heat can be lost. The beneficial effect of steam jackets proved to e-xist in low pressure counterflow cylinders must therefore be in evidence in a high degree in una-flow cylinders, since a loss of jacket heat is avoided. It is assumed, of course, that jacketing is limited to the cylinder head and that the cylinder is unjacketed (Fig. 1 and 2). The head jacket extends to the point where cut-off normally occurs so that the clearance surfaces of the first order are effectively heated from the outside, while the highly superheated com- pression steam prevents cooling from the inside. A further reduction of surface losses can be accomplished by enlarging the jacket surfaces and by the utmost reduction and machining of clearance surfaces. Even though the clearance surfaces in a una-flow engine are exposed during the whole cycle, they are not subject to the rush of exhaust steam passing them, nor to re-evaporation; and they therefore suffer little cooling on account of the com- parative tranquility of the steam molecules adjacent thereto, and furthermore have the benefit of both the jacket and compression heat. All these causes combined with the una-flow principle and una-flow construction produce an almost adia- batic action of the clearance surfaces. Fig. 2. The una-flow engine fundamentally avoids the thermal mixup characterizing the counterflow engine. The cylinder is composed of two single acting cylin- ders with the exhaust ends in common. On account of the long piston the stroke volumes of both cylinder ends are relatively displaced for a distance equal to the length of the piston. The two inlet ends are hot and remain hot, while the common exhaust is cold and remains cold, and the temperature changes gradually from the hot inlet to the cold exhaust. The jacketing, comprising the heating cham- bers at the inlet ends and the cold exhaust belt around the common exhaust, is in perfect accord with this. In a counterflow cylinder the two stroke volumes overlap. The exhaust end of one side more or less reaches to the admission end of the other, according to the length of the piston. From a thermal point of view the arrangement is obscure. The central exhaust port and belt in the middle of the cylinder effectively cool that part of it at which the piston attains its highest velocity. This favorable 8 action is further augmented by the omission of jackets on the adjacent part of the cylinder. Furthermore, the piston on account of its large area exerts a very low unit pressure. The cylinder is of very simple form and can be kept free from badly distributed material, thus avoiding local heating and warping. The large bearing surface of the piston, the cooling action of the exhaust belt, as well as the simplicity of the cylinder, which precludes the possibility of warping even at the highest steam temperatures, render the use of a tail rod unnecessary, provided the material used is suitable, the design correct, and proper lubrication supplied. (See una-flow locomotives and engines built by Sulzer Bros.) The piston has two sets of rings, comprising four or six altogether. During the period of high pressures both sets are in action, and the pressure has dropped to about 3 at. when one set has overrun the exhaust ports. The many una-flow engines in operation are proof that the piston is not the cause of any difficulties even with the highest degrees of superheat, and that with good workmanship a piston can be made perfectly tight. If, however, a cylinder should become scored, its simplicity allows it to be easily replaced without great expense. There can be no doubt about the possibility of employing, in una-flow engines of the kind described, steam temperatures far in excess of anything used at present. Even with the highest initial temperatures the cycle extends far into the moist region, thus insuring moderate working temperatures for cylinder and piston, the highly superheated steam being limited to the cylinder head. The una-flow engine, therefore, opens up further possibilities of development in the utilization of higher' degrees of superheat. It is all the more suitable for this because the superheat benefits the whole cycle, while in counterflow compound engines the high pressure cylinder receives too much and the low pressure cylinder too little superheat. This feature of the una-flow engine does not, however, contradict the fact that it is also suitable for saturated steam, excellent results being actually obtained with both kinds. The una-flow engine has the uni-directional flow, the hot inlet and the cold exhaust in common with the steam turbine. From a thermal point of view it forms the missing link between the reciprocating steam engine and the steam turbine. The opinion is frequently advanced that, in regard to their thermal action, the cylinder head and piston surfaces of a una-flow engine are merely interchanged. This leaves out of consideration the fact that the piston surface is protected against the action of the exhaust by a cone of stagnating steam. The existence of this phenomenon has been frequently proved beyond dispute in the case of air and water. The surface corresponding to the face of the slide valve of a counterflow engine is located in the una-flow engine at the circumference of the cylinder. The port necessary with a slide valve, together with the clearance and clearance sur- faces involved, are completely avoided in the una-flow engine. The surfaces of the una-flow exhaust ports are outside of the cylinder and therefore have no bearing upon the thermal conditions. Piston and steam have a different velocity only after the former has opened the exhaust ports, when the available pressure energy is transformed into kinetic energy. The steam, however, attains its full velocity only after it has passed the edge of the piston and therefore the cooling action of this steam upon the piston can only be small. The cooling action caused 9 by the low velocity of approach in the cylinder now remains to be examined. Considering the exhaust ports as nozzles bounded on one side by the piston edge, the steam at this narrowest point attains a mean critical velocity of about 410 m/sec. On examining cross-sections ahead of this smallest section or throat, very moderate velocities are found. Fig. 3 makes these conditions clear for a cylinder of 600 mm bore and 800 mm stroke, the piston being assumed to have overrun the ports for a distance of 20 mm. Calculation of the velocity at this narrowest point shows it to be 410 m/sec., while at a distance of 15 mm ahead of this point it is only 71 m/sec.; at 40 m it is 36 m/sec.; at 85 m it is 20 m/sec., and at 130 m it is 15 m/sec. It should be noted that in considering the various cross-sections a reduction of area due to the bridges has been assumed at the narrowest section (throat), and this does not apply to the cross-sections of approach. As the piston progressively uncovers the exhaust ports the proportions of the cross-sections are changed in such a way that the velocity of approach is increased. At the same time a reduction of pressure occurs, the effect of which is to reduce the velocity of approach, beginning at the point at which the critical pressure is reached. Most indicator cards of una-flow engines show clearly that when the piston reaches the dead center the pressure inside the cylinder has dropped to the back pressure; or in other words, the greater part of the steam has in this position been exhausted. Therefore, owing to the short duration and low intensity of the flow of exhaust steam along the piston surface and the small harmful exhaust surfaces, the resulting cooling action is small. The whole cylinder section acts as an approach to the exhaust nozzles. A further protection is given by the layer of stagnating steam, and there is finally a very intense heating action during compression and admission. The 10 heating of the piston surface by the hot live steam is so effective that this surface acts almost adiabatically during the following exhaust period. The very favorable steam consumption figures obtained with this type of engine are further proof that a simple interchange of the cylinder head and piston surfaces, in regard to their thermal behaviour, is out of the question. Such favorable economy can only result if the piston and its cooling action is negligible. This is further confirmed by tests of Prof. Nagel. A test with saturated steam of 184 C, and a cut-off at 12%, showed that the temperature of the piston surface at a point near its circumference was about 164.5 G. The total fluctuation at this point was only 1.3 G. This surprisingly high temperature, and moreover the small fluctuation, justifies a very favorable conclusion. These figures should be still better towards the center of the piston surface. The thermal, constructional, and operative advantages of this type of prime mover are such that in continuous operation the economies of compound and triple-expansion engines can be obtained with both saturated and superheated steam. Jacketing of the Cylinder. The firm of Sulzer Brothers, of Winterthur, Switzerland, constructed an ex- perimental engine of the una-flow type after such engines had been put on the market by the Erste B runner Maschinenfabrik Gesellschaft, who were the first to take up the manufacture of una-flow engines on the author's recommendation. Fig. 5. Sulzer Bros, entrusted the author with the design of the first una-flow cylinder of 600 mm bore, 800 mm stroke, and 155 r. p. m. All later Sulzer engines are built with only slight changes from this design, which is shown in Figs. 4 and 5. According to Sulzer Brothers' usual practice the engine was put on the testing floor and a great number of tests were carried out with the object of observing 11 its performance and determining its economy under the most varying conditions. An important part in the program was the study of the effect of the jackets. For this reason the author incorporated in this design not only head jackets, through which the live steam had to pass before entering the cylinder, but also jackets at the ends of the cylinder barrel, which were separated by a neutral zone from the exhaust belt. These cylinder jackets could be shut off separately. During the tests the head jackets were necessarily always in operation, but the cylinder jackets were either in service or shut off, as indicated in Fig. 6 by the words ,,with jacket" and ,, without jacket". The tests proved that the effect of the cylinder jacket decreases with increasing steam temperature. With saturated steam the difference was almost 1 kg/I HP-hour in favor of cylinder jackets, while it was barely 0.5 with steam of 265 C and only 0.2 kg with steam of 325 C. All figures refer- red to the most economical M. E. P. The steam pressure was 9.2 at. gage and the vacuum 66 cm. For the point of best operating economy, i. e., an M.E.P. of about 3 kg/sqcm, these differences change in such a way that a small increase results when running with cylinder jackets and with a steam tempe- rature of 325 G, while steam of 265 G shows a small decrease in steam consumption. For a steam temperature of kg/cm* 325 C the point of equality of steam consumption, when operating both with and without cylinder jackets, is found to correspond to an M.E.P. of 2.5 kg/sqcm. The corresponding point for steam of 265 G occurs at an M.E.P. of 3.4 kg/sqcm. For saturated steam this point moves towards a still higher M.E.P. This explains the customary omission of cylinder jackets for superheated steam. It is also noteworthy that the best economy with steam of 325 G very clo- sely approaches the value of 4 kg/I HP-hour. The results for saturated steam, especially with cylinder jackets, are extremely favorable. It must be considered, however, that the saturated steam had a very slight degree of superheat in order to make sure that it was actually dry. It should further be noted that this engine was well designed and well built. The clearance volume and clearance surface (the latter being machined) were moderate, and the whole engine was built with the high precision usual to Sulzer Brothers' shop practice. The measurements were made by means of a surface* condenser, which, as is well known, gives slightly lower but more accurate results than boiler feed water measurements. Comparing these curves with those of compound or triple-expansion engines, it will be found that the steam consumption of the una-flow engine is influenced 12 Fig. 7. by the load to a much smaller degree. This is shown particularly by the curves marked ,, without jackets'"', where there is hardly any change in the steam con- sumption between mean effective pressures of 1 and 3 kg/sqcm, especially with high superheat. Even with saturated steam little change is noticeable between mean effective pressures of 1 and 2,4 kg/sqcm. The curves of Fig. 6 justify the following conclusions in regard to cylinder jacketing: For highly superheated steam (300 C and more) and high mean effective pressures cylinder jacketing is useless (Fig. 7), and has a deleterious effect even at as low an M.E.P. as 3 kg/sqcm. For low mean effective pressures cylinder jackets may yet be expected to yield a small gain. For moderate steam temperatures of about 250 G a short cylinder jacket as shown in Fig. 8 is advisable. It will slightly im- prove the economy even for a high M.E.P. and produce considerable gain for a low M.E.P. With saturated steam or low de- grees of superheat a cylinder jacket separa- ted only by a narrow zone from the exhaust belt (Fig. 9) should be used under all circum- stances. Head jackets are essential in all cases. The separation of cylinder jacket and exhaust belt is advisable in order to avoid unnecessary loss of jacket heat and to pro- vide more favorable operating conditions for the piston, especially when using super- heated steam. The center part of the cy- linder, where the piston attains its highest speed, has the lowest temperature. Ex- cellent results can be secured with self- supporting pistons, even with very high temperatures of superheat, if the designer pays attention to these thermal conditions by providing the piston with bearing sur- faces at its center only, leaving its extremities to project in the manner of a plunger towards both ends of the cylinder without actually touching the walls. The head jackets do not impair the operation of the piston, since no rubbing surfaces are in contact with the former. Their heating action is very effective because they are continuously in contact with the working steam ; they heat harmful surfaces of the first order, and with proper lubrication the transmission of heat from them is not impeded by an oil film (Fig. 7). There is practically no loss of jacket heat to the exhaust. Conditions are not so good in Fig. 8 and still worse in Fig. 9. In these constructions the amount of jacket heat lost to the exhaust increases more and more because the exhaust steam to a certain extent flows past heated surfaces, although these may be partly protected by an oil film. Fig. 8. 13 2 a. Influence of the Clearance Volume upon the theoretical Steam Consumption (Volume Loss). (The Una- Flow System as a Means for Reducing the Volume Loss). A certain amount of steam admitted per stroke into a cylinder with clearance will produce an indicator card of less area than in an ideal cylinder without clea- rance. The difference in area may be termed volume loss. This volume loss is represented in Figs. 1 to 4 for different conditions. The diagrams corresponding to the cylinder with clearance are drawn in heavy lines, while those for the ideal cylinder are shown dashed. Volume losses can be expressed absolutely or relatively. In most cases it is convenient to figure the volume loss in per cent of the engine output. u-^ H Fig. 1. A comparison of the two areas AO PG and ABPG in Fig. 1 shows area BOP to be a loss. The parts of the diagrams lying below the line G P show a loss of the area GES and a gain of area PCV = GFQ T. By subtracting the latter, the resultant loss is represented by the area TQFES. In Fig. 2 the admission has been lengthened until the points F and .P of Fig. 1 coincide with the beginning and ending of the diagram. Consequently the shaded areas BOP and GES represent the loss for the diagram with clearance. In Fig. 3, with still longer admission, a comparison of the diagrams shows the loss to be equal to the shaded areas BOVC and GHS. 14 The reduction in diagram area produces a corresponding increase of the loss due to incomplete expansion. A special case is shown in Fig. 4, which represents a diagram without volume loss. This may occur for instance in high pressure cylinders of compound engines when expansion is carried to the back pressure and compression to the initial pressure. Although the diagram has no direct volume loss, the stroke volume of the cy- linder with clearance must be increased from F! to F 2 , with a consequent increase in the surface loss. \d O \r Fig. 3. Fig. 4. The diagrams clearly disclose the fact that the compression is a means of reducing the volume losses, since they would be essentially larger without com- pression. It is also clear that for a constant admitted steam volume cp the point of cut-off will vary with different lengths of compressions. There will therefore be a certain position of 9? for which the diagram area produced will be a maximum and the corresponding volume loss a minimum. \ X^ Fig. 5. a) Determination of the best position \ \ of 9?, if pv = constant, and /? 1? /? 2 , cp and ^ S Q are known (see Fig. 5). Diagram areae F = ABEDRF = = h + h-h-h=ABHG + BEKH FRJG RDKJ. \\ 15 The admitted steam volume

)pilog e = Pi(V E S ) + V E /?! loge -jr~ (VE i + Pi lo ge ~y -- V E -p 1 - v - - p l log e -- *- y E V E V H. *->0 Pz B 99 Dj A H

The quantities F c , p e , Zl i e , ^ i c and p fc can be obtained directly from the Mollier chart. The values plotted in the following diagrams were obtained by this method. In Figs. 10, 11, 12, 13, 14 and 15 the steam consumption for an initial pressure of 14 at. abs., 1 at. abs. back pressure, superheated steam of 300 G and clearance volumes of 5, 8 and 11%, has been plotted against the length of compression A;, for various values of constant values e and constant M. E.P. The construction of auxiliary diagrams (Fig. 9) is recommended for determining the curves of M. E.P., the abscissae representing M.E.P., and the ordi- nates the steam consumption. These curves are plotted for constant compression, and a vertical line intersecting them determines the values of the steam consumption for a given constant mean effective pressure. Meow e pir Fig. 9. 21 The great influence exerted by the compression upon the steam consumption is shown in Figs. 10, 11, 12, 13, 14 and 15. As indicated by previous results, the Fig. 10. Saturated steam 14 at. abs. Noncondensing, Clearance space 5/ . Fig. 11. Saturated steam 14 at abs. Noncondensing, Clearance space 8/ . best economy is obtained with long compressions for early cut-offs e and small mean effective pressures p t and vice versa. It will also be noticed that link valve gears and valve gears controlled by shaft governors acting upon both inlet and 22 exhaust have approximately correct variation of compression (Fig. 14). Valve gears operating with fixed compression can rightly be used in connection with small clearance volumes. Fig. 14 is especially interesting, since it also con- tains the curve of compression obtained with the standard Walschaert gear of the German State Railways for an ex- haust lap of 3 % mm. It is surprising to. find that this curve agrees closely with the calculated minimum values of the compression for constant M. E.P. The shortest cut-off of the link valve gear necessitates a large clearance vo- lume, the bad effect of which is only partly neutralized by a correct variation of compression. It would be more im- portant, however, considerably to reduce the clearance volume required by this type of gear, since the bad effect of large clearance volume far outweighs the correcting influence of the com- pression. The best steam consumption for constant mean effective pressure p t on the one hand, and constant cut-off e on the other, occur at different lengths of compression. If, as it must be, p t is considered the governing variable, then shorter compressions are arrived at. The smaller the clearance volume is, the shorter will be the compressions at which both these minima occur. Figs. 16 and 17 explain the different sets of curves more clearly. In Fig. 17 are repeated two curves from Fig. 15, one being a constant M. E.P. curve for Pi = 10 at., and the other a constant cut-off curve for e = 50%. The diagrams in Fig. 16 marked A (9% compression), and B (76% compression), have the same theoretical steam consumption of 8 kg/HP-hour, the cut-off being in both cases at 50% of the stroke. Dia- gram C, also with a cut-off at 50% but with 47% length of compression, has a steam consumption of only 7.85 kg, while still another diagram D, having the same M. E.P. (10 at.) as diagram C, but with a cut-off at 44% and a Fig. 12. Saturated steam 14 at. abs. Noncondensing, Clearance space 11 / . 23 compression of 19%, only has a steam consumption of 7.6 kg. This last diagram is represented by the lowest point of the curve for constant M.E.P. (Fig. 17). At this point D, however, the constant M.E.P. curve is intersected Fig. 13. Superheated steam 300 C, 14 at. abs. Fig. 14. Superheated steam 300 C, 14 at. abs. Noncondensing, Clearance space 5/ - Noncondensing, Clearance space 8/ . by a curve of constant cut-off at 44%. The lowest point of the latter in turn is intersected by another M.E.P. curve which also has a minimum. By following from minimum to minimum along the M.E.P. and cut-off curves a point is 24 10 finally reached at which the .two minima coincide. This point represents a dia- gram with complete expansion and with compression to the initial pressure, a dia- gram which, as previously proved (Fig. 4), has no volume loss and gives the lowest theoretically possible steam consumption for the assumed range of pressures. Similar curves are shown in Fig. 18 for condensing operation. They refer to superheated steam of 13 at. abs., a steam temperature of 300 C, a back pressure of 0.08 at., and a clearance volume of 2%. In this diagram also the M.E.P. curves essentially determine the best com- pression. For the M.E.P. of 2 to 3 at. ordinarily used it is evident that the best compression approximates 90% at this low back pressure, but even for considerably higher mean effec- tive pressures the difference in steam consumption between 90% and the best compression is negligible. The, difference gradually disappears as the back pressure approaches the absolute vacuum, since in this case com- pression naturally would have no in- fluence whatever upon the steam consumption. Nevertheless, for 2% clearance, superheated steam of 13 at. abs. having a temperature of 300 G, an M.E.P. of 2,8 at., and a back pressure of 0.044 at. abs., the best compression is 90%. These may be considered average conditions for con- densing una-flow engines. This proves conclusively that the long compression of the una-flow en- gine is in no way a necessary evil accompanying the use of piston-con- trolled exhaust ports. The flatness of the M. E. P. curves also indicates that it is permissible to keep the compression constant in the 20 W 60 <30 100 Compress/on //? %* Fig. 15. Superheated steam 300 C, 14 at. abs. Noncondensing, Clearance space 11/ . above case, which is a further argument for the correctness of the una-flow system. The, long and constant compression of 90% of the condensing una-flow engine is therefore correct and admissible. 25 Many authors discuss the ''high compression" of the una-flow engine in the sense of its being unavoidable or undesirable. "High compression" is evidently confused with "long compression". A compression line may be long with low terminal pressure, or short with high ter- minal pressure. Generally speaking, ter- minal compression pressures are too low in the majority f of condensing una-flow Fig. 16. 5,0 X I 3,5 5 'eat ^ 20 ffff 00 Fig. 17. Superheated steam 300 C, 14 at. abs. Noncondensing, Clearance space 11 / . Fig. 18. Superheated steam 300 C, 14 at abs. Condensing (0.08 at. abs.) Clearance space 2/o- engines, and should be considerably higher according to paragraph 9 of the summary. Steam consumption and compression curves Jfor saturated steam, correspond- ing to those for superheated steam shown in Fig. 18, would show only a slight deviation from the latter, with the effect that the best compressions are slightly 26 shorter throughout. The assumption should, however, be remembered that ex- pansion and compression are adiabatic and that the compression is thought to cf 18 Iff 3Md//C/?er flat/m in Fig. 19. Saturated steam 14 at. abs. Noncondensing. C/ect ranee s/icrce Fig. 20. Superheated steam 300 C, 14 at. abs. Noncondensing. begin with the quality of team resulting from continued adiabatic expansion during exhaust. Jacketing of the cylinder, especially in una-flow engines, requires shorter com- pressions, since the effect of the jacket is to increase the temperature of the resi- dual steam, thereby superheating it at an earlier stage and thus raising the com- pression line. The heavy, full line curves in Fig. 19 give steam consumptions plotted against clearance volumes, for different mean effective pressures, a constant length of Fig. 21. Saturated steam 13 at. abs. Noncondensing. Most favourable compression. d 10 Fig. 22. Saturated steam 13 at. abs. Condensing (0.1 at. abs.). Most favourable compression. compression of 90%, saturated steam at a pressure of 14 at. abs. and atmospheric exhaust. The heavy dashed line curves also give the theoretical steam consump- tions plotted against clearance volumes for various mean effective pressures, but for the best compression in each case. The dashed-and-dotted lines are lines of constant best compression. The points of their intersection with the dashed lines give the best compression for that particular combination of M.E.P. and clearance volume. For example, an M.E.P. of 10 at. requires a best compression of 20% for a clearance of 12%, the steam consumption being 9.4 kg. Also for 6% clearance, 28 an M. E.P. of 10 at., and a best compression of 10%, the steam consumption would be 8,8 kg. Fig. 20 shows similar curves for superheated steam of 14 at. abs., a tem- perature of 300 C and atmospheric exhaust. It should be observed that in Figs. 19 and 20 the dashed curves for an M. E.P. of 2 at. show a distinct minimum. At this point the expansion reaches the back pressure. Reduction of the clearance volume beyond the point in- dicated by the minimum of the M. E.P. curve, for a constant M.E.P. of 2 at., results in a loop on the indicator card with a corresponding increase in steam consumption. Fig. 23. Superheated steam 300 C, 13 at. abs. Noncondensing. Most favourable compression. Fig. 24. Superheated steam 300 C, 13 at. abs. Condensing (0.1 at. abs.) Most favourable compression. The dashed curves of Fig. 19 are repeated in Fig. 21. They give steam con- sumptions for different mean effective pressures plotted against clearance volumes, for saturated steam of 13 at. abs., atmospheric exhaust and best compression in each case. Fig. 22 shows similar curves for a back pressure of 0.1 at. abs. Figs. 23 and 24 show corresponding curves for superheated steam of 14 at. abs. and a temperature of 300 G, Fig. 23 being for atmospheric exhaust and Fig. 24 for condensing operation (back pressure 0.1 at. abs.). Apart from the curves for low mean effective pressures it is interesting to note that for atmospheric exhaust the steam consumption shows an almost linear 29 dependence upon the clearance, the best compression being assumed in each case. It is therefore possible to calculate the mean specific volume loss per 1 % clearance and per HP-hour. For dry saturated steam of 14 at. abs., this is found to be 0.0918 kg/HP-hour and 1% clearance for atmospheric exhaust, and 0.1715 kg. HP-hour and 1% clearance, for condensing operation. The corresponding figures for the mean specific volume loss for superheated steam of 14 at. abs. and a tem- perature of 300 G are 0.072 kg for atmospheric exhaust, and 0.12 kg for con- densing operation. For instance in the case of a single cylinder condensing engine running on saturated steam, an increase in clearance volume of 6% will raise the steam consumption by 1 kg/HP-hour. The curves of Figs. 21, 22, 23, 24 also contain the steam consumption of the ideal engine without clearance, which is given by the intersection of the M. E.P. curves with the zero clearance line. The distance of a horizontal line drawn through such points from the corresponding M. E.P. curves gives the volume loss for any particular clearance. It may therefore be stated that for the same initial and back pressure, the same M.E.P. and best compression, the. theoretical steam consumption and the volume loss increase almost linearly with the clearance volume. Excluding small clearances and mean effective pressures, this relation is strictly true for condensing operation and approximately so for atmospheric exhaust. A mathematical expression of the volume loss R can be based on the diffe- rence in area of the diagrams with and without clearance, for

ii /2/o 55?? 55 "KO/ " " " " " " 1 A / 3 /O5 ** /O ?5 ?? 55 55 HO/ 5? 55 55 55 55 55 O \ Q/ ' /05 Zi /O 55 55 55 55 Q O/ " " " " " " 07 O/ These figures also show an approximately linear relation between volume loss and clearance volume, except for very small clearances. As a final result of the foregoing discussions, Fig. 25 shows a simple rule which allows the best compression to be determined for any given case. For a given amount of steam (p the compression must evidently be correct if a displace- ment of the line (p by an amount dq> produces equal changes of area, shown shaded in Fig. 25, on both the expansion and compression sides of the diagram, in such a way that the following equation is satisfied: 427 (ij i e ) d

' w-1 " Pi ^ The bracketed part of the fifth term can be expressed as a series by the bino- mial theorem; considering the first three terms only, the expansion gives: Substituting this value in the equation for <9, , o ~ r /IT/ OP n 1 ft *-\V HI \ / ^/ ^ r- 1 /PW . /^\ 2 /

* * /y6 lilt/lclUlt; I */ **/ ** JU JL/ ,// w \ft/ \^v Hence equation (XV) can be expressed in the form: ax 2 c = x or x 3 -I a; = 0. a a This cubic equation may be written x 3 -}-px-{-q = Q where p = and Cv = , and may be solved by Cardani's formula as follows: \ With the help of equations XV, XVI and XVII, the critical back pressures p 2 have been figured for various initial pressures, clearance volumes, lengths of compression, and amounts of admitted steam. The results are plotted in Figs. 32, 33, 34 and 35, and show the influence of these different variables upon the critical back pressure. Fig. 32 gives the variation of the critical back pressure for different clearance volumes and initial pressures, for a constant amount of steam admitted (p = 10% and a constant length of compression of 100% (una-flow steam engine). The ordi- nates of the individual curves indicate that for a given clearance volume the cri- tical back pressure p z is proportional to the initial pressure p^ which is also evi- dent from the previous equation, 1__ 1 33 =-, or p z = p- n =p i x--*=p l * . (XVIII) Therefore, for the same clearance volume, the same output and the same length of compression, the critical back pressure is proportional to the initial pressure. For the same initial pressure, output and length of compression, the critical back pressure varies with the clearance volume at a steadily increasing rate, at first slowly, then more rapidly, until the rate of increase attains a linear maximum. The critical back pressure is zero for zero clearance volume, and is small for very small clearances. This is self-evident and also proved by the curves, but is frequently left out of consideration in the design of steam engines. The clea- rance volume is therefore the cause of all evil. If the clearance volume is made zero the critical back pressure as well as the volume loss are zero, and the reci- procating engine in this respect is put on the same level as the steam turbine. Noteworthy is the slow initial increase of the critical back pressure, which again calls attention to the necessity of small clearance volumes. 38 Fig. 33 gives the relation of the critical back pressure to the amount of steam admitted, for different clearance volumes, for a constant initial pressure of 13 at. abs. and a constant length of compression of 100%. The critical back pressure grows with decreasing amount of steam admitted. In this case also the necessity for small clearance volumes is evident, especially for early cut-offs, as for instance in condensing una-flow engines. Therefore: for a constant clearance volume, for the same initial pressure and length of compression, the critical back pressure increases with decreasing output. Fig. 32. Critical back pressure plotted against length of compression is shown in Fig. 34. Each curve refers to a different amount of steam admitted, the initial pressure being 13 at. abs. and the clearance volume 2% in all cases. The curves show first a rapid increase of the critical back pressure which attains a maximum at about 30 or 40% length of compression, followed by a steady decrease. The effect of the compression is the less, the higher the MEP or the greater the amount of steam admitted. The curves confirm the lengths of compression ordinarily used in counterflow engines for normal cut-offs of 30 to 40% and more, and the com- pression of una-flow engines (100%) for the usual admissions of 15 to 10% and less. On the other hand Figs. 32 and 33 seem to lend support to the long admis- sions and subdivided pressure ranges of counterflow engines. In the case of multi- stage engines the low pressure cylinder and receiver pressure are the determining factors for the critical back pressure. The compression has the least influence, especially for late cut-offs. It should be noted here that the scale of ordinates in Fig. 34 is twice that of Figs. 32, 33 and 35. 39 The interrelation of critical back pressure, length of compression and clea- rance volume is given in Fig. 35, the initial pressure being 13 at. abs. and the amount of steam admitted 10%. The curves show the immense influence of the clearance upon the critical back pressure, which latter seems to follow a geometrical progression with increasing clearance. The critical back pressure increases rapidly up to about 40% length of compression, the rate of increase being progressively larger for larger clearances, after which it shows a steady decrease with increasing length of compression. In the order of the influence exerted, the clearance volume ranks first, then initial pres- sure, then admission and finally length of com- pression. The need for small clearance volume 9.0 too Fig. 34. loo is imperative. This is fulfilled in the best pos- sible manner by the una-flow engine, especially if fitted with high lift single-beat poppet valves which allow a clearance of 1% to be realized. The length of compression (90%) combined with the short cut-offs as used in una-flow engines is also favorable, since Fig. 32 gives a critical back pressure of only 0.004 at. abs., for P! = 10 at. abs., 9 = 10% 7*= 100%, and S Q = 1 % ; a back pressure which is beyond even the most modern condensing equipment. At the same time the above combination also gives a very small volume loss. The question is not to produce an engine which has the smallest critical back pressure, but one which combines the latter with a small volume loss. It is possible to have a large volume loss and yet a critical back pressure equal to zero, for instance in the case of large clearance volume and compres- sion equal to zero. 40 On account of the adverse influence of high initial pressure combined with short admission, the single stage una-flow engine has to rely upon small clearance volumes which, however, can be realized without difficulty. Compound or triple expansion counterflow engines can be run with more liberal clearance volumes by reason of the long admission and the low initial or receiver pressure and still have a critical back pressure beyond that attainable with the best condensers. The case is very different for single stage counterflow engines which usually have very large clearance volumes. For instance, according to Fig. 35, a single stage engine with iS" =10%, p l = 13 at. abs., 99 = 10%, and F fc = 40%, has a cri- tical back pressure of about 0.5 at. abs. All the bad influences are cumulative; large clearance volume, short admission and the most unfavorable length of com- pression of 40%. It might not be impossible to find an engine combining all these points, in actual operation. For 20% clearance, 15 at. abs. initial pressure, 10% admission and 40% length of compression, the critical back pressure becomes 1 at. abs. To run such an engine condensing would be utterly wasteful. The use of several stages not only reduces the volume loss, but also lowers the critical back pressure, and to such a degree that other defects, such as large clearance volumes, lose their significance, to a certain extent. As can be seen from Figs. 32, 33, 34 and 35, the critical back pressure becomes zero for S = 0, ^ = 0, V k = and attains a maximum for zero admission. In all the above considerations, admission, mean effective pressure and output have the same meaning and may be used indiscriminately. The possibility of causing an increase in steam consumption by going beyond the critical back pressure, as well as the useless generation of too high a vacuum are out of the question in case of well designed una-flow engines. These condi- tions, however, sometimes occur in counterflow engines, even to such an extent that the engineer and fireman are able to notice the bad effect of too high a vacuum. Prof. Josse reports such a case in the Zeitschrift des Vereines deutscher Ingenieure, 1909, page 324. He states that "the economy of the engine improved until the back pressure fell to 0.2 at. abs. From this point onwards a further reduction in steam consumption due to increased vacuum was not noticeable". Such a result in this particular engine was caused not only by the critical pressure being exceeded, but also by increased losses of initial condensation and leakage due to the higher vacuum. The initial condensation was considerable in this case. Further more, the pressure difference between the engine cylinder and condenser increases with a high vacuum by reason of the usual deficiency in exhaust area. In una-flow engines the exhaust port areas can always be made sufficiently large; leakage and initial condensation are reduced because the engine has no exhaust valves and may be fitted with single-beat inlet valves, and furthermore has the benefit of the una-flow action. In well designed and well built una-flow engines the results of the above calculations, which implicitly contain the rule of equal heat changes, do not, therefore, require any appreciable corrections. For given initial and mean effective pressures, the lowest steam consumption is obtained when the clearance volume is zero and the back pressure is zero; the length of compression has then no influence. The critical back pressure, i. e. the most economical back pressure, and the length of compression become of impor- 41 tance as soon as the clearance volume has a definite value. According to Fig. 31 the length of compression can then be increased to 100% and the back pressure reduced until the changes of total heat during compression and expansion become equal, thus offering the best basis for low steam consumption. In other words: for a given initial pressure, given mean effective pressure and given clearance volume, the minimum steam consumption will be obtained when the length of compression is 100% and the back pressure is such that the change of total heat during expansion is equal to the change of total heat during compression. (Una-flow steam engine.) This rule may also be arrived at if the second wording of the fundamental law given on page 31, dealing with back pressure, is applied to una-flow engines. The minimum steam consumption requires the shortest possible cut-off and the longest possible compression of 100%, these being related to each other by the rule of equal heat changes. An examination of the common types of steam engines will reveal the fact that incorrectly designed engines are the rule and correctly designed engines the ex- ception. There is hardly a steam engine designer who is not guilty of some viola- tion in this respect. To begin with, the average designer is not aware of the harm- fulness of the clearance volume, which explains the carelessness with which un- necessarily large clearances are used. The latter are rendered necessary for short cut-offs, for instance in locomotives, marine engines, or non-condensing engines with shaft governor controlling both inlet and exhaust. In marine engines this is the case with the high and intermediate cylinders, while the low pressure cylin- ders usually have unnecessarily large clearances. The lengths of compression are frequently incorrect. These "necessarily" and "unnecessarily" large clearances can be avoided. A knowledge of the above rules is indispensable, as well as re- cognition of the fact that changes in load as well as change of rotation may be accomplished merely by the steam admission organs without change in the exhaust timing. The proper choice of steam distributing organs as well as their arrange- ment and mechanism are also important factors. For instance, a single-stage condensing una-flow marine engine with single-beat valves arranged in the cylinder heads fulfills all of the above conditions. (See Fig. 31, Chapter V.) This engine has the great advantage of a small clearance volume of less than 1 % ; the exhaust timing is independent of the inlet gear and the constant length of compression is correct and permissible. The same reasoning holds true for the stationary una- flow engine having single-beat valves. (See Fig. 6, Chapter VI.) Both types of engine therefore have very small volume losses despite the large pressure ranges. In the same way a considerable reduction of clearance volume in non-condensing una-flow engines can be accomplished by shortening the compression (large exhaust lead). This is shown in Fig. 32, Chapter III, including also the effect of an exhaust ejector, which produces a proper change of compression with the cut-off, thereby further reducing the volume losses. Summary. 1. The volume loss is determined by the clearance volume, initial pressure, back pressure, mean effective pressure and length of compression. It increases with increasing initial pressure and clearance volume, decreases with increasing 42 back pressure and mean effective pressure, and becomes a minimum for a certain length of compression. 2. Correct compression tends to reduce the volume loss; compression may be kept constant for small clearance volumes, but should be varied inversely with the cut-off in case of large clearances (Single cylinder engines). 3. Change of compression ,in case of high vacuum has no material effect upon the steam consumption. 4. The clearance volume loss is zero if expansion reaches the back pressure and compression rises to the initial pressure. 5. The theoretical steam consumption, for the same initial pressure, back pressure, mean effective pressure and best compression in each case, increases nearly linearly with the clearance volume. Apart from very small values of the clearance and mean effective pressure, this linear dependence is almost exact for condensing operation, and approximate for atmospheric exhaust. 6. For given initial pressure, back pressure, mean effective pressure and clea- rance volume, the length of compression must be such that the change of total heat during expansion is equal to the change of total heat during compression. 7. For given initial pressure, mean effective pressure, clearance volume and length of compression, the back pressure must be such that the change of total heat during expansion is equal to the change of total heat during compression. 8. For given back pressure, mean effective pressure, clearance volume and length of compression, the initial pressure must be such that the change of total heat during expansion is equal to the change of total heat during compression. 9. For given initial pressure, back pressure, mean effective pressure and length of compression, the clearance volume must be such that the change of pressure during expansion is equal to the change of pressure during compression. 10. For the same initial pressure, back pressure, the same terminal compres- sion pressure and terminal expansion pressure, and for equality of total heat changes, the lengths of the best compressions have the same ratio as the clearance volumes. (Different mean effective pressures.) 11. With proper proportioning of the length of compression, the clearance volume has to be kept as small as possible; this applies especially to single cylinder condensing engines and the low pressure cylinders of compound and triple expan- sion engines. 12. Subdivision into stages results in reduction of volume losses, the high pressure cylinder having the smallest and the low pressure cylinder the largest volume loss. Intermediate cylinders have a loss between both according to their relative size. 13. For given initial pressure, mean effective pressure and clearance volume, the lowest steam consumption is obtained if the length of compression is made 100% and the back pressure chosen so as to make the change of total heat during expansion equal to the change of total heat during compression (una-flow engine). 14. The critical back pressure is determined by the initial pressure, the clea- rance volume, the mean effective pressure, and the length of compression. It increases in proportion to the initial pressure and faster tlmn proportionally to 43 the clearance volume; it first increases with increasing length of compression, then decreases with increasing length of compression and increasing mean effec- tive pressure. It is zero for initial pressure = zero, clearance volume = zero, length of compression = zero, and attains a maximum for mean effective pres- sure = zero. The clearance volume has by far the greatest influence, the pressure range has less, and the length of compression and mean effective pressure have the least. 15. In compound engines the low pressure cylinder determines the critical back pressure. Under the same conditions compounding reduces the critical back pressure corresponding to the lower initial pressure of the low pressure cylinder. 16. It is not so important merely to achieve low critical back pressure alone as it is to obtain simultaneously small volume losses and low critical back pres- sure. The volume loss may be very large and the critical back pressure may still be zero. Of all single cylinder condensing engines, the single-beat poppet valve una-flow condensing engine has at the same time the smallest volume loss and a critical back pressure which is far below anything that can be reached even with the most modern condensing equipment, mainly on account of its small clea- rance of less than 1% and its favorable length of compression. 44 2b. Additional Clearance Space. Practically all condensing una-flow engines must be able to run non-condensing. In case of breakdown of the condenser, lack of cooling water, or during the winter months when the exhaust steam is used for heating purposes, the engine must be capable of operation with either atmospheric or higher back pressures. The Fig. 1. Fig. 2. simplest way of accomplishing this purpose is the provision of an additional clea- rance space. (See Figs. 1 and 2.) The amount of additional clearance depends upon the initial and back pressure. If the latter is for instance 1 at. abs., the initial pressure 13 at. abs., and the clearance for condensing operation 1.5%, then the additional clearance should be 14.75%, according to the tables to be given later. At the same time this increased clearance will cause a lengthening 45 of the cut-off from 8 to 12% for the same output with non-condensing operation (Fig. 3). The drop in pressure at the end of expansion amounts to 0.8 at. for con- densing and 1.0 at. for non-condensing operation. It is found that for other initial pressures, with approximately the same drops of pressure (0.8 or 1 at. abs.) at the end of expansion, the mean effective pressures produced are about equal. In the previous chapter it was demonstrated that the mean specific volume loss for saturated steam of 13 at. abs. and non-condensing operation was 0.0918 kg/HP- hour and per 1% clearance, and 0.072 kg/HP-hour, and per 1% clearance for superheated steam of 300 G, the other data being the same. For 14.75% clea- rance the total losses amount to 1 .35 and 1.06 kg respectively in the two cases. The general adoption of the additional clearance space despite this considerable increase in steam consumption is due firstly to its simplicity, and secondly to qualities which tend partly to coun- teract this heavy loss. The effect on the overall economy is negligible if an engine operates with additional clear- ance only for several days or hours during the course of a year. Fig. 3 also indicates that although the drop in pressure at the end of expansion is higher when using the additional clearance space, the loss due to incomplete expansion is less; and the condensing cylinder being rather large for non- condensing operation becomes more or less adapted to this condition. The additional clearance also preserves the una-flow principle, including the series arrangement of live steam space, inlet valve, piston and exhaust, which is such a valuable feature of the una-flow engine. Although it is possible to use auxiliary exhaust valves instead of additional clearance, and these valves being relieved of pressure at the time of opening can be of single beat or annular con- struction, no joint at all being preferable to even a tight joint or seat. Care must be taken that the clearance valves which control the additional clearance pocket do not materially add to the cylinder clearance for condensing operation (Figs. 4 and 5). In this respect it is advantageous to provide the clea- rance valves with projections which fill up the space between valve seat and cylin- der surface. The valve area of the clearance valves must be large enough to avoid throttling during expansion and compression. It is also advisable to arrange the additional clearance so that it will act as a kind of heat insulator when the engine is running condensing, which is especially of .value for the crank end of the cylinder. The clearance valve may also be designed in the form of a spring loaded safety valve, but then the above mentioned projections cannot be used. The safety valve action of the clearance valve is unnecessary when the main steam valve is not, or only partly balanced so that it can act as a safety valve. The "safety" inlet valve is preferable to the "safety" clearance valve since its weight and spring load are less. The inlet valve is designed for high speed and held closed by steam pressure, its spring being only strong enough to overcome inertia. The clearance valve on the other hand is heavy and its spring has to overcome the total steam pressure. In case of sudden failure of the vacuum this heavy spring load combined with the great weight of the valve cause an objectionable hammering, which can only be stopped by screwing the valves back. In Fig. 6 is reproduced a diagram such as is obtained from a una-flow engine running non-condensing and fitted with auxiliary exhaust valves, the clearance being the same (1%%) as for condensing operation. It is evident that the ratio Fig. 4. Fig. 5. of expansion is too high. A construction of this kind is shown in Fig. 7, having the auxiliary exhaust valves arranged in the cylinder heads. The increase in clea- rance volume due to these valves was not taken into consideration in the diagram of Fig. 6. The diagram indicates that, assuming the same mean effective pressure, the expansion line reaches the back pressure while the piston uncovers the exhaust ports. The loss due to incomplete expansion is zero. The shape of the diagram indicates the counter-flow action and proves that the cylinder is too large for non-condensing operation, especially for smaller loads, when the toe will change into a loop. This produces a backflow of exhaust steam into the cylinder and a corresponding increase in condensation losses. For loads higher than normal the exhaust action will be partly una-flow and partly counterflow. The loop at the end of expansion cannot occur in engines fitted with additional clearance spaces. 47 The worst feature of auxiliary exhaust valves, however, is their detrimental effect upon the condensing operation of the engine. They increase the clearance space and the harmful surfaces as well as the possibility of leakage, and sacrifice the very valuable series arrangement of live steam space, inlet valve, piston and exhaust. For 1% increase in clearance volume, an additional steam consumption of 0.12 kg/IHP-hour may be expected, superheated steam being assumed. This figure does not include the effect of leak- age and the surface losses caused by the valves and their pockets, nor additional losses due to the operation of these valves while the engine is running condensing. It is advisable to keep these valves in operation even while running condensing, Fi 6 since they are liable to stick after re- maining out of use for some time. If the auxiliary exhaust valves shorten the length of compression also for condensing operation, a larger volume loss results, because 48 an increase in clearance necessitates a corresponding lengthening of the compres- sion. The bad effect upon condensing operation appears all the more objectionable since it occurs during the whole working period; while on the other hand, for short periods of non-condensing service even a considerable increase in steam consump- tion due to additional clearance could be tolerated. For long periods of non-con- densing operation, as for instance during the winter, single-beat auxiliary exhaust valves are preferable to additional clearance. When auxiliary exhaust valves are used, they are usually placed below the engine room floor level, which renders their attendance difficult and the arrange- ment of the piping awkward. The amount of the necessary clearance is determined by the following rule: for a given length of compression, mean effective pressure, initial pressure and back pressure, in order to keep the volume loss as small as possible, the clearance volume should be made large enough to produce equal variation of pressure during expansion and compression (see chapter on volume loss). On an average, the ter- minal expansion pressure for non-condensing operation may be taken as 1 at. gage, and this implies a terminal compression pressure of 1 at. below initial pressure. The following table gives the total amount of clearance volume required, when operating non-condensing, for 90% length of compression, starting with a pressure of 1.03 at. abs. and ending 1 at. below initial pressure, with adiabatic compression and saturated dry steam. The figures are based on the latest Mollier chart. Initial Pressure 8 9 10 11 12 13 14 15 16 at. abs. Total Clearance 27.9 23.8 21.3 19.2 17.6 1625 15.15 142 13.4 O/ /o 49 3 a. Losses due to Throttling. Throttling is a change of state in which the total heat remains constant, the effect of which is to diminish the amount of heat and the pressure difference avail- able for utilization between boiler and condenser. Losses due to throttling may occur in the superheater, steam main from superheater to the engine, stop valves, inlet valves, piston-controlled exhaust ports or exhaust valves, and in the exhaust pipe between engine and condenser (Fig. 1). The losses due to throttling occuring in the superheater, steam pipe, stop valves, and inlet valves may be partly regained in connection with the subsequent expansion, although the greater part is lost. The percentage of this regain depends W Fig. 1. Fig. 2. upon the extent of the expansion. The temperature-entropy diagram in Fig. 2 shows the conversion of a certain quantity of heat at high temperature and small entropy into an equal quantity at lower temperature and larger entropy. The change is represented by the area GCDQKHG and is equal in area to the strip QKLWQ which results from increasing the entropy. The part KNVJK, falling within the area of expansion will be regained, while the part NLWVN below the line of terminal expansion pressure is definitely lost. Throttling losses occuring in the exhaust valves or exhaust pipe are irretrievable, for which reason they must be restricted to the smallest possible amount. They are especially harmful because S/ump/, The una-flow steam engine. 4 50 their effect extends through the whole compression stroke. (See chapter on the relation of the una-flow engine and the condenser.) The heat losses and throttling losses in the steam pipe cannot be separated and are usually combined in one figure; a loss of 0.5 to 1.0% is considered an average and corresponds to a steam velocity of about 40 to 50 m/sec, calculated on the total amount of steam flowing through. The throttling losses occuring between the stages of multistage engines are eli- minated in the una-flow engine. In order to estimate the throttling losses in the inlet and exhaust valves it is necessary to know the relation between effect (losses) and cause (valve area), to FA which the following calculations refer. f ^5 A P: Fig. 3. Determination of inlet valve areas. In Fig. 3, p t and v t represent the pressure and volume of the steam at the dead center position of the piston; piston travel x 1 = and corresponding crank angle 6^ = 0. p 2 , v 2 are the pressure and volume at the point of valve closing, for a piston travel = x 2 and crank angle = (5 2 . p and v are the pressure and volume at any intermediate point where the piston travel = x and crank angle <5; w represents the velocity of the entering steam corresponding to the pre- vailing pressure difference. F is the valve area in sqm,

w-F'dt n D"- H (x + s) :-D z -H(x dx 0,2125 y w-F'dd x-{- s A (x + s) (2) For a small drop in pressure it may be assumed with sufficient accuracy that w = i * g (Pi P) ! In order to facilitate the calculations, the admission line or rather the curve representing the change in pressure plotted against crank angle or time will be replaced by a parabola. It will be shown later that this assumption is admissible if the object of the calculation is not the shape of the admission line but the final drop of pressure at the end of admission. This final pressure drop, however, is to form the basis of the determination of the inlet valve areas. Therefore we may write p l p = a d- and p 1 p z = a d z 2 or 2 and therefore combining equations (3) and (2) dp dx 0.2125- + Jdp . C dx P h j s-h 2 .9) p 1 (p p t ) F-6-dd 0.2125 . t ( F 6- * + dd 57 has to be diminished by a certain percentage according to the cam profile. It must also be 10" i\ \ ft 8 Fig. 12. considered whether the valve remains stationary during part of the time it is open. (See Figs. 12 and 13.) In case the steam valve has a certain amount of lead (Fig. 14), the integra- tion of the .F curve must apply only to the area after the dead center, since the Fig. 14. Fig. 15. purpose of lead is merely to fill up the clearance space, a condition which was assumed from the start. C L = ^2 C Vi' The values of C are inversely proportional to (p, and Further, ^ maXj = - . nr ~ ^ or 9>1 Having found .F max for a certain cut-off and pressure drop, the question arises of the pressure drop for different cut-offs. For instance, at 12.5% cut-off and h z = 2 at. pressure drop, the value of .F max as taken from the curves in Fig. 11 is 1.325 sqcm. Assuming a direct drive from the eccentric, the values obtained for 58 the pressure drop for different cut-offs are given in the following table and plotted in Fig. 15. * 2 = 5/ 10% 12.5 / 15% 20% 30% 40% 50% fc 2 = 2.55 2.2 2.0 1.8 1.55 1.15 0.8 0.7. Fig. 15 also contains two more curves for higher pressure drops during normal admission. These curves show that, if for a condensing una-flow engine the pres- sure drop is normal for rated cut-off, then a larger cut-off will show a smaller pres- sure drop. For a small cut-off the pressure drop increases still further, while in case of a higher pressure drop at normal cut-off this gradually tends to be a maxi- mum. For condensing una-flow engines it is sufficient to calculate ^ max f r normal cut-off. Example. The necessary valve area of a single-beat valve for a una-flow stationary engine is to be calculated for the following conditions. Cylinder diameter D = 0.4 m, stroke H = 0.5 m, r p m = 150, steam pressure p = 13 at. abs., steam tem- perature ^ = 300 centigrade, assumed pressure drop ='2 at. for 12.5% valve gear cut-off, or z 2 = 0.125 and <5 2 = 41.4,

001 3 56sqm 2 :=:1 3 i 56 sqcm 2 > If the maximum valve lift for the assumed valve gear cut-off of 12.5% is equal to Vio tne valve diameter, then.F max = n - d 0.1 d = 13.56 sqcm, d = 6.6 cm, and A max = 0.66cm. The common empirical formula, based on mean piston velo- c m city would give a steam velocity w m = -- m = 92 m/sec. These calculated values of F' max and ft max can be realized without difficulty if the single-beat valve is ope- rated by a lay shaft gear running at twice the engine speed (See final chapter). Permissibility of the use of the parabola. It is still to be proved that the actual diagram admission line plotted against crank angle or time may rightly be replaced by a parabola. For this purpose a diagram is laid out with the crank angles d as abscissae and the valve openings as ordinates. This results in a curve such as that shown in Fig. 16. The total crank angle is now divided into a large number of parts or intervals and for each 59 part the mean valve area is determined (F^ F 2 , F s etc.). It is assumed that p v = const. In the above example it was assumed that p 1 = 130000 kg/sqin, ^ = 300, i>! = 0.2 cbm/kg and /v ^ = 26000. y Volume of clearance space Fj = s\ weight of steam Q l = - ; stroke volume -f w 2 w 2 ^ * ~~2~ v 2 = H\ '.= = O.S- dt = for i intervals. The n 360 i piston is first considered to be moved forward a distance corresponding to the V 1 26000 first interval without admission of steam, so that p 2 = p r y ', v 2 = ; ;> 2 and y m = -^ 2 > ('o These values represent the state of the steam at the end of the first interval, produced by expansion only, without the admission of live steam. w m and y m being also known, it is now possible to calculate the additional weight of steam Fig. 16. admitted, which is dQ =

= coefficient of velocity, w = the velocity of the exhaust steam in m/sec, F = instantaneous value of the exhaust port area in sqm, y = specific weight and t = duration of exhaust. Even with highly superheated live steam, the exhaust steam of condensing engines is always, and that of non-condensing engines in most cases, saturated. The change of state within the cylinder can therefore be assumed to follow Mariotte's law: /?! t>j = pv = p 2 v z = const. Pii y i5 7i represents the state of the steam at beginning of exhaust, p, y, y the state of the steam at any intermediate point, Pzi v zi 72 the state of the steam in the exhaust pipe (condenser, atmo- sphere, etc.), Pe, v e, 7e the state of steam at the narrowest place of exhaust port. x represents the piston travel in % of the stroke measured from the admission end. As long as - < 0.577 is p e = 0.577 p. The value 0.577 remains about P the same for any steam wetness. If p < 1.735 p 2 then p e = p 2 . The change of the cylinder volume during exhaust is neglected. Since the weight of steam present in the cylinder is proportional to the absolute pressure, p dp Q dQ _ T . , n , J - = - PC ; Q =V -y; dQ = w-(o- F y e -dt p Q * dp d Q y (Q-F y e d t .. ~p"~ = ~Q~ ~^ T r~ t = - ; dt = pr (<5 = crank angle corresponding *- 360 ' 6 to time t) dp -F-dd. 7e v _^. D2 , Ht(x , s} . A jp.n.j ~Y S-n-V-y -4 D ' H '( x + s >> A C* 1 1) . i _ I w \*s j. \fit v f g /O \ IP"- A.(X+ S ).Y~ Range of High Pressures. In this case /?> p cr = 1.735 p 2 , w = 3.23 y p v = 3.23 j/c . For practical purposes it is sufficiently exact to replace the variable quantity (x + .9) of equation 2 by the constant quantity (5 + 1 0.5 a), where a = exhaust lead (the critical pressure is reached approximately at the dead center). 62 0.2124. y./c- 3.23 .(s + l 0.5 a). 1.62 1.62 , per Ocr (Xp _ 0.423. y.]/7 f J p ^ (s + l_0.5)J Fig. 18. The quantities F for the range of high pressures may be plotted against <5 according to Fig. 18 and *r f J Then log e p x \_ 0.423 . y f c r ^1 (s + 1 0.5 a) A 9? - 0.5 a) . 5.45 - lo glo c 1.735/?2 (3) in m 2 , ^4 in m 3 /min, c in kg/m 2 m 3 /kg. For condensing una-flow engines with double-beat valves the average clearance may be assumed to be s = 0,03 and y = 0.9 for drilled exhaust ports with well rounded edges; also A = 1 in m 3 /min and c = ip 1 v 1 = 15000 FmH - dcr = 0.0495 (1.03 0.5 a) Iog 10 in m 2 . . . (4) For non-condensing una-flow engines also, the steam is in most cases saturated at exhaust and therefore F mH d c> . = 0.0457 (1.11 0.5 a) lo glo with /?!! = = 17500, 5 0.11, 9> = 0.9, ^4=1. ' n . . . (5) Range of Low Pressures. If p has been reduced to p cr = 1.735 /? 2 tnen Pe Pz and w = Therefore ~ 0.2124- I-5I p Jl - "1 P I F-dd p A (x + s) Inserting 1+5 0.5 a for x + s and c for p.y the last equation may be written as follows: 63 dp 0.2124 P' P. A (s + 1 0.5 a) 0.942- We j_ A (5 + 1 0.5 a) r x- F. For J/: d d = F mN . ((5 2 <5 e( .) will be der F mN (<5 2 ; 1,735 Pftypi < 1-735 p 2 ti-e- =i.6 ) } the value (5a) should \ ^2 / be integrated in smaller limits (i. e. 1.6 and 1) (5. upper corner in Fig. 20). In accordance herewith the second values of the brackets of equation (9) and (10) should be neglected, if they become negative. An examination of engines with piston-controlled exhaust ports (not overrun by the piston) shows that r-, * -^max "max & . . j-, r = ^y approximately. F m ""m * The following table as well as Fig. 20 gives the values of F m and -F max (diffe- rent scales) for the exhaust through the high and low pressure ranges for various values of exhaust lead a and pressure ratios , A being =1, live steam pressure Pz p l = 13 at. abs., ^ = 300,

max Pz = 129 sqcm. With a port diameter of 12.5 cm, the area of one port would be 122.7 sqcm and . a single port would be almost suffi- cient. 4. For the same engine running non-con- densing, the conditions may be p t = 3 at. abs. and p 2 = 1 at. abs., hence = 3, a =20%. Pz Then according to Fig. 20, -F max = 48.2 sqcm, and one port of 10cm diameter having an area of 78.5 sqcm is therefore already too large. The formula commonly used, based Fig. 22. upon mean piston speed, would give the velocities w =. 13.8 m/sec in case 1, w = 9.4 in case 2, w 24.3 in case 3, and w = 65 in case 4. This proves the inadequacy of this formula. In Fig. 22 is shown a diagram in which the admission and exhaust lines were calculated point by point after the areas for inlet and exhaust had been found by the above method, a drop of pressure at the inlet from 13 to 11 at. abs. and at the exhaust from 1.2 to 0.05 at. abs., 13% valve gear cut-off and 10% exhaust lead being assumed. 68 3b. The Relation of the Una-Flow Engine to the Condenser. A high vacuum is of great advantage to the operation of una-flow engines. Fig. 1 shows compression curves for different back pressures and the same ter- minal pressure, the clearance volumes being correspondingly changed. These curves indicate how appreciably the diagram area increases with better vacuum. At the same time it is possible to keep the compression up to the desired value by properly proportioning the clearance volume. For a high vacuum the clearance Fig. i. volume used may be very small. The limit is usually determined by the design, 2% being considered an average figure. The duration of the exhaust of a una-flow engine with 10% exhaust lead and 90% length of compression is only about one half of the time available for the exhaust of a corresponding counterflow engine. The working steam of the una-flow cylinder must therefore be exhausted into the condenser in one-half the time. It is a fact that in the usual design of counterflow engines there exists a considerable pressure difference between the interior of the cylinder and the con- denser, which is used to overcome the resistances in the usually too narrow ex- haust passages. The shortening of the duration of the exhaust in the una-flow engine is all the more a reason for diminishing to the utmost the resistance between condenser and engine cylinder, and this can be accomplished by short passages 69 of large area. Furthermore, the exhaust port area of the una-flow cylinder can easily be made three times as large as the exhaust valve area of the ordinary counterflow engine. If now the remaining cross-sections have sufficient area to harmonize with these large exhaust port areas, and the length of the passages is kept down to the minimum, then complete pressure equalization will result. This is proved by experience as well as theory (See end of this chapter). In Figs. 2 and 3 are shown a longitudinal and a cross section of a una-flow cylinder where the exhaust belt connects over its full width to the jet condenser placed immediately below it. The injection water enters by means of a perforated tube placed horizontally across the condenser. As may be seen from these illustra- tions, the exhaust passages are extremely short and wide so that there is practically no resistance. Fig. 2. Fig. 3. Figs. 4 and 5 show the application of a jet condenser of the Westinghouse- Leblanc type to a una-flow cylinder. This condenser and a similar one developed by the A.E. G. are based upon a principle which formed the substance of a patent issued to the author. The condenser body in this case forms the support for the engine cylinder. As in Figs. 2 and 3, this gives a very short connection and large transfer area, thus insuring equalization of pressure between the cylinder and condenser. On account of this complete equalization, the compression begins at the lowest possible pressure with the result of a considerable gain of diagram area, a corre- sponding reduction of clearance volume and clearance surfaces, as well as increased thermal efficiency (Fig. 1). The short duration of the exhaust period due to the piston-controlled exhaust correspondingly reduces the cooling action of the con- denser upon the interior of the cylinder. As soon as the exhaust ports are covered on the return stroke, the connection with the condenser is cut off, any further cooling is prevented and the heating effect of the steam jacket at the cylinder end comes into full play without any adverse influence due to the exhaust. It is fundamentally wrong to interpose oil separators, change-over valves, feed water heaters or elbows in the connection between engine cylinder and con- 70 denser. Such accessories cause very large resistances to the flow of steam and should be avoided unless their use is rendered necessary by other important considera- tions. The atmospheric exhaust pipe should be connected to the condenser body. If the connection between the condenser and air pump is shut off, the former Fig. 4. Fig. 5. then acts as a kind of exhaust muffler or silencer (See Fig. 2). This silencer action should be assisted not only by the volume of the condenser but also by a change of direction of the steam flow. If no such provision is made, the loud exhaust will be very objectionable, as is shown by experience. <. <^ ^ CQ PdJP uoj sy/^i/o cr-p 71 4. Losses due to Friction. (Mechanical Efficiency.) Dimensioning of Driving Parts. Very complete data are available for the dimensioning of driving parts of stationary una-flow engines. From these data have been compiled the curves shown in Fig. 1, which apply to steam pressures of from 10 to 12 at. gage, and condensing operation. In Fig. 1 may be seen the weight of the reciprocating parts including two- thirds of the weight of the connecting rod, plotted against cylinder diameter. The average values may be represented by a curve according to the equation r __ (Cylinder dia. in cm) 2 - 5 &VT -26- The weight of the piston is about 3 / 10 , that of the connecting rod 2 /7 f the total weight of the reciprocating parts. Since the ratio of stroke to cylinder bore is the determining factor for the reciprocating weights, the ratio of stroke to cylinder bore is also shown in this chart. It will be observed that small engines have a proportionally long stroke, while large engines have a proportionally shorter stroke. Since the average buyer of engines generally has a prejudice against what may be called high speed in the sense of high number of revolutions per minute, regardless of piston speed, small engines are therefore built with a comparatively long stroke. For large engines a high number of revolutions is usually demanded, and since the majority of builders have a similar dislike for high piston speeds, an engine of short stroke is the result. It seems strange, however, that the type of frame used does not appear to have any bearing whatever upon the bore and stroke ratio although some designers are inclined to make side crank engines with long, and center crank engines with short strokes (See Fig. 1). The piston speed based on cylinder diameter shows a more rapid increase for the smaller sizes than for the larger ones. The maximum continuous load rating of una-flow engines usually corre- sponds to a mean effective pressure of about 4.5 kg/sqcm. If the mechanical effi- ciency for this load is assumed to be 0.94, then the figures for the HP output obtained agree closely with those given by the makers. The normal rating usually corresponds to a mean effective pressure of 3 kg/sqcm based on brake HP. This figure is evidently a compromise between high economy and low initial cost. The lowest curve in Fig. 1 represents weight of reciprocating parts divided by rated brake HP. The values show small variation (3.45 to 4.15 kg/BHP) but increase gradually with the cylinder bore. The curve between 3.6 and 5.6 kg/sqcm represents inertia of reciprocating parts for an infinite length of connecting rod. 72 Fig. 2 gives first an indicator card having an MEP of 4.3 kg/sqcm corre- sponding to the maximum continuous load. From this card are developed three net pressure diagrams containing inertia curves plotted for a length of connecting rod equal to five times the crank radius, for three different cylinder bores of 500 > Net pressure and Inertia force Curves. MEP = 4,3 kg/ cm 10- Cylinder dia. 500 mm Crank pin Main bearing, for 50% balancing Cylinder dia. 900 mm Piston rod Crosshead pin Fig. 2. 900 and 1300 mm. For any piston position the vertical distance between inertia curve and net pressure line represents the load upon that driving part to which the inertia curve refers. The dashed and dotted lines apply to the load on the piston rod, the dashed lines to the crosshead pin, and the full lines to the crank pin. The dotted curves in the same way give the load on the main bearings, 50% of the weight of the reciprocating parts being balanced. For center crank shafts two equally loaded main bearings of equal size are assumed, while for side crank 73 shafts the main bearing loads have been increased by about 20% on account of the overhang. In regard to bearing load, apart from impact, it would be more advantageous if the inertia forces for the smaller engines would correspond to those of the larger size in the last diagram of Fig. 2; and this could be accom- plished by increasing the speed, for an engine of 500 mm cylinder bore, from 162 r. p. m. to say 188 r. p. m. The proportions of piston rod and tail rod as well as diameters of the diffe- rent bearings are plotted as functions of the cylinder bore. It will be noted that the ratio of piston rod diameter to cylinder bore is slightly less for engines of large size, by reason of the proportionally shorter stroke of the latter. The dis- tance from rear end of piston to center of crosshead pin is usually about 3.33 times the stroke. The factor of safety against buckling of the piston rod, based on the loads represented in the diagram of Fig. 2, is 10 for small engines and 9 for larger engines. Average ratios. Crosshead pin diameter to cylinder bore 0.265 Side crank, crank pin diameter to cylinder bore 0.33 Side crank, main bearing diameter to cylinder bore . . . 0.5 Center crank, crank pin diameter to cylinder bore . . . 0.425 Center crank, main bearing diameter to cylinder bore . . 0.425 Length of crosshead pin to its diameter 1.3 1.6 Side crank, crank pin length to its diameter . ... . .1.0- 1.2 Center crank, crank pin length to its diameter 0.9 1.0 Side crank, main bearing length to its diameter 1.3 1.8 The crank pin diameter of center crank shafts will be found only in rare cases to be larger than the diameter of the main bearing, and the main bearing at the flywheel side longer than the opposite main bearing. F Fig. 1 further shows the ratios of piston area to bearing areas which (I ' a) give the following averages: Crosshead pin 8, side crank, crank pin 7. 7, center crank, crank pin 4.25, side crank, main bearing 1.95, and each main bearing of center crank shafts 1.5. Combining these data with the specific loading taken from the diagrams in Fig. 2, the resultant bearing pressures were calculated and are shown at the top of Fig. 1. These values refer to maximum continuous load and smallest dead center inertia, horizontal forces only being considered. Strictly speaking, the additional forces due to flywheel weight, belt pull, etc. should be combined with the horizontal forces, but this would not materially alter the results. It will be seen that crank pin and main bearings of side crank shafts sustain about 50% higher loading than the corresponding bearings of center crank shafts. The highest bearing pressures given in Fig. 1 are undoubtedly permissible in case of force feed lubri- cation. On side crank shafts, excessively large leverages, i. e. distances from the con- necting rod center to the main bearing center, cause increased bending, higher 74 bearing pressure and on account of the deflection of the crank shaft, increased pressure at the inside edge of the main bearing. This tendency can be reduced by shortening the leverage or the use of self aligning bearings, or both (Fig. 3). Since, there are no secondary forces acting on the connecting rod in the horizontal plane, the factor of safety against buckling in this plane need not be more than 5, as against a factor of 9 or 10 in the vertical plane. This condition can be easily met by flattening the otherwise circular rod section. The crank hub should be placed as close as possible to the connecting rod and should not be wider than 0,65 times the shaft dia- meter for large engines and 0.75 for small engines. On some large Belgian engines this figure is even cut down to 0.6. The projecting part of the crank pin bearing cor- responds to the projecting part of the crank hub. The rear side of the crank in this case becomes flat. The crank is frequently pressed in place on the shaft, although a shrink fit is preferable by reason of the lesser chance of damage to the structure of the material. A key, although frequently used, is unnecessary. The same applies to the connection of the crank pin to the crank arm, and the former should be ground after assembly. Fi 8- 3 - UsssL^J A still shorter over- hang may be obtained by casting crank, crank pin and crank shaft in one piece of cast steel. < By designing the crank in the form of a disc, as shown in Fig. 4, an especially ,large reduction in the overhang may be obtained, with a corresponding decrease in the shaft diameter. The present state of foundry prac- tice allows such a construction to be used without anxiety. Constructional data of a stationary una-flow engine built by Sulzer Bros, and installed in a cotton spinning mill at Grefeld are as follows: Cylinder bore | 23000 Kc) 75 1100 mm, stroke 1200 mm, speed 110 r. p. m., steam pressure 12 at. gage, steam temperature 320. The engine is of the center crank type. Main bearings, 475 mm dia. by 650 mm long Crank pin, 475 mm dia. by 380 mm long Crosshead pin, 300 mm dia. by 430 mm long Piston rod 220 mm diameter, tail rod 170 mm diameter Weight of piston 2125 kg, weight of piston rod 1500 kg Weight of crosshead 1532 kg, 2 / 3 connecting rod 1780 kg 50% of the reciprocating parts are balanced by counterweights fastened to the crank cheeks.j \ The friction HP of this engine at 110 r. p. m. with a smooth flywheel was 113,6. The correspon- ding mechanical efficiency for a rated load of 1700 1 HP is therefore 0.933, a figure which disproves the opinion frequently advanced that the mechanical efficiency of una-flow engines is low. The assumption that the engine friction must be nearly independent of the HP output is based on the fact that the load on the driving parts is practically the same for idling as it is for K the rated HP Starting with the '- -j ^ ^ ^ s s/ s 1 \ ^ ^r 1: ^ 1 1 1 ^ ti OOx SS> ^s xxX ^ ^ ^N: ^t mi - ** engine idling, and gradually in- creasing the output, the gross \ ^ load on the driving parts first A decreases slightly, at rated output ^ reaches the same value as for \^ idling, and then becomes some- what greater for larger output. The engine at Crefeld, for instance, has center inertia load of approximately 6 ] a corresponding mean pressure of the ine gram of 3 kg/sqcm, and a useful rated me; tive pressure of 3 kg/sqcm. It follows 1 engine friction for idling and rated outp be the same. Further, the weight of the flywheel, crank shaft and the rest of the parts as well as the centrifugal force oftheco rod end, crank and counterweights are res for a constant portion of the total friction The above will be further illustrated following test results obtained by Sulzer B \ ^ / ^^ / /'s a dead cg/sqcm, rtia dia- m effec- ,hat the ut must piston, driving rmecting ponsible by the ros. Fi /s I ->v 1 j g- 4. A una-flow engine of 700 mm cylinder bore and 900 mm stroke, having a rope flywheel of 4000 mm diameter, gave the following friction at different speeds (without ropes). 133 112 100 85 68 r. p. m. 66 53 46 38 28 I HP friction. 76 The corresponding rated load would be 820 690 615 520 420 I HP, so that the friction HP in % would be 8 7.7 7.5 7.3 6.7%. Another engine of 600 mm cylinder bore and 725 mm stroke, fitted with a rope flywheel of 2400 mm diameter, with 14 grooves, gave the following results: 150 100 50 34 r. p. m. 41 22 8.5 4.5 I HP friction. The corresponding rated load in this case would be 475 317 158 108 I HP, so that the friction HP in % is 8.65 7 5.4 4.15%. In reducing the engine speed from 150 to 50 r. p. m. the friction HP should diminish from 41 HP to 1 / 9 of the same, or 4.5 HP. It actually was 8.5 HP, which reflects the influence of the weight of the driving parts. The weight and inertia of the latter have an equalizing effect, so that for constant speed the friction HP remains nearly constant independently of the instantaneous output. Dimensions of Driving Parts of Other Una-Flow Engines- (Sulzer Bros, design.) Steam pressure at admission valves 12 at. gage. All bearings have force feed lubrication. 1. 550 mm cylinder bore, 650 mm stroke, 158 r. p. m. Two main bearings 250 mm dia. by 360 mm long Crank pin . . , . . 250 200 Grosshead pin 150 220 Grosshead shoes 520 long 300 wide Piston rod 110 dia. Connecting rod length .... 5.5 times the crank radius. 2. 500 mm cylinder bore, 600 mm stroke, 165 r. p. m. Two main bearings ..... 230 mm dia. by 330 mm long Crank pin 230 180 Crosshead pin 140 200 Crosshead shoes 480 long 275 wide Piston rod 100 dia. Connecting rod length .... 5.5 times the crank radius. 3. 850 mm cylinder bore, 1000 mm stroke, 125 r. p. m. Two main bearings 375 mm dia. by 550 mm long Crank pin 375 310 Crosshead pin 240 ,, 350 ,, Crosshead shoes 800 long 500 wide Piston rod 150 dia. Connecting rod length .... 5.5 times the crank radius. 77 Sulzer Bros, also mention the fact that they have found the friction HP of their una-flow engines equal or slightly less than the friction HP of their tandem compound engines of equal power. The driving parts of a una-flow engine should naturally cause more friction than the parts of a tandem compound engine since the size, or rather diameter, of the bearings of a una-flow engine is larger on account of the higher piston load. In a una-flow engine the single piston carries live steam pressure, while the low pressure piston of a tandem compound engine only carries receiver pressure and the much smaller high pressure piston sustains the difference between the live steam and receiver pressures. On the other hand, the single una-flow cylinder with its one piston and one or two piston rod packings will cause less friction than the two cylinders with two pistons and three or four rod packings of the tandem counterflow engine. Steam cylinders arranged in tandem are furthermore subject to misalignment with accompanying binding of the moving parts, and the friction Fig. 5. caused by such misalignment may be considerable. The piston system in una- flow engines can always be supported at two points only, for instance by means of a crosshead and self-supporting piston, or on crosshead and tail rod support with floating piston. It is assumed of course that the metallic packing used is of such design as to permit of lateral movement of the piston rod. Furthermore, the tandem counterflow engine requires four times the number of steam distri- buting elements as the una-flow engine (8 valves against 2 valves), with a corre- sponding increase in friction. A comparison between a una-flow and a cross-com- pound engine will still further emphasize the advantages of the una-flow system. According to what is said above, the lesser friction of the una-flow cylinder must outweigh the increased friction of the una-flow driving parts, if the experience of Sulzer Bros, is accepted as having general application. That part of the engine friction caused by the piston, especially if self-sup- porting, may be considerable. This friction may be reduced by fitting the piston with shoes of bronze, babbitt or Allan Metal. The friction is least with a floating piston, i. e. a piston supported by its rod, despite the additional friction of the second stuffing box and tail rod support. 78 The una-flow piston should be made as light as possible, and this may be accomplished by constructing it in two parts of cast steel (See Fig. 5), thereby reducing friction, inertia and impact. Next to the piston, the main bearing, crank pin and crosshead pin contribute the largest share to the total friction. The friction of high grade metallic packings is extremely small, as is also the friction of the poppet valves and their gear. This applies especially to una-flow engines with only two valves and one packing. The friction of the driving parts can be considerably reduced by a proper oiling system, especially by means of force feed lubrication. The latter type also reduces impact. The provision of force feed lubrication for the condenser pump driving parts and the use of a housing around the flywheel are further mea- sures in the right direction. Short stroke engines with bearings of large diameter naturally have a higher friction loss than long stroke engines. In engines having steam jackets on the cylinder barrel the friction is usually greater if the jackets are shut off. In the same way a new engine while being run in will have more friction than later, and the friction of an engine immediately after starting will be larger than when in regular operation, especially if it has not been warmed up previously. Taking it altogether one may say that the una-flow engine has a slightly better mechanical efficiency than the ordinary tandem compound engine. 79 5. Losses due to Leakage. Valves, Pistons, Piston Rod Packings. Tight steam distributing organs are a rare exception. Slide valves are gene- rally considered to be tighter than piston valves, this being the reason for the practice of many concerns to use piston valves for the high pressure and slide valves for the low pressure cylinders, a practice which is also supported by pres- sure and temperature considerations. Corliss valves are fairly tight, but they are far from being absolutely tight. Well made piston valves fitted with snap rings may be considered fairly tight. Piston valves without rings should only be used in small sizes for saturated steam and must be made a good fit. Larger piston valves for use with superheated steam should always be equipped with snap rings on account of the necessary clearance required for expansion. Even then a certain amount of leakage .must be expected in those positions in which none or one ring only is active, in addition to the constant amount of leakage past the ring joints. In case of highly superheated steam, carbonized oil may be the cause of increasing leakage. Double beat valves are usually leaky. The leakiness increases with the amount of balance, the pressure and the temperature. With all types of valves leakiness will be enhanced with increasing superheat, on account of warping and the increasing fluidity of the steam. The body of double-beat valves as shown in Figs, 1, 2 and 3 will sustain a heavy load in the direction of the axis during the expansion and exhaust periods. The corresponding deflection will cause the lower face of the valve to leave its seat and leak. The radial forces can be neglected if the seats are made flat. In the same way, if the temperature of the valve is higher than the tempe- rature of the material forming its housing and seat, then the valve body will expand more than the latter, and the upper valve face will lose contact and start to leak. The above temperature difference may have several causes. In a valve design as shown in Fig. 1, in which valve and seat have the same height and the same thickness of material, equal expansion, in the most favorable case, will Fig. 1. Fig. 2. 80 occur only if the material of both parts has the same coefficient of expansion. This can be realized by due attention to the work of the foundry. Both parts are exposed to live steam temperature on one side and to the varying temperature of the cylinder steam on the other. Conditions are not so good in the design shown in Fig. 2, in which a valve cage of the ordinary type is used. Valve and seat are of different thicknesses and therefore expand unequally, especially during the first period after starting. Fur- thermore, the two parts will have different temperatures during operation, since the valve is exposed on one side to live steam, while the cage is entirely surrounded by cylinder steam. The most unfavorable design is the one shown in Fig. 1, page 4, which has the valve seats cast in one piece with the cylinder head. The difference in expansion between valve and seat, especially just after starting up, must be considerable. The coefficients of expansion and the mean tempera- ture of both parts will certainly be different and the corresponding leakage will therefore be considerable. Fig. 3. Fig. 4. The valve will leak the more, the higher it is. The first rule in poppet valve design is therefore to make the valve as low as possible. For this reason, valve gears should be avoided which at late cut-offs give unnecessarily large valve lifts and therefore require high valves, unless a restriction of the upper passage is not objectionable. It is further advisable in cases where valve cages or similar con- structions are objected to on the ground of the number of tight joints required, to use a kind of saucer to form the lower valve seat, thus at the same time reducing the height of the valve. (See Fig. 3.) In this design the vertical forces as well as the deflection of the valve body are reduced on account of its small height and the small radial width of the valve ring. By grinding in at the operating temperature, a close approximation to complete tightness for one particular pres- sure and temperature will be obtained. During the first period after starting and at any other pressure and temperature the valve will leak. Perfect tightness under 81 all conditions can be accomplished with a resilient valve such as is shown in Fig. 4. The lower valve face is smaller in diameter than the upper one. This difference produces a force in addition to the spring load, which tends to press the lower rigid face as well as the resilient upper face against the corresponding seats. The upper resilient face also takes care of unequal expansion. In order to make the valve low and to reduce its deflection, a saucer forming the lower valve seat may be used to advantage. Dirt in the steam, however, may cause leakage even with this type of valve. Theory of the Resilient Valve. The following forces act on the valve: Downwards: 1. The spring pressure minus the steam pressure on the stem, where p a is the absolute pressure in the valve chest. 2. The pressure on the upper side of the resilient annular ring, where p t is the absolute pressure in the cylinder. Upwards : 3. The pressure on the lower side of the lower annulus (r* Q *)n(p a ~ Pi ). 4. The upward reaction of the upper valve seat W v 5. The upward reaction of the lower valve seat W z . The horizontal forces balance each other. The sum of the vertical forces must be zero, so that P + (& Q z )n( Pa Pi )-(r* Q*)n(p a - Pi ) W, W 2 = . (1) If the radii, the steam pressures and the spring pressure are known, only the bearing reactions W 1 and W z remain unknown. W 1 may be found from the condition that the deflection f of the resilient ring due to the steam pressure (measured at the middle of its face), must be equal to the deflection / 2 caused by the seat reaction W^ provided the lower valve face is to remain on its seat. Therefore f l = / 2 . Imagining a piece cut out radially from the valve with the angle dcp at the center, then the steam pressure acting on this element of the resilient surface is and the corresponding deflection is *>(*-<>) (p. -p0 .... (2) where E = the modulus of elasticity and J = the moment of inertia (approxi- mately constant) of the cross-section at right angles to the plane of bending. Stumpf, The una-flow steam engine. 6 82 The seat reaction on the radial element under consideration is W 1 H^-. and 2n the corresponding deflection is ^ = Wl ' In' 3E-J ' ' ' (3) Then on equating / x = / 2 W l = -Q- n (R -\- Q) (R Q) (p a p^ (4) From equation (1) Substituting in (5) the value of W as found in (4), then In the limit, when the valve just rests on the lower seat, W 2 = 0. Neglecting P, the excess of the spring pressure over the steam pressure on the valve stem, the highest permissible value for Q may be obtained from the following equation: "8" '~ r + ~s^'- < 7 ) Denoting R by Q -f- a, and r by Q -f- 6, equation (7) becomes 10 or since a-\-1q and 6 -f- 2p are approximately equal, If the valve expands by the amount A I in excess of the casing, in consequence of unequal temperatures and coefficients of expansion, the resilient ring must deflect by the same amount in order to remain steam tight. In this case /, / 2 is not zero but is equal to 41 _ ./ / /cn //t /i /a ......... (v) The seat reaction W-^ for any given A I is obtained by inserting the values for /j (from equation 2) and / 2 (from equation 3) in equation 9; hence For the valve to be tight, W 1 must be positive, for which purpose a pressure difference (p a Pi ) is necessary, as may be seen from equation (10). The mini- mum pressure difference required for tightness may be obtained from equation (10) by making W^ 0, i. e. d 83 From this equation it will be seen that the pressure difference is proportional to Al and consequently to 1. In order to keep (p a />) small, the valve must be low. The dimension d is determined by the strength of the material, and the values of R and Q by the desired steam velocity. To find the bending stresses of the resilient ring during operating condi- tions; let ^ = the temperature of the valve, Q! = the coefficient of expansion of the valve material, t 2 = the temperature of the surrounding casing, a 2 = the coefficient of expansion of the casing, I = the distance between valve seats at normal temperature t . Then Al = l\a (t t) a (t Ml (12) Substituting this value in equation (11), the pressure difference (p a p^ will be found at which the valve will commence to be tight. At all greater pressure differences the valve will be perfectly tight. For example, a valve is taken in which / = 30 mm, R = 125 mm, Q = 104 mm, d 3 mm, p a = 12 at, and p t = at. Assume also that t =15 *! - 300 % = 0.000012 a 2 = 0.000011. The following table gives the relative expansions Al of this valve for certain temperature differences (t Z 2 ), calculated according to equation 12. In this table are also found the stresses K^ and the pressure differences (p a p^ neces- sary for tightness calculated by means of equation 11. t l t z 50 100 150 200 Jl = in mm 0.009 0.0255 0.042 0.0585 0.075 Pa ~ Pi = in at abs. 1.35 3.8 6.3 8.75 11.25 K = 230 645 1070 1490 1920 From these figures it will be seen that many prevalent valve constructions are far from being steam tight. The excessive height of many valves, neglecting other constructional errors, makes steam tightness absolutely impossible. Even in a valve only 30 mm high, with a temperature difference between valve and seat of 200 G., a steam pressure of 11.25 kg/sqcm is necessary in order to obtain steam tightness; i. e. at all lower pressures there will be leakage. 6* 84 Calculations for a Resilient Valve of 160 mm mean diameter, for a condensing engine working with steam of 9 at. abs. (Fig. 5). The following two condi- tions are assumed for the cal- culations of W l and W 2 : 1. The actual closure oc- curs at the mean circum- ference in both the top and bottom seats (Radii R m and r m ). 2. In the most unfavorable case closure occurs at the inner edge of the upper seat (Ri) and the outer edge of the lower seat (r a ). In the latter case W 2 on the lower valve seat is to be zero. In Fig. 5, Fig. 5. = 81.5 mm = 76 mm Ri = 80 mm r a 77.5 mm 6*71 Q = 67.5 mm d = 17 mm. The steam pressure on the valve stem is (p a 1) =18 kg. Wj, according to equation (4) above, is 221 kg, for R m and r m 195.5 kg, for R t and r a . W 2 = in case 2. The spring pressure F is then calculated from equa- tion (5), (Pa-1) ( Pa - Pi ) (R? n-r*. n] W l = 18 + 195.5 107 = 165.5 kg. With this spring pressure in case (1) W 2 = P +(p a Pi) (R m * n r m 2 n = 88.5 -f 245 221 = 112.5 kg. The thickness d of the resilient ring is calculated for the case in which the valve is opened at the greatest pressure difference p a p,, the valve and seat being assumed to have the same temperature. Again taking a radial element d

= 10.5, and the time for 1 of crank angle t 1 / 600 sec. For any other speed n the valve velocity v must be multiplied by TT-T- and the acceleration by Time or crank angle is taken as the abscissa for all calculations. The investi- gation will extend only to the point at which the cam lever reverses, since the j-i Fig. 11. subsequent closure of the valve is an exact repetition of the phases during lifting. The necessary dimensions of the valve gear parts are assumed to be known or are determined according to Figs. 12, 13 and 14. From Figs. 13 and 14 the travel of the cam lever (a = KK'} can be found corresponding to the crank angle a. 90 \ Fig. 12. This corresponds to the part MM' of the valve lift curve (equidistant from the cam profile through the center of the cam roller) and the valve lift h. The cam lever therefore changes a I into h. The line / K in Figs. 12 and 13 corresponds to the z-axis in Fig. 14. The distance a = KK' corresponds to the angular displacement /? of the cam lever. Imagine the roller to move along the cam (the latter supposed held stationary) from M to M', corresponding to the angle /5. ML is the center line of the valve stem. Drawing M' L' tangentially to an arc struck from center N with radius N L, the base circle MM' will be intersected at a point whose distance from M' =h= valve lift, corresponding to a and p. The valve lift curve M M'Q is obtained from the cam profile by increasing or decreasing the radius of the latter by the roller radius. O^M is usually from 3 to 10 mm, and O^M' = roller radius -f- 3 to 10 mm. The cam profile as well as the valve lift curve consist of two circular arcs with a common tangent between them. As long as the roller remains on the arc of the cam with center O^ it is advisable for the sake of accuracy to take the in- crements of the crank angle a equal to 2, while, later on, increments of 5 are sufficient. Fig. 15 shows the valve lifts thus determined plotted against the crank angle a. ^ max cor- responds to the dead center position of the eccentric. This curve is also important for the determination of the admission line (see chapter on losses due to throttling). Instead of now differentiating this valve lift curve to find the valve velocities, the latter may be determined in a more exact manner from the in- stantaneous angular velocity and the corresponding lever 91 arms. The velocity of the eccentric rod H K in Fig. 12 is found from Fig. 14 to be >! = V - 1 = r 2 n r n bO = co co r-, The angular velocity of the cam lever is o^ = - (Fig. 13). The "roller r 2 contact line" M'0 2 P always intersects the roller center and either of the centers O : or <9 2 , or i g at right angles to the straight middle piece of the valve lift curve. The velocity along the roller contact line is v' = o^ r 3 '. For the velocity along the relative valve stem axis M' L' (valve velocity y), 7*3 has to be considered instead of r 3 ', since the right-angled triangles NSP and M' TU are similar, and v : v' = r 3 : r 3 '. Therefore the valve velocity v r 3 = co r 3 . The distances r-,, r 9 and r, are to be measured in meters. The roller contact line M'0 Z P coin- cides with MN for the point of valve opening and therefore r s = 0. The same occurs when the roller reaches the upper land and M' falls on Q, in which case also r 3 = 0. The valve velocities thus determined are plotted in Fig. 15 on a crank angle basis, giving the curve y, which indicates a rapid increase of the velocity up IT Fig. 14. Fig. 15. to the point 7\, corresponding to the point at which the lifting curve and straight portions merge, after which it decreases until it reaches zero for the dead center position of the eccentric, or for the point at which the upper curve and upper land run together. The velocity u =. corresponds to the dead center position of the eccentric after which the whole procedure is repeated with the signs reversed. In case the roller reaches the upper land or runs a certain distance on it, a corresponding part of the velocity curve coincides with the #-axis. The maximum valve acceleration p = corresponds to the steepest tangent Wv 7\ T 2 of the velocity curve. Receding from point T z a distance equal to 6 crank angle or 1 / 100 sec, an ordinate at this point will represent the change of velocity in Vioo sec - In this case the latter amounts to v x = 0.205 m/sec. The accelera- 92 tion is consequently p =- - ==20.5 m/sec. If the scale of velocity is chosen /ioo so that 100 mm = 1 m/sec, then the ordinates in mm give directly the accelera- tions in m/sec. The accelerating force during the period of increasing velocity is exerted by the eccentric through the cams, and for decreasing velocity the retarding force must be produced by the valve spring. The opposite holds true for the closing period of the valve. The force to be exerted by the valve spring depends upon: 1. The maximum acceleration p max to be produced by the spring. 2. The weight G of the valve and the parts connected to the same. 3. The friction of the valve stem, valve and roller head. 4. The steam pressure upon the valve stem area. The weight G is generally assumed to be balanced by the friction and there- fore neglected. g The accelerating force P = jz-rrr p max . An additional 10 to 20% are neces- y.oj. sary to take care of inaccuracies in the construction of the valve gear and the increased friction during the first period of operation. The steam pressure on the valve stem area may either relieve or oppose the spring and must in every case be considered. Torsional stress of the spring material k d = about 3500 kg/sqcm. Pistons. Pistons may be classified as: 1. Self-supporting pistons. 2. Floating pistons. 3. A form intermediate between the other two. All questions concerning the design, construction and operation of the piston should receive thorough consideration. The self-supporting piston is undoubtedly the most difficult, and the floating piston the easiest to deal with. The self- supporting piston together with its cylinder form a bearing. The first condition for satisfactory operation is therefore sufficient difference in the properties of the materials of which the two parts are made. For instance, a steel shaft runs quite satisfactorily in a babbitt, phosphor bronze, brass or cast iron bearing. All these combinations present sufficient difference in the properties of the materials employed. An exception exists in the combination of hardened steel on hardened steel which may frequently be found in valve gear joints. Babbit on babbitt, bronze on bronze, or mild steel on mild steel never work together satisfactorily. Cast iron pistons, however, can be made to work well in cast iron cylinders, as is shown by the una-flow engines built by Sulzer Bros. Cast iron is a collective designation which includes materials of very heterogeneous composition. Sulzer Bros, after extensive experiments have found the proper mixtures for the cylinder and piston, which possess sufficient difference in properties to work together safely and satis- factorily. If a designer lacks sufficient faith in his foundry he will do well to equip his piston with a babbitt or bronze mounting in order to provide materials with 93 sufficient difference in properties. A cast steel piston should always be equipped with such a bearing surface. There are still, however, concerns who try to make piston and cylinder from the same mixture, and in addition to this cardinal error commit others of equal consequence which result in certain failure. There are also materials which do not work well together despite sufficient difference in their properties, such as for instance, cast steel on cast iron. Even with a bronze mounting the difference in expansion between these two materials must be taken into con- sideration. A journal and bearing must ordinarily have sufficient clearance to allow room for the oil film. This condition applies all the more to piston and cylinder since the dimensions are larger and the temperature difference is greater. A clearance of 3,5 to 4 thousandths of the diameter between piston and cylinder bore have been found satisfactory. Machining the piston by first turning it to the exact cylinder diameter and afterwards turning it off eccentrically so as to produce a bearing surface for about 90 to 120 has not proved satisfactory for superheated steam. This method gives enough clearance at the top but on account of the higher temperature of the piston and its greater expansion, the weight concentrates at the edges of the bearing surface and the piston is likely to seize. A proper way of finishing the cylinder is to bore it barrel-shaped, or machine it while heating the ends, for instance by passing steam through the jackets and cooling the center by blowing air through the exhaust belt. The heads of a una-flow piston expand more than the center by reason of their higher temperature, and should therefore be of smaller diameter than the center. They should be out of contact with the cylinder and only act as plungers. The part of the piston forming the bearing surface should be rounded off liberally at its ends to prevent it from scraping the oil off the cylinder wall. Briefly, the endeavor must be to produce a piston and cylinder with exact cylindrical surfaces and sufficient clearance, and to maintain this condition at high temperatures. If this can be accomplished for the long una- flow piston with its large bearing surface and temperature difference, the most difficult part of the problem is solved. The case of the floating piston which is carried by the piston rod is much simpler. A radial clearance of 2 to 3 mm, according to the size of the cylinders, may be used, so that no consideration of bearing action is necessary. Only the piston rings project beyond the piston surface and are in contact with the cylinder wall. The tail rod can either have a stationary bearing behind the stuffing box or be carried on a crosshead. The great length of the piston rod between the bearings caused by the long piston, makes a light cast steel construction and a rod of large diameter a necessity. From a thermal point of view the floating piston is to be preferred, since only the piston rings transmit heat from the hot to the cold por- tions of the cylinder, while in the case of the self-supporting piston the large bearing surface also takes part in this action. When using a stationary bearing for the tail rod behind the stuffing box there will be a rise and fall of the piston at every stroke, which must be considered. If the cast steel used is very soft, then the cast iron rings are liable to seize. The long and heavy piston together with the long span of the rod usually bring about a condition which is a mean between the floating and self-supporting 94 constructions. Part of the weight is then carried by the cylinder wall and part by the piston rod. The wear of the cylinder tends to alter the weight distribution in such a way as to increase the part carried by the piston rod and thus relieves the piston and cylinder bearing surfaces. The piston rod also resists possible forces acting on the outside of the piston. If for instance, the piston rests in the cylinder and by reason of leaky rings the steam obtains access to the clearance space above the piston, then a heavy downward load will result. Floating pistons offer greater safety against this possibility, this safety being imparted by the piston rod. This applies especially to vertical engines where the piston, if not guided by a tail rod, is in a condition of unstable equilibrium and is liable to slap. This may even be occasioned by the piston overrunning the inlet ports, which should be fundamentally avoided. Such overrunning can only be permitted in the case of floating pistons, although even then the possibility of vibration should be reckoned with. The greater number of mistakes which give rise to lateral forces are incurred in the arrangement and construction of the piston rings, which latter should pro- tect the piston surface against such forces. For this reason they should be placed as far as possible towards the ends of the piston, in order to leave as little area as possible for the formation of lateral forces. Even if this is done there still remains some possibility of an unbalanced load, especially with the large surface of a una-flow piston. If, for instance, the rings of a horizontal engine are not secured against creeping, their center of gravity, being located eccentrically oppo- site the joints, will move to the lowest possible position and all the joints will fall in line at the top. The steam leaking through them will then undoubtedly exert a heavy pressure upon the large surface, thus forcing the piston downwards and causing rapid wear. This cannot happen with a floating piston since the pressures will equalize in the annular space and temporary lateral forces will- be resisted by the piston rod. The ring joints of floating pistons and of pistons having the rings on the plunger heads should therefore be equally spaced over the circumference; for instance, where three rings are used at each end, the joints should be at 120. In self-supporting pistons without plunger heads, the ring joints should be kept within the bearing area; thus for three rings one joint may be arranged in the center and one each at 30 to the right and left. The bearing surface then pro- tects the ring joints against the steam. With such an arrangement complete tight- ness may be attained if the workmanship is good and the rings are sufficient in number. In floating pistons the action of the rings is similar to a labyrinth packing which always passes a certain amount of steam, since the ring joints can never be made absolutely tight. The pressure ratio, in case of una-flow engines always being above the critical value, the weight of steam flowing past the ring joints may be calculated by means of the formula in which / denotes the free area of the joint, z the number of joints in series, p^ and v l the absolute pressure and the specific volume of the steam inside the cylinder. In Fig. 16 are illustrated five different types of ring joint fastenings which can only partly render the joints tight, but perform the important addi- 95 to 96 tional function of securely locking the rings against creeping. In a una-flow piston where the rings are usually mounted on the piston heads which expand conside- rably, the locking elements must in no case project to the cylinder wall. The slot at the ring joints must be from 2 mm to 5 mm wide, according to the diameter of the rings, in order to allow for expansion. Friction will cause the rings to assume a higher temperature, especially with superheated steam and poor lubrication. If the clearance provided is insufficient, the joints will close and the rings expand against the cylinder wall, so that increased heating, greater expansion and a heavier pressure result, which may lead to a complete destruction of the cylinder surface and rings. Dimensions of Concentric Cast Iron Piston Rings in mm. Cylinder Bore] Piston Ring Length of piece cut out Thickness Width 300 12 12 24 ' 400 15 15 35 600 20 20 GO 800 22 22 84 1000 25 25 108 1200 28 28 132 1400 30 30 155 1600 32 32 180 1800 34 34 205 2000 36 36 230 The different phases in the manufacture of ordinary piston rings are pre- sented in Fig. 17. First is shown the rough casting of sufficient width to hold it in the lathe, rough turning and cutting off follow; a piece is then cut out and the ring closed up for finish turning to the correct diameter. It is not possible to obtain a uniformly distributed pressure with concentric rings, and only a very rough approximation to this condition can be reached with excentric rings whose thickness increases from the joint to the opposite side. The concentric type, however, is preferable in order to avoid a one-sided center of gravity and a large clearance in the groove behind the ring. Attention may also be drawn to what are known as hammered rings. These are made of Swedish iron, turned to correct diameter and width in one operation, are then split, and afterwards given the required tension by hammering. Approxi- mately uniform pressure distribution, greater strength and reliability in operation are obtainable with this form. Even in small sizes they may be sprung sufficiently to be slipped into the piston. Mention should also be made of the piston rings designed by Schmeck (Fig. 18), which are made in sections whose joints are secured by spring-loaded plugs which prevent them from creeping and force them against the cylinder wall. Rings of this type have the advantage that they are practically tight, especially at the joints, that their bearing pressure is uniform, they can be easily assembled and disassembled, and adapt themselves to warped cylinders. The cast iron plugs 97 and ring sections are finished in such a way as to fit the cylinder bore. This type of ring is much used by the Hannoversche Maschinenfabrik vorm. Egestorff and is reported to have given complete satisfaction. n. V,06S'- i ro V5 -19 60 f- Fig. 17. It is advisable not to permit the outer ring to overrun the cylinder bore. Such overrunning exposes part of the ring surface to the steam pressure, thus causing the ring to collapse and destroying its function of tightness for at least a certain distance near the dead center. An ordinary cast iron ring, especially if the de- Fig. 18. flections are large, will not withstand the stresses produced thereby for any length of time. The clearance behind the rings should be made as small as possible, so as to reduce the deflection and leave as little space as possible for the accumula- Stump/, The una-flow steam engine. 7 98 tion of steam which would press the ring against the cylinder wall and cause con- siderable wear both on the ring and cylinder, especially in the middle of the latter. With superheat and dirt in the steam a ring may under these conditions lose as much as 5 to 10 mm in thickness in a few weeks, and the cylinder bore several millimeters, especially in the center. A wide ring will be the most subject to this destructive action. Small width, high grade material, no overrunning, small clearance behind the rings, well secured joints, and a good fit in the grooves are therefore advisable. If the cylinder is made of a fairly hard, close-grained cast iron, then no ridges will form, thus eliminating any reason for al- lowing the rings to overrun. All the foregoing is supported by the experience which the author gained with a piston packing of the type shown in Fig. 19. Both rings overran the cylinder bore. The rings, in collapsing, had to push the spring along their sloping surfaces, with the effect that both rings and spring went to pieces. Fig. 20 shows the pieces which the author found in the cavity of the piston. The spots where wear occurred on the springs are clearly visible in Fig. 2Q. The number of rings used should vary according to the pressure range. The Fig. 19. the rings at the ends of a una-flow piston, with both sets active during the first part of the stroke, fulfills this, requirement in the best possible manner. A una-flow piston should always be ma- de in two or three parts, held together by piston rod and nut. The castings in this case are simple, light and without core plugs, which will be especially appreciated in the case of cast steel. .It is advisable to test the piston with water pressure if the foundry cannot be relied upon. A cast steel piston of two-piece construction is shown in Fig. 5, page 77. Each half carries a bronze shoe fastened to it with copper rivets, covering an angle of 120. The two halves, fitted with three rings each, have radial clearance over their whole circumference and are made as light as possible in order to reduce inertia. Fig. 20. 99 The ring joints have a labyrinth effect. Fig. 40, page 157, shows an older piston of one-piece design having grooves fitted with Allan metal rings. The latter are made to project about 1 mm above the surface of the piston when new, and during the first period of operation part of this metal is transferred to and fills out the pores of the cylinder surface. Both these rings should be placed in the middle of the piston. Una-flow pistons for locomotives are always made in three pieces, i. e., two heads carrying the piston rings, having clearance all around, and a center sup- porting piece. It was formerly customary to make the latter of hard steel, but Swedish iron is now used with better results. The heads which are made of cast or forged steel, expand considerably under the action of superheated steam. This expansion is transmitted to the center piece and must be considered in designing the latter. In order to reduce the weight of the reciprocating parts to a minimum, Fig. 21. locomotive pistons are always made without tail rods and have given satisfaction except for minor troubles. These have been due to errors in the composition of the material and to disregard of the effect of the expansion of the heads upon the center supporting piece. The experience obtained with such pistons, as well as the favorable observations of Sulzer Bros., indicate that the problem may be solved with self-supporting pistons, provided the important requirement of a reliable lubrication system is satisfied. One oil feed on top, and one each in or below the horizontal plane on both sides of the cylinder, every feed being connected to a separate force feed pump, will give satisfactory results with the proper kind of oil. The pump plungers should be timed to deliver oil only during the periods when the corresponding orifices are covered by the piston. If the feeds are con- nected to the cylinder ends, carbonization of the oil is to be feared; and if the oil is fed to the center of the cylinder at the time of exhaust, it is likely to be blown through the ports. Both these conditions lead to a high oil consumption which is sometimes complained of in connection with una-flow engines. The time during 7* 100 Q which the piston covers the oil feeds is longest when the latter are in the middle of the cylinder, so that even with pumps having a continuous feed there is a reaso- nable certainty of oil being car- ried between the rubbing surfaces. In order to avoid losses due to the exhaust, it is permissible to arrange the three feeds close to one side of the exhaust belt in- stead of in the center. In regard to lubrication also, the floating piston offers greater safety, since, if correctly constructed, the rub- bing surface is limited to the rings. If, however, the steam obtains access to the spaces behind the rings, heavy friction and large oil consumption may result. With self-supporting pistons the sanje effect may be caused by lateral forces. Everything considered, it may be stated that the self- supporting piston requires greater care in regard to design, choice of material, lubrication and ope- ration, but has the advantage of not requiring a tail rod with its bearing or crosshead. The floating piston has greater reliability but is more complicated and increases the floor space required. Self- supporting pistons may however be made to operate satisfactorily. Piston Rod Packings. Soft packing is used only in small and cheap engines, while metallic packing is the rule with steam of high pressure and superheat. The Lentz packing shown in Fig. 21 consists of a plurality of one-piece cast iron rings, whose number varies according to the pressure to be carried. These 0, 3 .S o> 02 6X3 s 101 rings are fitted to the rod and work between the ground surfaces of a corresponding number of housings, thus producing a labyrinth effect. The individual housings form metal to metal joints and are pressed against the bottom of the packing space by means of the outside gland, sufficient clearance being provided to allow the cast iron rings to move laterally. The last chamber collects the water of con- densation so that it may be drained away. With pure steam, absence of dust, and good lubrication (the oil being preferably forced into the packing under pres- sure), satisfactory results should be permanently obtainable. The one-piece rings may sometimes be found unhandy in assembling. The American type of packing shown in Fig. 22 is better in this respect, since it may easily be assembled and disassembled and offers greater freedom in the design of adjacent parts. Each ring is made in four pieces, the two oppo- site segments with their babbitt lining being pressed against the piston rod by means of springs. The second ring of similar construction is set at 90 to the first one. The two remaining segments of each ring are forced by springs against the other segments. Botli rings are properly fitted to the housing at their joints and relatively to each other. The whole packing system is held by means of axial springs against a spherical seat, thereby accommodating itself to inclined positions of the piston rod, while any lateral movement of the same is provided for by the sliding fit of the rings in their housings. This packing is occasionally stated to be unsatisfactory for vacuum, although this criticism may be unjust. The Duplex packing shown in Fig. 23 which is equipped with an additional set of conical bab- bitt rings, is equally satisfactory for both vacuum and high pressure. Different 102 Fig. 24. f C Ct f7 Fig. 25. Fig. 26. 103 springs are supplied for various pressures. Good workmanship is claimed for this- packing. Pure steam, regular and ample lubrication, as well as frequent use of the drain cocks, especially in running in, are essential for success. Packings of a similar type of Simplex and Duplex construction are offered by the firm of Max Dreyer & Co., of Magdeburg. The Proell packing (Fig. 24) is based on a similar principle. Each cast iron ring is cut into six parts which "are held together by means of a coil spring, the joints of one ring being staggered in relation to those of the adjacent one. A pair of such rings is contained in each housing, which is easily removable by means of an internal lip. These housings come together on metal to metal joints and thus form chambers in which the rings have suffi- cient play to enable them to move late- rally. The whole packing is held together and against the bottom of the packing space by the outside gland. The oil is either fed onto the piston rod or forced between the middle rings. This packing, however, does not accommodale itself to inclined positions of the piston rod. This packing is furnished in Simplex, Duplex or Triplex forms, according to the number of rings employed. The Kranz packing (Fig. 25) made by the Elementenwerk Kranz, of Ludwigs- hafen, also employs cast iron rings in pairs, each cut into three parts, sur- rounded by sectional primary housings held together by coil springs. These pri- mary housings are radially movable between the ground surfaces of the secondary hou- sings, provision being also made for axial expansion of the former. Oil may be fed under pressure to the packing, although it is claimed that a drip feed to the rod its satisfactory. The use of three pairs of rings with a large number of cavities results in a thorough labyrinth effect. The water of condensation is caught by a further pair of rings on the outside, so that it may be drained away. 104 The packing designed by Wilh. Schmidt (Figs. 26 and 27) takes care of in- clined positions of the piston rod in the best possible manner. The packing as a whole is inserted between two rings having spherical surfaces with a common center, these in turn being held between two flat surfaces so that both lateral and rotative movements are rendered possible. A deep recess insures further flexibility as well as a cooling effect. The segmental babbitt rings of conical sec- tion are held together by a powerful spring. A special fitting containing a felt ring and having an oil connection serves to lubricate the rod as well as to keep out dirt. This packing has been very successful on locomotives. 105 6. Losses due to Radiation and Convection. It will be seen from a comparison of the una-flow with the ordinary com- pound counterflow. engine that the former can have only small radiation and convection losses. The radiating surface of the counterflow engine with its two cylinders, receiver and accessories is two or three times as large as that of the una-flow engine, with correspondingly higher losses. The loss due to radiation of the una-flow cylinder is very small compared with the radiation losses of the steam pipe, for which reason the latter will be dealt with first. Assuming a flow of superheated steam at a very small velocity through a pipe having a length of 100 m, then the steam at the far end will have a lower temperature and a cor- respondingly larger specific weight, (y x : v 2 = T l : T z ) but practically the same pressure. The highest point E in the temperature-entropy diagram shown in Fig. 2, chapter I, 3 a, corresponds to the state of the steam at the entrance of the pipe, while the lower point Q on the same pressure line represents the state of the steam at the far end. The narrow vertical strip EFWQE below the part EQ of the pressure line, extending down to the line of zero temperature ( 273), represents the total amount of heat lost; but as the heat represented by the area of the diagram below the back pressure line cannot be utilized, the actual radia- tion loss is represented by the strip ER VQE. Insulating and lagging the pipe can therefore only result in a mere reduction of this loss. A further radiation loss occurs in the cylinder and will show itself in a very slight deviation to the left of the vertical adiabatic line. Insulating the cylinder can therefore only tend to reduce this slight loss. Much more important is sufficient insulation around the cylinder heads, which really form part of the steam pipe. The live steam pipe in high grade plants is always covered while the cylinder is provided not only with insulation, but lagging as well. The latter supplements the effect of the insulating material in an efficient manner and may, if constructed of several casings one within another, with a bright inner surface, entirely take its place.- All flanges should also be covered. The materials used for insulating steam pipes and cylin- ders are Kieselguhr, asbestos, magnesia, cork, or glass or textile waste. The more porous the material is, and the thicker the layer, the better will be the insulating effect. Part of the heat is lost by radiation and part by convection. The process is so complicated that mathematical treatment fails completely and actual tests have to be relied upon. The following table will help to clear up matters. 106 Maximum drop in temperature per 1 m length of pipe, for 14 at. gage; Outside steam pipe insulated, without lagging. diameter Thickness of 300 "C 350C 400C of steam Insulation steam temperature steam temperature steam temperature 80 60 40 80 60 40 80 60 40 mm mm m/sec m/sec m/sec m/sec m/sec m/sec m/sec m/sec m/sec 60 0.08 0.11 0.14 0.10 0.14 0.18 0.13 0.19 0.25 108 80 0.06 0.09 0.12 0.08 0.12 0.16 0.11 0.17 0.22 100 0.05 0.08 0.11 0.07 0.11 0.15 0.10 0.16 0.20 60 0.06 0.09 0.11 0.08 0.12 0.15 0.10 0.16 0.20 133 80 0.05 0.08 0.10 0.07 0.10 0.13 0.09 0.13 0.17 100 0.045 0.07 0.09 0.06 0.09 0.11 0.08 0.12 0.15 60 0.045 0.075 0.09 0.06 0.09 0.12 0.08 0.12 0.16 159 80 0.04 0.065 0.08 0.055 0.08 0.11 0.07 0.11 0.14 100 0.035 0.055 0.07 0.05 0.075 0.10 0.065 0.10 0.12 60 0.04 0.06 0.08 0.05 0.075 0.105 0.07 0.105 0.13 191 80 0.035 0.05 0.07 0.045 0.07 0.09 0.065 0.95 0.12 100 0.03 0.045 0.06 0.04 0.06 0.08 0.055 0.85 0.11 60 0.033 0.05 0.065 0.045 0.07 0.09 0.06 0.09 0.12 216 80 0.03 0.045 0.06 0.04 0.06 0.08 0.055 0.08 0.11 100 0.025 0.04 0.05 0.035 0.055 0.07 0.05 0.075 0.10 60 0.03 0.045 0.06 0.04 0.06 0.08 0.055 0.085 0.11 241 80 0.025 0.04 0.05 0.035 0.055 0.07 0.05 0.075 0.10 100 0.023 0.035 0.045 0.03 0.05 0.06 0.045 0.07 0.09 60 0.028 0.042 0.055 0.038 0.057 0.075 0.053 0.08 0.10 267 80 0.023 0.035 0.046 0.033 0.05 0.065 0.045 0.07 0.09 100 0.02 0.03 0.04 0.028 0.042 0.055 0.04 0.06 0.08 60 0.025 0.038 0.05 0.033 0.05 0.065 0.045 0.07 0.09 292 80 0.023 0.034 0.042 0.03 0.045 0.060 0.04 0.06 0.08 100 0.020 0.028 0.036 0.025 0.044 0.050 0.035 0.053 0.07 60 0.022 0.033 0.043 0.03 0.045 0.06 0.042 0.06 O.OS3 318 80 0.020 0.030 0.028 0.026 0.039 0.052 0.038 0.056 0.075 100 0.018 0.027 0.034 0.023 0.034 0.045 0.034 0.05 0.068 60 0.02 0.03 0.04 0.027 0.041 0.055 0.037 0.058 0.073 343 80 0.018 0.027 0.035 0.024 0.036 0.048 0.033 0.05 0.065 100 0.015 0.023 0.03 0.022 0.032 0.041 0.029 0.043 0.057 60 0.018 0.028 0.037 0,025 0.038 0.055 0.035 0.053 0.07 368 80 0.016 0.025 0.032 0.023 0.034 0.045 0.031 0.047 0.061 100 0.014 0.022 0.028 0.021 0.031 0.04 0.027 0.042 0.054 60 0.016 0.024 0.032 0.023 0.034 0,046 0.031 0.047 0.062 394 80 0.014 0.021 0.028 0.02 0.03 0.04 0.027 0.041 0.055 100 0.012 0.018 0.024 0.017 0.026 0.034 0.024 0.036 0.048 60 0.015 0.023 0.031 0.022 0.033 0.044 0.03 0.045 0.060 420 80 0.0135 0.020 0.027 0.010 0.028 0.037 0.026 0.039 0.52 100 0.012 0.018 0.024 0.016 0.024 0.032 0.023 0.035 0.44 107 It will be observed from this table that the heat loss for average thickness of insulating material is inversely proportional to the steam velocity. Higher velo- cities of course result in smaller diameter, circumference and surface of the steam pipe, thus also reducing the heat loss. Higher velocities are therefore advisable up to the point where the throttling losses become excessive (See chapter I, 3 a, especially Fig. 2). Heavy insulation (80 to 100 mm thick) is to be recommended. Assuming an outer temperature of 0., then the heat losses increase faster than the temperature gradient, according to the law of Stephan Boltzmann. If the steam cylinder is regarded as a pipe, the steam may be considered to flow through it with a velocity equal to the mean piston speed. Applying the above rule of the inverse variation of the heat loss with the steam velocity, it will be found that in view of the lower temperature the radiation losses of the cylinder must be extremely small, especially as the oil film and the lagging form part of the insulation. These losses in fact are so small as to appear negligible in com- parison with the other losses. 108 7. Losses due to incomplete Expansion. A loss -of diagram area within the limits of the piston stroke is caused by the fact that the exhaust begins with a certain exhaust lead or distance / before dead center (Fig. 1), and this loss increases as the exhaust lead f v and terminal expan- sion, pressure p e increase. For small exhaust lead and low terminal pressure this loss is negligible. Even in non-condensing una-flow engines in which a large ex- haust lead is used in order to soften the exhaust puffs, the loss of diagram area within the limits of the piston stroke is insignificant when compared with the lost work represented by the toe of the diagram, shown shaded at D. The pro- blem of finding a means of utilizing this work without increasing the cylinder dimensions is well worth while. The solution to be described later has the effect of reducing the pressure p u at which compression begins, with a consequent lower Fig. 1. terminal pressure. A smaller clearance volume may therefore be used, thus dimi- nishing the volume loss. Without going into calculations, it will be seen that the gain F at the compression line will be approximately proportional to the shaded area D, or in other words, the higher the terminal expansion pressure is, or the longer the cut-off, the lower will be the terminal compression pressure. This implies an increasing pressure difference during compression for an increasing pressure difference during expansion, a combination which has been proved desirable in the chapter on volume loss. This rule would be fulfilled in its entirety if the pres- sure changes on both sides, i. e. expansion and compression, were equal. Further- more, the use of a longer exhaust lead / would then become permissible, since the lost area within the limits of the piston stroke now forms part of the toe of the diagram, and co-operates in lowering the back pressure at the time compression begins. There can therefore be no objection to making f v large, since by increasing the duration of the exhaust, the compression is shortened and the exhaust puffs are softened. If this is done, the number of the exhaust ports will be so far reduced that the exhaust belt may eventually be dispensed with and only one port remains which connects directly to the exhaust pipe. Piston and cylinder also become 109 considerably shorter. The relation between length of cylinder L, length of piston 4, exhaust lead /, stroke /, and exhaust port diameter d, is as follows: J t = d + I If, L = 2 For instance, for a stroke of 660 mm, an exhaust lead of 25% and exhaust ports of 120 mm diameter, the piston length will be 450 mm, and the length of the cylinder 1100 mm as compared with a piston length of 594 mm and a cylinder length of 1254 mm for 10% exhaust lead. The distance f v is limited by the maxi- mum cut-off, since direct exhaust of live steam must be avoided. For locomotives the value of f v must be limited to 25%, while for locomobiles it may be taken as large as 30 to 35%. The utilization of the energy represented in the toe of the diagram is based upon its complete conversion into kinetic energy by means of conical nozzles such as are commonly used in steam turbines. Each exhaust puff would therefore act as a kind of wad or plug moving with a high velocity through the exhaust ru* Fig. 2 (Diise = Nozzle). pipe and finally creating behind itself a partial vacuum whose absolute pressure is p u . The exhaust pipe must therefore be long enough so that there will always be at least one such plug moving within it, thus preventing atmospheric pressure from reaching the nozzle and destroying the vacuum. The end of the exhaust pipe must form a diffusor to change the kinetic energy into pressure energy at atmospheric pressure. Fig. 2 shows such an exhaust pipe diagrammatically. The work corresponding to the shaded areas D and E (Fig. 1) is determined for various pressures p u . The weights of steam G e and G u , corresponding to the pressure p e and p u can be found from the diagram and the dimensions of the engine. Using the equation of work: G f G,, w 2 The velocities w are calculated for various values of the pressure p u and plotted against the latter, as shown in Fig. 3. Each value of p u has associated with it a certain velocity w d which must exist at the point of entrance into the diffusor in order that atmospheric pressure may be overcome. This velocity therefore cor- responds to a pressure difference (1 p u ) and may be easily obtained from the Mollier chart and also plotted in Fig. 3. Of course the steam when leaving the diffusor must in practice still have a certain velocity, and the pressure difference should therefore be reckoned not from the atmosphere, but from a slightly higher pressure corresponding to this velocity. This shifts the diffusor velocity curve 110 into a somewhat higher position, and the intersection S of this curve with the nozzle velocity curve, which determines the obtainable pressure p u , is moved slightly towards the right corre- sponding to a higher value of p u . Friction losses in the long exhaust pipe cause a further loss of velo- city >!, between the nozzle and -^>. ~ the diffusor, and this again re- x ^^; ^i fi t " suits in a shift of the point of r\^4Jlo7~"" intersection to the right, cor- responding to a still higher pres- sure p u . The ejector effect of the exhaust puffs will eventually become less and less for higher steam velocities. This will be the Fig. 3. (Duse = Nozzle). case especially in single cylinder engines, because the length of exhaust pipe required is very great. For instance in an engine running at 180 r. p. m. which corresponds to 6 exhaust puffs per second, and for a steam velocity o- 540 m/sec, the length of the exhaust pipe must be 90 m, or better 100 m. Calf culating the loss of pressure required to overcome the resistances by means, of Eberle's equation: , t.o in which A p represents the loss of pressure in kg/sqm, I is the length of the ex- haust pipe = 100 m, d its diameter = 0,1 m, w the steam velocity = 540 m/sec, y the specific weight = 0.58 kg/cbm and $ a constant = 10.5 x 10~ 4 , then the loss of pressure will be found to be 11.1 at. For d = 0.05 m the loss would be 35.5 at. However, Eberle's tests from which the above formula was obtained, covered only velocities up to 150 m/sec, as did similar tests by Fritsche, Ombeck, Lorenz and Ritschel, so that the results are not directly applicable to velocities higher than the critical value. Such higher velocities will result in still greater losses. In reality the above results will be smaller since the assumed steam velocity will not be constant but decreasing. It would seem from the foregoing that the direct exhaust ejector principle would not have great prospects if based on complete conversion of pressure energy into velocity. However, as reported by Giildner 2 ), partial vacua have been observed in long exhaust pipes of gas engines, although no special provision had been made to cause and sustain them and a great part of the energy was lost in valves, sharp edges and elbows. There must therefore still remain a possibility of solving this problem in another way. It is, however, not easy of solution, since it involves the calculation of the friction of accelerating and expanding steam, which is very difficult to express in a mathematical form. Actual tests must therefore be relied upon. x ) Stodola, ,,Die Dampfturbinen". 3d. Edition, p. 55. 2 ) Giildner, ,,Entwerfen von Verbrennungskraftmaschinen. 3d. Edition, p. 40. Ill Although this problem may seem difficult in connection with a single cylinder, it is very simple for multi-cylinder engines, of which the locomotive is the chief representative. Even in an engine having two cranks at right angles, an exhaust lead of 25% will produce sufficient overlap of the exhaust periods so that the exhaust of one cylinder begins before the other has ceased (Fig. 5). If now the exhaust pipes are joined at an acute angle, a jet ejector action is obtained. This effect is well known and widely used in locomotive practice, where the combina- tion of blast pipe and stack also form an ejector, thus producing a partial vacuum in the smoke box which serves to draw off the flue gases. The theory of the blast pipe was first developed by Zeuner in his classical treatise on the subject 1 ). An equation in a somewhat simpler form based on this theory, is found in v. Ihering's book, ,,Die Geblase". The development of the formula for the ratio G 2 : G of the quantities of the ejected to the ejecting steam is very lengthy and will not be repeated here. The formula is based upon the principle of the continuity of flow and includes a number of assumptions and simplifications, the most important of which are that friction losses are not con- sidered, but the unfavorable assumption is made instead that the velocity w z is entirely lost and w 3 = 0. (See Fig. 4.) The following formula then results: m m m ^ n 2 2 I u. in which 72 The efficiency of upon the losses due as the impact losses the two streams mix. regarded as absolutely ejector friction depends as well an to at the point where The impact is to be inelastic. Assuming *ft Fig. 4. F sectional areas w velocities p pressures / specific weight These symbols used without index refer to the junction of the nozzles and the steam at that point; Index 2 refers to the variable port opening at the cylinder; Index 1, to the throat of the blast pipe and the mixture; Index 0, to the final section of the blast pipe; w 3 = the velocity of the ejected steam at the junction. W = then the efficiency r] s = i Since G 2 is small compared with G+G 2 the ejecting stream loses little of its energy, the impact loss is small and Tis = 0.75 to 0.80. The efficiency of the blast pipe is considerably lower, being only about 0.28, because the weight of the ejected air is about 2.6 times the weight of the ejecting steam. In addition to this there is a further loss equal to the energy contained in the steam at the final section of the stack. A similar loss will occur at the end of the blast pipe in regard to the exhaust ejector action, since the energy contained in the steam at the final section P\ cannot be utilized for steam ejecting. This finds its expression inequation 1, where A increases with decreasing *) Zeuner, ; ,Das Lokomotiven-Blasrohr". 1863. Zurich. Meyer-Zeller. 112 F . The area F should therefore be made as large as possible, but is limited by considerations in regard to the blast action on the flue gases. Strahl 1 ) has deve- loped a formula for the blast pipe area, which is based on Zeuner's treatise and is a- R as follows : F = -y=, where F is the blast pipe area, F^ the smallest stack area, y xA R the grate area, A the coefficient of divergence of the stack, x the coefficient of flue gas friction from ash pan to stack; a = f (m); m = A F : : F - a is nearly con- stant, and S 0.03 for m = 13 to 19. The weight ratio of the ejected air L to the ejecting steam D is L ~D = A large blast pipe area produces a lower pressure in the cylinder but insuf- ficient vacuum in the smoke box and therefore unsatisfactory steaming of the locomotive. It is self-evident that with a given amount of work available in the toe of the diagram only a certain total of ejector action for cylinder and boiler together can be produced, the distribution of which depends essentially upon the blast pipe #rea. In addition to the blast pipe area, the dimensions of the stack are also important. Too small a stack area leads to a large velocity loss at the outlet, and too large an area causes a considerable impact in ejecting the flue gases. It follows from this that there must be a best stack area for which, in consequence of the losses being a minimum, the blast pipe area is a maximum. Ex- pressed mathematically, a = f (m) is a maximum for m = 15.5 approxi- mately, as is demonstrated in the above-mentioned paper by Strahl. These best stack dimensions must be strictly adhered to in a locomotive in which the ejector effect of the exhaust is utilized, and this leads to considerable difficulties in large locomotives of the present day. The length of the stack is limited to such an extent by the loading gage and the high location of the boiler, that the expansion of the jet leaving the blast pipe may not completely fill out the whole area of the stack, so that air may enter from above through the remaining area and thus partly destroy the vacuum. Actual tests, however, have shown that such loss of vacuum can easily be avoided even with a large stack area. a ) Strahl, ,,Untersuchung und Berechnung der Blasrohre und Schornsteine an Loko- motiven, 1912'', Wiesbaden, C. W. Kreidel. Fig. 5. 113 The further calculations are based on a 010 freight locomotive of the German State Railways as an example, which is described on page 242. This engine is a two cylinder superheater locomotive with a cylinder bore of 630 mm, a stroke of 660 mm, a driving wheel diameter of 1400 mm and a steam pressure of 12 at. gage. Assumed is an evaporation of 7000 kg/hour, a loss of pressure of 1 at. from boiler to valve, adiabatic expansion and compression at an entropy of 1.7, and three different speeds of 20, 40 and 60 km/hr, which are referred to below as cases I, II and III. The exhaust port of the cylinder, having a'diameter of 120 mm, is placed 7 mm off center to compensate for the angularity of the connecting rod of 2600 mm length. The diagram shown in Fig. 5 is based on these figures and the cross-shaded areas represent the periods of overlapping exhaust. The ejector action is effective only during part of the exhaust period and therefore only the part A of the shaded area in Fig. 6 can be utilized for ejector action, the part B being lost. An increase in the exhaust lead would of course have the effect of increasing A by a part or the whole of B; but considerations of evenness and strength of the draft forbid this. Of the total work represented by the area A, a part Fig. 6. is lost in producing the draft in the smoke box by the rush of the steam at high velocity from the nozzle (blast nozzle loss) and a part is lost by throttling and friction in the pipes (pipe loss). Finally, after subtracting the stack loss, there remains the effective gain of work C at the compression line, corresponding to an absolute pressure p u . The investigation begins with the determination of the free exhaust port areas for various piston positions within the limits of the exhaust lead /. The secondary nozzle delivering steam to the feed water heater must also be included in this consideration, since the steam passing through it acts in the same way as in the main exhaust pipe. Fig. 7 illustrates the profiles of the main and secon- dary nozzles, as well as the piston positions at crank angle intervals of 5. In order to find the smallest areas of opening of the nozzles, the outlines of which are shown shaded in Fig. 7 for 55 before dead center, their plan projections are first laid out and their areas multiplied by V cos I n calculating the weight of steam flowing through the nozzles, the velocity loss must be taken into considera- tion by using a velocity coefficient y = 0.9. On account of the very unfavorable nozzle profiles when first uncovered, with consequent turbulence, it was consi- dered necessary to use a smaller value of

o 115 and exhaust lines I, II and III shown in Fig. 9 were obtained point by point in this manner, corresponding to speeds of 20, 40 and 60 km/hour. The part of the exhaust line where the ejector action comes into effect was also determined point by point, the weights of steam ejected G 2 being calculated by means of equa- tion (1). As long as the pressure in the cylinder is above the critical value p k = 1,73 at. abs., the pressure energy can only be completely converted into kinetic energy by the use of conical nozzles; otherwise the jet will still possess some pressure, and the suction effect will be diminished on account of the lower velocity conse- quent on the decrease in specific volume. In equation (1) this is taken care of by an increase in y. The theoretical as well as the actual factors of divergence i^ and if}' were therefore calculated according to the rules of steam turbine design, I I E J3mi 6,8* S.95 2.i Fig. 9. and are also tabulated in Fig. 7. According to these figures the actual divergence is insufficient in case I for crank angles of 50 to 30. A correction was there- fore made in the calculated values of G 2 . In case II the actual divergence is al- most, and in Case III exactly correct, so that G 2 need not be corrected. The compression lines in Fig. 9 were laid out in accordance with these results, and it is found that the initial compression pressures are respectively 0.7, 0.97 and 1.0 at. abs. for cases I, II and III. The very small pressure reduction in cases II and III makes it desirable to analyze the losses during the exhaust period. The blast pipe loss can easily be calculated since we know the weights of steam exhausted during the time the piston travels the distances f d and / f , as well as the blast pipe area and the specific volume corresponding to atmospheric pressure. With these data the mean steam velocity at the outlet of the blast pipe and the corresponding energy may be calculated. To be exact, instead of taking the mean velocity w, it would be neces- 8* 116 sary to find the value of the integral w*dd but for a rough approximation this o is not necessary. The amount of work represented by the areas A, B and C may be found with a planimeter, and the efficiency of the ejector is determined by the ratio of G 2 : G, so that the remainder represents the pipe loss. The major part of the latter is caused by the throttling of the steam when entering the nozzle. A high vacuum is formed in the exhaust pipe just before and after beginning of compression but it only partly reaches the cylinder. This is taken into consideration in equation (1) by taking a lower value of n. The following table gives the relative amounts of the different losses. Case I II III Area A: Blast nozzle loss 40 25 14 Pipe Loss 14 30 56 Impact loss during ejection . Gain of compression Area B: Blast nozzle loss 8 27 1 4 15 9 22 Pipe loss 10 17 8 Total . . 100 ' 100 100 It will be seen from this table that the blast nozzle loss (part A and B taken together) is approximately constant and amounts to from 41 to 36%. The pipe loss increases from 24% at low speeds to 44% at high speeds, and the work re- presented by area B from 11% to 30% respectively. It therefore follows that at high speeds the bad effect of area B must be eliminated, and this can be easily accomplished with a three cylinder locomotive. In such an engine, having cranks at 120 and an exhaust lead of 25%, there will always be two cylinders exhausting at the same time. The ejector effect begins at the dead center and proceeds with greater nozzle areas in the cylinder wall, with the result that the throttling loss and pipe loss are also reduced. The loss of work may therefore be divided in all three cases as follows: Blast nozzle loss 40% Pipe loss . 24% Impact loss ....*.... 8% Useful work. . . ... . . 28% Assuming this division of losses, and taking the mean effective pressures from the indicator cards, the following figures were obtained for two and three cylinder locomotives. The three cylinder locomotive also shows only a surprisingly slight gain due to the ejector effect at high speeds or early cut-offs. This is explained by the fact that the utilization of the toe of the diagram is equivalent to an enlargement of the cylinder. If the cut-off is early, then a further increase in expansion will not produce much gain. The steam consumption figures are already so low that 117 I II III Two Cylinder Locomotive. Mean effective pressure, with ejector effect, kg/sqcm Mean effective pressure without ejector effect kg/5qcm 6.84 6.1 3.93 3.83 2.83 2.83 Gain due to ejector effect /o 12.2 2.6 Indicated HP 940 1080 1170 Steam consumption (7000 kg/hr) divided by IHP. Three Cylinder Locomotive. Loss area A -j- B kg/sqcm 7.45 2.7 6.47 0.67 6.2 0.33 Compression gain 0,28 (A-\- B) .... kg/sqcm Mean effective pressure, with ejector effect kg/sqcm Gain due to ejector effect /o 0.75 6.85 12.3 0.19 4.02 5.0 0093 2.923 3.0 Indicated HP 940 1110 1200 Steam consumption (7000 kg/hr) divided by IHP. 7.45 6.3 5.8 not much more could be desired. The saving of 12% for late cut-offs is noteworthy, because it is based on very conservative assumptions. Furthermore, the exhaust ejector effect allows of a considerable reduction of clearance volume; for instance, in the case under consideration, from 17 to 11%. Taking into account the saving due to the una-flow principle a total saving of 15% may be expected with certainty. This saving is all the more important since it occurs at heavy loads and therefore increases the hauling power of the locomotive by this amount. Although in describing the exhaust ejector principle locomotives were con- sidered exclusively, its field of usefulness is not limited to the latter. It may also be found advantageous in street railway locomotives, road rollers and stationary engines, as well as locomobiles which are still frequently built with two cylinders so that they may be started from any crank position, which is desirable for instance in peat pressing plants. 118 8. Prof. Dr. Nagel's Experiments. A series of extremely interesting experiments on the temperature conditions in a una-flow cylinder were conducted by Prof. Nagel in the engineering labora- tory of the Technische Hochschule, Dresden, and described by him in the Zeit- schrift des Vereines deutscher Ingenieure Vol. 1913, No. 27, July 5. Part of his report is as follows: "After Prof. Stumpf had published his first communications on the character and success of the una-flow steam engine a number of years ago, 119 Dr. Mollier and the author approached the Verein deutscher Ingenieure with the request for an appropriation for investigating the temperature conditions in a una-flow cylinder. This request was granted in the most whole-hearted manner. At the same time the Saxon Government provided considerable sums for the completion of the testing plant. The una-flow cylinder used for this purpose in the engineering laboratory of the Technische Hochschule, Dresden, was built by the Ntirnberg Works of the Maschinenfabrik Augsburg- Ntirnberg, and took the place of the low pressure cylinder of the existing triple- expansion engine. It was put into operation during September 1911. The cylinder as shown in Fig. 1, has a bore of 450 mm and a stroke of 650 mm, and the engine runs at 150 r. p. m. In order to determine the thermal peculiarities of the Stumpf cycle it was planned to measure the temperature changes of the working steam at different points in the cylinder. It was later also found desirable to measure ffo/ben OecM Fig. 2. (Kolben = piston; Deckel = cover). the temperature variation of the cylinder wall. The determination of the steam temperatures offered great difficulties. It was at first attempted to use thermo- couples of copper and constantan wire of 0.2 mm diameter. Tests of a similar nature on a counterflow engine in the laboratory, using the same elements, had been started five years ago, but a critical examination showed that their sensitiveness is by no means sufficient to follow the changes of temperature with the necessary speed. After long and futile experiments with thermo- couples composed of thinner wires down to 0.07 mm diameter, it was found, according to a test report published in an American periodical, that wires of so small a diameter as 0.01 mm were required for the thermocouple to be suffi- ciently sensitive. It seemed impossible to produce thermocouples with wires of this thinness on account of the difficulty in making the junction, and for this reason the use of electric resistance thermometers was decided upon early in 1912. The material for the latter was obtained in the form of drawn tungsten filaments as used in electric lamps. A wire of about 50 mm length was wound in zigzags upon a glass frame provided with platinum hooks for this purpose, as shown in Fig. 2. The main difficulty was a satisfactory connection of the resistance unit to the lead wires in order to enable the termometers so. con- 120 structed to resist the effect of the steam currents inside the cylinder. The measurement of wall temperatures was rendered difficult by the fact that the insertion of the measuring unit into the cylinder wall necessitates the drilling of a hole which more or less disturbs the heat flow. This may be the cause of an erroneous temperature indication. In order to reduce this possibility to the utmost, the arrangement illustrated in Fig. 3 to 5 was employed. A hole of 15 mm diameter was drilled in the cylinder wall, into which was closely fitted a cast iron plug having a hole of 9 mm diameter bored to within 0.5 mm of the bottom. Into this hole was fitted a second cast iron plug having two drilled holes of 2 mm diameter from end to end. These holes contained the copper and constantan wires of 0.1 mm diameter insulated by small glass tubes, their Fig. 5. ends being embedded in grooves at the bottom surface of the cast iron plug. A thermocouple of similar construction was also fitted to the piston, as indi- cated at ft, in Fig. 1, and in Figs. 6 and 7. The leads of this thermocouple were carried through the hollow tail rod, provided with a porcelain lining for this purpose. For measuring the change in voltage corresponding to the changes in temperature, a galvanometer made by Edelmann in Munich was employed. According to Fig. 8, it consists of a powerful electromagnet which is supplied with current from a storage battery. In the magnetic field is stretched a fila- ment of gold or platinum having a diameter of from 0.002 to 0.005 mm, which carries the current to be measured. The displacement of this wire due to electro- 121 magnetic forces is a measure of the current flowing through the circuit, and therefore also a measure of the temperature. This displacement, although amounting to only a fraction of a millimeter, is projected on an enlarged scale onto the focal plane of a camera by means of a beam of light from a source L and a system of microscope lenses. The photographic plate is moved pro- portionately to the piston travel or crank angle behind a slit in the focal plane, Fig. 6 u. 7. thus producing a photographic record of the changes of temperature with stroke or time. Special methods were devised for rapid calibration of the dis- placement of the filament. In Figs. 9 and 10 are reproduced two records of steam and wall temperature based on time. The remarkable 'feature about Bilet- eberie Fig. 8. (Bildebene = Focal plane.) the temperature change of the working steam is the fact that at the end of compression the latter attains temperatures of such a magnitude as were hitherto thought impossible. A terminal compression temperature of about 500 was observed when running with saturated steam of 10 at. gage. The wall temperature was measured at the points a, 6, c, d, e, /, g and /c, indicated in Fig. 1. The change of temperature at the end of the cylinder barrel at point b is of considerable significance. A series of tests made with constant cut-off 122 Temperature time diagram of steam close to cylinder head surface. (Point of measurement a.) tintauchttqfe Q 1 Umdrehung Fig. 9. Temperature time diagram of cylinder wall at a depth of 0,5 mm from inside surface. (Point of measurement /.) Fig. 10. of 10% and saturated as well as superheated steam of different temperatures showed that the highest mean temperature at this point was reached when operating with saturated steam; even superheated steam of 350 did not produce the same high wall temperature. The sensitiveness of the thermocouples was raised to such a degree that the passing of every piston ring over a point of measurement produced a clearly discernible wave. The moment of passage of the several rings over the thermocouple is clearly indicated in Fig. 10 by the shading between the various lines; the diagram was taken at point d" The above report also includes descriptions of the several instruments which were used during the tests and for the analysis of their results. Among others, there are mentioned a harmonic analyser by Mader, an instrument fitted with a microscope for measuring indicator cards, made by H. Maihak, of Hamburg, and an apparatus furnished by Steinmuller, of Gummersbach, for automatically measuring the condensate, which proved to be very exact. The temperature diagram in the above report by Prof. Nagel merits parti- cular attention. Instead of a terminal compression temperature of 500 for 3.3% clearance, a final temperature of 900 should be obtainable with a clearance volume of 1%. While on the one hand the terminal compression temperature was 500 for saturated steam, it decreased to 480 or 450 for increasing degrees of super- heat. This may probably be attributed to the more energetic heating action of the steam jacket in the case of saturated steam. Prof. Nagel further states that the temperature at the point b of the cylinder wall, for 12% cut-off and saturated steam of 184 in the jacket, was 128, which fell to 111 for the same cut-off and superheated steam of 220; and again slowly rose to 118 for a further increase of the initial steam temperature to 350. The temperature of the inner cylinder head surface was 177 for an initial or jacket steam temperature of 184, the total variation during one revolution being only 0,5. The temperatures at the points *, c, d, (Fig. 1) were found to be 128, 102 and 83, with a total fluctuation of 3, 3 and 2.8 respectively. At the point k on the piston, distant 36 mm from the cylinder wall, the temperature was found to be 164.5 with a total fluctuation of 1.3. Attention is especially called to the latter figures, since they prove the statements previously made concerning the favorable thermal action of the piston head surface. At the points of measurement the heat had to penetrate a metal thickness of 0.65 mm. It will also be noticed in Fig. 9 that a very pronounced kink occurs in the temperature curve where the steam changes from the saturated to the superheated state, and also that an abrupt drop in temperature takes place at the moment of admission, from the high terminal compression temperature of about 530 to that of the live steam. The contrary effect of heating by the jacket steam and cooling by the cylinder steam at the point a is also evident in Fig. 9, as well as the corresponding small temperature fluctuation at this point during the complete cycle, considered apart from the sudden rise due to the heat of compression. The comparatively high mean temperature and small fluctuation at the point d are also noticeable in Fig. 10. 124 In Fig. 11 is shown an especially clear temperature diagram in which the kinks in the compression and expansion lines corresponding to the change from saturated to superheated steam are clearly noticeable. There is a surprisingly high temperature during the last part of expansion, the first part of compression and especially during exhaust (about 100), although the engine was operated with a vacuum of 98%. As this card was taken at the cylinder side of the cover, the explanation is easily found in the great flow of heat from the cover to the working steam during that time. 3. C. Cooer sieff . I0f J).C. Cr Fig. 11. The drop of temperature through the cylinder head wall was found to be 7 to 8 for saturated steam, 15 for steam of 250, and 25 for steam of 350 initial or jacket temperature. A close study of the temperature diagrams given in Figs. 9 and 10 and 11 has as its final result a confirmation of the thermal advantages of the una-flow principle and the jacketing of the heads. This is still more emphasized by the comparison of the una-flow temperature diagram (Fig. 11) taken by Prof. Nagel with a counter- flow temperature diagram taken by E. T. Adams & T. Hall from a common slide valve engine of the Sibley College-Cornell University, as shown in Fig. 12. The comparison elucidates the striking thermal difference between both engines. Whereas the una-flow engine shows the highest temperature at the inlet end and the lowest at the exhaust end, the counter-flow engine shows quite a thermal mixture distributed over both strokes. Interesting is the postponement of the phases of high metal temperature caused by the preceding phases of high steam temperature in the counter-flow engine. - 125 126 II. 1. The Una-Flow Stationary Engine. The una-flow engine has found a very wide use as a stationary prime mover mainly by reason of its simplicity, its straight line construction, its high economy and its adaptability to changing load requirements. The tandem counterflow engine which still comes occasionally into competition with it, is at a disadvantage on account of its two cylinders, its two pistons, its piping and the inaccessibility of its exhaust valves. The conditions of close regulation required of stationary engines, in which may be included engines for electric current generation, are satisfied in the una- flow engine in the best possible manner since the action of the governor is direct and is not impeded by steam already contained in the engine, as is the case in multiple expansion engines where the effect of such steam on the regulation makes itself unpleasantly noticeable. The range of cut-off in una-flow' engines is usually from to 25%, although cut-offs are found up to 40%, or even 50 or 60% for instance in rolling mill engines. The range of the governor must include zero cut-off, in which case the inlet valve does not open at all. The lead of the steam valves at all cut-offs must be kept down to the minimum or reduced to nothing if possible, since large lead causes condensing engines to knock badly, especially if the clearance is large and the vacuum high. Non-condensing engines with large clearance and long compression always run quietly and can therefore stand more lead. It is easily possible to start engines having 25% maximum cut-off even under load and with a directly driven air pump, more particularly if an additional clearance space is provided and opened up during the first strokes until sufficient vacuum is generated. The best location for these additional clearance pockets is in the cylinder head opposite the end of the cylinder, so that for condensing service they will act as a very effective insulation between cylinder head and frame, while pro- viding double the amount of cover jacket surface for non-condensing operation. Every condensing stationary una-flow engine should be equipped with addi- tional clearance spaces in order to facilitate starting if the air pump is directly driven, and to allow of running the engine without the condenser. The clearance volume averages about 1,5 to 2% for condensing operation and high vacuum, and 13 to 28.% if the additional clearance spaces are opened up for non-condensing service (see page 48). In non-condensing engines the necessary clearance may be arranged in the cupped ends of the piston. On account of the large work of compression such engines require comparatively heavy flywheels. Since the una-flow engine has only two inlet valves, the use of a lay-shaft is unnecessary. As is shown in Figs. 1 to 3 of this chapter, and in Figs. 4 and 5 (page 10), the inlet valves may be driven from an eccentric on the crankshaft 127 W/////////////////////////////////'/'///"' Fig. 1. Fig. 2. Fig. 3. 128 acted on by a shaft governor, by means of a rocker arm and cam mechanism (Stumpf gear). A lay-shaft with its bevel gears and bearings is thus dispensed with. Con- sideration must be given to the expansion of the cylinder. If the latter is pro- vided with steam jackets receiving their supply from a connection to the steam pipe ahead of the main stop valve, then the cylinder may be warmed up prior to starting, and the valve gear, if set correctly for the hot engine, will give proper distribution from the very start. If the cylinder is unjacketed, then the steam distribution will be incorrect for a while after starting until the cylinder has reached its expanded condition. It is always advisable to design the valve gear with out- side admission, or in other words to arrange cams and rollers so as to make their action conform to the steam lap of a slide valve, so that while the cylinder is still insufficiently heated, the head end valve will open late instead of too early. This negative lead combined with a simultaneous earlier cut-off will do less harm than an early opening of the valve, which may cause the engine to knock. Fig. 4. There remains another possibility of correcting the bad influence of the expan- sion of the cylinder even if the latter is not provided with steam jackets, by ex- tending the cylinder lagging around the rod between the valve bonnets (Fig. 4), thus heating it to approximately the mean cylinder temperature. It is also possible to drive the head end valve through an equal-armed rocker mounted at the center line of the exhaust belt. This insures permanent correct motion for the head end valve. If a lay-shaft is used, the influence of the cylinder expansion is eliminated, and the steam distribution must always be correct. Fig. 5 illustrates details of a valve bonnet as used with the Stumpf gear. The cam is connected to the valve crosshead, and the reciprocating slide is grooved to accommodate the roller and at the same time form an oil bath. The guide for the reciprocating slide is long enough so that the groove never runs beyond it, the loss of oil by splashing and the entrance of dust thus being prevented. The oil collecting in the groove is transferred by the roller to the cam, so that perfect lubrication and reliable operation of these important parts is insured. 129 Fig. 6 shows a twin una-flow engine with Stumpf gear, in which a jack shaft having two crank throws is driven by a pair of eccentrics on the crank shaft set at 90. This jack shaft carries a shaft governor acting upon an eccentric on each side of it, which operates the valve mechanism of its corresponding cylinder through a rocker arm. In this way the valve gears are positively connected, so that both of them always give the same cut-off. The short vertical eccentric rod also helps to equalize the cut-offs of both cylinder ends, and the small diameter of the jack shaft facilitates the design of the governor. This engine possesses great reserve power since each half is able to carry the whole load. The elimination of exhaust valves and their gear will be found very convenient in horizontal engines, since it leaves the whole space underneath the cylinder free for piping and permits of a close arrangement of the condenser. (See Figs. 2 to 5, chapter I, 3b, p. 6970.) The Erste B runner Maschinenfabrik- gesellschaft was the first concern to take up the una-flow engine, and decided to re- build an old 80 HP. single cylinder con- densing engine with a forked frame by fitting it with a una-flow cylinder, designed by the author, having a bore of 400 mm and a stroke of 420 mm (Fig. 7). On the free end of the crank shaft was mounted a shaft governor acting on the eccentric ope- rating the inlet valves by means of a rocker arm on the exhaust belt and a pair of Lentz cam mechanisms. Although this first design was susceptible of improvement in many respects, its economy even with rather low vacuum was 'equal to that of a compound engine of the same size. Fig. 8 shows another engine built by the same Company. Shortly after the latter had taken up this work, the Elsassische Maschinenfabrik decided on a large scale experiment. Their first una-flow engine, built to the author's design, had a cylinder bore of 640 mm and a stroke of 1000 mm, with a rated load of 500 HP. (Figs. 1 and 9). This engine, which was directly con- nected to an electric generator, was tested by the Elsassische Verein der Dampf- kessel-Besitzer (Alsatian Association of Steam Boiler Owners), on February 21, 1909. The result of a trial of four hours and eight minutes duration showed a steam consumption of 4.6 kg/I HP-hour for an initial steam pressure of 12.6 at. gage and a temperature of 331 G, at a speed of 121 r. p. m. This is a very creditable Stumpf, The una-flow steam engine. 9 Fig. 5. 130 131 result if it is borne in mind that the engine did not derive the full benefit from the vacuum on account of too small an exhaust pipe and the use of an oil separator between cylinder and condenser. (Back pressure 0,145 at. abs.) By correct design Fig. 7. of the condensing equipment in the way previously suggested, by jacketing of the cylinder, and the use of tighter valves, the steam consumption could be conside- rably diminished, as proved by later engines built by the same makers. Fig. 8. Figs. 2 and 3 show a una-flow engine of 900 HP rated load built by the same firm. Noteworthy is the heavy frame, the engine being of center crank construc- 9* 132 tion which is now used by several concerns. The center crank type is advantageous where the forces on the moving parts are heavy, especially for large short stroke engines. In Figs. 10 and 11 is shown still another engine by the same makers, as well as its governor and valve gear parts. Fig. 9. A una-flow engine built by Burmeister & Wain, of Copenhagen, and designed by the author, is shown in Figs. 12 and 13. The direct connection of the con- denser to the cylinder should be noted, as well as the method of supporting the Fig. 10. rear end of the latter on two adjustable rods, and the simple air pump drive. The cylinder is left unjacketed on account of the use of superheated steam (Fig. 14). The head jackets, however, are carried up to the point of normal cut-off. The 133 Fig. 11. Fie. 12. 134 piston is turned to a smaller diameter for a corresponding distance to provide for expansion. The additional clearance pockets are fitted with two clearance valves, one one each side of the inlet, one of which is sufficient for starting, while the Fig. 14. second one has to be opened for non-condensing operation at full speed (Fig. 15). The area of contact between cylinder head and frame is kept as small as possible in order to reduce the conduction of heat to a minimum. The cylinder head casting is perfectly symmetrical so as to increase its range of usefulness. 135 Fig. 15. Fig. 16. 136 Fig. 17. Fig. 18. 137 Tests made with this engine by Mr. Bacher, Professor at the Technical Hochschule in Copenhagen, showed the following results. Load Pressure kg/sqcm Steam Tempe- rature C Vacuum in /. of 760 mm r. p.m. Steam Con- sumption kg/hr. Steam Consumption KW BHP IHP KW/nr. BHP/hr. IHP/hr. 64.50 98.7 116.0 9.90 352 94.0 179.0 479.0 7.44 4.86 4.12 86.46 130.0 149.0 9.87 354 93.8 175.0 632.0 7.32 4.86 4.24 108.66 163.0 184.5 9.84 353 93.5 176.5 798.4 7.36 4.90 4.34 131.24 197.0 222.0 9.80 353 92.6 173.5 976.7 7.45 4.97 4.40 109.00 164.0 186.0 9.75 dry saturated 93.0 178.0 1150.6 10.55 7.03 6.20 In view of the omission of the cylinder jackets, these results are in close agreement with the tests of a 300 HP una-flow engine given on p. 11. A single-acting una-flow engine built by Burmeister & Wain is shown in Fig. 16. Engines of this type are widely used in Danish dairies. The horizontal valve is operated by a cam mechanism directly connected to a shifting eccentric on the crank shaft. (Figs. 17 and 18.) A stationary engine built by Ehrhardt & Sehmer, of Saarbriicken, for the power plant of the Saar Valley Railway, is shown in Fig. 19. This engine has Fi?. 19. a cylinder bore of 650 mm, a stroke of 1000 mm, and a rated load of 500 HP at a speed of 130 r. p. m. It has run for long periods at an overload of nearly 100%. The average cut-off of una-flow engines at rated load being only about 10%, their 138 139 capacity for overload is far greater than that of any other type of engine. If the dimensions of the driving parts are based upon the initial pressure less the inertia, then even a heavy overload does not materially increase the load upon them. In Fig. 20 is shown the largest una-flow en- gine so far built, con- structed by Ehrhardt & Sehmer for driving a rol- ling mill at the steel works of Gebrtider Roch- ling, atVolklingen, having acylinderboreoflVOOmm and a stroke of 1400 mm, the speed being 110 to 130 r. p. m. The cylinder was made with this large bore on account of the low steam pressure avail- able at the time, and is provided with two inlet valves at each end for this reason. Later on, when new high pressure boilers have been instal led, it is intended to substitute for the present cylinder a smaller one with only one valve at each end. Another una - flow engine built by the same firm and delivered to the Aplerbeck steel works, is illustrated in Figs. 21 and 22. The cylinder bore is 1450 mm, the stroke 1500 mm, the speed 100 r. p. m., and the steam pressure 8 at. gage. 140 141 142 143 This engine has nearly the same dimensions of driving parts as the one just de- scribed. The diameter of the piston rod is 250 mm, that of the tail rod 225 mm, the crosshead pin is 400 mm diameter by 600 mm long, the crank pin 550 mm diameter by 600 mm long, and the main bearing 730 mm diameter by 1200 mm long. The use of a side crank in such a large engine of short stroke is noteworthy. In order to reduce the overhang, the crank and crank pin are of cast steel in one piece, with a hub length of only 450 mm (hub length : shaft = 0.62). The side crank construction, together with its corresponding type of frame, makes the engine simple and inexpensive. The same cannot perhaps be said of the Zvonicek valve gear employed on this engine, but it has the advantage of giving the late cut- offs essential for rolling mill engines, and of permitting the use of a standard governor which is in many cases preferred to a shaft governor on account of its simplicity and accessibility. The Zvonicek gear consists of a fixed eccentric, the strap of which Fig. 25. is provided with a cam profile and held at its armlike extension under the control of the governor. The combined motion of eccentric and cam is transmitted to the valve bonnet cam mechanism by a reach rod provided with a roller at its lower end. Figs. 23 and 24 illustrate clearly the trend of development due to the una- flow engine and show the replacement of the two cylinders of an old tandem counterflow engine by a single una-flow cylinder. A number of such reconstruc- tions have been carried out by Ehrhardt & Sehmer and other firms. An engine built by Musgrave & Sons, Ltd. of Globe Iron Works, Bolton, England, is shown in Fig. 25. This firm is credited with the introduction of the una-flow engine on a large scale in Great Britain and colonies, the Stumpf valve gear being employed exclusively. A test carried out by Mr. F. Thomas on one of their engines, having a cylinder bore of 685.8 mm and a stroke of 914.4 mm gave the following results : Steam pressure 10.67 at. gage at the throttle, superheat 10, speed 129 r. p. m., vacuum at the cylinder 66 cm, load 317 I HP, and steam consumption 4,98 kg/I HP-hour. The cylinder barrel was unjacketed. 144 Stork & Co., of Hengelo, Holland, also employ only the Stumpf gear on their engines, one of which is shown in Fig. 26. This firm has been very successful in in- troducing the una-flow engine in Holland and the Dutch colonies. Stork & Go. report that a test of one of their engines (650 mm bore by 900 mm stroke, speed 125 r. p. m.) showed a steam consumption of 4.86 kg/I HP-hour, the steam pressure being 8.24 at. gage, and the temperature 248 G. The cylinder barrel was un jacketed. The type of una-flow engine built by the Maschinenfabrik Augsburg- Niirnberg is shown in Figs. 27 and 28. The inlet valves are placed horizontally and are ope- rated by a rocking shaft and cam mechanism. This cam receives its motion from a short jack shaft which in turn is driven by the governor eccentric on the lay shaft. The clearance valve is situated opposite the steam valve and is arranged to act automatically in case of sudden failure of the vacuum. Partially or entirely unbalanced inlet and spring-loaded clearance valves may be made to serve the same purpose. One engine of this type (903 mm cylinder bore, 1000 mm stroke) was fur- nished to J. P. Stieber, at Roth near Niirnberg and was tested by the Bayerische Revisionsverein on February 23, 1912. Duration of test hours 4.17 4.05 8.06 Boiler pressure at. gage 13.3 13.3 13.2 Steam temp, at engine 255 291 310 Steam pressure in cylinder at. gage 9.3 10.4 11.4 Vacuum in cylinder % 90 90 90 Vacuum in condenser % 93 92 92 Actual cut-off . ~ % 3 6 12 Speed . r. p. m. 125.5 123.3 126.3 Indicated horse power HP 473 793 1109 Steam consumption in kg/I HP-hour including con- densate from steam pipe 4.69 4.74 4.71 Heat consumption in Gal/I HP-hour based on total heat of steam entering the engine 3320 3440 3460 Thermal efficiency % 19 18.4 18.3 This engine was designed for a steam temperature of 330 G for which reason the cylinder barrel was left unjacketed. With jackets the steam consumption would in this case have been considerably lower on account of the beneficial effect of cylinder jackets for small cut-offs and low initial temperatures. On the other hand, the results once more demonstrate the small variation in steam consumption for large ranges of load (473 to 1109 HP) when no cylinder jackets are employed. An engine built by the Gorlitzer Maschinenbauanstalt is shown in Fig. 29. The cylinder has a bore of 1100 mm, a stroke of 1300 mm, and the engine runs at 91.6 r. p. m. The valve gear comprises a lay shaft with governor and shifting eccentric acting on short rocking shafts alongside the cylinder. The ends of the cylinder are jacketed. The air pump is driven from an extension of the tail rod. Fig. 30 shows one of the latest engines built by this Company. There is only one governor eccentric, and the valves are operated through Lentz cam mechanisms from a rocking shaft on the cylinder, having two levers set at 180. The air pump 145 bb s Stumpf, The una-flovv steam engine. 10 146 tc 147 10* 149 150 is again driven from the tail rod. In the single-eccentric type of valve gear the lay-shaft as well as the hole in the latter for the synchronizing device are shorter, and the governor may be placed close to the rear bearing. The basic idea of this gear is similar to the one used by the Maschinenfabrik Augsburg- Nurnberg. The single-eccentric gear is also described in the Z. d. V. d. I., 1914, No. 19, page 729. The valves are placed on the cylinder and consequently have somewhat larger clearance volume and surfaces. The jacketing is excellent; the arrangement of the condenser, however, is not free from objections. The firm reports the following steam consumption results. Mean Engine Rated Load Stroke Bore Steam Pressure at. gage Steam Temp. r. p.'m. Steam Consumption kg/IHP-hr. Measurements based on" HP mm mm C 1 550 1000 750 11.8 250 127.2 4.8 1 2 810 1300 850 9.5 269.7 91.6 - 4.99 Boiler 3 300 800 650 11.7 253 15.67 488 [ feed water 4 523 800 600 11.5 231 152 4.88 J 5 140 | 600 1 375 9.7 340 200 4.38 Condensate The steam temperatures in engines 1 and 5 were measured at the entrance to the cylinder head, and in engines 2, 3 and 4 in the middle of the same. Fig. 31. The greatest credit for the commercial introduction of the una-flow engine is due to Sulzer Bros., of Winterthur and Ludwigshafen. The cylinder of their first engine, which is^ in operation in the brass rolling mill of Wieland Bros., at Ulm, was designed by the author (Fig. 4, chap I, Ib, p. 10). The design of later engines was based on this first one, the only change being the substitution for the Stumpf gear of a lay-shaft gear, having two eccentrics which operate the valves by means of cams and roller levers pivoted in the valve bonnets (Fig. 31). The reciprocating roller slide of the Stumpf gear has therefore been replaced by the pivoted roller lever. The governor is placed close to the rear lay-shaft bearing in order to reduce the deflection of the shaft and also to shorten the bore required for the synchronizing de- vice. All of the driving parts, including those of the air pump in the basement, are comple- tely enclosed. A gear pump on the lay-shaft supplies oil under a pressure of about 1 at. to the bearings of engine and air pump. The oil collects in a reservoir in the base- ment, where it is fil- tered and again enters the circulating system. Low oil consumption and smooth running due to the oil cushion in the bearings are ad- vantages of this sy- stem. The consump- tion of cylinder oil is also low since only cne cylinder and generally only one piston rod packing have to be lubricated, as against two cylinders and se- veral packings in an ordinary multi - stage engine. The low oil consumption is also proved by the high mechanical efficiency. An engine furnished by Sulzer Bros, to the firm of Junker & Ruh, of Karlsruhe, having a cylinder bore of 675 mm, a stroke of 800 mm, and a speed 152 of/150 r. p. m., showed that for a steam pressure of 11 to 12 at. gage and a tem- perature of 250, 12 to 18 kg of cylinder oil were consumed weekly, and 8 to 10 kg of bearing oil were added to the circulation when running 10 hours daily and six CO tB co be days per week. The whole of the oil in circulation, amounting to about 1 barrel, is replaced every 6 to 9 months. A hand pump is provided to supply the bearings with oil before starting. As shown in the sectional drawing Fig. 32, the cylinder design incorporates all the essentials previously mentioned. It is, however, to be 153 60 154 CO bo 155 regretted that in many cases cylinder jackets are omitted when their use should be 'dictated by low steam temperatures. . Most of the engines built by Sulzer Bros, are fitted with a valve design the purpose of which could be otherwise accomplished simpler and better. Extensive experiments have enabled Sulzer Bros, to find the proper mixtures for cylinder and piston castings, whereby reliable operation of these parts is in- sured without the use of a tail rod. The cylinders are bored barrel-shaped so that the cylinder surface becomes almost exactly cylindrical under operating conditions. To these precautions, in combination with a thoroughly reliable lubricating system, must be ascribed the fact that Sulzer Bros, have never had piston troubles. All of their una-flow engines have therefore been built with self-supporting pistons, except the engine shown in Figs. 33 and 34, supplied to the Crefeld Cotton Spinning Mill, and a series of engines supplied to the Badische Anilin- and Soda-Fabrik, where a tail rod was used to meet the purchaser's wishes. A Sulzer stationary engine of standard design is shown in Fig. 35 (350 BHP at 150 r. p. m.), while Fig. 36 shows two Sulzer una-flow engines of 450 BHP each, supplied to the Hafod Copper Works, Swansea, South Wales. The Maschinenfabrik Esslingen employs a particularly effective method of boring una-flow cylinders under temperature conditions closely approaching those of actual operation. The cylinder ends are heated to a high temperature by ad- mitting live steam to the jackets, and the middle is cooled approximately to con- denser temperature by a blast of air through the exhaust belt. The cylinder is then bored cylindrically, and the piston is turned smaller than the cylinder bore with a correct allowance, a difference of four thousandths of the diameter being usually sufficient. Since the piston expands more than the cylinder, and is of great length, ample bearing surface will be obtained. The piston heads should be turned somewhat smaller in order to allow for their greater expansion. A piston as shown in Fig. 5, chap. I, 4, p. 77, fitted with bronze shoes, offers still greater safety against seizing, and this is true to a still greater degree of the floating piston having clearance all around. Piston troubles are caused in many cases by a wrong method of supplying oil to the cylinder. The force pump should be timed in such a manner that deli- very takes place only while the feed orifice in the cylinder is covered by the piston. A good distribution of oil to the piston and cylinder wall will then be obtained. The oil feeds should preferably be placed in the center of or close to the exhaust belt, where the cylinder has the lowest temperature. One feed should be arranged on the vertical center line and one each at either side in or below the horizontal plane, each feed being supplied by a separate plunger. Complaints which are sometimes made regarding the high oil consumption of una-flow engines frequently arise from defective methods of introducing the oil. It is fundamentally wrong to supply two or more feeds from the same pump plunger. A neat arrangement of piping is obtainable if the steam pipes are placed on one side of the engine and the exhaust pipe, air pump, and the cooling water and discharge pipes on the other. The following is a report of economy tests on the una-flow engine of the Cre- feld Cotton Spinning Mill, built by Sulzer Bros. 156 Steam was generated by four Lancashire (twin furnace) boilers, having a total heating surface of 400 sqm. A fifth boiler, the steam and feed lines of which were blanked off from the others, supplied steam for heating purposes. The feed water Fig. 38. was weighed, transferred to a large tank and fed to the boilers by means of a cen- trifugal pump. Indicator cards were taken every 10 minutes, and the steam pres- sure, superheat and vacuum were recorded at the same intervals. The guarantees given for this engine were: 157 Maximum load 2340 I HP: 10% cut-off, 300 temperature 1590 I HP 4.45 kg/IHP-hr 10% 350% 1530 ,, 4.15 13% 300 1920 4.65 13% 350 1860 4.35 These guarantees applied to steam of 11.5 at. gage pressure and condenser cooling water of 15 C. The duration of the test was from 8 : 32 A. M. to 4 : 06 P. M. a total of 454 minutes. The load varied from 1477 to 1752 I HP with an average of 1632.8 I HP at a speed of 109.5 r. p. m. Steam pressure at the cylin- der was 11.1 to 12.1, average 11.6 at. gage; steam tempe- rature at cylinder was 260 to 300, average 282.6 G; vacuum was 71.3 to 73, 1cm, average 72.2 cm. Barometer reading 76.4 cm. The temperature of the cooling water was 11.5 and that of the air pump discharge 30.1 G. Since the load was higher than the guaranteed figure of 1590 HP, and the steam temperature less than 300, a corresponding correction of the test results was necessary. According to the guarantees, the steam consumption increases from 4.45 kg to 4.65 kg or 0.2 kg for an increase of load from 1590 to 1920 HP, or 330 HP; therefore for an increase in load of 1632.8 1590 =42.8 HP, the permissible 2 42 8 increase may be - - = 0.026 kg and the steam consumption may be 4.45 ooU -f- 0.026 4.476 kg, and still be within the guarantee. Tests have shown that a reduction of the initial temperature from 300 to 282.6 produces an increase in steam consumption of 3%. The permissible steam consumption may therefore be 4.476 x 1.03 = 4.61 kg, and yet be within the guarantee. The total steam consumption was 56410 kg, or 56410-60 or 454 7455.06 1632.8 7455.06 kg/hour == 4.56 kg/I HP-hour. 158 159 The guaranteed figure was therefore satisfied without taking advantage of the permitted allowance of 5%. The above report was made by the Association for the Inspection of Steam Boilers, of Miinchen-Gladbach, Crefeld Branch Office, on October 4, 1913, and signed by Mr. Rhenius. In examining thik result it must be borne in mind that the cylinder barrel of this engine was not jacketed. A una-flow engine designed by the author for the Soumy Machine Works is shown in Figs. 37 to 39. It has valve gear of the Stumpf type and is designed to be used with saturated steam of 7 at. gage. The cylinder has a bore and stroke of 450 and 600 mm respectively, and the speed is 150 r. p. m. The resilient inlet Fig. 41. valves and the clearance valves are designed and arranged in such a manner that the total clearance volume amounts to only 1.24% for a linear piston clearance of 3 mm. The nut is flush with the piston so as to avoid the clearance volume of about 0.5% resulting from a projecting nut. The two-piece cast steel self-sup- porting piston is fitted with a bronze shoe fastened to it with copper rivets; the rest of the piston has several millimeters clearance all over. Each half carries three somewhat narrow rings, none of which overruns the cylinder bore. The harmful surfaces are small, and are jacketed and machined in addition. The ends of the cylinder are provided with jackets since saturated steam is used. The suction of the air pump takes place through ports, and the discharge valves are arranged in the heads, so that the clearance is small and the suction effect a maximum. The condenser is placed immediately under the cylinder with a connection of large area. The cylinder jackets are supplied with steam through a separate pipe con- nected to the steam main ahead of the stop valve. This allows the cylinder to 160 be warmed up before starting, and the valve gear therefore gives correct distri- bution from the very beginning. The eccentric rod is shortened and guided by a swinging link interposed in the valve gear, thus compensating the angularity of the connecting rod. This equalization of cut-offs and valve lifts at the two cylinder ends is very complete for all cut-offs, which range from to 25%. Fig. 40 shows another similar engine with Stumpf gear designed by the author for a company in Finland. The lower seats, instead of the valves, are resilient, and the piston is fitted with Allan metal rings to prevent seizing. The clearance volume is 1,25%. Details of una-flow engines built by the Ames Iron Works, of Oswego, N. Y., are given in Figs. 41 to 44. The head and the cylinder ends are jacketed and the additional clearance spaces are arranged in the cylinder heads. The upper resilient seat of the valves is made of steel and shrunk in place on the cast iron valve body. For non-condensing service the engine is fitted with a piston having cupped ends and the length of compression is 90%. The following test results were verified by Mr. F. R. Low, editor of "Power" (S. page 160). In Fig. 45 is shown a small vertical una-flow engine of 30 HP at 400 r. p. m. for marine lighting service. Fig. 42. A better design is shown in Fig. 46, illustrating a similar engine of 30 HP designed by the author for an English firm. The cylinder bore is 220 mm, the stroke 160 mm, and the speed 400 r. p. m. The governor eccentric oscillates a roller lever acting on a triangular cam which transmits the motion to the valves. The whole cam mechanism is enclosed in a separate housing filled with oil. The cylinder ends are jacketed, and the additional clearance pockets are formed in the cylinder heads and arranged to be heated by live steam when operating condensing. The perfect tightness of- the single-beat valves employed, the small clearance space and 'clearance surfaces, the generous jacketing, and the ample exhaust port area all combine with the una-flow action to insure high economy. The single-beat valves provide absolute safety against damage from water which may be trapped in the cylinder. The design of the two cylinder vertical single-acting stationary engine shown in Fig. 47 is worthy of notice. The inlet valves are single-beat and are placed in 161 the center of the cylinder heads. The valve gear consists of a cam mechanism with reciprocating slides operated through a bell crank by an eccentric and shaft governor located at the free end of the crank shaft. The cranks are set at 180 in order to obtain proper balance at the high speed at which this engine is t<5 run. Single-beat valves are permissible because the high compression balances the pressure against which they open. The valve gear parts may, however, be easily made strong enough to withstand the load if the valve should be lifted when there is no compression to balance the Fig. 43. pressure upon it. The cylinder head proper is a thin dished steel plate, and the cylin- der is provided with a forged steel liner. The cylinder head is jacketed and the upper end of the cylinder is also heated by live steam admitted to a number of turned grooves. Each groove communicates with the adjacent ones at opposite sides so that a continuous flow may take place through the grooves. The engine is intended for use with saturated steam, for which reason the unjacketed part of the cylinder next to the exhaust belt is short. The forged steel cylinder liner should be made of hard ma- terial in order to insure satisfactory ser- vice of the cast iron piston rings. The top surface of the piston is arched to Slumpf, The una-flow steam engine. 1 r^S - s K* rri CT> C^ CO ^ 00 -! s| o oo oo 10 t"* r^ > 2 rt "> ! *H fi > i" S s s 3 i TH * CN CO CO < g "? 2 S -o 1 o to o TH ^ co 10 O to _ CO 00 Jj tO d <^ J5 ^ 10 ^ c o CO <^ O^ s ! ^"^ C"^ CO 2S CD E # M) nj ~ co to ^, CO ^ 00 o -^ 10 10 ^ cs eo c^ *-> '5 o> 13 sfl *-> oo io e* "* o OO (M 10 CO "? CD 5) 3 g o> , ^o oo t^ o I-S c Ul "^ <7 to J? 10 -H 6 S 1 2 o ^ri O v3< m 10 r^ "? "S. E 11 M> O C eg *-< '3 > . g M W KH " ^ C< ' OO m rH O tO X ^ CO !> CO O ^H CO CO C^ *-3* in cd a ip u ,g> . " MH P< A ' S 6 1 g * . . ^ be i . cC iaD K 3 a CD o t/2 5 'I H ' 1 ' ' ' 1 J DH *j Of Q : : i : ; I* : I Ctf 0> o * ^ * ^> *^ c c g ^ g g = -^'Sre'^si 3 T u;rtS oqLl S2 >- o Indicated H Steam const 162 be 163 provide sufficient strength with a light section. The small thickness of cylinder head and cylinder liner is intended to bring the temperature of the harmful sur- faces as close as possible to that of the live steam, in order to reduce their tempera- ture variation. This design allows the additional harmful surface to be reduced Fig. 45. to as small an amount as 8%. The flow of steam through this cylinder takes place in such a perfect manner as cannot be attained by any other cylinder design. The steam enters centrally at the top, spreads out in all directions in the narrow space between cylinder head and piston, and leaves in a similarly even manner at the ll* 164 bottom of the cylinder. The water of condensation collects on the piston surface during the outstroke and is completely removed by the rush of exhausting steam, this action being assisted by the arched form of the piston head. The una-flo\v principle, together with the very favorable flow conditions, the thorough draining of the cylinder at each stroke, the ample jacketing, the perfect tightness of the 165 inlet valve, the small clearance volume and surfaces will insure a very low steam consumption with this type of engine. Proper attention must be paid to the design of the lower part of the piston which forms a seal against vacuum. Fig. 47. Fig. 48 shows an interesting cylinder design with automatic auxiliary exhaust valves. (See also chapters on withdrawal of sfeteam and on locomobiles.) This engine is a una-flow engine in a restricted sense only, since at light loads the exhaust steam leaves through the valves only, while for longer cut-offs part of it exhausts also through the ports. The ports in the cylinder leading to the auxiliary exhaust valves are overrun by the piston, thus determining the compression. The auxiliary 166 exhaust valves in this design are operated automatically by the working steam of the cylinder; they may, however, be opened and closed by a separate valve mo- tion. On the stem of each exhaust valve is mounted a piston working in a cylinder Fig. 48. forming part of the valve bonnet, the upper side of which is connected to the corresponding end of the engine cylinder. When the main piston is near the dead center, the high pressure in the engine cylinder also acts on the piston attached to the auxiliary exhaust valve and thus holds the latter closed until the main exhaust ports are uncovered and the pressure is released. The spring on the upper end of the stem then opens the valve and holds it Fig. 49. 167 open until at the end of the stroke the steam pressure rises sufficiently to close it. Exhaust therefore takes place through the valves until the piston overruns the auxiliary exhaust ports. The clearance space in this design is not larger than that of an ordinary condensing engine, and the length of compression can be chosen to suit any required conditions. The volume loss will be considerably reduced, although there will Fig. 51. Fig. 50. Fig. 52. Fig. 53. be a small increase in the surface loss. In order to keep the latter down to a minimum, the auxiliary exhaust valves should be placed on the cylinder barrel instead of in the heads, and arranged so that in the dead center position of the piston at least one, or preferably two rings seal the auxiliary exhaust ports. 168 If this is done, the compression steam, assisted later by the live steam, will close the valves before the piston uncovers their ports on the expansion stroke, whereby direct exhaust of live steam is avoided. The auxiliary exhaust has no value for condensing engines, but some value for non-condensing una-flow engines with high back pressure and low initial pressure, where it will result in a reduction in steam consumption, especially in the case of jacketed cylinders, with the further advantage of higher mean effective pressures, and therefore smaller cylinders, driving parts and flywheels. Figs. 49, 50, 51, 52, 53, 54, 55 refer to una-flow non-condensing engines built by the Skinner Engine Company, of Erie, Pa, U.S.A. The diagram (Fig. 49) be be and the line of stress. Provision is also made in the design for the collection of the oil and its return to the lubricating system. The moving parts of the engine are oiled by a gravity overhead lubricating system. The cylinder is lubricated by a mechanical force feed lubricator distri- buting oil positively to the proper points. 177 The main bearings (Figs. 65 and 66) are of the quarter-box type, lined with babbitt metal. The cap forms a strong tie and is relieved in the middle so that it exerts pressure directly over the quarter boxes. In the construction shown in Fig. 65, the adjustment is effected by means of heavy set screws, which are provided with steel contact blocks to prevent them from wearing into the quarter boxes, while in Fig. 66 the same effect is obtained by vertical wedges operated by screws passing through the cap. The Nordberg Mfg. Co. also builds una-flow engines with auxiliary exhaust valves which may be put in or out of operation thus making the engines suitable for both condensing and non-condensing service. Fig. 72. Owing to careful use of those principles which have proved successful in Europe, especially those developed by the author, the una-flow engines built by the Nord- berg Mfg. Co. may be considered among the best American machines of this type. Their first engine was fitted with Corliss valves as described in chapter 2 2, page 184. Still further improvement in these engines might be made by the application of high lift single beat valves without cages, thereby reducing the clearance, the harmful surfaces, and leakage, so that the water rate might be still further im- proved. The dual clearance una-flow engine built by the Harrisburg Foundry & Ma- chine Works, of Harrisburg, Pa. (Fig. 67) is of especial interest. As the name Stumpf, The una-flow steam engine. 12 178 implies, this engine has two clearances, the first, or cylinder clearance, between the piston and the valve, and the second (EGE, Fig. 68) connecting the outer ends of the valve. In the earlier design this connection was by an external pipe (Fig. 68) but a hollow valve is now used (Fig. 71). The latter is an ordinary piston valve with snap-rings, moving in a ribbed bushing and having inside admission, as in locomotive practice with superheated steam. As indicated in the figure, the cy- linder clearance is kept as small as possible. The residual steam is first compressed into both clearance spaces together, which are at that time in connection through the piston valve. Towards the end of the stroke the auxiliary clearance is shut off, so that the steam is compressed into the cylinder clearance alone. The valve then automatically connects the auxiliary clearance with the opposite end of the cylinder so that the steam compressed into the clearance now mixes with, and expands with the working steam on the return storke. This arrangement is of course only justified for non-condensing service, so as to avoid a large clearance or auxiliary exhaust valves. The una-flow principle is fully adhered to, since the exhaust steam only leaves through the central exhaust ports. The series arrange- ment of live steam, inlet valve, piston and exhaust is also retained, so that any leakage of steam past the piston valve cannot pass directly to the exhaust. The effect of this dual clearance principle is to raise the expansion line and lower the compression line (Fig. 69). Consequently the mean effective pressure, output and uniformity of speed are somewhat increased and the necessary flywheel weight is decreased. Conditions in an engine of this type for condensing service are somewhat different. The piston heads are flat, but in spite of this there remains a clearance of from 5/ to 7%. In this case also the auxiliary clearance connects the valve head pockets. In addition, a further clearance pocket is arranged to be connected to the cylinder by a spring-loaded valve, which may be opened or closed by hand, or operated automatically, thereby adapting the condensing cylinder to non-conden- sing service (Fig. 72). Increased compression will cause the valve to open against the spring, thus relieving the pressure by admitting steam into the pocket. The heads and ends of the cylinder barrels are jacketed for both condensing and non-condensing service. The piston valve is actuated in both cases by an eccentric on the crankshaft controlled by a shaft governor. The Filer & Stowell Co., of Milwaukee, Wis., successfully built a una-flow engine with drop piston valves and a valve gear resembling a Corliss gear, the valves being located at the side of the cylinder barrel. In the later type, the valves were placed in the heads on the cylinder center line, thus decreasing the clearance and making a better jacket arrangement possible. For higher engine speeds, poppet valves were adopted later, operated from eccentrics on a lay shaft, with a positive opening and closing motion, the eccentrics being controlled by a lay shaft governor placed between them. The result is an engine similar in many respects to the Nord- berg construction, the chief points of difference being the valve gear and bonnet design. The Filer & Stowell Go. evidently consider thorough jacketing of much importance, and therefore their claim of a consumption of 13 7 /s Iks. f saturated steam per I HP per hour for a una-flow engine having a 16 X 30" cylinder, 125 Ibs./sq. in initial pressure, 25" vacuum and 150 r. p. m. may well be credited. For CO r^ tic 12' 181 these engines also the next logical step would be to adopt the single-beat valves actuated from a double-speed lay shaft. Fig. 73 illustrates a Filer & Stowell 20 X 22" una-flow engine driving a 200 kW direct current generator. This engine is arranged for condensing and non- condensing service, auxiliary valves being employed in preference to clearance spaces because of more or less extended periods of operation with as high a back pressure as 5 Ibs/sq in. The clearance space formed in each head by the auxiliary exhaust valve pocket may be closed off by a hand-operated clearance valve, so that no clearance is added when running condensing. If, for some reason, the vacuum should drop to 20" or 22", the hand-operated clearance may be opened, thus adding the clearance formed by the exhaust valve pocket, the valve itself remaining closed. If the back pressure is further increased, the mechanism actu- ating the exhaust valves may be put in action and the length of compression varied to suit the back pressure while the engine is in operation. Fig. 74 shows a group of five 18 X 42" una-flow cylinders which are part of an order of six 18 X 42" and six 20 X 48" cylinders. Four of the 18 X 42" cylinders are to take the place of those of two 18" and 36 X 42" cross-compound corliss engines driving generators, and the two others are to replace those of two ammonia boosters. The six 20 X 48" cylinders are to supplant cross-compound cylinders driving ice machines. All these engines are installed at Swift & Company's plant at La Plata, Argentine Republic, and are to operate with 175 Ibs/sq in. steam pressure and 150 superheat. All these cylinders have additional clearance spaces and clearance valves for non-condensing service. The trend of development is thus the same as in Europe, where many old compound cylinders are being replaced by single-stage una-flow cylinders. 182 2. The Corliss Una-Flow Engine. Figs. 3 and 1 show a side elevation, a vertical section and an indicator card of a condensing Corliss una-flow engine. The eccentric on the crank shaft directly operates the two valves placed in the cylinder heads. The release is effected in the ordinary way by a cam under control of the governor. The valves are closed by oil-vacuum dash pots (Fig. 2) which have absolutely no rebound after the valve has closed. The oscillation which is so common with air dash pots is entirely avoided, since air which is a com- pressible medium is replaced by oil which is incompressible. For this reason the lap of the valve can be made extremely small Q Fig. 1. Fig. 2. and the latch only takes hold and begins to move the valve at a time when the latter is already partly balanced by the compression (Fig. 1). Further experience is required to find the smallest allowable lap of the valve in combination with the proper construction of a reliable oil vacuum dash pot which exactly locates the 184 valve in its end position, in order to reduce as far as possible its movement when unbalanced. In this manner it should be possible to obtain reliable operation with high pressures and superheat. Provision must be made in the valve gear to close the valve positively in case the force of the oil dash pot should be insufficient for proper closure, owing to very small cut-offs or other reasons. A negative angle of advance of 45 gives an ample range of cut-off and also insures release under all conditions, if the governor connections are properly adjusted. The omission of the wrist plate and the attainment of extremely small clearance volumes and surfaces are further valuable features of this design. Fig. 4 shows a Corliss una-flow cylinder as constructed by the Nordberg Manufacturing Company, of Milwaukee, Wis. In accordance with the above prin- ciples, the inlet valves have very small lap and are multiported in addition, so Fig. 4. that four edges open simultaneously. In this way the rotative movement of the valve is reduced and high speeds are rendered possible. The additional clearance volume and clearance valves for non-condensing operation are placed in the lower part of the cylinder heads. In Fig. 5 is shown a non-condensing Corliss engine with the inlet valves arranged in the heads, the auxiliary exhaust valves at the ends of the cylinder barrel, and the una-flow exhaust ports in the center. The eccentric on the crankshaft has a negative angle of advance of 15 and drives the inlet valves directly through an ordinary Corliss mechanism, while the motion of the exhaust valves is derived from that of the inlet valve levers. The inlet valves have the ordinary releasing gear. Admission and cut-off are deter- mined by the steam valves, the beginning of exhaust is fixed by the piston unco- vering the central exhaust ports and compression is delayed until the auxiliary exhaust ports are overrun on the return stroke. The valve gear proper therefore 185 Fig. 5. 186 has no influence whatever upon the exhaust phases. This renders possible the use of a negative angle of advance and therewith a range of cut-off of from to more, than 60%. The length of compression may be reduced to about 8% of the stroke in connection with the- extremely small clearance volume realized with this design. The auxiliary exhaust valves are protected by the piston rings against high pressures and temperatures during a considerable part of the admission period. They are subject to pressure only after the piston has uncovered the auxiliary ports, and do not open until the outer dead center is reached, when the main exhaust ports are wide open. Closure of the valves takes place after the piston has again covered the auxiliary ports. The actual closure of the auxiliary exhaust is therefore so rapid that the indicator card shows sharp corners at this point. Since the inlet valves also close quickly, and as the clearance volume and surfaces are small, a low steam consumption may be expected with this type of engine. 187 3. The Una-Flow Engine arranged for Bleeding. Steam may be withdrawn or bled from a una-flow cylinder by means of check valves placed at suitable points, as shown in Fig. 1. For instance, by providing ports at a distance of say 10% of the piston stroke from the opening edge of the main exhaust ports, the steam withdrawn through the former may be used in heating systems, to drive exhaust steam turbines or engines, or for other purposes. It is possible, for example, to withdraw steam at a pressure of 0,5 to 1,0 at. abs. from the cylinder of a condensing engine and to use it in an exhaust steam turbine driving a rotary air pump. The bleeder valves could also be placed closer to the - -1- \ I Fig. 1. 188 cylinder ends in order to withdraw steam at a higher pressure for heating purposes and the like. A plurality of such heating systems may thus be arranged in series, as shown in Fig. 1, for heating the feed water to a high temperature. In this case, however, a more or less noticeable loss of diagram area must be expected. Fig. 2. An example of such withdrawal of steam is shown in Fig. 2, which illustrates a una-flow locomotive cylinder fitted with automatic bleeder valves. In locomo- tives this method of withdrawing steam may very well be considered for train heating or feed water heating. The diagrams of Figs. 3 and 4 give the quanti- g\100 SO SO VO 2O iff, 25% Fig. 3. ties which may be withdrawn at different locomotive speeds with valves placed at 35 and 50% of the stroke before the dead center. The quantities of steam with- drawn are the larger, the greater the area of the valves, the greater their distance 189 from the exhaust end, the lower the speed and the lower the required pressure. The following table gives the quantities of steam withdrawn for different engine speeds and pressures of the heating steam, the bleeder valves being situated either 35 or 50% of the stroke from the outer dead center. The figures denote quantities of steam withdrawn in percent of the total working steam in the cylinder. Steam quantities withdrawn: Pressure in receiver kgs. per cm 1 Cut-off at quarter of the stroke 45 km per hour 60 km per hour 75 km per hour 90 km per hour Withdrawal valve 35% before the end of the stroke: 1.25 59.0% 55.0% 52.0% 43.0 % 1.5 51.0 48.5 46.0 41.0 2.0 37.5 35.2 32.5 30.0 2.5 25.0 22.5 20.0 17.0 Withdrawal valve midway in the stroke: 2.5 41.8 39.4 37.0 33.6 3.0 29.4 27.0 25.1 22.8 3.5 20.1 18.7 17.0 15.3 4.0 11.3 10.3 9.2 7.9 clearance s/iace fff.zs Fig. 4. 190 The amount of steam which can be withdrawn from the cylinder of a una- flow condensing engine will of course be considerably smaller than in a non-con- densing locomotive. This very important problem of bleeding steam may also be solved by means of a compound una-flow engine with a tandem arrangement as shown in Fig. 5. The piston of the high pressure una-flow cylinder is fitted with a piston valve driven from the main connecting rod, which provides the necessary short compression. The same effect may be obtained by the use of automatic auxiliary exhaust valves operated by the cylinder steam and placed near the ends of the cylinder, their ports being controlled by the piston (Fig. 6). The low pressure cylinder is of stan- dard una-flow construction, although its clearance volume must be increased to about 7 to 10% according to the receiver pressure. The cut-off of the low pressure cylinder may be controlled by a pressure regulator under the influence of the re- ceiver pressure (Fig. 7). This regulator consists of a cast iron housing partly filled Fig. 5. with mercury carrying an iron float, the upper end of which is connected in a sui- table way to the inlet gear of the low pressure cylinder. The position of this float changes with the height of the mercury column displaced by the receiver pressure. The connection is made in such a way that the regulator shortens the cut-off of the low pressure cylinder for a decreasing receiver pressure. In the arrangement shown in Fig. 7 the pressure regulator acts upon an intermediate pin in the drive from the eccentric to the inlet valves, in such a way as to displace the rods from a straight line to a position on either side of it, thus producing the desired change of cut-off, within certain limits, with a permissible change in lead. The piston of the auxiliary exhaust valve (Fig. 6) is under the influence of the cylinder pres- sure. The valve will therefore be closed whenever the pressure inside the cylinder is high. It is opened by a spring in the valve bonnet when the main piston uncovers the exhaust ports and also when the expansion reaches the back pressure before this occurs. The losses accompanying a loop in the exhaust line of the indicator card for small cut-offs are therefore avoided. It is also possible to use large high pressure cylinders even for long cut-offs and yet avoid a large pressure drop at the end of expansion. The opening and closing of these automatic valves is rendered noiseless by an adjustable double-acting oil dash pot (Fig. 48, ch. II, 1, p. 166). The lower resilient seat insures tightness of the valve. The design permits of ample valve lift and valve areas, quiet operation and small clearance, as well as very favorable sealing conditions during the first and last part of the stroke, when the piston rings come between the inlet and auxiliary exhaust valves. Further valuable features of this design are found in the arrangement of all the valves and their gear on top of the cylinder, and the unhampered disposal of all the piping together with the condenser underneath the same, so that every pipe flange is easily accessible. Fig. 6. The exhaust valves of the high pressure cylinder may also be arranged in the more usual way with separate valve gear, possibly with omission of the central exhaust. In Fig. 8 is shown an arrangement which makes it possible to withdraw steam during both the expansion and compression strokes. The una-flow cylinder of standard design is fitted at each end with an automatic auxiliary exhaust valve controlled by the cylinder steam as previously described. The upper end of the stem of the balanced valve carries a piston working in a cylinder the upper side of which is connected by a pipe with the corresponding end of the engine cylinder. This pipe connection is fitted with a reducing valve which opens towards the valve cylinder when a certain pressure is reached. The same end of the valve cylinder also has a second pipe connection leading to a pilot valve communicating with the 192 condenser, which is operated by a mechanism driven by the layshaft. When this pilot valve is lifted, pressure release occurs above the piston of the auxiliary exhaust valve, and the latter is opened by its spring, thus admitting steam from the engine cylinder into the heating connection. When the pressure inside the cylinder falls to or below that carried in the heating system, a return flow of steam is pre- vented by the closure of a number of automatic metal strip flap valves disposed around the auxiliary exhaust valve. The latter, however, remains open until after the auxiliary exhaust ports are covered by the main piston on its return, when the rising compression pressure acts upon the valve piston through the pipe con- nection previously mentioned, and thus closes the valves. The opening or timing of the small pilot valve, which is operated by the layshaft, is controlled by the pressure in the heating system by means of an apparatus containing an iron float carried on mercury. The lower mercury level is exposed to the pressure in the heating system, and a fall of pres- sure in the latter will therefore lower the mercury column and change the position of the float, thus causing the auxiliary valve to open earlier and allowing more steam to be withdrawn from the engine cylinder. Conversely, if the pressure in the heating system rises, the mercury column will also rise and the float in its new position will open the pilot and exhaust valves later so that less steam will be withdrawn. The middle position of the float corresponds to a horizontal position of the shor^ link connecting the upper end of the float rod with a small crank, and the motion of the latter will thus be the same whether the float rises or falls. The crank will therefore be in the extreme left Fig. 7. position for the middle position of the float, and in the extreme right position for either the highest or lowest position of the float. This crank rocks an eccentric pivot upon which is mounted the double-armed lever operating the small pilot valve. These levers are moved by eccentrics keyed to the lay-shaft. The operation of the whole mechanism will be clear from a study of the series of indicator cards reproduced in Fig. 8. Starting with the highest float position, the point of opening of the auxiliary exhaust valve moves more and more towards the left while the float falls. This continues until the float reaches its middle position, for which the cut-off in the engine cylinder determined by the load is so short that hardly any steam can be withdrawn during expansion. (See cards No. 2 and 1.) The regulating mechanism follows the change of cut-off. If much heating steam is required when the engine is operating with small loads the float 193 , oo be Stamp/, The una-flow steam engine. 13 194 will fall still further, the small crank again moves towards the right; and, since no more steam is available during expansion, an arrangement comes into play by which steam is withdrawn during compression. This is done by raising the back pressure and therewith the compression line, by admitting air into the con- denser through a snifting valve actuated by a connecting link from the upper end of the float. The compression steam then soon attains the heating pressure and escapes through the still open auxiliary exhaust valve, and past its check valves (see diagram 8). If much steam is required while the engine is running with heavy loads, then withdrawal occurs during both expansion and compression, as is shown in dia- grams 4, 5 and 6. The cut-off in this case is late enough so that even for the lowest float positions steam will be withdrawn during expansion. The diagrams in a general way show that the action of this mechanism tends to produce the smallest possible loss due to incomplete expansion. Instead of using a crank mechanism for operating the small pilot valves, a cam may be fitted on the upper end of the float rod which acts upon a spring-loaded roller and thus adjusts the position of the fulcrum of the double-armed lever. This combination has the advantage that by properly designing the cam profile any desired depen- dence between float travel and exhaust valve timing may be realized. In order to obtain sufficient valve area for the withdrawal of steam during expansion, each end of the cylinder is provided with two of the automatic exhaust or bleeder valves, each surrounded by a nest of check valves. The end of tlie lay-shaft carries two small eccentrics which operate the pilot valves through the above mentioned double-armed levers controlled by a common float. The governor on the lay-shaft controls the engine output entirely independently of the heating requirements. So far the problem of withdrawing steam for heating purposes has been solved mostly by installing a tandem engine with a counterflow high pressure and una- flow low pressure cylinder, the heating steam being taken from the receiver as was described above. 195 4. The Una-Flow Rolling Mill Engine. Much credit for their work in this field is due to the firm of Ehrhardt & Sehmer, who originated the design shown in Fig. 1. The latter is noteworthy for the use of valve gear of the Zvonicek type, which has the advantage of ample valve opening at short cut-offs as well as a large range of admission without excessive valve lifts at late cut-offs. The governor adjusts the cut-off by moving the eccentric strap which carries a cam profile at its upper side for the eccentric rod roller to work upon. For rolling mill engines a maximum cut-off of 50 to 60% is absolutely essential. This is easily obtainable with the Zvonicek gear, the only disadvantage of which is its complication, although this has never proved to be a source of complaint. The single stage una-flow engine gives considerably more power than the tandem compound and it will pull through where the latter would stall. This feature is highly appreciated by rolling mill engineers on account of the varying resistance of the roljs, and is the main reason for the rapid intro- duction of the una-flow engine in rolling mill practice. This preference has even led to the replacement of several old tandem cylinders by cylinders of the una- flow type. Fig. 2 shows a flywheel rolling mill engine of medium size built by Ehrhardt & Sehmer. The engine has Stumpf valve gear, and the condenser air pump is driven by the tail rod. Figs. 3 and 4 show the cylinder oLa larger una-flow rolling mill engine with Zvonicek valve gear also built by Ehrnardt & Sehmer. Fig. 5 shows an ordinary tandem reversing rolling mill engine built by Ehr- hardt & Sehmer. The three pairs of cylinders act on three cranks set at 120, which arrange- ment requires less maximum admission than cranks at 90. This shorter cut-off allows of greater expansion during the rolling process. The crank shaft consists of three similar pieces coupled by flanges, so that any one of them may be easily replaced in case of failure. The eccentric shaft is carried on the frame and is driven from the crank shaft by means of spur gears. This allows of the use of smaller eccentrics, renders the valve gear more accessible on top of the engine frame and brings the center lines of the units closer together. The piston valves are arranged on top of the cylinders to one side of the center lines in such a way that the high and low pressure valves of one tandem unit are in line and are both driven by the same Stephenson link gear. The latter has 13* 196 crossed eccentric rods as is usual in rolling mill practice. A stop valve is provided on each of the six cylinders, these valves being operated by an auxiliary power cylinder which at the same time con- trols the position of the Stephenson link in such a manner that for long cut-offs the steam is throttled, while for short cut-offs the stop val- ves are fully opened. The stop valves of the low pressure cy- linders thus allow a certain amount of steam to accumu- late in the receivers when stopping the engine. The problem was to adapt this very satisfactory design for -use with una-flow cylinders while retaining as far as pos- sible all its valuable features. This was a change which it would pay to undertake, since three cylinders with their distance pieces, valve gear and accessories could be dis- pensed with. The driving parts of course must be strengthened to take the higher pisto-n loads of the una-flow engine. Such a design by Ehr- hardt & Sehmer is shown in Figs. 6 and 7. The general structure, i. e. frame, valve gear, and the use of piston valves has been retained. The triple arrangement of this en- gine, as in the previous case, will also give the advantage of early cut-offs and long ex- pansions during rolling. To this end the maximum cut-off of the valve gear is made somewhat short, but in order 197 198 199 200 201 202 to insure plenty of power for starting, the piston valve bushing contains an auxiliary port which gives a considerable increase of cut-off and reduction of lap. This arrangement would, however, result in too much lead ; and in order to prevent this the auxiliary port is connected to the cylinder at such a distance from the end of the latter that at the time of steam admission the port has been overrun by the piston and is straddled by a pair of rings. If for instance the maximum cut-off of the main valve is 35% and the cut-off of the auxiliary port 70%, then at slow speeds shortly after starting, or when gripping the ingot, the cut-off of 70% will be effective. As soon as the engine comes up to speed, however, and especially when it begins to race after the passage of the ingot through the rolls, the auxiliary port cannot supply suffi- cient steam to make itself noticeable, bo and the cut-off falls to practically 35%. A con- siderable saving of steam and increased safety of operation are the result. The limitation of the maximum cut-off of the main valve has the advantage that for early cut-offs the port openings are considerably improved, which is especially important in view of the essentially unfavorable valve opening consequent on the use of crossed eccentric rods. The use of a piston valve with a large exhaust lap, in combination with the link valve gear, permits of a reduction in the length of com- pression. The same purpose is served by an auxiliary valve mounted in the main piston valve, by means of which the compression may be reduced or almost entirely eliminated by an adjustment made from the operating platform. Each cylinder is fitted with a separate stop valve for throttling or cutting off the steam supply from the platform, if operating conditions require it. The throttling which takes place when running with late cut-offs permits of gentle and gradual starting of the engine. The steam must in fact be throttled to a greater extent in a una-flow engine owing to its greater starting torque as compared with that of a tandem com- pound. The steam consumption during starting will therefore be correspondingly less. Even with such throttling, the early maximum cut-offs used in this triple engine will produce appreciable expansion. If the ingot should stall the engine, it is a simple matter to exhaust the steam within the cylinders by reversing the gear; and on again admitting steam, the ingot will be freed from the rolls. The above described valve gear will therefore safeguard the operation of the engine even against such eventualities. Separate auxiliary exhaust ports are not necessary since the inlet ports are used for this purpose. The piston valves are preferably designed with inside admission. The steam space between stop valve and piston valve should be made as small as possible. A comparison between this una-flow rolling mill engine and a tandem com- 203 205 fi 03 PH e o u be be fi s o bo 206 pound for the same purpose will demonstrate the fact that considerable simpli- fication, cheaper construction and a reduction in floor space are attainable with this design. Such an engine is also essentially more powerful. The best engine for rolling mill service is the one which consumes the least steam during the comparatively long and frequent periods of idling. The claim made by Ehrhardt & Sehmer that the una-flow rolling mill engine is the only one which satisfies this condition is justified, since it alone has a theoretically correct no-load diagram and the least no-load steam consumption owing to the una-flow exhaust, to the long compression and the comparatively large inlet areas in consequence of the rather early maximum cut-off. In compound engines the steam distribution becomes very poor when using early cut-offs below 20%, and it is advisable to use the throttle instead, to adjust the output to the load. In contrast to the unfavorable changes of exhaust lead and compression in the compound engine, these exhaust phases are always the same in the una-flow engine, since they are determined by the exhaust ports. The no-load diagram must accordingly always be correct. In Fig. 8 is reproduced a series of continous indicator diagrams taken from a flywheel una-flow rolling mill engine, which shows clearly the no-load cards as well as the rapid succession of no-load and full load cut-offs, thus demonstrating the excellent governing characteristics of the engine. This of course applies equally to the reversing engine. A very interesting design of a flywheel una-flow rolling mill engine, 40 X 48". 110 R. p. M. max, built by the Mesta Machine Co., Pittsburgh Pa., is shown in Fig. 9 and 10. The power of the engine is transmitted by spur-gearing on the roller shaft. The live steam enters below into the jackets on the ends of the cylinder barrel, feeding also the hollow head covers. The exhaust belt is separated by two neutral divisions from the jacketed ends of the cylinder. The steam enters the cylinder through resilient poppet valves placed on the cylinder barrel. The hollow piston is carried by a heavy hollow piston rod, supported by the crosshead and a slipper. A common governor controls the cut-off by shifting a small crosshead on the operating eccentric. From this crosshead the motion is transfered by a rod with cam and roller on the inlet valves. The whole design is heavy, strong, reliable and especially adapted for rolling mill work, and all precaution is taken as by safety valves, draining valves, railing, stairs, lagging, enclosures a. s. f. for securing best efficiency and maintenance of the engine. Attention may be called to the comparatively flat steam consumption curve (Fig. 9) for the una-flow engine, which not only shows a lower steam consumption than the compound engine, but also a more nearly uniform steam consumption over wide ranges of load. This latter feature especially recommends the una-flow engine for rolling mill service, where great and suddem variations of the lead are the rule. 207 Fig. 1. 5. The Una-Flow Hoisting Engine. 1. For Condensing Service. The una-flow engine is suitable for use as a hoisting engine if means are pro- vided to eliminate the compression during the periods of starting and stopping, in order to facilitate the exact stoppage of the cage. An auxiliary valve (see Fig. 1) may be employed for this purpose, having sufficient lap to reduce the com- pression to almost nothing when starting or when the cage is being brought into the desired position, while during hoisting the full compression of about 90% is effective. During the hoisting period the auxiliary valve is inoperative as regards the compression and the latter is deter- mined entirely by the piston-controlled exhaust ports of the cylinder. As shown in the diagram of Fig. 2, the compression is respectively about 25, 65 and 90% for cut-offs of 80, 50 and 40%. High economy in the utilization of steam on the one hand, and excellent maneuvering abilities on the other, are thus combined in the best possible manner. When the main inlet valves are slide or piston valves, they may also control the auxiliary exhaust, and the exhaust lap should then be proportioned from the above standpoint. Figs. 3 and 4 show a design recommended for a small hoisting engine. Instead of the auxiliary piston valve shown in Fig. 1, two poppet valves are provided at the ends of the cylinder, which come into ope- ration only at late cut-offs. This arrangement per- mits of a considerable reduction of the clearance volume and surfaces. The valve gear is of the Gooch type and the motion is transmitted to both the inlet and auxiliary exhaust valves by means of a cam mechanism. The valve diagrams for this engine are similar to those shown in Fig. 2. With this construction it becomes possible to run the engine non-condensing for short periods if the exhaust valves and exhaust lap are pro- portioned accordingly. For this temporary condition it is permissible to use rather small valves with correspondingly high steam velocities. The force necessary to operate this gear is so very small that reversing and notching up may be carried out by hand, thus dispensing with a power maneuvering Fig. 2. 208 cylinder. For this reason it is advisable to use the Gooch valve gear, in which only the slide block and valve rod have to be lifted, instead of the link and eccentric rods. The four valves may also be operated by means of the ordinary tapered cam gear (Figs. 5 and 6). The cams are preferably arranged so as to give only a small inlet valve lift for cut-offs of 80 to 90%, and to hold open the comparatively small auxiliary valves during almost the whole of the compression stroke. This Fig. 3. Fig. 4. enables the cage to be stopped with accuracy in the desired position. A late cut- off is also used at the commencement of hoisting, with the auxiliary valves like- wise in action. During the greater part of the hoisting period the cut-off is early and the auxiliary valves are out of action, so that the engine operates as a true una-flow engine under the most favorable conditions. The tapered cam gear may also be operated by hand in most cases without the use of an auxiliary power cylinder. Since the clearance space for 90% com- 209 pression may be properly proportioned to give sufficient compression for any condenser pressure, the maneuvering of the engine may be freely and easily ac- complished without depending on safety devices such as spring-loaded valves to prevent excessive compression. The condenser air pump should be preferably independently driven. Fig. 5. Fig. 6. 2. The Una-flow Hoisting Engine, Exhausting to Atmosphere or to a Low Pressure Turbine. Fig. 7 shows a una-flow hoisting engine for non-condensing service built by the Gutehoffnungshiitte Works for the Vondern Colliery, shaft No. 2. The engine is designed to hoist eight tubs of 500 kg net each, from a depth of 600 m, with SLumpf, The una-flow steam engine. 14 210 a maximum velocity of 20 m/sec. It is fitted with two rope drums each of 6400 mm diameter and 1900 mm width, which are adjustably coupled for hoisting from different levels. The engine at present works non-condensing, with steam of 8 at. gauge pressure. Both cylinders have a bore of 1100 mm and a stroke of 1600 mm. The cylinders are unjacketed plain cylindrical castings supported at their ends on feet resting on base plates. The inlet is controlled by piston valves which are arranged to give a supplementary exhaust while the cage is being brought to rest Fig. 7. and during the first' few strokes after starting. The compression is thus almost entirely relieved, thereby rendering possible an exact stoppage of the cage and easy starting of the engine. The clearance volume is 10% with an additional clearance pocket of 8%. The valves are indirectly operated by tapered cams mounted on a short cross shaft placed at right angles to the cylinder at its middle and driven from the crank shaft by means of bevel gears and a lay shaft. The cams act on pilot valves which control auxiliary pistons coupled to the main piston valves. The tapered cams are shifted by means of the reversing lever without the use of a power cylinder. This hoisting engine is equipped with a combined depth gauge and safety device of the Gutehoffnungshiitte type which controls the hoisting operation from 211 14* 212 beginning to end. It prevents overspeeding and slows the engine down automatically and in a predetermined manner when the cage approaches its stopping place. The safety device not only controls the cut-off as the engine gets under way, but also causes the steam-operated brake to come into action gradually, according to the amount of overspeed, and to release the brake when the speed again falls. If the cage passes its stopping point, the brake is applied with full power. In- creased safety is thus imparted to both hoisting and stopping. The engine has given complete satisfaction as regards service, but not in respect to steam consumption, and this is fully accounted for by the low initial pressure, the large clearance volume and the use of piston valves. The una-flow engine is particularly well suited for the work of a hoisting engine on account of the direct action of the steam. The greater the number of expansion stages, the more sluggish the engine will be; and conversely, the smaller the number of stages the more lively it will be in its action. For this reason, and also on account of the higher diagram factor, the stroke volume of the una-flow cylinder can be made considerably smaller than that of the low pressure cylinder of a compound hoisting engine, both in respect to stopping the cage and for the accelerating period. In the latter connection it may be mentioned that the una- flow engine will run at higher mean effective pressures with the same steam con- sumption as the compound engine. With the una-flow engine there is no need to worry over the maintenance of the compound effect for the various changes of load, or when the engine is tem- porarily at rest. The radiation losses will also be considerably smaller as com- pared with those of the four cylinders and receivers of a twin tandem compound. These radiation losses are especially large in the latter since the receiver pressure must be maintained while the engine is temporarily at rest. On the other hand, the una-flow hoisting engine may use more steam for maneuvering. Apart from its thermal superiority, the una-flow hoisting engine has obvious constructional advantages. In order to make these clear, a comparison has been made in Fig. 8 between a twin tandem compound hoisting engine of usual design having cylinders of 900 and 1400 mm bore by 1800 mm stroke, and a una-flow hoisting engine of the same power, the cylinders of which would have a diameter of about 1250 mm. The length of the una-flow cylinder casting would be 3000 mm as against 2900 of the low pressure cylinder. The overall length of the una-flow engine would be 6 m less than the length of the tandem compound engine. The engine house and the foundation would be shorter by the same amount. Two complete cylinders with valve gear, two distance pieces and two receivers are dis- pensed with. The oil consumption will be correspondingly smaller and the whole engine will be cheaper, simpler and more reliable, notwithstanding its heavier driving parts. 213 6. Una-Flow Engines for driving Air Compressors, Pumps, etc. The constructional simplification of the una-flow engine is of particular advantage in connection with air compressors, pumps or blowing tubs. Two- stage engines are usually built as cross-compounds, which works out satisfactorily in many cases. The una-flow engine, however, permits of a straight line con- struction; and although this can also be employed in two-stage engines in a tan- dem arrangement, it is somewhat inconvenient and has, therefore, not found much favor. The una-flow engine is of course also suitable for a twin' arrangement as shown in Fig. 1. This illustration represents a una-flow pumping engine built by the firm of Gustav List in Moscow for a municipality in Central Russia. The point of importance in this case was reserve power, with first cost as a somewhat secondary consideration. This requirement is satisfactorily met in this engine, since each side is a complete unit and can be operated independently, although with a different flywheel effect. A surface condenser is arranged crosswise under- neath the cylinders and either of the latter may be blanked off from it by means of a slip flange, inserted between the connecting flanges. Each cylinder is also provided with an independent change-over valve and exhaust pipe for non-con- densing operation. The pumps are placed in a well and driven from the engine tail rod by means of a bell-crank which also drives the condenser air pumps. The surface condenser is cooled by the water delivered by the main pumps, so that no circulating pumps are necessary. In Fig. 2 is shown an air compressor in which the air and steam cylinders are combined. The steam cylinder is single-acting and the inlet valve is of the single-beat type forged in one piece with the stem. The valve is actuated by a rolling lever mechanism enclosed in a housing and running in oil. This mechanism comprises a rocking spindle operated by the eccentric, which spindle carries a curved lever acting upon one arm of a rolling lever, which in turn opens the valve. By keeping the whole mechanism running in oil, the wear of the moving parts is almost entirely eliminated. The clearance volume of the steam end is ex- tremely small, since the valve stem may be brought close to the cylinder barrel. An auxiliary exhaust valve is provided to reduce the compression. In order to shorten the cylinder as much as possible, an exhaust lead of 20/ is used and the port area is correspondingly reduced. This larg ; exhaust lead permits of the use of a partial exhaust belt and allows the cylinder and piston to be shortened. The air end is of standard design. The suction valves consist of a split spring steel band or ring covering the suction ports which are drilled in the cylinder flange, and the cylinder head forms the valve guard. The whole cylinder head surface is available for the arrangement of the discharge valves, which is of particular advantage in small compressors. Fig. 1. 216 Fig. 3. Fig. 5. 217 so 218 The discharge valve is also in the form of a split spring steel band mounted in the cylinder head so as to cover the delivery holes communicating with the cylinder. The clearance volume of the air end also is very small. Very favorable steam consumption results may be expected with this construction, in view of the ex- cellent test results obtained with a similar cylinder design given in the chapter on locomobiles. The cut-off may be changed by buckling the eccentric rod by means of a handwheel and link. The central exhaust ports half way between the steam and air ends, with the outlet at the lowest point of the cylinder, will make it impossible for water to get into the compressor side. Simplicity, low first cost and accessibility of all important parts are advantages of this design. One such engine as shown in Fig. 3 has been built by the firm of A. L. G. Dehne, of Halle a. d. S. For larger units a two-crank arrangement would offer certain advantages, one side comprising a standard horizontal or vertical una- flow engine and the other a standard horizontal or vertical blowing tub, air com- pressor, etc. It would be desirable in this case to arrange the cranks in such ,a way as to reduce the flywheel weight to a minimum. Fig. 4 shows a una-flow driven straight line, two-stage air compressor built by the Linke- Hoffmann Works, of Breslau. The cylinders -are arranged in tandem, the air cylinder being next to the frame, with a distance piece between the air and steam cylinders. The differential air piston at the same time performs the function of a crosshead. The high pressure stage is at the crank end, and the low pressure stage at the head end of the air piston. The intercooler is placed in the foundation below the air cylinder. The steam valves of the una-flow cylinder are operated by a Stumpf gear and a governor on the crank shaft. Auxiliary exhaust valves are also provided, likewise actuated by a cam mechanism driven from the crankshaft, by means of which the length of compression may be reduced to about one-half of that given by the central exhaust ports. The engine operates condensing without the use of the auxiliary exhaust valves, and their valve gear should therefore be arranged so that it may be disconnected. A high speed pumping engine built by the Worthington Pump & Machinery Corporation, of New York City, is shown in Figs. 5, 6 and 7. In this case the una- flow cylinder is placed next to the frame; and the double-acting pump, arranged in tandem with it, is connected to the steam end by tie rods running from the pump body to the crank end cylinder head. The steam cylinder bore is 13%", pump plunger diameter 11", stroke 21", speed 210 r. p. m., steam pressure 220 Ibs/sqin. gauge, steam temperature 562.4 F. and total head including suction lift 153 ft. The steam valves are horizontal and are actuated by tapered cams on the lay-shaft, which runs on ball bearings and is driven by spiral gears from the crank shaft. The section of the lay-shaft carrying the cams is shifted endwise by a speed governor so as to control the maximum speed. A pressure regulator acting in a similar manner is mounted at the end of the lay-shaft and governs the engine for constant water pressure. The pump is provided with nests of auto- matic valves of special design, which are accessible from b'oth sides by means of hinged manhole covers. An interesting feature is the complete enclosure of 219 220 t>0 221 Fig. 8. 222 Fig 9. 223 all working parts, which even extends to the lay-shaft, governor, piston rod, crank shaft and flywheel, so that no moving part is visible. The automatic lubricating system includes all bearings, pins and glands. The engine has proved very satis- factory in service. In Fig. 8 and 9 is shown the application of a una-flow cylinder to a recipro- cating tube pump, an indicator card of which is reproduced in Fig. 10. This type of pump is being developed by the Humphrey Gas Pump Company, of Syracuse, N. Y., and consists of a tube provided with a foot valve, which is reciprocated in a well or casing by direct attachment to the piston of a una-flow cylinder mounted at the well head. The weight and energy of the tube and its contents are utilized in connection with the functions of the power medium, which in this machine is steam, by permitting the latter to be used expansively. This is in marked con- trast to the usual type of direct- acting well pump, where late cut-offs are necessary, thus re- sulting in excessive steam con- sumption. The cylinder has a bore of 12", with a stroke of approxi- mately 10", the speed being from 150 to 200 cycles per mi- nute. The cylinder is of course Fig. 10. single-acting and its upper end is closed to form a cushion chamber which serves to retard the reciprocating parts and tube towards the end of the upstroke when the exhaust ports are uncovered, while the water continues to move through the tube. Hence practically the whole of the kinetic energy of the moving parts is stored, and thus becomes available for accelerating them on the downstroke. This energy, as well as that due to the fall of the tube, is then used in the compression of the residual steam in accordance with the una-flow cycle. While this is taking place the water is still in motion relatively to the tube, and this continues until admission begins at the commencement of the upstroke, when the tube is again accelerated and the foot valve closes. The valve gear consists of a swinging link and lever operated from the cross- head to which the rods carrying the tube are attached. Since the length of the stroke is not positively determined by mechanical means, a delayed action of the inlet valve is provided for by causing the valve gear lever to operate a pilot valve, which in turn controls the admission and release of steam behind a piston forming part of the double-beat inlet valve. The latter is of cast iron and works on a resi- lient lower steel seat. The valve is cushioned both on opening and closing by a double-acting adjustable oil dash pot, which is insulated as far as possible from the valve chest cover. This valve gear has proved satisfactory and quiet in action, but contains a somewhat large number of parts. A direct acting piston valve mechanism has also been built. The volumetric efficiency is considerably over 100%. A great una-flow blowing engine, built by the Mesta Machine Co., Pittsburgh Pa, is shown in Fig. 11. 224 Fig. 11. Una-Flow Blowing Engine built by the Mesta Machine Company Pittsburgh. 225 III. The Una-Flow Locomotive. Fig. 1 shows a superheater freight engine which was built for the Moscow- Kasan Railway by the Kolomna Engine Works, of Kolomna, near Moscow. This was the first locomotive built upon recommendation of Mr. Noltein, the manager of the railway, and was fitted with una-flow cylinders designed by the author. They were originally fitted with small auxiliary exhaust valves which were removed later, so that the engine now operates as a true una-flow. Fig. 1. Fig. 2. Fig. 2 shows one of a pair of superheater freight locomotives built for the Prussian State Railways by the Stettiner Maschinenbau - Aktiengesellschaft ,,Vulkan". In Fig. 3 is given a longitudinal and cross section of the locomotive, as well as a longitudinal section of the cylinder. Figs. 4 and 5 show the simpli- city of the cylinder and valve gear parts. All the experience obtained with the above-mentioned Russian locomotive was utilized in this design. These engines were put in heavy continuous day and night service and proved so successful that a number of engines of the same design were ordered. Slumpf, The una-flow steam ensine. 15 226 The valves and live steam spaces are arranged in the heads while the exhaust ports and exhaust chamber or belt are at the middle of the cylinder. This strict separation of hot live steam from the cold exhaust steam is not only advisable for thermal reasons but also from an operating standpoint, since the parts exposed to high temperatures cannot affect the operating conditions of -136O Fig. 3. the piston or cause warping of the cylinder, and the central exhaust belt effectively cools the middle part of the latter where the piston attains its highest velocity. For the non-condensing service of locomotives it was necessary to provide a large clearance volume, in this case of 17%%, for superheated steam of 12 at. This large clearance is the weak point of the design, and in later engines 227 it was materially reduced. The greater part of the clearance space is disposed in the concave ends of the piston (Fig. 6). The piston heads thus take the shape of spherical caps, great strength and stiffness being thereby obtained. They are made of cast steel and are fitted with two piston rings each. A distance piece or drum, made of hard forged steel 7 mm in thickness, is placed between the heads and the whole is clamped together by means of the piston rod and nut, for which purpose a certain amount of clearance is left between the hubs of the piston heads. The supporting drum has a clearance of three thousandths of the diameter, so that for a cylinder bore of 1000 mm its diameter would be 997 mm. This would bring the piston center line 1 % mm below the cylinder center. The allowance takes care of the expansion and distortion of the cylinder and drum. Special con- sideration must be paid to the expansion of the heads, since they are exposed to the full live steam tem- perature during admission. They therefore expand more than the supporting drum and thus distend the ends of the latter, whereby scoring of the cylinder may be caused. This was a source of difficulty in some of the earlier pistons, but was overcome by using a sphere of smaller radius for the piston heads and provi- ding a larger allowance for the ends of the supporting drum. Lubrication is effected, not by introducing oil into the steam chest, but by feeding it to a number of points in the cylinder wall, and thus bringing it directly to the working surfaces. Each end of the cylinder has three Fig. 4. oil feeds, one on top and one on each side on the horizontal center line, each feed being supplied by an independent plunger. Even with this arrangement the oil may still carbonize and it is therefore better to place the feeds closer to the middle of the cylinder. A tail rod with its attendant stuffing box was not used on these locomotives. A gain in weight thus results, but the long piston, the length of which is given by the stroke less the exhaust lead, generally works out heavier than the standard piston with tail rod. The necessary clearance volume depends upon the pressure and temperature of the live steam. Assuming the back pressure to be 1,1 at. abs. (0,1 at. being added for the resistance in the blast pipe), 90% length of compression (adiabatic), quality 15* 228 of steam = 1, and terminal compression pressure = live steam pressure, then for different steam pressures the following clearances are necessary: Steam pressure, at. gage 11 12 13 14 15 16 17 18 19 Clearance volume . . % 16.9 15.8 14.6 13.9 13.1 12.6 12.1 11.6 11.1 In using this table it must be borne in mind that a pressure loss of about 1 at. occurs in the superheater, and that the clearance should be increased by such an amount that for normal cut-off the difference between the terminal corn- Fig. 5. pression and live steam pressure is equal to the difference between the terminal expansion and back pressure, according to the rules given in the chapter on vo- lume loss. The una-flow action is not limited to the steam, but applies also to any foreign matter contained in it, such as scale, mud, cinders or soot. The latter are swept out by the exhaust steam through the central ports and escape through an orifice or drain at the lowest point. The una-flow action therefore permanently maintains the interior of the cylinder in a clean state. The drain just mentioned also insures the removal of water, and thus elimi- nates a difficulty occurring in all ordinary locomotives. The cylinder in the latter forms the lowest point of the system, the live steam enters from the top, and the exhaust steam leaves the cylinder also at the top. Damage due to water, such as fractured cylinders and covers, as well as breakage of driving parts, are possible with this arrangement. Nothing of this kind can happen with a una-flow locomo- tive, since the water is effectively cleared from the cylinder by the exhaust steam 229 and passes away through the drain. It is surprising to see how much water is ejected through this drain when starting with a cold cylinder. A different kind of water hammer may be caused by the kinetic energy of water, apart from its being trapped in the cylinder, and this may of course happen in a una-flow engine. Fig. 3 also shows a mechanism by means of which both valves may be lifted off their seats from the cab, thus putting the two sides of the piston in communication with each other through the inlet pipe, and relieving the compression when coasting. It is advisable to provide considerable lift for the Fig. 6. valves, as otherwise when running down a long grade the temperature of the steam pulsating to and fro between the cylinder ends will become so high in consequence of friction and wiredrawing that the oil may carbonize and cause piston troubles. The admission of cool air to the cylinder through a valve opened simultaneously with the lifting of the steam valves is also recommended. When the steam valves are lifted together with their cams, the rollers should run clear of the latter so as to prevent unnecessary wear. In contrast to the special by-pass arrangements used in ordinary piston valve cylinders, the valve gear of the una- flow locomotive, with only minor additions, provides a by-pass for coasting without increasing the clearance volume and harmful surfaces. The exhaust lead is usually taken at 10%, thus fixing the length of com- pression at 90%. A rapid uncovering of too large a port area may produce an 230 abrupt exhaust and a pulsating draft in the fire box, whereas a uniform draft is desirable for good combustion. This may be realized by using a large exhaust lead in combination with a correspondingly small area of ports and of the blast pipe, as well as by interposing some form of receiver or exhaust chamber, thus allowing of better expansion of the exhaust and serving to muffle the noise. In ordinary locomotives the steam distribution for very early cut-offs is un- satisfactory, and for this reason throttling of the steam is usually employed instead of making the cut-off* early er. In contrast to this, the una-flow locomotive can be run entirely without throttling, and allows the power requirements to be regu- lated entirely by means of the valve gear. The constant compression on the other hand is a disadvantage, particularly in regard to the large clearance volume, and makes itself especially felt when running with large cut-offs. Fig. 7 shows a resilient valve made of forged steel, for a passenger locomotive. The valves are operated by a cam and roller mecha- nism similar to that used with the Stumpf valve gear for stationary engines (Fig. 8). The cam rollers are carried in milled grooves in the reciprocating slides, the grooves at the same time serving as oil retainers. The guides are long enough to prevent the grooves from overrunning them, thereby Fig. 7. preventing loss of oil and the entrance of dust. The oil lubricating the cam crosshead or guide collects mostly in these grooves, and is transferred by the rollers to the cams so that proper lubrication of these impor- tant parts is insured. The guides themselves are provided with wick oilers. The roller slides are operated by a standard Walschaert gear without any changes from that used on existing counterflow locomotives. Both roller slides have screw adjustments. The valve spring is arranged to be adjustable and must be powerful enough to allow of running with late cut-offs at high speeds. Fig. 9 shows a double-seated automatic compression release valve which is in communication with both ends of the cylinder through two pipe connections. The entrance to each pipe is controlled by a valve which may be operated from the cab. When these valves are opened the live steam from one cylinder end closes the corresponding side of the compression release valve and opens the other, so that the compression steam of the opposite cylinder end may escape through the 231 passage between the two seats of the auxiliary valve. When admission occurs at the opposite end of the cylinder, the auxiliary valve changes its position, so that compression is now relieved on the other side. The passage between the seats of the valve may be connected with the exhaust belt. This device therefore allows compression to be entirely eliminated, so that a great reserve of power becomes available for the difficult starting period. The shut-off valves at the cylin- 232 ders are closed as soon as the train is in motion, and the release valve thus be- comes inoperative. Experience has shown, however, that the device is not abso- lutely essential. Fig. 10 illustrates a una-flow freight locomotive exhibited at the Brussels International Exposition of 1910, cross-sections of which are shown in Fig. 3. In order to compensate for the increased weight of the una-flow cylinders, the boiler has been moved back on the frame by a small amount for better distri- bution of the axle loads. Cylinder bore 600 mm Stroke 660 Driving wheel dia. . . 1350 ,, Steam pressure, gauge . 12 at. Boiler heating surface . 140.42 sqm Superheater surface . . 38.97 sqm Grate area 2.35 ,, Weight, empty 52125kg 57750 , Service weight Fig. 9. Fig. 11 shows a una-flow freight locomotive built by the Schweizerische Lokomotivfabrik, of Winterthur, for the Swiss Federal Railways. Cylinder bore 570 mm Stroke 640 Driving wheel dia. . . 1330 Leading truck wheel dia. 850 Steam pressure, gauge . 12 at. Boiler heating surface . 143.7 sqm Superheater surface . . 37.6 Grate area 2.44 Weight, empty . . . .60875 kg Service weight .... 67 700 The general design of the cylinders is similar to those previously described, with the exception that a muffler has been incorporated in the saddle supporting the smoke box, between the exhaust belt and the blast pipe (Fig. 12). Fig. 14 shows a una-flow passenger locomotive built by the Maschinenbau- anstalt Breslau for the German State Railways. An engine of this type was exhibited at the Turin Exposition. 233 234 Cylinder bore . . . . . 550 mm Stroke . 630 Steam pressure, gauge . 12 at. Boiler heating surface . 136.98 sqm Superheater surface . . 40.32 Total heating surface . 177.30 Driving wheel dia. Truck wheel dia. Grate area . . . Weight, empty . Service weight . Adhesive weight 2100 mm 1000 2.31 sqm 56900 kg 62500 35000 The Northern Railway of France had a superheater freight locomotive fitted with una-flow cylinders which were designed by the author in a similar way to those previously described, the existing Stephenson link valve motion being retained. In Fig. 14 is shown a una-flow passenger locomotive for saturated steam, two of which were built by the Kolomna Engine Works for the Russian State Rail- ways. Two other engines of the same design were fitted with superheaters. The former run with 14, and the latter with 12 at. gage pressure. The extra pressure Fig. 12. carried by the engines using saturated steam was rendered possible for the same axle loads by putting the weight of the superheater into the heavier boiler plates. Since there is a loss of about 1 at. in the superheater, there is a gain of 3 at. in favor of the locomotive using saturated steam. All previous experiences and test results were utilized in the design of the cylinders. Attention is directed especially to the careful jacketing of the heads and ends of the cylinders for saturated steam. The condensate from the jackets is returned to the boiler by a small pump with suction ports, operated by a cam on the roller rod. Main dimensions of the engines \ising saturated steam: Cylinder bore 500 mm Driving wheel dia 1700 Steam pressure, gauge . . 14 at. Stroke .... Heating .surface Grate area 650 mm 166.57 sqm 2.45 A series of comparative tests were made by Prof. Lomonosoff on the Russian State Railways on one each of the saturated and superheater una- flow locomotives, in competition with two ordinary saturated steam com- 260 Una-Flow Locomoti f the German State Railways. 235 236 pounds and one superheater compound locomotive, are as follows: The results of these tests 2 6 two cylinder loco- Saturated Steam Superheated Steam 1700mm Three-axle tender Compound Una-flow Una-flow Compound Train weight 350 tons I Tver to Moscow j V G 51.5 35.6 50.3 37.7 50.5 35.7 54.3 30.8 49.6 29.2 Z 42.8 36.9 41.5 45.8 37.4 Train weight 270 tons I Moscow to Tver | V G 67.3 43.9 71.0 38.8 73.3 32.7 76.3 29.4 69.0 32.0 Z 41.1 42.1 41.3 42.5 38.1 Train weight 424 tons I Moscow to Tula 1 V G 49.4 28.3 51.6 31.8 43.3 35.3 47.9 25.0 44.6 23.7 Z 39.8 42.9 49.5 42.5 34.7 Train weight 270 tons J Tula to Moscow j V G 56.8 40.1 56.8 38.5 55.4 36.4 59.9 28.0 57.9 30.1 Z 40.5 36.0 34.5 37.5 33.7 V = Mean speed in km/hour. G == Naphta consumption per 1 ton-km. Z = Evaporation per 1 sqm of boiler heating surface. These figures permit of a fair comparison of the economy of the locomotives; since for one run the total number of ton-km is the same for all locomotives, their output must also be practi- caly equal, except for changes in the train resistance caused by wind and other weather conditions. The evaporation Z per 1 sqm, however, is different for every locomotive and no conclusion can there- fore be drawn from the same. The line from Moscow to Tver is almost level, but that from Moscow to Tula has long and heavy grades in both directions. The conclusions which can be drawn from these tests are as follows: The una-flow locomotive shows better economy than the compound for small loads, while at higher loads its fuel consumption is higher than that of the latter. This can be easily explained by the effect of the long constant compression and the large clearance volume (see chapter on volume loss). The una-flow locomotive working with saturated steam shows in general a higher economy than the compound except for long cut-offs. Larger cylinders would be of advantage in this case. Fig. 15. 080 Una-Flow Locomoti : the Moscow Narrow Gage Railway. 16. 237 The superheater una-flow locomotive is at least on a par with the super- heater compound, although even here the former has a slightly higher fuel con- sumption for heavy loads. Larger cylinders are of course more feasible with the una-flow system than with the compound engine. Fig. 15 shows a una-flow cylinder with horizontal valves for a small locomotive built by the Kolomna Engine Works. The mechanism operating these valves is of in- terest; it consists of a double armed lever, the lower end of which works against the valve stems while its upper end receives its motion from a rocking lever fitted with two cam profiles which are alternately in rolling contact with it. The rocking lever is driven by a Marshall gear. Attention is drawn to the accessibility of the valves which are removable after taking off the valve chest covers, without disturbing any other part of the gear. Fig. 17. It will be noted that the Kolomna En- gine Works consistently adhere in all their designs to the true una-flow arrangement. Much value is placed by these builders on the series arrangement of live steam space, inlet valve, piston and exhaust. 238 In Fig. 17 are given the main sections of a una-flow cylinder for a narrow gage locomotive (Fig. 16) using saturated steam, built by the Kolomna Engine Works, which was shown at the Turin Exposition. In view of the small size of the cylinders, the latter are fitted with slide valves, which are separate for each cylinder end. Since these valves are of small area compared with that of the ordinary D- valve, the load upon them, as well as the resistance which the valve gear has to overcome, is considerably diminished. Reliable operation is insured by feeding oil under pressure to their working faces. Cylinder bore 355 mm Stroke 350 Driving wheel dia. . . 750 ,, Steam pressure .... 12 at. gage Boiler heating surface Grate area Weight empty . . . Service weight . . . 54.26 sqm 0.93 19.2 tons 21.2 Fig. 18. Fig. 19. Fig. 20. It is the policy of the Kolomna Engine Works, in cases where superheating is not acceptable, to obtain improved economy by applying the una-flow system. In this way they rebuilt with una-flow cylinders eleven saturated steam loco- motives of the tank type of the Warsaw narrow gage railway during the years 1913 1914. These una-flow cylinders are fitted with piston valves and have the cylinder ends jacketed, in addition to the heads. 239 The use of two rows of central exhaust ports with a simultaneous increase of the exhaust lead from 10 to 30% considerably shortens the cylinder and piston, as shown in Fig. 18. The second set of exhaust ports provides a very effective increase in port area shortly before dead center. If this advantage is not con- sidered essential, then the use of a single row of ports with the same exhaust lead of 30% will still further reduce the length of the cylinder and piston (Fig. 19). It is important to reduce the port area as the exhaust lead is lengthened in order to keep the loss of diagram area at the end of the stroke within reasonable limits. (See diagram Fig. 20.) The long exhaust lead at the same time shortens Fig. 21. the compression from 90% to 70% and reduces the clearance volume from 16,2% to 13,6%. This reduction of the length of the cylinder and the saving in weight of both cylinder and piston should prove of value particularly for locomotives. The design of the una-flow locomotive cylinder shown in Fig. 21 differs from those previously described in that the piston valve gives-a supplementary exhaust for long cut-offs and the engine operates on the true una-flow cycle only for early cut-offs. As is shown by the indicator diagrams, the true una-flow cycle is main- tained for cut-offs up to 30%. For longer cut-offs an auxiliary (counter- flow) exhaust comes into effect, which shortens the compression with increasing cut-off. In this way starting is facilitated for long cut-offs, and the advantage of the full una-flow cycle for steady running is yet retained. The compression release device shown in Fig. 9 is thus dispensed with in this case. The varying length of the compression considerably reduces the volume loss at late cut-offs, but this is ob- tained at the expense of the series arrangement of live steam, inlet valve, piston and 240 exhaust. The piston valve has inside steam admission, and the outside lap con- trols the supplementary exhaust. The inlet and exhaust at each end of the piston valve are in parallel with the piston, and steam leaking by the valve will pass directly into the exhaust. The supplementary exhaust is conducted from the ends of the piston valve housing to the exhaust belt through separate pipe connec- tions. The piston is built up of two cast steel pieces, each of which is fitted with a bronze shoe. Fig. 22 shows a similar piston valve design which was used for two passenger locomotives of the German State Railways. A corresponding design was employed for the cylinders of a una-flow locomotive of the North Eastern Railway of England (Fig. 23), and for those of the locomotives of the Neuruppin- Kremmen-Wittstock Railway (Fig. 24). This same cylinder and piston valve design was also used for the three cylinder passenger superheater locomotives (Fig. 25) built for the German State Rail- ways by the Vulkan Engine Works of Stettin. The three cranks are set at 120. Fig. 22. The two outside cylinders drive the second coupled axle, and the inside cylinder, which is placed slightly ahead of the others, operates the first driving axle. By distributing the piston loads upon two axles a longer life of the center crank axle is expected, especially since the forging can be made with easy curves, without disturbing the natural fibers of the material. The motion of the valve of the inside cylinder is derived from the two outside Walschaert gears, the movement of which is combined according to the parallelogram of velocities. The head ends as well as the crank ends of the three cylinders may be connected by means of special valves in order to relieve compression when coasting. An examination of the exhaust timing of a three cylinder locomotive shows that the exhaust periods of the individual cylinders overlap; a very uniform draft in the fire box is thus obtained, particularly if an auxiliary exhaust is provided for long cut-offs. The three cylinders constitute an excellent reinforcement of the locomotive frame. The clearance volume of each cylinder is 11%, the bore 500 mm and the stroke 630 mm. 241 Fig. 23. Fig. 24. Stumpf, The una-flow steam engine. Fig. 25. 16 242 fcC 243 The piston valve bushings are designed in such a way as to prevent catching of the rings when removing or reassembling the valves. The pistons are fitted with bronze shoes in the middle of their length so as to avoid contact with the hot part of the cylinder. Noteworthy is the small clearance volume of only 11%, as well as the manner of arranging the clearance spaces so as to reduce the harmful surfaces to a mini- Fig. 27. mum. For this reason the piston valve is made single-ported instead of the more common double-ported construction, and this is compensated for by the long travel of 190 mm. The locomotives are equipped with superheaters of the Schmidt type and exhaust steam feed water heaters. In the first locomotives described in this chapter, the true una-flow principle was adhered to by the author in order to obtain a minimum surface loss in addition to the other advantages mentioned. This minimum surface loss, however, is accompanied by a large volume loss. When working with superheated steam a small 16* 244 J610 Fig. 28. 245 surface loss can also be realized in the counterflow construction if the whole cycle takes place in the superheated region, as is nearly always the case in modern superheater locomotives. A simultaneous reduction of the surface and volume losses to a minimum is possible with una-flow locomotives in the following manner. A reduction of the volume loss while retaining the full una-flow cycle is pos- sible by raising the initial, or lowering the back pressure. The latter way avoids the difficulties arising from a considerable increase in the boiler pressure and is based on the utilization of the large amount of energy still contained in the toe of the diagram, for the purpose of reducing the back pressure. This principle was described in chapter I, 7. Its first application is found on a superheater freight locomotive of the German State Railways, built in 1920 by A. Borsig of Berlin, according to designs furnished by the author. The main dimensions of the locomotive, which is illustrated in Figs. 26 to 28, are as follows: Cylinder bore 630 mm Stroke 660 Driving wheel dia. . . 1400 ,, Maximum speed .... 60 km/hour Steam pressure .... 12 at. gage Grate area 2.62 sqm Boiler heating surface . . 149.65 sqm Superheater heating surface 53.00 ,, Total heating surface . . 202.65 Feed water heater surface 15.0 ,, Weight empty . . . . ? 65.5 tons Service weight 72.0 This brings the una-flow locomotive into a new phase of development, since the lower back pressure reduces the compression pressures and permits of the use of smaller clearance volumes. The exhaust ejector action also produces an approximately correct variation of the compression with the cut-off, since the energy available in the large toe of late cut-off cards produces a strong ejector effect, with a correspondingly low back pressure and a low compression pressure; while at early cut-offs the ejector action is less pronounced and the back pressure and terminal compression pressure are higher. In order to obtain the exhaust ejector effect a large exhaust lead is essential, and the latter at the same time shortens piston and cylinder. With this long duration of the exhaust only a small port area is required, with the result that the exhaust belt can be dispensed with and a considerable reduction in the weight of the cylinder and piston thus results. A comparison of Figs. 27 and 28 with the previous designs indicates how much more compact this construction has become. This is in part due to the use of horizontal single-beat poppet valves which were employed for the first time on this locomotive. This type of valve, although simple and perfectly steam tight, has so far not been favorably received because it requires a high lift and a large force to raise it. With the high compression of the una-flow engine, however, the pressure on the valve is balanced to a large extent and the high lift is obtained by arranging the cam roller between the valve stem and the fulcrum of the valve lever. The lift of the cam, which is 14 mm radially, is thus increased to 24 mm at the valve. For cut-offs up to 50% the effective inlet areas of the single-beat valve are equivalent to the areas of a standard piston valve of 220 mm diameter. The fact that beyond this cut-off the valve area remains constant must be consi- dered a further advantage. The small cam lift permits of a cam profile of very gentle curvature, thus insuring smooth lifting and seating of the valve. The whole 247 cam mechanism is very substantially constructed and swinging levers were used instead of sliding parts wherever possible. It should therefore stand up well in service. The single beat valve is made of chrome-nickel steel and works on a remo- vable steel seat expanded into the cylinder casting (Fig. 30). If this seat should become damaged by scale or other foreign matter it can be easily resurfaced or renewed. The valve stem has a diameter of 25 mm and is supplied with oil under pressure. The common center of gravity of the valve head and spring retainer is located at about the center of the guide so that good working conditions are assured. The valve stem furthermore is not exposed to the live steam but to the varying pressure and temperature of the cylinder steam. It is entirely independent of the cam mechanism except for the tappet contact, and is free to follow any slight distortion of the cylinder casting. Considering the success of the horizontal valves in the Lanz locomobile, which are double-beat in addition, the con- clusion is justified that this is a very reliable construction. When coasting, the valves may be lifted off their seats by com- pressed air admitted between small pistons formed on the valve tappets, so that the rollers clear the cam. Special means fcr connecting the cylinder ends are therefore not re- quired, and the usual relief valves may be omitted, since" the inlet valves act as such. They also relieve the high compression which may occur when the throttle is nearly closed. The automatic compression release device also may become superfluous since the late cut-offs at starting produce a strong exhaust ejector effect and the com- pression is therefore considerably shortened. Attention may be drawn to the accessibility of the valves; for their renewal it is only necessary to take off the valve chest cover and disconnect the valve spring, the spring cap lock being a split spherical washer. Comparing this with the procedure of taking out an ordinary piston valve, which requires frequent removal of carbonized oil, the great simplification due to the single beat valve will be appreciated. The driving parts and the Walschaert gear are the same as those used on counterflow locomotives. On account of its greater length the cylinder was moved forward by 180 mm. The una-flow cylinder is not heavier than the corresponding counterflow cylinder, since the piston valve chest with its large exhaust chamber, as well as the tail rod and its guide are omitted. This allows the piston rod of 95 mm diameter to be bored out to a diameter of 60 mm, thus also saving weight. The forged steel piston heads, which are only slightly dished, hold between them a cast iron supporting drum cast from a special soft mixture, while the cylinder is made of a hard quality of cast iron. The supporting drum is turned smaller Fig. 30. 248 than the cylinder bore by 2,2 mm on a length of 140 mm at its middle, which allowance increases to 5 mm towards the ends. Each piston head carries three rings. The greater part of the total clearance volume of 12% is taken up by a linear clearance of 40 mm between piston and cylinder head, and this also results in very small harmful surfaces. The pressure oil feeds are arranged at the middle of the cylinder, where the temperature is lowest and little possibility of carbonizing exists. One feed is placed on top and one on each side at 45 below the horizontal center line. It may be a matter of surprise that hardly anything is ever heard of attempts to utilize the energy of the exhaust steam of one cylinder to produce a vacuum in another. Such experiments have been made, for instance in connection with Kordina's blast pipe, but they were bound to fail in counterflow locomotives. As was shown in chapter I, 7, in dealing with the exhaust ejector effect, a large part of the available energy is require'd for producing the draft in Ihe fire box, and if the energy is used in a wasteful fashion there will be nothing left for the reduction of back pressure. The ordi- nary counterflow cylinder has exhaust passages of such an uneven character that the steam continuously suffers changes of direction and velocity which naturally dissipate part of its energy. This is. well shown by a comparison of Fig. 31 with Fig. 28. The una-flow cy- linder is designed on the principle of conserving the exhaust energy, while in the counterflow design it might be thought that a dissipation of energy was aimed at. However, no blame for this can be laid on' the designer, since tortuous and uneven exhaust passages are inseparable from the use of piston valves. In consequence of these conditions, a certain pressure difference at the blast nozzle becomes necessary, since velocity must again be generated. It may be con- sidered an excellent result if a pressure difference of 0,1 at. gage at the blast nozzle is found sufficient. In many cases, however, the kinetic energy is dissipated to such an extent that there is not enough left to cover the blast pipe loss. The latter can only be diminished by a reduction of the velocity. Since the blast pipe area is fixed, this reduction of velocity can only be accomplished by a diminution of the volume, which thus leads to an increase of pressure. In this way the back pres- sure in piston valve cylinders occasionally rises to 0,5 at. In contrast to this con- dition, a back pressure of 0,5 at. below atmosphere is aimed at with the ejector effect of the una-flow exhaust for late cut-offs. Fig. 31. 249 Although the possibility of the utilization of the exhaust energy for the pur- pase of reducing the back pressure has no inherent connection with the una-flow principle, it is limited to the latter by virtue of the conditions of exhaust whereby the energy is conserved, and it must therefore be considered an integral part of the same. The exhaust opening in the cylinder wall is in the shape of a nozzle which is dimensioned in such a way that in combination with the rounded edge of the piston an approximately correct nozzle area for any particular pressure drop is obtained, when assuming average pressure and speed conditions. The remaining pressure energy is therefore transformed into kinetic energy and thus reaches the blast nozzle with the least possible friction losses. Even for full opening there is a certain amount of divergence in the exhaust nozzle. At the junction of the I I E ]3mi 6.8* i, 2. * Fig. 32. two exhaust pipes the section suddenly doubles, and from this point on the pipe diverges to the blast nozzle and thus acts as a kind of diffusor. The steam for the feed water heater is withdrawn from the cylinders by separate nozzles, and the pipes from them lead to a common ejector nozzle similarly to the main ex- haust pipes. It was of course important to make the blast pipe section as large as pos- sible in order to reduce the blast pipe loss. The stack was therefore designed with the most favorable dimensions, its diameters being 460 and 610 mm as compared with 410 and 460 mm of that of the standard engine. The possibility therefore arose that the jet leaving the blast nozzle, which was dimensioned to obtain a diffusor action, would not spread out sufficiently to fill out the stack section. This danger is especially great in this case, since the jet has very little internal pressure and spreads out to only a comparatively small extent. If the stack is not completely filled with steam, air enters the smoke box from above and thus 250 partially destroys the vacuum. In order to prevent this, the jet may be divided by ribs in the nozzle into smaller diverging jets, which spread out and again join in the stack (Fig. 28). In actual operation however the ribs proved to be superfluous. The effect of the ejector action is shown in the diagrams of Fig. 32, which were drawn for an evaporation of 7000 kg/hour and for speeds of 20, 40 and 60 km/hour, under very conservative assumptions. The clearance volume was assumed to be 12%, so that the compression would not exceed the initial pressure even without the exhaust ejector action. It will be seen from the diagram that the ejector effect only begins after a piston travel of 6,7%, since in a two cylinder locomotive the cylinders are in connection only during a short part of the stroke, and therefore only part of the exhaust energy can be utilized for producing the ejector effect. It was shown in chapter I, 7, how much greater the gain would be with a three cylinder locomotive, especially when working with high mean effective pressures, since in this case with proper length of exhaust lead the whole of the energy of the diagram toe can be utilized for the exhaust ejector action. In connection with this, it may be mentioned that the trend of modern loco- motive design is toward an increasing adoption of the three cylinder locomotive. The exhaust ejector principle, however, is only one of the means for reducing the volume loss, its effect being the lowering of the back pressure. The next step will consist in raising the upper limit of the steam pressure. With the customary design of fire box, steam pressures up to 16 at. gage are possible, although the number and size of the stays becomes excessive. For still higher pressures a diffe- rent design of fire-box would be necessary, such as for instance a box of the Brotan type, which permits of steam pressures of 20 at. gauge. The following table shows that by raising the steam pressure from 12 to 20 at. gage, the amount of heat which can be converted into work increases from 116 to 146 cal. per 1 kg steam. This represents a gain of about 26%. Boiler pressure at gauge Live steam Exhaust steam Heat con- verted into work Cal. Satur- ation point at. abs. Clear- ance volume Cylinder pressure at gauee Temp, at Cylinder C Total heat Cal. Back pressure at. abs. Total heat Cal. Counter-flow 12 11 320 740 1,15 624 116 2,2 10 Una-flow 20 19 320 734 0.85 588 146 4.75 7 This remarkable result can only be attained by employing the una-flow prin- ciple, since the pressure at which the steam becomes saturated increases from 2.2 to 4.75 at. abs. and during a considerable part of the expansion moisture is therefore formed in the cylinder. This does not have much effect upon the eco- nomy of the una-flow engine but is very detrimental to that of the counterflow cylinder where the presence of moisture causes large surface losses. The practice with counterflow locomotives is therefore to use higher superheat with higher initial pressure. The increase in temperature, however, is the cause 251 of many difficulties with piston valves and piston rod packings. Furthermore, the superheater elements must be shortened so that the flue gases do not exert a cooling effect upon the superheated steam. This in turn leads to a great number of superheater elements and inefficient utilization of space. The una-flow engine can of course be adapted to meet this condition in a better manner, since its design makes it more suitable for high temperatures than the customary counterflow cylinder with piston valves. On the other hand there is no necessity for using these high temperatures in the high pressure una-flow locomotive, since the una- flow action corrects the bad influence of moisture in the steam. The calculated gain of 26% of the high pressure una-flow locomotive with exhaust ejector action will probably be exceeded in practice, since it does not include the benefit due to the single beat poppet valves, the reduction of the clearance volume to 7%, nor even the gain due to the una-flow principle itself. The future line of progress of the locomotive is therefore clear. It leads natu- rally from the two-cylinder to the three-cylinder engine with una-flow cylinders having small clearance volumes, to the use of single-beat poppet valves and the utilization of the ejector action of the exhaust, in combination with high pressures and high superheat. 252 IV. The Una-Flow Locomobile and Portable Engine. The una-flow engine is .especially suitable for locomobiles and portable engines. Simplicity, lightness, cheapness and economical use of steam are demanded of this type of engine, light weight being particularly required for portable and self pro- pelled machines. All these requirements are satisfied by the una-flow engine. Figs. 1 to 3 illustrate the construction employed by the Maschinenfabrik Ba- denia vorm. Wm. Platz Sohne, A.-G., Weinheim (Baden). In comparison with the usual design of locomobile with a tandem engine, there is a saving due to the omission of one cylinder with all its accessories. Com- pared with the usual type of compound poppet valve locomobile, one complete driving unit and two valves on the remaining cylinder are dispensed with. While the una-flow locomobile engine has only two valves, the compound locomobile requires eight. The omission of exhaust valves is of particular advantage. The cylinder may therefore be mounted directly on the boiler, and the two inlet valves may be arranged vertically in the heads (Fig. 4). These valves are operated by an eccentric mounted on the crank shaft next to the flywheel, controlled by the flywheel governor. The motion is transmitted to the valves in the same manner as on stationary una-flow engines. Bolted to the cylinder is a frame of the forked type which carries the center crank shaft. On the free ends of the shaft are mounted two overhung flywheels, one of which carries the governor and the other is pro- vided with teeth for the barring device (Fig. 5). Provision is made in the design for interchanging the two flywheels so that the condenser may be placed either to the right or the left of the engine. All the parts concerned are constructed in such a way as to make them suitable for either method of setting up, and a good basis for large scale production is therefore obtained. The air pump is placed vertically and is driven from a crank pin on the flywheel. Una-flow locomobiles are built for condensing service as well as atmospheric exhaust. In the former case, non-condensing operation is provided for by the employment of auxiliary clearance pockets. The reciprocating masses are partially balanced by counterweights on the crank cheeks and in the flywheels. The governor shown in Figs. 6 and 7 has two weights pivoted on two pins fixed on the flywheel. These weights are in the form of bell crank levers, the short arms of which carry rollers bearing on the thrust washer of a central spring common to both. The long arm of each weight is provided with a pin at its extreme end, but only one of them is used for the connection to the shifting eccentric. The free weight, however, also takes part in moving the eccentric to the extent that it bears more strongly on the spring plate and thus relieves the second weight, 253 Fig. l. Fig. 2. 254 Fig. 4. 255 which therefore has a larger force available for shifting the eccentric. The primary eccentric is provided with tapped holes so that it may be turned through an angle of 180 2 6 (6 being the angle of advance), and bolted in this position. If the shifting eccentric is at the same time swung around and its connecting link attached to the other weight, the governor is then changed over for a left hand engine. This change can therefore be made without any alteration in the parts concerned. This governor is obviously only suited for high speeds, since the natural oscillations due to the weights themselves would be detrimental to regu- lation at low speeds. This type of governor, although now no longer used, is mentioned here in order to demonstrate the conditions which must be satisfied by a locomobile governor from the point of view of its suitability for a variety of uses. The movement of the shifting eccentric is transmitted to the valves by a rocking lever and cam and roller motion in the manner previously described. It is advisable to make the rocking lever in one piece of cast steel. In Fig. 8 is shown the assembly of a 100 HP locomobile, also built by the Badenia Company. The cylinder is bolted to the boiler and is in rigid connection with the main bearings through the forked frame. The bearing housings rest on half-round pieces of steel, so that free expansion of the boiler and correct align- ment of the shaft are assured. This feature has been patented by the builders. The bearings are of double construction, so that the inner halves on the one hand closely support the crank .arms and take the steam load, while the outer halves carry the flywheel load. Chain lubrication is provided for in the center of each double bearing. One of the flywheels is cast with teeth for barring pur- poses and the other carries the shaft governor described above. The additional clearance pockets which are provided to allow the engine to be run non-condensing are arranged in the cylinder heads. The exhaust belt of the cylinder has an ex- tension forming a base which is bolted to the boiler. The exhaust belt has an opening on either side for the connection to the condenser. The flywheels, frame, bearings, cylinders and heads, condenser, all the valve gear parts, governor, barring device, feed water heater, change-over valves, and all accessories are constructed in such a manner as to allow of either a right or left hand arrangement, to provide for condensing or non-condensing service, and to suit either direction of rotation, according to the purchaser's wishes. The chief features which render the una- flow engine of particular value for locomobiles are the central mounting of the cylinder on the boiler, the arrangement of the valve gear on the cylinder, the well balanced connection between the shaft bearings through the forked frame to the cylinder, the freedom of relative expansion between boiler and frame, and the simplicity of the entire construction. The excellent superheater designs of the Maschinenfabrik Badenia shown in Figs. 9 and 10 are worthy of note. The advantages of this design are small thrott- ling losses, large capacity per unit of surface, and accessibility of the superheater and boiler tubes. The superheater is mounted at the smoke box end of the boiler in such a manner that the area in front of the flue tubes is left free for access to the latter (Fig. 9). 256 Fig. 5. Fig. 6. Fig. 7. 257 Stump/, The una-flow steam engine. 17 258 Fig. 9. Fig. 10. 259 In the construction shown in Fig. 10 the same result is attained by arranging the superheater tubes in a winding fashion back and forth between the flues, so that the latter are still accessible. Figs. 11, 12 and 13 illustrate later designs by the same builders which show greater care in the support of the crosshead guides and the main bearings, as weli bD as increased area of the exhaust openings. The unbalanced vertical centrifugal forces of the counterweights are transmitted from the main bearings through supporting colums to the foundations. More latitude is therefore allowed in the proportioning of the counterweights. 19* 260 Fig. 12. Fig. 13. 261 Fig. 14. Fig. 15. 262 In Fig. 14 is shown a una-flow semi-portable engine or locomobile, built by Robey & Co., Ltd., of Lincoln, England. Figs. 15 and 16 illustrate a una-flow locomobile constructed by the Erste Briinner Maschinenfabrik, of Briinn, Czecho- Slovakia. In Figs. 17 to 19 are shown the general arrangement, cylinder design and valve gear diagrams of an agricultural una-flow portable engine built by the Kolomna Fig. 16. Engine Works, of Kolomna, near Moscow. This type of engine must necessarily be of cheap and simple construction, and yet answer all the demands put upon it. With this in view, the single-beat valves are arranged horizontally, with an ope- rating mechanism common to both. This consists of a three-armed rocking cam lever, one arm of which projects between the ends of the valve stems. The others are formed with cam profiles, by means of which it is rocked back and forth by contact with the roller of a swinging lever mounted on a shaft above it. The 263 latter is operated by a lever and link from a shifting eccentric controlled by a shaft governor. The whole of the cam mechanism runs in oil and is enclosed in a housing cast onto the exhaust belt. The cover of the housing is adapted to support the smoke stack of the boiler when it is not in use. In this small type of engine, un- balanced single-beat valves are used, and these are made in one piece with their rn stems. The horizontal arrangement of the valves was adopted in order to obtain a simple drive common to both of them. The cylinder is cast in one piece with the crank end head. This engine was designed for use with saturated steam of 10 at. pressure. In accordance with the results of previous tests, the heads as well as the ends of the cylinder barrel were well jacketed. A neutral, unheated zone extends between the cylinder jackets and the central exhaust belt. 264 The cylinder rests on a casting (Fig. 20) which at the same time forms a housing for the stop valve and distributes the steam to the ends of the cylinder. The passages in this casting are so arranged that the water of condensation from the jackets may run back to the boiler. The piston is made in two parts, with bowl - shaped ends to accommodate the necessary clearance space. In Figs. 21 to 26 are shown details of the crosshead guides, stuffing box, crosshead, connec- ting rod, main bearings and crank shaft. The boiler is provided with a fire box of large size, so that straw and wood may be used for fuel. The assembly drawing of Fig. 17 and the half-tone illu- stration of Fig. 27 show that the construction is extremely simple and well adapted to ser- vice conditions. All the previous experience and results of tests have been fully utilized in order to obtain a very low steam consumption. A similar portable engine was built by the Engine Works of the Hungarian State Rail- ways in Budapest. A pair of in- dicator cards of this engine are reproduced in Fig. 28, and the cylinder is shown in Fig. 30. To permit of the use of a small clearance volume, steam operated auxiliary exhaust valves (Fig. 29) are provided near the ends of the piston travel. To the end of each valve spindle is attached a spring loaded piston, the outer side of which is connected by a pipe to the clearance space of the corresponding cylinder end. During admission and expansion, the high steam pressure acting on the valve piston holds the valve closed, but as soon as the pressure is relieved through the exhaust ports, the valve 265 is opened by its spring. The steam remaining in the cylinder is then swept out by the piston until the latter overruns the passage leading to the auxiliary exhaust valve, and compression begins. The compression pressure, assisted by the sub- sequent admission of live steam, then closes the valve. Steam dashpots are used to make these valves quiet in operation. The piston rod stuffing box is also worthy of notice. All the details of this engine, including the auxiliary exhaust valves, have proved very satisfactory in service. ' The action of the exhaust valves, as HSv.H Mean position of eccentric rod Fig. 19. well as the effect of too large an exhaust port area, is distinctly noticeable in the indicator cards of Fig. 28. The steam consumption is low, as is shown by the follow- ing test results: Date Feb. 25, 1914 Duration of test . min 420 Feed water used, total kg 1424 per hour 203.42 ,, ,, ,, ,, ,, and per 1 sqm heating surface 24.1 Temperature in feed water tank, mean C 25.5 Goal used, total kg 400 ,, ,, per hour 57.14 ,, ,. ,, ,, & per 1 sqm grate area, mean . . ,, 168 266 Fig. 20. Fig. 21. 267 Fig. 2' Fig. 25. Fig. 26. 268 Fig 27. Ash, clinkers, etc., total kg 46.5 % of coal used % 11.6 Steam pressure, kg/sqcm gage . , 10 Temperature of air at fire door C 22 ,, ,, gases in smoke box, mean ,, 540 Draft in smoke box, mean mm of water 10.2 Evaporation per kg coal kg 3.56 Steam consumption, total ,, 1424 ,, per hour, mean ,, 203.42 Revolutions per minute 249.5 Brake load kg 93 Lever arm of brake load mm 560 Indicated HP 21.60 Brake HP 18.13 Mechanical efficiency 0.84 Coal consumption per B HP-hour kg 3.15 Steam BHP- 11.22 Steam JHP- 9.40 The more than ample exhaust port area of this and of other cylinders designed by the author finally led him to the calculation of the exhaust and inlet areas, which was given in chapter I, 3 a, in dealing with throttling losses. The great 269 influence of the back pressure of the exhaust and of the lead on the port area was thus recognized. A non-condensing engine requires a much smaller port area than a condensing engine. In the design of a non- condensing cylinder with an Fig. 28. exhaust lead of 25 to 30%, the exhaust belt shrinks to two narrow slits formed as nozzles as shown in Fig. 31, to which are connected the two exhaust pipes. These appear surprisingly small, but have sufficient area (Fig. 32). The large ex- haust lead gives a shorter length of com- pression, and therewith a smaller clea- rance volume, so that the auxiliary exhaust valves of Fig. 29 and 30 may be dis- pensed with, especially if the parts are so dimensioned that a suction effect with a consequent reduction of back pressure is obtained. The proportions are chosen in such a way that the toe of the diagram becomes rounded off instead of being sharp like that of Fig. 28. The piston and cylinder thus become considerably shorter, and the cylinder is supported on a single foot only. The joint between cylinder and cover is also arranged in a better manner, and the accessibility of the valves is improved by the use of a slipper type of crosshead. A better draft in the firebox is obtained by the ex- Fig. 29. pansion and diffusion of the exhaust, and finally all the good constructional features of the engine previously described are retained, especially the valve gear and valve arrangement. 270 Fig. 30. Fig. 31. 271 Fig. 32. 272 Fig. 33 shows a four cylinder V type una-flow engine for automotive purposes designed by the Stumpf Una-Flow Engine Company, Inc., of Syracuse, N. Y., three of which have been built in different sizes. The cylinders are cast in pairs, which are arranged at 90 with one another. The two cranks are set at 180 and two connecting rods work side by side on the same crank pin, the cylinder blocks being displaced axially for this purpose. The cylinder bore is 85 mm (3 3 / 8 ") and the stroke 95 mm (3%"-). The single-beat valves are operated by two cam shafts which are movable endways. These cam shafts are provided with a neutral cam, and cams for 9%, 25% and 80% forward cut-off, and one for 80% cut-off for reversing. All the working parts are enclosed. 273 V. The Una-Flow Marine Engine. In recent marine practice, serious endeavors have been made to introduce superheating and to employ balanced lift or poppet valves for steam distribution. The una-flow engine is especially adapted to meet this modern tendency, since it is well suited for such conditions. It is also apparent that the advantages which balanced poppet valves have over slide or piston valves are much enhanced when exhaust valves are dispensed with altogether, as is the case with una-flow engines. At the same time this avoids the undesirable complication of the valve gear, which has been the great stumbling block in the introduction of balanced valves in marine engines. The simplification associated with the una-flow engine is of especial advantage when superheated steam is to be used, since it is particularly suited for this. Owing to the unequal distribution of superheat in the case of multi-stage engines, many difficulties have arisen in practice, chiefly in connection with high pressure cylinders. Such troubles are not likely to occur in the una- flow engine, because the superheat benefits the whole working cycle, despite the fact that the latter extends into the saturated region. The increase in reliability consequent thereon is of particular importance from a marine standpoint. In the case of the first few una-flow marine engines which were constructed, the decision to employ this type was governed principally by the fact that the problem of introducing superheated steam into marine practice could be solved in the simplest and surest manner by its adoption. Since that time, other experiments have shown that the una-flow engine is also well adapted for use with saturated steam, and it should therefore satisfy the natural conservatism of those ship owners and engineers who still regard the introduction of superheating with much scepticism. Such engineers always refer to the absolute necessity of thorough cylinder lubrication which is indispensible with high superheats, and which endangers the safe and efficient operation of the boiler. This is especially the case in multi-stage engines, where the most diffi- cult working conditions occur in the high pressure cylinder, which must be espe- cially well lubricated. On the other hand when saturated steam is employed, cylinder lubrication is commonly dispensed with altogether, or else oil is fed very sparingly and mostly at the beginning and end of a trip. Such a practice would be better justified in a una-flow engine working with saturated steam, where moreover, the balanced inlet valves do not require lubrication. The first una-flow marine engine, intended for a steam trawler, was built by J. Frerichs & Co., A. G., of Osterholz-Scharmbeck. This engine is of 450 BHP, with two cranks at 90, so that there is a gain of space for fish storage purposes corresponding to that occupied by one of the cylinders of a triple expansion engine, which had so far been the type usually employed for this purpose. Saeuber- Slumvf, The una-flow steam engine. 18 274 lich's patent valve gear was used, which gives up to 80% maximum cut-off for maneuvering in addition to ample valve lifts at normal cut-offs. The engine works with highly superheated steam and has satisfied every demand put upon it. Fig. l. The Stettiner Maschinenbau-A.-G. Vulkan, of Stettin-Bredow, next decided to construct a una-flow engine and install it in a steamer of their own build (see Figs. 2 to 4). This engine has two cranks at 90, and a cylinder bore of 580 mm and a stroke of 600 mm. It develops 400 BHP when using steam of 12 at. gage pressure. The boilers are fitted with superheaters, and a mixing tube is provided so that saturated steam may be mixed with superheated steam so as to obtain 275 a fairly wide range of working superheats. A Klug type of valve gear is employed, in which the motion of the end of the eccentric rod is communicated to the hori- zontal valves in the cylinder heads by means of a curved rod and a cam and roller mechanism. The gear was designed for a maximum cut-off of only 26% so as to obtain large opening of the valves at early cut-off. In order to provide for starting and maneuvering, an auxiliary piston valve giving a maximum cut-off of 90% Fig. 2. is mounted on* the exhaust belt, and is operated from a second pin, which in this particular case coincides with the point of suspension of the arm of the eccentric. This valve at the same time controls a set of auxiliary jexhaust ports to permit of relieving the compression when starting up with no vacuum in the condenser, since the air pump is directly driven from the engine. This type of auxiliary valve gear is shown diagrammatically in Fig. 5, which represents the arrangement em- ployed on a una-flow marine engine installed in the steamship "Strassburg" owned by the Hamburg- American Line. It should be noted that the pilot valve which admits live steam to the auxiliary piston valve, as well as the cylinder valves which control the connections from the latter to the ends of the working cylinder, are actuated automatically by the valve gear, so that no extra manual operation is required for maneuvering. When the gear is in either of its outermost positions 18* 276 for ahead or astern running, the pilot valve and cylinder valves are opened, while in intermediate positions of the gear all these valves are closed. When maneuvering, it is only necessary to turn the reversing wheel until the engine responds. If the main gear does not start the engine, then the auxiliary gear will come into action. As soon as the engine begins to turn over, the gear is brought back to the normal running cut-off of 10%. In this position the auxiliary valve gear is completely cut out. The elimination of hand-operated valves thus greatly simplifies maneuvering. Fig. 3. The air pump and auxiliaries are directly driven from the main engine, and it is for this reason that the auxiliary piston valve is also arranged to relieve the compression when starting and maneuvering. The condenser is incorporated in the rear columns in the customary manner. The entire valve gear is mounted at the front of the engine where all parts are acces- sible. As shown in Fig. 6 the crank cheeks are formed as eccentrics. The engine is designed to work ordinarily with superheated steam of a temperature of only 250 G and the ends of the cylinder barrels are therefore steam jacketed, in addi- 277 Fig. 4. tion to the heads. The cylinders are bolted together along their exhaust belts, where the temperature is lowest; the rigidity of the w r hole structure is therefore considerably increased without appreciable changes of alignment due to expansion. Fig. 7 shows the steam valve to- gether w r ith its valve bonnet and cam mechanism. In Fig, 8 is shown an out- line view of the cargo steamer "Vulkan", which was fitted with the engine just described. The Hamburg- American Line also decided to fit two una-flow engines to the twin-screw steamer "Strassburg", which plies between Hamburg and Co- logne. This boat, shown in Fig. 9, was built in the yards of Gebriider Sachsen- berg A.-G., of Deutz near Cologne. Eaeh propeller shaft is driven by a two-cylin- der vertical una-flow engine, the cylin- ders of which have a bore of 440 mm and a stroke of 450 mm. Each engine develops 250 I HP at 175 r. p. m. when using steam of 12 at. pressure, at a tem- perature of 325 C. As will be seen from the photograph reproduced in Fig. 10, and from the drawings given in Fig. 11, these engines are built on the same lines as the "Vulkan" engine just described. They are likewise fitted with the Klug type of valve gear with auxiliary gear as described above. On account of the high degree of superheat, the steam jackets on the ends of the cylinder barrel were omitted. The firm of Burmeister & Wain, of Copenhagen, also decided to introduce the una-flow engine on two single-screw steamers ordered by the United Steam- ship Co., of Copenhagen. Each engine is of 1000 BHP and has three cylinders. The Klug type of valve gear is employed, but the auxiliary mechanism is omitted since with three cranks at 120 the maximum cut-off necessary for maneuvering is only about 40%. With this longest cut-off the inlet valves still give sufficient opening at the normal cut-off of 10%. Each cylinder has independent steam and condenser connections, and it thus becomes possible to cut out any one in case of need. By using a longer cut-off in the two remaining cylinders, the full running power may then be obtained. Fig. 5. 279 Fig. 6. Fig. 7- 280 Half side views of these engines are shown in Figs. 12 and 13. These photo- graphs clearly show the compact arrangement of the Klug gear at the front end of the engines, the three operating rods being arranged to rock three telescopic shafts which transmit the motion to the valves of the individual cylinders. The Fig. 8. Fig. 9. latter have a bore of 635 mm, with a stroke of 915 mm, the speed being 84 r p. m. The cylinders are designed to give a very compact assembly, so that a total saving of 1.75 m in the length of the engine results. This arrangement has the advantage of giving a better static balance of the reciprocating parts, so that the cut-offs at both head and crank ends of each cylinder may be made equal. This results in very smooth running. The consideration of dynamic balance is unimportaut at the low speed in question. The air pump is directly driven from the main engine, and a small auxiliary pump is provided for creating a vacuum before starting 281 the engine. This may also be accomplished by using an ejector or by opening the blow-off cocks. All the cylinders are provided with bleeder valves, the steam withdrawn being used for heating the feed water. Fig. 10. The connection between the exhaust belts of the cylinders and the condenser is especially worthy of notice. Each cylinder has an individual connection, but since the exhaust belts are all interconnected, there is always ample area of pas- sage from each cylinder to the condenser. 282 Despite the high superheat used, there is no special provision for cylinder lubrication. This is due to the fact that with single-stage expansion, even with high superheats, the cycle extends into the saturated region, so that the average temperature of the working surfaces is low. The cylinders are lubricated sparingly only at the beginning and end of a trip, while in ordinary running they are not oiled at all. For this reason it was not thought necessary to provide an oil separator. One test, which was not merely an exhibition, but a test under actual working conditions, showed a coal consumption of 0.6 kg/I HP-hour. The coal used was Newcastle coal having a calorific value of 7300 cal/kg. The steam pressure was 11 at. and the temperature 220 C. The steam used by the auxiliary machinery was included in figuring the above result. It should be noted that no forced draft or means of preheating the air is provided, and that the ends of the cylinder barrels are not jacketed. The boilers are fitted with Jorgensen patent superheaters which have proved very reliable in service. The steamer proved most satisfactory to both the purchasers and builders, and it was a pleasant surprise to both parties to find that the guaranteed speed Fig. 11. 283 Fig. 12. 284 Fig. 13. 285 was obtained with 800 HP instead of 1000 HP as mentioned in the specification. The excellent indicator cards taken during the trial trip are shown in Fig. 14. Fig. 15 shows a design of a una-flow marine engine embodying a Walschaert valve gear, which has been so successful on locomotives. In Fig. 16 is shown the assembly of the compound engine of the steamer "Wera", owned by the Orient Co., of Petrograd. On account of the excellent steam consumption results obtainable with the una-flow engine, this company decided to replace the compound engine by a two cylinder una-flow machine having a cylinder bore of 600 mm and a stroke of 711 mm. Superheating had been tried experimentally on this ship, but the old engine had proved unsuitable for use with it. For this reason it was decided to change over to una-flow cylin- ders, the design being shown in Figs. 17 and 18. The steamer is fitted with two engines, each of which is capable of developing 500 HP at a maximum speed of 125 r. p. m. The ends of the cylinder barrels are steam jacketed, so that the engine is suitable for working with saturated steam if the occasion should arise. In marine practice there is also need for a valve gear which will give proper valve lifts for normal cut-offs without having excessive movement for cut-offs of 70 to 80%. From this point of view was developed the valve gear shown in Fig. 19. In all the common valve and reversing gears the angle of advance as well as the throw of the resultant eccentric are varied, with the effect that for small cut-offs the throw of the eccentric is also small, as is the valve opening. All parts of the engine, however, have to be proportioned to accommodate the maximum eccentric throw. Obviously, the valve opening for small cut-offs could be mate- rially improved if instead of changing the eccentricity and angle of advance, the former is kept constant and only the latter is changed. This method of course necessitates a change of the lap lines as is shown in Fig. 22, in which the left hand diagram is drawn for valve gears of the Klug or Walschaert type while the right hand diagram is for a gear incorporating this new principle. The right hand dia- gram is drawn so that the constant throw of the eccentric is equal to the maxi- mum throw in the case of the left hand diagram. The valve opening "A" for maximum cut-off is equal in the two cases. For normal cut-off, however, the valve opening "B" of the new type of valve gear is seen to be considerably larger than that given by the standard valve gears. In the design shown in Fig. 19, a bevel gear is keyed to the end of the crank shaft and meshes with a pinion mounted in a rotatable housing, which in turn drives another bevel gear similar to the first, but in the opposite direction. The latter is keyed to a sleeve together with four eccentrics, each of which operates one of the valves of the two cylinders. These eccentrics are connected to rocking levers mounted on an eccentric spindle which is geared to, and is turned simul- taneously with the gear housing. This gear housing is turned by means of the reversing wheel, and in order to turn the eccentrics and the eccentric rocker spindle through the same angle, the gear ratio, on account of the differential action, must be two to one. The main eccentrics as well as the eccentric rocker spindle are moved in the same direction and in such a relationship that with one of the cranks in its dead center the corresponding cam roller remains stationary, thereby keeping 286 288 o ^* si 289 the point of opening constant. When the main eccentrics are thrown from full ahead to full astern, the eccentric rocker spindle is moved through the same angle. The resultant eccentric curve for this gear is therefore not a straight line but a circle drawn about the center of the shaft. The throw of the eccentric is always the same and the lap is altered in proportion to the angle of advance. When the arms of the rocking levers have the ratio 1:1, the eccentricity of the spindle must be one-half that of the main eccentrics. The eccentric rods are very short in order to compensate for the angularity of the connecting rods. In addition, the rocking levers are proportioned to give a later cut-off at the crank ends of the cylinders and an earlier cut-off at the head ends for forward running, so as to eliminate the effect of the weight of the reciprocating parts. The motion of the valves is derived from that of the rocking levers by means of reach rods and cam and roller mechanisms. The rollers of the head and crank end bonnets are arranged above the cams, and the eccentrics are set so that the rocking levers for the same cylinder move in opposite directions. An interesting modification of this gear is shown in Fig. 20. In this case the eccentrics operate the four valves of the two cylinders through telescopic shafts. The latter are mounted on a long spindle mounted eccentrically in two bea- rings arranged on the exhaust belt. This spindle and the bevel gear housing are interconnected by a vertical spindle and two worm gears in such a manner that the center of the telescopic shafts is swung through the same angle as the main eccen- trics. When the latter are moved from the full ahead to the full astern position, the center of the telescopic shafts is displaced by the same angle. The throw of the main eccentrics and the eccentricity of the spindle carrying the telescopic shafts are so proportioned that the lap of the inlet valves is changed in conformity with the alteration in the angle of advance of the eccentrics. In this case, as in the other form of this valve gear described above, no auxiliary gear is required, since cut-offs up to 85% are easily obtainable. A diagrammatic outline of this gear is shown in Fig. 21. In Fig. 23 are shown a side elevation and plan of a una-flow marine engine for a paddle steamer plying on the river Volga in Russia. This engine has a cylinder bore of 600 mm, a stroke of 800 mm and develops 180 BHP at 26 r. p. m. Stumpf, The una-flow steam engine. 19 Fig. 17. 290 Details of the cylinders and heads are given in Figs. 24 and 25. The crank end cylinder heads are tied to the main bearings by cast steel rods of square section which also serve as crosshead guides. The main bearing housings are cast in one piece with their supports. The valve gear is arranged at one side of the engine and consists of two sets of Klug type reversing motions which ope- rate the valve cam mechanisms through the medium of a pair of telescopic shafts mounted transversely on the exhaust belts. The arms of the eccentric straps, extend downwards and are connected at intermediate points by links to a yoke piece which may be adjusted by hand through screw gearing. The main valve gears are designed for a maximum cut-off of 25%, the head and crank end Fig. 18. cut-offs being equalized as far as possible. An auxiliary valve gear is driven from a pin on each of the short eccentric arms, which in this case coincides with the point of attachment of the swinging link. This auxiliary gear gives a cut-off up to 90% of the stroke, and thereby permits of easy maneuvering. The auxiliary valves are operated by means of a cross shaft and rocking levers supported at the crank end of the cylinders. The auxiliary gear is cut out by a suitable mechanism actuated by a sleeve mounted on the cross shaft, this mechanism being connected to the main valve gear yoke in such a manner that the entire control during maneuvering is effected from the main gear. 291 o> *-H tc 19* 292 In Fig. 26 is shown a modified design of this engine incorporating a gear similar to that shown in Fig. 20. The telescopic shafts are supported on a long eccentric spindle arranged on the exhaust belts on top of the cylinders. o M t'c In Fig. 27 is illustrated a two-cylinder una-flow marine engine built by the Kingsford Foundry & Machine Works, of Oswego, N. Y. This engine is fitted to 293 a tug boat working on the New York State barge canal, and has a cylinder bore of 18" with a stroke of 18". The valve gear is arranged at the front of the engine and consists of a revolving cam shaft driven from the crank shaft by two sets of spiral gears and an intermediate vertical shaft. The cam shaft is enclosed in a housing bolted to the exhaust belt, and operates the ver- tical valves in the cylinder heads by means of roller levers and tappets. A set of five cams is provided for each cylinder, which give cut-offs of 75% astern, neutral, 75% ahead for maneu- vering, full load ahead, and half load ahead. The cam shaft is shifted bodily endways by le- vers actuated by a steam cylinder, the valve of which is controlled by a hand lever having a follow-up motion. Equal cut-offs at the head and crank ends of the cylinders are obtained by placing the roller for the head end valve 4% ahead of its normal 180 position, the average difference in crank angle for equal cut-offs being 9. The tappet of the head end valve has a clearance sufficient to make 4% of the cam inoperative, so that the total time of opening of the head end valve corresponds to the cam angle minus twice the angle of offset, or 9. The cams as well as the rollers are beveled off to facilitate the endwise movement of the cam shaft. This en- gine has proved thoroughly reliable in operation. Fig. 21. Fig. 22. 294 In Figs. 28, 29 and 30 is shown a four cylinder single-acting una-flow marine engine having the condenser incorporated in the frame. This engine was built by the firm of Karl Schmid of Landsberg for the steamer "Koriolan", and has a cylinder bore of 470 mm, a stroke of 350 mm, and develops 400 HP at 250 r. p. m. For better balancing, the cranks of each pair of adjacent cylinders are set at 180, and the pairs of cranks are in turn set at 90. A separate crosshead guide is pro- vided, apart from the cylinder bore, and the piston is accordingly of stepped con- struction. Any water of condensation dripping from the pistons is thus kept away from the lubricating oil in the crank pit. The cranks have scoops formed upon them which, dip into the oil and deliver it to the crank pins. A further advantage claimed by the builders for this type of construction is the lower working tempe- rature of the crosshead piston and its removal from the hot cylinder walls. This type has proved very reliable, but the same may be said of the straight piston design shown in Fig. 31, if a different method of lubrication is employed. The Fig. 24. Fig. 25. valves are located centrally in the heads and are operated by cams through bell crank levers mounted on eccentric pivots by means of which the cut-off may be changed or the valves made inoperative (see Fig. 30). Corresponding to the posi- tions of the hand lever, the cam has separate steps for 60% cut-off astern, neutral, 60%, 20%, 10% and 5% cut-off ahead. The cam shaft runs at half engine speed, for which reason it is provided with two sets of cam profiles for each cut-off. By means of the eccentric adjustment the rollers may be moved horizontally and be brought into proper engagement with the cams or swung clear of them. The cam shaft is shifted by means of a handwheel and rack and pinion. Each cylinder is provided with a separate stop valve. The general construction of the engine and some of the details of the valve gear recall those of a marine oil engine. The surface condenser is incorporated in the frame, and this arrangement results in a considerable saving of space and a reduction of the back pressure, although it is probably only suitable for small engines. This machine has given excellent satisfaction. The four cylinder marine engine shown in Fig. 31 is of similar design, but in this case straight pistons are used. The valve gear shaft is also driven by spiral Fig. 23. 295 Fig. 27. gearing, but in contrast to that of the engine just described, it runs at double the speed of the crank shaft. Single-beat high lift valves are arranged centrally in the cylinder heads and are operated by Lentz cam mechanisms by means of rocking levers mounted on eccentric pins, these levers being actuated by eccentrics on 296 the main valve gear shaft. The latter is driven through a differential gear, the housing of which may be turned by a worm and hand wheel for altering the phase relation to the shaft, whereby the angle of advance and consequently the cut-off is changed and the engine may be reversed. The necessary lap of the valve cams Fig. 28. is obtained by communicating the rotative motion of the differential gear housing to the eccentrics on which the rocking levers are mounted, in such a way that the angle of advance of the latter always corresponds to that of the main eccen- trics. The whole of the valve gear is enclosed in a housing. In consequence of the double speed of the valve gear shaft, the angle of cut-off is doubled and the valve 297 lift quadrupled, which thus permits of the use of a very small single-beat valve. The latter is unaffected by pressure and temperature changes and will therefore remain permanently tight. The clearance space and harmful surfaces are also materially reduced. This design has been developed in connection with the single- acting vertical two cylinder engine shown in Fig. 47 of Ch. II, 1, p. 165, under the heading of stationary engines. (See also the following chapter.) In Fig. 32 is shown a design of a single-acting vertical six cylinder marine engine developed by the Stumpf Una-Flow Engine Company, Inc., of Syracuse, N. Y., in which straight pistons are also used. The double-beat valves are arranged horizontally in the heads and are operated by bell crank levers actuated by tapered cams mounted on a cam shaft running at engine speed. The valves are easily removable from the opposite side of the cylinders, and each cylinder head is detachable with the valve in place without disconnecting the gear, so that the piston is easily accessible. The cylinders are all bolted together at their exhaust Fig. 29. belts, where the temperature is lowest, and are tied to the cast steel base plate by substantial bolts with tubular distance pieces. Cast steel side members are arranged to take the side thrust of the piston and also permit of the attachment of side plates for enclosing the driving parts. The cam shaft with its tapered cams may be moved bodily endways by means of a hand wheel and worm gear, whereby the cut-off may be changed and the engine reversed. The six cylinders have a bore of 22" with a stroke of 24" and develop 5000 HP at 250 r. p. m. All problems concerning manufacture and operation have been satisfactorily solved in this design. In all single-acting engines of this kind the greatest care must be exercised in the design and manufacture of the piston and rings, which must make a va- cuum-tight joint with the cylinder. Air leakage past the piston into the vacuum touches the engine at a vulnerable spot. In the case of high powered engines, if it is desired to use the Schlick balan- cing arrangement, this can be carried out with a four or six crank una-flow engine. A four cylinder engine of this kind is shown in Fig. 33. Hollow pistons can be 298 made very light, as was mentioned in dealing with locomotive details. Such light pistons might be employed for the outside units, while the necessary additional weight can be easily added in the cavities of the pistons of the middle cylinders without alteration of any other parts. The effect of the inertia forces of the reciprocating masses upon the loads on the driving parts is more favorable in large una-flow engines than in the modern triple or quadruple expansion engines, where the inertia and steam pressures are Fig. 30. additive in the latter part of the stroke. The actual maximum stresses occurring in regular running are much smaller in large una-flow engines. At the same time piston speeds of 5% m/sec may be employed. Investigation proves that for cut-offs later than normal, the load distribu- tion on the driving parts of a una-flow engine is more even, while for early cut-off and at low speeds of revolution the multi-stage engine is better in this respect. For small and medium sizes and low speeds, the three cylinder arrangement has many advantages, such as better load distribution on the driving parts, more uniform torque and a smaller shaft diameter. 299 With three, four, or more cylinders, the una-flow marine engine offers a great reserve of power. Since each cylinder and set of driving parts forms a complete unit in itself, any one or more of these units may be disconnected if requiring repairs, while the cut-off in the remaining ones may be increased to make up for the loss of the one out of action. For instance, it is possible to cut out two units of the four cylinder engine just described, and to increase the cut-off in the re- maining ones from 10 to 20% to make up for the difference. Fig. 31. In comparison with that of the multi-stage engine, the reversing gear of the una-flow marine engine is much more simple and reliable. In the latter the pro- cess of reversing only applies to the inlet valves. Since in this type of engine there are no intermediate receiver pressures to be taken into account when reversing, and as the compression is always constant, the difficulties caused by excessive compression pressures in the first cylinders of multi-stage engines are avoided. Reversing takes place much more smoothly, especially since the balanced inlet valves offer very little resistance. Consequently the valve gear parts are subject to very little wear, as is proved also by experience. 300 In a quadruple expansion engine the diagram factor of the indicator card shown in Fig. 34 may be taken as an average of 55 to 60%. The remainder is Fig. 32. lost by throttling in the valves and pipes, and by condensation losses. On the other hand, the diagram factor of the indicator card (Fig. 35) of a una-flow engine may reach 80% with a good vacuum, i. e., a difference of 20 to 25% in favor of 301 302 the una-flow engine. This explains in part the essentially smaller dimensions of a una-flow cylinder in comparison with the low pressure cylinder of a multi-stage engine. This is also to some extent the reason for the fact that the steam con- sumption of a una-flow engine is not greater than that of a quadruple expansion engine of equal power, both for saturated and superheated steam. - Fig. 34. Fig. 35. By distributing the steam flow to several cylinders, smaller inlet valve dimensions are obtained, in contrast to the bulky valves necessary with multi-stage engines, in which the total working steam has to pass from one cylinder to the next in series. Another valuable feature of the una-flow marine engine is the small number of spare parts required, since each of them may be used with, any one of the cylinder units. 303 VI. The Una-Flow Engine with Single-Beat Valves and Double or Triple-Speed Lay Shaft. The usual types of valve gears with fixed lap necessarily give very small valve lifts at early cut-offs. P'or instance, at 10% cut-off the valve opening a' is only 0.065 r (see Fig. 1). In order to obtain the necessary valve area when slide valve gears are used, a large travel and considerable lap must be provided and this results in large friction losses and leakage. In poppet valve gears very steep cams become necessary. In order to alter the cut-off, the resultant eccentricity has to be changed; and since cut-offs up to 50% must be provided for in many cases, the eccentric travel for early cut-offs is short and therefore the resultant valve lift is small, i. e., the inadequate effect of the above measures is partly or wholly nullified by the necessity for a long range of cut-off. A solution of the pro- blem on this basis is im- possible, since the power necessary to lift the valve through the required height in a given time cannot be applied to it in this way. Instead of Fig. 1. trying to accomplish the work with small valve lifts and large forces, it would be better to use large valve lifts and small forces. If therefore the lay shaft is arranged to run at double the engine speed, then the period of valve opening extends through twice the angle a arid the opening increases from a' to a, as shown in Fig. 1. Since a = 2r sin 2 , therefore a = 4 a' approximately within fairly wide limits. By doubling the lay shaft speed, the opening of the valve is thus quadrupled. In order to prevent the valve from opening twice during one revolution of the engine, a cut-out eccentric is interposed in the valve gear, running at engine speed, which increases the useful stroke of the main eccentric and makes every second stroke of the latter inoperative. In Fig. 2 is shown a Zeuner diagram for such a valve gear, in which the two outstrokes of the resultant eccentric are shown in full lines while the instrokes are dashed. Every alternate outstroke 304 23 produces a valve opening of 19 mm while the following stroke falls short of the lap line by 4 mm. The constructional simplicity of this gear is evident from Fig. 3. The double-speed lay shaft carries a shaft governor which varies the throw and angle of advance of the main eccentric in the usual manner. The latter ope- rates the cam lever in the valve bonnet indirectly through a double armed lever, pivoted on one of the cut- out eccentrics which are forged in one piece with their shaft and revolve at engine speed. The position of the gear shown in Fig. 3 corresponds to the lifting stroke, in which the cut- out eccentric magnifies the motion of the main eccen- tric. After the crank has turned through 180, the main eccentric is again in the same position, but the cut-out eccentric has also turned through 180 and thus counteracts the motion of the main eccentric with the effect that the valve re- mains closed. Figs. 3 and 4 respectively show the gear and the valve of a una-flow engine having a cylinder bore of 400 mm, a stroke of 500mm, and running at 150 r. p. m. The main di- mensions of the valve gear are given in the following table: - Main Eccentric Shaft (Double speed) Cut-out Eccentric Shaft Crank End Head End Throw of eccentric . . . . . . mm 37 14 12 Angle of advance . . . dpi? 30 60 90 Lao . mm 30 Opening a mm 20 17 Maximum cut-off . . . . . / n 21 24-5 The cut-out eccentric shaft is driven from the lay shaft by a train of spur gears; and where auxiliary exhaust valves are employed they may be operated from the former. Since the governor runs at double the speed of that of an ordi- nary lay shaft engine, its regulating force is quadrupled, and it will therefore hardly be affected by any valve gear reaction. The ratio between minimum and maximum eccentric travel of the shifting eccentric of an ordinary valve gear may be designed to give a maximum cut-off of 75%. If the same ratio is used for the main eccentric which runs at double speed, the corresponding crank angle is 305 equivalent to a maximum cut-off of only 25%. This may be improved upon by a suitable change in the angle of advance of the cut-out eccentric, and the maximum cut-off may thus be increased to about 35%. For ordinary stationary engines a maximum cut-off of 25% would seem to be sufficient, since with a normal cut-off of 10% at rated load an overload of more than double this amount may be carried. In special cases the cut-off may be Fig. 3. materially increased by running the main eccentric shaft at engine speed and the cut-out eccentric shaft at double speed (Fig. 11). In this manner it becomes possible to reach a maximum cut-off of 70%, but in this case the valve lift will be only doubled instead of quadrupled. Apart from the increase in the weight of the valves, the forces necessary to accelerate them will consequently be doubled. Since these forces only amount to a part of the total valve gear reaction, the gover- nor, now revolving only at engine speed, should still be able to handle them satis- factorily, provided that there is sufficient frictional resistance in the shifting eccentric. A design comprising a secondary eccentric rotatably mounted upon a primary eccentric is suitable for this purpose. Stumpf, The una-flow steam engine. 20 306 The most satisfactory arrangement, however, is to operate the governor shaft at twice the engine speed, since this quadruples the regulating forces as well as the valve lift. The high lift obtainable in this manner allows of the use of a single- beat valve, the diameter of which need only be one half that of the equivalent double- beat valve. Such a small single-beat valve is extremely light and permits of the re- duction of the clearance vo- lume to a very small amount. The compression pressure will therefore run up to a high fi- gure so that the steam pres- sure on the single-beat valve will be almost balanced at the time of opening. This con- sideration will show that a small clearance volume is ab- solutely essential to the use of single-beat valves and that Fig. 4. the latter can only be em- ployed in combination with the double-speed valve gear. A good vacuum is desirable in view of the small clearance space. It also demonstrates why previous experiments with single-beat valves were bound to fail. Figs. 4 and 5 show the great simplicity of the single- beat valve as compared with the equivalent double-beat valve, both figures being Fig. 5. 307 drawn to the same scale. Fig. 6 shows the plain character of the cylinder head castings, as well as the short and simple steam passages and the small amount of clearance volume and harmful surfaces. The reduction of surface loss, volume loss and leakage, as well as the losses due to throttling may be expected to improve the steam consumption by 0.4 to 0.5 kg/HP-hour. The single-beat valve has only half the diameter of the double-beat valve. Despite the small dimensions of the valve, the pressure drop during admission will be less on account of the direct flow of steam with the least possible changes of area and direction, and the well rounded corners. A large nozzle, in general, has less friction losses than a small nozzle, since the friction of the walls is relatively less in comparison with the quantity of steam passing through it. For this reason the friction losses of the single-beat valve with its more compact steam jet must be less than those of the double-beat valve Fig. 6. where the flow is split up. The even profile of the cross-section of approach to the valve seat and the following diffusor-like enlargement of the steam passage will also cause a gradual increase in kinetic energy, with a subsequent change of the same into pressure energy, so that the total amount of work changed into heat due to friction will be small. The reduction of throttling losses also benefits the governor action, since the pressure difference at the valve during the latter part of admission constitutes the major part of the load to be handled by the governor. In comparison with this the forces necessary to lift the valve are small, mainly because the clearance space is filled with steam at a pressure almost equal to that of the live steam, so that an infinitely small lift of the valve is sufficient to allow the pressures to equalize fully. On the other hand, the increasing pressure difference at the valve during closing imposes an increasing load upon the governor, if it is not checked by frictional resistance in the mechanism. Since the forces on the valve depend 90* 308 upon the lift, the diameter, and the pressure difference, there will be a best dia- meter for which the loads on the valve become a minimum. A further part of the valve load is caused by the valve spring, which should therefore not be made heavier than necessary. The single-beat valve, on account of its light weight, also improves conditions considerably, since the forces neces- sary to accelerate it are smaller than those of the equivalent double-beat valve. A valve spring calculation has already been given in Chapter I, 5, but the method there employed is not applicable in this case, since the combined motions of two eccentrics revolving at different speeds have to be dealt with. A sketch of the valve gear mechanism is shown in Fig. 7 and corresponds to the arrangement shown in Fig. 3. Noteworthy is the gentle rise of the cam profile L, which in this o / case has a lifting radius of 35 ; - = 18 mm. The first step is to draw up the valve lift curve. For this purpose the move- ment of the whole gear is determined point by point for crank angles of 10 each. This corresponds to an angle of 20 at the main eccentric on account of the double speed of the latter. The paths of the points B and D are obtained in this manner. In the present case it will be found that when the crank has turned through about 60, the center K of the roller is again in the same position as at the start, and the investigation will therefore be restricted to a crank angle of from to 6(R The valve lift h for any position of the gear is the distance, measured parallel to the valve stem center line, between the curve L and an arc described with the radius J K. The curve L is drawn through the center of the roller equidistant to the 309 310 cam profile. The valve lifts thus found are plotted in Fig. 8, giving the curve marked h. The valve velocities are determined essentially in the same manner as before, being calculated from the angular velocities of the shafts and the instantaneous lever arms r 1? r 2 , /- 3 , r 4 , r 5 . The latter are obtained from the full size drawing. The velocity of the rod DH is combined of the two velocities imparted to it by the two eccentrics. First assuming the cut-out shaft to be stationary, then the velocity of the rod DH is v 2 = co , where co 2 is the angular velocity of the main eccentric shaft and r 1? r 2 , and r 3 are the lever arms. The second component of the velocity of the rod DH, which is produced by the cut-out eccentric, may Fig. 8. be found by assuming the point B to be held stationary. The instantaneous effec- tive lever arm r 6 of the cut-out eccentric O^C is the distance of the point 0-^ from the line CG which bisects the included angle between the lines CE and CF drawn parallel to the rods A B and DH. Very approximately, v l r. It should be noted that r 6 changes its sign between the crank angles and 60. The real velocity of the rod DH is v 5 = v + v 2 . The velocity of the valve is v = v 5 311 In the practical application of this method it will be found that increments of the crank angle of 10 are too coarse to permit of accurate determination of the valve movement during the lifting period. It is advisable to determine the rapidly varying lever arm r 5 for every 2 mm of roller travel, as is shown at the right in Fig. 8. It will also be found best to continue the r 5 line back to the zero ordinate as shown dashed. The relation between the roller travel and crank angle is next determined, giving the curve a at the right of Fig. 8 and the calculated velocities v are then plotted in the diagram at the left. The points of change of curvature of curve r 5 and of the velocity curve v in this case correspond to a crank angle of 5. The imaginary extension of the r 5 curve to the zero ordinate permits of the continuation of the velocity curve to zero crank angle, as indicated by a dashed line, and thus facilitates the location of a tangent. Fig. 9. The tangent T-T drawn at the steepest part of the velocity curve gives the maximum retardation of the valve to be taken care of by the spring. In the present case, with the engine running at 150 r. p. m., a crank angle of 10 corresponds to 60 10 1 : - -r^r = sec. From the diagram the change of velocity for a crank angle lou ot)U yu of 10 is found to be 0,56 m/sec, and therefore the acceleration = 0.56 X 90 = 50.4 m/sec 2 . The corresponding force required to accelerate a valve having 50.4 a weight of G kg is P = G - - = 5.14 G. 9.81 The calculation of the valve spring is carried out as shown in Chapter I, 5, but it should be noted that for crank angles of 5 and 55 60, the inertia, spring pressure and pressure on the valve due to throttling act in the same direc- tion, and are opposed only by the steam pressure on the valve stem area. The friction is assumed to be balanced by the weight. The retardation during the 312 closing period of the valve must therefore be also calculated by means of tangents, and in order to keep it small, the lifting radius of the cam should be made rather large. In Fig. 9 is shown a una-flow cylinder with auxiliary exhaust valves placed near the ends of the cylinder barrel, both exhaust and steam valves being of the Fig. 10. single-beat type. The exhaust valves are opened after the steam pressure is relieved by the piston uncovering the main exhaust ports and are closed after the exhaust valve ports are overrun by the piston. 313 The operation of the auxiliary exhaust valves is thus simplified, and the single- beat form becomes permissible. They may be operated by an eccentric placed 90 ahead of, or behind the main crank. In Fig. 10 the exhaust eccentric is shown mounted on the cut-out shaft, and in Fig. 1 1 on the main lay shaft, both shafts running at engine Fig. 11. speed. The arrangement shown in Fig. 10, with a double-speed lay shaft and single- speed cut-out shaft, will give a range of cut-off up to 35%, while that of Fig. 11, with a single speed lay shaft and double-speed cut-out shaft, has a range up to 70%. 314 Fig. 12. 315 An interesting design of the self-contained type is shown in Fig. 12, where the crank shaft is used as lay shaft, the shifting eccentric being controlled by a flywheel governor. The auxiliary eccentric shaft is driven from the crank shaft at double engine speed, by spur gears placed at one side of the crank, while the exhaust eccentric is placed on the other side. The movement of the main eccentric is transmitted to both single-beat inlet valves by rocking levers as previously described, the latter being pivoted on eccentrics at 180 on the double-speed shaft, thus magnifying the useful movement. The lower end of the second lever is moved by the lower end of the first one by a pin with bushing and sleeve. The single-beat exhaust valves are operated directly by the exhaust ec- centric. The gear shown in Fig. 11 with its long range of cut-off is suitable for engines intended for ordinary driving purposes, while that shown in Fig. 10, having a more limited range of cut-off, is more useful for those driving pumps and compressors. 316 Summary. At the beginning of this book the different losses of the steam engine were analyzed. The question which arises at the end is, how is the engine to be designed in order to have a minimum total of all these losses ? Such an engine in the form of a una-flow engine with single-beat valves is presented in Fig. 6 of Ch. VI, p. 307. The losses in the steam engine are: 1. Surface loss. 2. Volume loss. 3. Friction loss. 4. Throttling loss. 5. Leakage loss. 6. Loss due to radiation and convection. 7. Loss due to incomplete expansion. It was demonstrated that the una-flow engine with single-beat valves as shown in Fig. 6 of Chapter VI has the smallest surface loss. In the first place the extent of the harmful surfaces is extremely small. The additional harmful surface consi- sting of the short and narrow steam inlet passage amounts to only about 5% of the smallest theoretical harmful surface, i. e. twice the area of the cylinder bore. In vertical engines, the additional harmful surface will be still smaller, as is evi- dent from the two and four cylinder single-acting engines shown in Fig. 47, Ch. II, 1, p. 165, and Fig. 31, Ch. V, p. 299. Furthermore, the harmful surfaces may in this case be easily machined and thereby still further reduced. The part of the harmful surface which needs jacketing the most, namely, the area of the cylinder head, is exposed to the heating action of the highly superheated live steam in the best possible manner. The extremely small clearance volume pro- duces a high compression with a terminal compression temperature of about 900 C. The harmful surface is therefore thermally prepared for steam admission in the best possible 'manner. The surface loss must consequently be very small. It was also demonstrated that the una-flow engine with single-beat valves has the smallest volume loss. There is no better way to reduce the volume loss than by keeping down the clearance volume to a minimum. The clearance volume of the engine shown in Fig. 6 of Chapter VI is only 1%, which reduces to about %% f r larger engines and to about %% in the case of the vertical engine with single-beat valves. The long compression also assists in further reducing the volume loss. For these reasons the latter will be extremely small despite the use of single stage expansion. With the small clearance volumes just given, and for the usual range of pressure drop, the critical back pressure of this type of engine, i. e. the 317 back pressure, below which the steam consumption increases, is far below anything that can be reached with even the best condensing equipment. It was also shown that the una-flow engine with single-beat valves, assuming the same lubrication and operating conditions, has less friction losses than the equivalent counterflow tandem engine, since there is only -one piston instead of two, one piston rod packing instead of three, and two valves instead of eight. It was proved also that the throttling loss of the una-flow engine with single- beat valves is smaller than that of any other steam engine. In the first place, the throttling losses occurring between the cylinders of multi-stage engines are eli- minated altogether. The piston-controlled exhaust permits of a sufficient port area even with very small exhaust lead, so that the pressures between engine cylinder and condenser may equalize fully. The throttling loss at the toe of the diagram is therefore reduced to a minimum, and the indirect loss due to thrott- ling, consisting of a loss of diagram area along the compression line, is eliminated. The early cut-offs employed result in small throttling losses in any case and these are still further reduced by the use of single-beat valves which produce a com- pact stream instead of the split-up stream of the ordinary double-beat valve, and furthermore permit a fairly correct nozzle diffusor action to be obtained. It was also shown that the leakage losses of the una-flow engine with single- beat valves must be very small. In the first place the number of points of pos- sible leakage is reduced to a minimum. While on the one hand the ordinary tandem counterflow engine has three piston rod packings, eight valve stems, two piston seals and sixteen valve seats, in the una-flow engine with single-beat valves these are reduced to one piston rod packing, two valve stems,' one piston seal and two valve seats. Piston rod stuffing boxes can be made perfectly tight by use of metallic packings of modern design. Similarly, leakage past the valve stems can be completely prevented if they are properly fitted. A self-supporting piston can be made perfectly tight if properly designed, i. e. if the outer rings do not overrun the cylinder bore, if a sufficient number of rings is employed, if they are secured against creeping, and their joints are placed at the lowest point so that the part of the piston in contact with the cylinder wall prevents the steam from reaching the joints. From Fig. 6, Chapter VI, it will also be seen that during the periods of high pressures all six rings are active in forming a seal, and later on when the pressure has fallen appreciably three rings still remain active. Finally the small high-lift single-beat valve (see Fig. 4, Chapter VI) will remain perfectly tight, since its diameter is small, there is only one seat, and the sealing pressure is proportional to the pressure difference. Here the superiority in regard to tightness of the una-flow engine with single-beat valves will find its strongest expression. Last but not least, the series arrangement of live steam space, inlet valve, piston and exhaust is also a very valuable feature of this type of engine. It should therefore be possible to attain complete tightness at all points. It was shown that the una-flow engine with single-beat valves has the smallest radiation and convection losses. Since the una-flow engine with single-beat valves possesses the smallest losses due to surface, volume, friction, throttling, leakage and radiation, there- fore very small mean effective pressures are theoretically permissible. 318 It is perfectly clear that the favorable thermal action of the surfaces of the una-flow engine with single-beat valves makes small mean effective pressures economically possible. Exchanging in rule 7 on page 42 the words "back pressure" and "mean effective pressure", then the rule reads: "For a given amount of initial pressure^ back pressure, clearance volume and length of compression, the mean effective pres- sure must be chosen in such a way as to make the change of total heat during com- pression equal to the change of total heat during expansion." If now the clearance volume of the una-flow engine with single-beat valves is about 1%, then the terminal compression pressure will very closely approach the initial pressure, and in accordance with the above rule this requires expansion to the back pressure and mean effective pressure equal to zero. No other type of engine gives such fine, sharp-cornered, no-load cards free from throttling as the una-flow engine with single-beat valves. This type of engine is therefore especially advantageous in cases where long periods of idle running are unavoidable, as for instance in rolling mill engines. The small thrott- ling losses therefore also favor low mean effective pressures. It is clear without further comment that small friction losses make small mean effective pressures permissible. Valve leakage in una-flow engines produces a certain increase in terminal pres*- sure at the end of expansion and compression and a corresponding loss in area of the diagram. Piston leakage on the other hand results in a loss of pressure at the end of expansion and compression, also with an equivalent loss of area. Both kinds of losses increase with increasing ratio of expansion or compression, or decreasing mean effective pressure. The perfectly tight steam distributing ele- ments of the una-flow engine with single-beat valves therefore make small mean effective pressures feasible. It is also obvious that small losses due to radiation and convection permit the use of low mean effective pressures. Hence the reduction to a minimum of all the six losses so far discussed makes it possible to work with low mean effective pressures. Since a low mean effective pressure is accompanied by a small loss due to incomplete expansion, this leads to the following statement: The reduction to a minimum of the first six losses has as its consequence a minimum of the seventh loss, i. e. a minimum loss due to incom- plete expansion. Small mean effective pressures, however, result in larger cylinder dimensions, therefore higher piston loads and higher first cost, the latter to a greater extent than in multi-stage engines. This finally leads to a compromise between the re- quirements for high economy and low initial cost, i. e. the use of higher mean effective pressures in practice, with a somewhat larger loss due to incomplete ex- pansion, which is on the whole greater than that of multi-stage engines. In regard to the loss due to incomplete expansion, the una-flow engine with single-beat valves therefore ranks high theoretically, but practical considerations forbid the full realization of this advantage. In the chapter on the loss due to incomplete expansion ways and means were indicated by which this loss may be considerably reduced with a simultaneous 319 increase in the mean effective pressure. This has led to the successful develop- ment of the una-flow engine with exhaust ejector action as typified by the una- flow locomotive with single-beat valves, in which the exhaust energy of one cylinder is used to create a vacuum in the second cylinder. This proves that in the case of multi-cylinder una-flow engines the loss due to incomplete expansion may be minimized, and this leaves hope that the same result may also be accomplished for the other kinds of service for which the una-flow engine is so well adapted. Finally it may be claimed that in the una-flow engine with single-beat valves and double-speed lay shaft, all the losses are a minimum except that due to incom- plete expansion, with the possibility that in the future this loss also may be reduced to a minimum. Since the first cost of a una-flow engine is usually 15% lower than that of the equivalent tandem compound engine, it would be correct to reduce the mean effective pressure of the former by such an amount that this difference in first cost is wiped out. In most cases, however, the striving after a reduction in first cost restrains the designer from availing himself of this possibility. The uni-directional flow, single stage expansion, piston-controlled exhaust and single-beat inlet valves are common features of both the new una-flow steam engine and the two stroke internal combustion engine, while the uni- directional flow is also a feature of the steam turbine. Thus a certain similarity is established in the design and performance of the two stroke internal com- bustion engine and the new una-flow steam engine, thereby proving conclusively that the latter rests upon sound principles. THIS BOOK IS DUE ON THE LAST DATE -~TIT> BELOW WILL BE ASSESSED FOR FA.UURE TO RETURN IS BOOK ON THE DATE DUE. THE PENALTY WILL INCREASE TO 5O CENTS ON THE FOURTH OAY AND TO $,.OO ON THE SEVENTH PAY OVERDUE. 01149 2- UNIVERSITY OF CALIFORNIA UBRARY