US 2496041 A
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Description (OCR text may contain errors)
hm, M, 1950 P. s. DICKEY 2,496,041
LOCOMOTIVE POWER PLANT Filed Feb. 15, 1945 2 Sheets-Sheet 1 FIG. 3
II I STEAM a )I l500psi-9OO F |Q AIR INLET FAN ELECTRIC STEAM VAPOR GENERATOR TURBINE COMPR L I J VAPOR 6 CONDENSERS 6 STEAM B I coN0 E NsER AMMONIA VAPORIZER 5. FAN
A HOT wlk D WELL E1 LIQUID/ 1-21 8 FIEcE vERI I PRESS. TEMP A LIQUID I53 80 I55 B VAPOR I53 80 631 C VAPOR 2I2 I20 649 D LIQUID 2I2 80 I55 l'mnentor PAUL 3,. DICKEY Jan. 3], IQEO P, s, DlCKEY ,496fi4fl LOCOMOTIVE POWER PLANT Filed Feb. 15, l945 2 Sheets-Sheet 2 PRODUCTS COMBUSTION FIG. 4
GENERATOR STEAM TURBINE LIQUID RECEIVER OIL COOLER Inventor PAUL S. DICKEY K A/Q l Cittomeg ELECTRIC GENERATOR Patented Jan. 31, 1950 LOCOMOTIVE POWER PLANT Paul S. Dickey, East Cleveland, Ohio, assignor to Bailey Meter Company, a corporation of Dela- TV are Application February 15, 1945, Serial No. 578,009
I 1 Claim. (Cl. 60-95) This invention relates to elastic fluid power plants and particularly to the apparatus of such plants as used in mobile service, such as for rail way locomotives.
A plant of this type should preferably utilize a steam turbine due to its greater economy and reliability, and it should also use vapor generators which will supply the turbine with high pressure and high temperature steam to obtain the greatest possible economy, from the steam cycle. The use of the high pressure and high temperature steam requires thatthe vapor generator be fed with relatively pure and oxygen free feed water, which means that it is almost essential to use a closed steam cycle so that feed water is obtained entirely from condensate from the steam turbine. It is not practical, however, for such mobile plants to carry condensing water for the steam condenser. In existing mobiles steam turbine plants an air cooled steam condenser has been used, but the results with equipment of this type are not entirely satisfactory.
A principal diificulty encountered with aircooled condensers on mobile steam power plants has been the wide variation in ambient temperature of the air encountered by the locomotive and used in condensing the steam. Such a locomotive may be subjected to air temperatures of 110 F. to 40 F. Since the air condenser, in order to obtain suficient surface, must consist of many small tubes, it is extremely difficult to arrange the condenser sothat the .water in these small tubes will not freeze .when the condenser is subjected to extremely low ambient temperatures. It has been found that the only possible way that freezing can be prevented, particularly if there is any air mixed with the condensing steam inside the condenser, is to keep a relatively high velocity throughout these small passages. Obviously, such high velocity with the large volumes of steam which must be handled requires pressure losses which limits the amount of vacuum available at the turbine outlet.
Furthermore, since the heat transfer per unit of surface is relatively low many small parallel passages must be used, and these passages being subjected to high velocity present a very serious problem of distribution of flow. It is apparent that if the flow is not properly distributed freezing will occur in one section of the condenser, and since adjacent tubes are subjected to lower air temperatures by virtue of the freezing the tendency is for adjacent tubes to likewise freeze, i. e., for the freezing to migrate across an entire condenser section.
The high velocities which must be used in air cooled condenser passages in order to prevent freezing make the problem of removing the condensate extremely dimcult, as the tendency is for water condensed to be carried along by the steam 2 rather than to drain into suitable receivers for collecting the condensate. This problem is still further complicated by virtue of the many sections required for condensing the steam and the diiierent pressures encountered in these sections to obtain the desired velocity. Even if suitable collection of condensate could be obtained in each section of the condenser it is difficult to get this condensate to fiowto one common point properly, since all of these receivers are likely to be at different pressures, locations and elevations.
The air cooled condenser also makes the problem of air removal quite difiicult, since there is no common point where condensate may be collected and deaerated as is the case in the conventional design of liquid cooled steam condensers.
Furthermore, the air cooled surfaces must be located along the exterior walls of the locomotive, usually not adjacent the turbine, thus necessitating a complicated and obstructing system of the large pipes necessary to conduct the exhaust steam to the condenser and of the smaller conclensate returns.
If freezing temperatures, or wide variations in ambient temperature, are not to be encountered then the field of travel of the mobile unit is greatly restricted. For example, trans-continental runs, encountering desert temperatures and mountain snows, must be avoided.
At the other temperature extreme, namely, continuous operation encountering high ambient temperatures, excessively large heat transfer surfaces are required for air cooling a steam condenser, with consequent increase in piping and fan power.
It is a principal object of my invention to provide improvements in apparatus and method of operationwhereby the disadvantages of known air cooled steam condensers are obviated and a higher heat cycle elficiency is obtained for any given vapor generator-vapor turbine installation.
A particular object is to provide an improved cooling system for condensing the steam ex lgausted from aturbine in mobile service where the usual arrangements of condenser cooling water are not available.
In the drawings: Fig. 1 is a diagrammatic showing of one embodiment of my invention.
wFig. 2 is a tabulation of fluid conditions in connection with Fig. 1.
Fig. 3 is a diagrammatic elevation of a portion of a locomotive.
Fig. 4 is a diagrammatic representation of anotherembodimeht' ofrrii, invention.
The'conomy of a mobile steam turbine een operating at 1500 p. s. hand-900 F total tem perature and having air cooled condensers, operating under most favorable conditions of a clean condenser and relatively low to F.) air temperatures, may be little better than that of a non-condensing steam engine cycle. If suitable vacuum could be maintained at the turbine exhaust the horsepower output of the plant would be materially increased and at a water rate approaching best stationary practice.
I have mentioned that in an attempt to avoid freezing of an air cooled condenser the economy of the total plant suffers through pressure drop velocities and other attendant factors. One known way of eliminating the freezing problem is to use a bifluid condenser which would consist of a steam condenser of the conventional type connected to the turbine exhaust and using a coolant of the non-freezing type, such as Prestone or any or" the other common coolant-s used in gasoline engines. The coolant would then flow to the air cooled heat exchanger where the heat taken up by the coolant in the steam condenser is removed by the air.
Such a system has the obvious advantage that difficulties due to freezing are eliminated, since the steam condenser can be placed in a protected zone underneath the turbine and the coolant, which is non-freezing, is the only medium subjected to low ambient air temperatures. It has the further advantage that the long largedlameter (because of low pressure and large volume) lines which must be used in the case of the air cooled steam condenser are eliminated.
However, this known system has several disadvantages which makes the scheme rather impractical. Since the system must work on a low temperature dii'ference suitable vacuum is to be maintained at the turbine exhaust, a very quantity of coolant is required. On ordinary steam condensers the quantity of cooling water at relatively low temperature 60 F.) for vacuums of the order of 28 inches of mercury average from sixty to one hundred times by it the amount of steam condensed. This me; that under the most favorable conditions a plant using 50,000 lb. of steam per hour would have to recirculate about 5,000,000 lb. of coolant per hour. Furthermore, the temperature difference must be used twice. That is, if the plant is designed to use a 20 temperature difference in the heat exchangers and an air temperature of 100 F. were encountered, a minimum coolant temperature entering the steam condenser would be 120 F. and the minimum condensate temperature would be 140 F., which corresponds to a vacuum of only 23 inches in place of the 28 inches desired.
A further disadvantage is that the heat transfer rate in the air cooled heat exchanger is low, so that extremely large surfaces would be required in fact considerably larger than would be required for the direct air cooled steam condenser.
The bifluid condenser system however does have the advantage of elimination of the freezing problem and elimination of the problem of hanling condensate and entrained air, and also the elimination of extremely large pipe lines between the turbine and air condenser, which must of necessity be located some distance away in order to arrange the surface so that the cooling air can be passed through it. Against these advantages are the distinct disadvantages of the tremendous quantity of coolant which must be pump circulated, the size of the air cooled heat exchanger, and the loss-of vacuum when higher ambient air temperatures are encountered.
My present invention provides a combination of apparatus and method of operation overcoming the mentioned disadvantages of the air cooled steam condenser or coolant steam condenser. It is in simplest terms the application of a refrigoration cycle to the condensing system of a high pressure steam power plant. Fig. 1 shows diagrammatically the principal apparatus of such a system as well as the flow piping for both the steam cycle and the refrigerant cycle. Fig. 2 is a tabulation of refrigerant values which may exist at locations indicated on Fig. 1.
In the particular example which I have taken for consideration I utilize an electrically propelled locomotive having an electric generator I driven by a steam turbine 2. Approximately 50,000 lb./hr. of steam at 1500 p. s. i. and 900 F. total temperature is supplied the steam turbine. If the turbine were operating at a back pressure of 28 inches of vacuum, such a plant should produce approximately 6500 H. P. at a water rate of '7 .5 lb. per horsepower. I will show that through the use of my invention the system will approach this output, and even after discounting the increased power required for auxiliaries there remains an increase in output horsepower of 60 92 to 70% over that which may be expected with an air cooled steam condenser as previously discussed herein.
The turbine 2 exhausts to a steam condenser 3 having a hot well 4 from which the condensate is returned as feed water to the vapor generator supplying the turbine. In this arrangement the steam condenser is located immediately adjacent the steam turbine, thus eliminating the considerable runs of very large piping previously necessary to carry the exhaust steam to conventional side wall air cooled condensers usually located on opposite sides of the locomotive, as indicated at 5 in Fig. 3. Such air cooled condensers 5 are normally provided with a centrally located exhaust fan 6 discharging through the roof of the locomotive through louvres l. The power plant of this invention is particularly adapted for installation in a locomotive of the type illustrated in Fig. 3.
For heat transfer in the steam condenser I utilize a refrigerant, preferably ammonia. While it is apparent that any of the common refrigeration cycles, such as dense air compression, vapor compression, or continuous absorption types may be used, the vapor compression system which is used extensively for refrigeration work at the present time seems to best apply to this problem, and I have chosen as an example the use of ammonia. Thus in my preferred embodiment of Fig. l the steam condenser 3 becomes an ammonia vaporizer.
While there is a wide range of refrigerants available for the range of temperatures which would be used in this application, many of these refrigerants have obvious disadvantages, such as fire hazard or corrosive action. However, it appears that ammonia, methyl ohloride and Freon all satisfy the conditions of the cycle desired and do not present any particularly severe handling problem. Only very slight modifications of an ordinary steam condenser will be necessary in order to prevent pollution of the condensate with the refrigerant in case leaky tubes should develop, but this does not present any serious difficulty. It is entirely possible that a refrigerant operating at low pressure could be used if desired, but it appears that the most economical refrigerant is one which operates at somewhat higher pressure, and thus lower specific volume.
The media generally used in compression machines, ammonia, sulphur dioxide, carbon dioxide, etc., exist only as a gas or vapor at atmospheric pressures and ordinary temperatures, but they are liquefied When compressed to a sufiiciently high pressure and cooled. The heat absorbed in reevaporating the liquid at a reduced pressure constitutes the refrigerating effect. To periodically return the refrigerating medium to its original liquid state the system must include the following parts:
1. Evaporating space wherein the liquid is evaporated, absorbing heat from its surroundings and producing the refrigerating effect.
2. A compressor in which vapor from the evaporating space is compressed and supplied the 'condenser at a terminal pressure corresponding to the temperature of the saturated vapor obtainable with the cooling effect available.
3. The condenser in which the latent heat and heat of compression is removed and the vapor is liquefied by air passing over the condenser tubes.
ammonia from a liquid receiver 8 passes through an adjustable expansion valve 9 to the tubes of the steam condenser 3 in which the ammonia liquid is vaporized.
The ammonia vapor leaving the steam condenser 3 passes to a vapor compressor l0 driven by the steam turbine. In the compressor the ammonia vapor has its pressure increased and gains a slight amount of heat of compression. The vapor at higher pressure passes to one or more heat exchangers H which are air-cooled and are, in fact, ammonia vapor condensers. Here the ammonia vapor at substantially the same elevated pressure loses its heat of vaporization plus the superheat it gained in the compressor and passes as a liquid to the liquid receiver. The tabulation Fig. 2 gives general values of pressure, temperature and heat content for the liquid or vapor ammonia which may be expected at locations A, B, C and D in the refrigerating cycle.
The system has, among others, the following advantages over the air-cooled steam condenser or the bifiuid condenser systems previously mentioned:
l. The problem of freezing in the air cooled stream condenser is definitely eliminated, as the steam condenser is here placed in a protected location adjacent the turbine and the fluidpassing through the air-cooled heat exchanger will not freeze under any ambient temperature conditions encountered.
2. The pressure loss which is encountered in the exhaust system of the air cooled steam condenser is eliminated so that good vacuum is available at th biee 3. A conventional design of steam condenser may be used which eliminates the problems of air removal and condensate drainage.
4. The design of the air cooled heat exchanger is very materially simplified as the problem of cooling the refrigerant is not complicated with the problems of maintaining vacuum, removal of noncondensibles, and similar problems inherent in the steam condenser. A possibility is present of dividing the air cooled heat exchanger into a number ofsections and placing these atmost convenient locations around the power plant. The piping for the vapor or liquid refrigerant to and from the air cooled heat exchangers is very decidedly smaller than to conduct the exhauststeam to an air cooled steam condenser.
5. This cycle takes advantage of the high coefiicient of heat transfer from a boiling or con.- densing fluid and has the advantage of providing a satisfactory temperature difference in both heat exchangers without loss of vacuum. at..the turbine exhaust. i 1 1.
6. A relatively small quantity of refrigerant must be recirculated since the latent heat of vaporization is utilized. Instead of. requiring more than one hundred pounds of coolant per pound of steam condensed, as is .the=case. of .the bifluid condenser, the system will require only two to ten pounds of refrigerant per pound' of steam condensed.
7. The smaller quantity of refrigerant required and the reduced volume of the refrigerantrequires very much smaller piping to-andfrom the heat exchangers. For example. in the .air cooled steam condenser with the turbine operating at 28 inches of vacuum,-- roughly 400,000 C. F. M. of air must be'passed tothe air cooled heat exchanger. In the bifluid condenser ap proximately 5,000,000 lbs. per hour of liquid coolant must be recirculated throughvthe air cooled heat exchanger. -In the proposed refrigeration cycle only about 3600 C. F: of ammonia vapor must be circulated through the heat exchanger:
8. Stoppage ihany portion of the air cooled heat exchanger does not seriously affect the performance of the system.
9. The oil coolers for the turbine-generator lubricating oil may be built into the refrigeration system, which is an advantage inasmuch as the latter would normally be located near the steam condenser and the oil and refrigerant piping would be quite simple.
10. Control of the refrigerating system to obtain the maximum possible vacuum with a minimum amount of compressor horsepower could be very easily accomplished with a thermally operated valve in the refrigerant line.
It will be appreciated that the various values which I have used in this description are by way of example only. I have not attempted to give a complete heat balance for any of the. cycles or systems discussed, as this appears to be unnecessary and would obviously represent only a particular set of conditions in any event. I have arbitrarily chosen to base my discussion upon a locomotive having a vapor generator capacity of approximately 50,000 lb./hr. of steam at 1500 p. s. i. and 900 F. total temperature. Ordinarily this should produce a turbine output of approximately 6500 H. P. with a steam condenser operating at '28 inches of vacuum. Allowing even as much as 500 H. P.for the exhaust fans of the air cooledheatexchangers and as high as 1000 for the aeemr ssea: mean gain in horsepower is 50-70% over a non-condensing steam engine locomotive or an air-cooled steam condenser turbine installation.
While in Fig. l I have diagrammatically indicated the vapor compressor as directly driven from the steam turbine, it will be appreciated. that this may be through the necessary speed reduction gears, but the location of the compressor will obviously be'adjacent the steam turbine and steam condenser to take advantage of the shortest possible runs of piping for the ammonia vapor. Also the compressor being driven by the main turbine increases the size of the main turbine, thus making it more efficient.
This preferred cycle takes advantage of the high coeihcient of heat transfer from the b-iling or condensing fluid, that is the heat transfer between condensing steam and boiling ammonia.
In Fig. 4 I follow the same general arrange ment as in Fig. 1, except that I have indicated that a portion of the air cooled heat exchanger surface may be located in the path of the forced draft supply for the vapor generator. The vapor generator is indicated at [2 and includes a conventional combustion chamber. It is supplied with any convenient fuel, such as oil or pu1- verized coal, and with air for combustion produced by a fan is passing air under pressure through a conduit M. In the conduit i4 is located a secondary heat exchanger l5 through which a portion of the high pressure ammonia vapor is passed for condensing the same. The heat taken from the ammonia vapor, namely, the latent heat of vaporization plus superheat due to the compressor is given oif to the combustion air, thus aiding the efficiency of the combustion process within the vapor generator H2. The heat exchangers i l comprise primary heat exchange or condensing means.
For cooling the bearing lubricating oil of the steam turbine I indicate that the oil may be passed through a coil I5 located in the liquid receiver 8. At this location the ammonia liquid (for the present example) is under pressure of approximately 212 p. s. i., a temperature of 80 F. and a B. t. u. content of 155 B. t. u. per pound. The oil temperaure being in the neighborhood of 140 F. and of only a fraction volume flow rate compared with the volume flow rate of the liquid ammonia will be cooled toward 80 F. with no very appreciable rise in temperature or B. t. u. content of the ammonia liquid. Here again the location is ideal, as the liquid receiver is preferably located with substantially nothing other than the expansion valve 9A between the liquid receiver and the inlet to the ammonia vaporizer 3. This location in immediate proximity to the steam condenser 3 and steam turbine 2 mini mizes the piping for turbine bearing oil to and from the oil cooling coil I6.
Control of the system to obtain maximum possible vacuum with the minimum amount of compressor horsepower may readily be accomplished with a thermally operated expansion valve 9A in the liquid ammonia pipe leading from the liquid receiver 8 to the ammonia vaporizer 3. A temperature sensitive bulb ll automatically actuates the expansion valve 9A. If desired, this may be a pressure responsive expansion valve sensitive to pressure within the hot well 4, which pressure is, of course, a function of temperature under these absolute pressure values of high vacuum.
In applying the refrigerating cycle to the steam condenser of such a power plant there is the further possibility of making this cycle work as a binary vapor engine, as well as producing the necessar cooling for the steam condenser. If the compressor were designed so that it would function either as a compressor driven from the main turbine, or as a turbine putting power back into the main unit, the system itself could be arranged so that it would either add to or subtract from the power of the main turbine. It might be desirable to have an ammonia vapor turbine and an ammonia vapor compressor as separate units, with the necessary valving.
If the ambient air temperature were high it would then be necessary to furnish power from the main turbine to the compressor in order to produce the necessary refrigerating action to give the desired temperature difierence both in t e steam condenser and in the air cooled heat exchanger.
If, however, the ambient temperature were low enough, then the reirigerating act.0n would not be necessary and a lower pressure could be maintained in the air cooled heat exchanger than in the steam condenser, and the action would be for the vapor generator in the steam condenser to run the compressor and furnish power back to the main unit, this vapor being condensed in the air cooled heat exchanger. A liquid pump would obviously be necessary in this case, and this pump could be connected in the liquid line and run only at times when the ambient temperature is low enough to use the compressor as a power generator. The horsepower required in the liquid pump would be quite small, since the quantities circulated are not large in liquid phase.
While I have chosen to illustrate and describe only a preferred embodiment, it will be understood that this is by way of example only. The type of refrigerant suggested, as well as the various vaiues of quantities, temperatures, pressures, etc, are given by Way of example rather than as indicative of a definite heat balance.
What I claim as new, and desire to secure by Letters Patent of the United States, is:
A turbo-electric railway locomotive comprising a steam generator having a combustion chamber, a steam turbine connected with said steam generator, a steam condenser arranged immediately adjacent said turbine and connected to receive exhaust steam therefrom, said condenser being arranged within the locomotive for protection from the exterior temperature conditions, a closed cycle volatile liquid refrigerant system comprising primary refrigerant condensing means subjected to temperature conditions at the exterior of the locomotive, secondar refrigerant condensing means disposed in the path of the draft supply to the combustion chamber for the steam generator, passage means connecting both said refrigerant condensing means in parallel with said steam condenser, a vapor compressor connected in said passage means at the intake side of said refrigerant condensing means, a liquid receiver connected in said passage means at the discharge side of said refrigerant condensing means, and an expansion valve arranged in said passage means between said liquid receiver and said steam condenser.
PAUL S. DICKEY.
(References on following page) REFERENCES CITED The following references are of record in the file of this patent:
UNITED STATES PATENTS Number Name Date Dodge Jan. 20, 1891 Murray Aug. 30, 1898 Patten July 9, 1912 Watson Oct. 30, 1934 Fisher Mar. 14, 1939 Larrecq June 18, 1940 Number Sarco Temperature Regulator, in
10 Name Date Schwarz Nov. 4, 1941 Price Aug. 25, 1942 FOREIGN PATENTS Country Date Switzerland Apr. 1, 1924 OTHER REFERENCES The Chemical Age, issue of Nov. 4, 1944, pages 447 and 448.