US2678531A - Gas turbine process with addition of steam - Google Patents

Description

May 18, 1954 B. MILLER GAS TURBINE PROCESS WITH ADDITION OF STEAM 3 Sheets-Sheei'I l Filed Feb. 21, 1951 ATTORNEY May 18, 1954 B. MILLER GAS TURBINE PROCESS WITH ADDITION oF STEAM Filed Feb. 21, 1951 3 Sheets-Sheet 2 INVENTOR ATTORNEY B. MILLER May 18, 1954 GAS TURBINE PROCESS WITH ADDITION OF STEAM Filed Feb. 21, 1951 6 Sheets-Sheet 3 ATTORNEY Patented May 18, 1954 `GAS TURBINE PROCESS WITH ADDITION or STEAM Benjamin Miller,-0zone Park, N. Y., assignor to The `Chemical Foundation, Incorporated, 1a. New York membership corporation lApplication February 21, 1951, Serial No. 212,060

Thisinvention relates to amethod of increas ing `the power :output of continuous ow Vlgas turbines.

' The `principal :purpose ofthe invention is to provide novel methods and apparatus for `producing power .by the expansion through turbines of compressed fgases :heated by internal combustion, such novel methods and apparatus being especially :adapted to `reduce the `cost per `horsepower :of `the machines.. as` compared' to those of `current types, whenfdesigned .to obtain comparafble thermal efiicienoy. l l .The typicalinternal combustion turbine cornprises essentiallythree elements. "The iirst element is a compressor, which takes in atmospheric Aair andforces it under Asubstantial supera-'omospheric fpressure "into the second element, a com `bustion :chamber `or .combustor in which the `air meets fuel :and combustion takes Lplace, Aproducing hot gases which'passto andthroughlthe third element, a turbine, in which Lexpansionof` the hot gases takes place with the generation of mechanical energy, part of which drives the lcompressor while 'the balance is available to drive a load.

The invention can be best` appreciated and A:evaluated by considering `first the operation of conventional apparatus now employed. As will Fb'e :explained hereafter, the .novel apparatus of the invention, for equal thermal eiciency pro- -duces approximately twice` as much". power as currently employed `.units of about .thefsame `size :and operating under the same `temperature and pressure conditions; this reduces the cost per horsepower 'to 40% less than that yof` comparable `:prior :art units.

In `a typical `internal combustion turbine currently employed, `an Aair 'flow rate of 90 pounds perv second is used to produce 6,140 horsepower for the vload with .a thermal eiciencyo'f 18.2 per cent. `The air enters the intake duct `at `80 E'. and 14.1 pounds per square inch absolute ip. rs. i. a), assuming an altitude `of. 1000 ft., and suffers a pressure Adrop oi 0.2 p. s. i. a. in passing through the `air washer and intake duct, reaching thecompressor at 80 'F. `and 13.9 p. rs. i. a. The axial-now compressor having an efliciency of 8B `per cent raises the pressure to "76."5 p. s. i. la. and the temperature to 470 F., absorbing 12,050 horsepower and putting 30.73

.million `B. t, `u. per hour into the. 11,192 moles i? 10 Claims. (Cl. (S0-39.05)

2 moles per hour, which is'equivalent to 35.95 million B. t.u. (lower heating value) per hour. The combustion is not perfect, and there is someheat lost to the surroundings, so that the enthalpy increase is only 83.45 million B. t. u. per hour, and the hot gases reach `the turbine at 1400 F. and r71.8 p. s. i. a., the pressure loss throughthe combustor and connectionsbeing 4.7 p.\s.i. a.`

In passing through the turbine, lthe pressure falls `to 14.2 p, s. i. a., and the temperature to 868 the eciency of the turbine being 85 per cent. The enthalpy drop is 48.11 million B. tfu. per hour, of which 1.73 million B. t. u. per hour are lost to the surroundings, 1includii'igdrect loss and frictionin the turbine andcompressor. The

balance, 46.38 million B. t. u. per hour, is-convertedinto used mechanical energy; this is 18,'190 horsepower, of which the compressor takes 12,050

fto return to `the air, Vwhile .6,140 horsepower remain for the load.

The thermal efficiency `is 18.2 per cent; the combustor loss is 2.9 per cent, and the .losses `in turbine and compressor total 2.0 per cent, while '76;9 per cent of the heatof the fuel remains `as sensible heat in the gases `which leave Vthe turbine at 868 F. and 14.1 p. s. i. a., the pressure drop through the exhaust duct being0.1 p. s. i. a.

Some of the heatin the exhaust'gases may be used. in various ways, as for heating water or generating `low-pressure steam; a part maybe `returned to the cycle by means of an indirect `heat exchanger, or regenerator, having two flow paths separated by `a diaphragm, one path being inserted between the turbine `exhaust, and the stack, 'and the other path betweenthe compresso discharge and the combustor.

The fraction of the `heat in the exhaust gases which can be returned `to the cycle by means of a regenerator depends upon the temperature at which the gases `leave the turbine, the temperature at which the air leaves the compressor, and the number of square feet of heat transfer surface in the regenerator. With a regenerator having `50,000 square feet of heat transfer surface addedto the equipment described abovesui cient heat can be transferred from the exhaust gases to the air to raise Vthe thermal efficiency to 25.0 per cent, but only at the expense of a substantial decrease in power output, from 6140 horsepower to 5550 horsepower.

The vdecrease in power output is caused chiefly by the added pressure drops which comewith `the insertion of the regenerator. The added pressure drop between the 'compressor and the combustor acts to decrease the pressure at `the turbine inlet from 71.8 p. s. i. a. to 70.5 p. s. i. a.. While the added pressure drop between the turbine exhaust and the stack acts to increase the pressure at the turbine exhaust from 14.2 p. s. i. a. to 14.6 p. s. i. a. Thus the entropy rise due to the pressure drop, which is proportional to the logarithm of the ratio of inlet pressure to exhaust pressure, is made less, and the corresponding enthalpy decrease is made less also. In this case the temperature at the turbine exhaust becomes 880 F. rather than 868 F., which decreases the turbines used power output from 18,190 horsepower to 17,600; since the air flow rate and the pressure at the compressor discharge is not changed, the power taken'by the compressor is not changed either, so that all of the decrease in turbine output is suiered by the'load.

The fuel rate is decreased from 85.95 million B. t. u. per hour to 56.61 million B. t. u. per hour,

or from 249 to 164 moles of methane per hour.

The decrease in total moles through the turbine from 11,441 Vto 11,356 is partially responsible for the decrease in power output, but the principal cause is the decrease in the ratio of inlet pressure to exhaust pressure.

The regenerator adds about per cent to the cost of the gas turbine and reduces its power output Aabout 10 per cent, so that the increase in thermal efliciency from 18.2 per cent to 25.0 percent is accomplished at the expense of an increase of about one-third in cost per horsepower of output.

By increasing the temperature of the gases at the turbine inlet both power output and thermal efliciency can be increased. However, at the present time the practical maximum temperature is 1500 F., this limit being imposed by materials of construction, particularly for the turbine rotor.

With the turbine inlet temperature limited to 1500 F., increased thermal efciency can be obtained by increasing the complexity and cost of the apparatus.

InY one such suggested method the compression is carriedY out in two stages, with an intercooler between the two stages. pression is thereby reduced for the same pressure rise, or a greater pressure rise may be obtained with the same work. The heat abstracted by the intercooler cannot in general be recovered, but a large part of it can be replaced bythe regenerator, since the temperature of the air leaving 'the compressor is less than it would otherwise have been, and the lower the temperature of the 4air leaving the compressor, the greater the quantity of heat which the regenerator can transfer to it.

In a practical operation of such a plural compressor unit, an air flow rate of 72 pounds per second is used tc produce 7,130 horsepower for the load with a thermal eiiiciency of 29.8 per cent. The air enters the intake duct at 80 F. and 13.9 p. s. i. a., the altitude being 1500l feet, and suffers a pressure drop of 0.2 p. s. i. a. in passing through the air washer and intake duct, reaching the first compressor at 80 F. and 13.7 p.V s. i. a. The 86 per cent eiiioient rst compressor raises the pressure to 42.3 p. s. i. a. and the temperature to 316 F., absorbing 5800 horsepower and putting 14.78 million B. t. u. per hour into the 8,953 moles of air.

The air then iiows through an indirect heat exchanger, or intercooler, where it gives up 13.84 million B. t. u. per hour in cooling. to 95 F. The heat is absorbed by water ilowing through the intercooler at the rate of 800 gallons per minute The work of com- 4 and being heated from F. to 105 F. The intercooler has 21,400 square feet of heat exchange surface.

In passing through the intercooler and ducts the air suiers a pressure drop of 1.4 p. s. i. a., and enters the second compressor at 40.9 p. s. i. a. and 95 F. The 86 per cent efficient second compressor raises the pressure to 131.3 p. s. i. a. and the temperature to 347 F., absorbing 6180 horsepower and putting 15.77 million B. t. u. per hour into the air. Y

From the second compressor the air flows through the regenerator to the combustor. The regenerator, which has 26,200 square feet of heat exchange surface, puts 22.23 million B. t. u. per hour into the air raising its temperature to 690 F. In the combustor the air meets fuel, supplied at the rate of 61.09. million B. t. u. (lower heating value) per hour, or 177 moles of methane per hour. There is a loss of 1.60 million B. t. u. per hour in the combustor, and a pressure drop'of 11.2 p. s. i. a. through the regenerator, the ducts, and the combustor, so the gases at the rate of 9,130 moles per hour reach the turbine at 1500 F. and 120.1 p. s. i. a.

In passing through the turbine the pressure falls to 14.4 p. s. i. a. and the temperature to 850 F., the efficiency of the turbine being per cent. The enthalpy drop is 50.24 million B. t. u. per hour, of which 1.51 million B. t. u. per hour are lost, and 48.73 million B. t. u. per hour are converted to used mechanical energy; this is 19,110 horsepower, of which the compressors take 11,980 to return to the air, leaving 7130 horsepower for the load.

From the turbine the gases pass through the regenerator and the exhaust duct to the stack, which they reach with a temperature of 475 F. and a pressure of 13.9 p. s. i. a., having given up 22.23 million B. t.' u. per hour to the air passing to the combustor, and still carrying 25.83 million B. t; u. per hour. I

Thus the total heat of 61.09 million'B. t. u. per hour supplied as fuel is accounted for as follows: 18.18 million or 29.8% to the load; 25.83 million to the'stack; 13.84 to the cooling water; 1.60 lost by the combustor; 1.51 lost by the turbineand compressor, and 0.13 lost by the regenerator and ducts.

By adding a second intercooler, a thirdv compressor, a second combustor and a second turbine-the power output per pound of air handled and the thermal eiliciency can be further increased. Such plural compressor plural combustor cycle produces 13,240 horsepower with a How rate of 72 pounds of air per second at a thermal eiiciency of 35.3 per cent. The second intercooler, like the first, has 21,400 square 'feet of surface, and it cools the air from 347 F. to 98 F., removing 15.59 million B. t. u. per hour from the air and heating 780 gallons of water per minute from 70 F. to 110 F. In passing through the second intercooler and the ducts the pressure drops 4.3 p. s. i. a., so the air enters the third compressor at 127 p. s. i. a. and 98 F. The 86 per cent efficient compressor raises the pressure to 331 p. s. i. a. and the temperature to 336 absorbing 5850 horsepower and putting 14.92 million B. t. u. per hour into the air.

From the third compressor the air iiows through the regenerator to the first combustor. The regenerator, which has 39,300 square feet of heat transfer surface, puts 37.58 million B. t. u. per hour into the air, raising its temperature to 906 F. Burning of 132 moles of methane per 'spondingly The steam may be generated at a higher pressure and temperature, or it may be superheated in an unnred superheater, or the exhaust gases may pass, through a regenerator which heats the air from the compressor and then through the unred boiler. For example, generating the steam at 320 F. instead of 310 F. would increase the thermal efficiency to 25.0 per cent, but would require more heat exchange surface, particularly in the water heater.

Comparing the cycle and apparatus of Fig. 1 with prior installations employing a single compressor with a compressed air preheater, it may be seen that the novel cycle of Fig. 1 produces 117 per cent more power at about the same thermal efficiency with the same air compressor, but passes about 34 per cent more volume through the combustor and turbine, requires the turbine to generate and transmit about 37 per cent more power, and requires about 25 per cent less heat transfer surface. The additional cost considering combustor, turbine, and heat exchange surface is about 20 per cent, so that the cost per horsepower is decreased by more than 40 per cent.

It will be noted that the boiler 9 and water lheater l of Fig. 1 have each but one path in the gas circuit. Thus the boiler and water heater can be designed for low pressure drop in the gas circuit and high pressure drop in the water-steam circuit; this improves the power output of the turbine with but a small additional load on the feed water pump, which in any event takes only a few horsepower. It will be noted also that as compared to prior installations where the maximum temperature of the heat transfer surface in the regenerator is comparable to the tempera ture at which the gases leave the turbine, in the novel system of the invention the maximum temperature of the heat transfer surface in the boiler 9 is' not much greater than the temperature of the boiling fluid.' By reason of the high allowable pressure drop in the water-steam ilow path and the lower maximum temperature, the cost of the boiler and water heater, per square foot of surface, can be substantially less than that of the regenerators used in prior art systems.

The division of the compressor into two or more stages, with intercooling between the stages and the division of the turbine into stages of expansion, with reheat between stages, can be employed to improve thermal eniciency. It has been found desirable, particularly in connection with the improved cycle, to use adiabatic cooling between compressor stages, the cooling being accomplished by the injection of liquid water which vaporizes.

The first of these improvements is illustratedI lby the flow diagram of Fig. 2. As there shown the system comprises the rst compressors 2l and second compressor 22 mounted on the shaft 23. The turbine 24 similarly is mounted on shaft 22 and drives the two compressors and load 25. As in the rst described embodiment, air is taken from an air supply 26 and is passed through the air washer 21 and thence to the rlrst compressor 2l. Air compressed in the first compressor as 4shown is passed through the adiabatic cooler and thence through the second compressor 22 to the combustor 29. The cooler 28 preferably is a spray chamber type; this is much less costly than an indirect cooler and imposes a much smaller pressure drop than an indirect cooler although it may not be quite so effective in reducing the Awork of the compressor 22.

AS in the embodiment illustrated in Fig. 1,

steam is generated in a steam generator, such as the unred boiler 30, byv vaporizingpreheated water by indirect heat exchange with hot exhaust gases from the turbine. These gases thus pass to the water heater 3l to preheat the water supplied to steam generator and are discharged to the stack.

In the combustor the air from the second compressor in admixture with steam from generator 30 meets the stream of fuel from fuel supply 32 and the products of combustion are expanded in the turbine 24.

In such a cycle the air flow rate to the -compressor is 72 lbs. per second. The cooler 28 is designed to introduce 466 moles of water per hour, and the mixture of air and water vapor enters the second compressor at 170 F. and 41.7 p. s. i. a. The second compressor raises the pressure to 133.8 p. s. i. a. and the temperature to 456 F. absorbing 7530 horsepower and introducing 19.31 million Bf t. u. per hour into the air and steam.

In the combustor the air-steam mixture from compressor 22 meets fuel flowing at the rate of 109.2 million B. t. u. (317 moles of methane) per hour, and steam from generator 30 flowing at the rate ofi-'2334 moles per hour. There is a loss of 2.72 million B. t. u. per hour, and the mixture of combustion products and added steam enter the turbine 24 at 1500 F. The pressure drop in the combustor and connections is 8.4 p. s. i. a., so that the pressure at the turbine inlet is 125.4 p. s. i. a.

The total flow through the turbine is 12,070 moles per hour, and the turbine exhaust pressure is 14.4 p. s. i. a.; the turbine einciency being 85 per cent, the turbine exhaust temperature is 822 F.

The enthalpy drop is 68.35 million B. t. u. per hour, of which 1.88 million B. t. u. per hour are lost, the remaining 66.47 million B. t. u. per hour being converted into used mechanical energy; this is 26,070 horsepower, of which the two comu pressors take 13,350 horsepower to return to the air, while 12,740 horsepower remain for the load. Thus the thermal efficiency is 29.7 percent.

From the turbine the gases pass, asexplained, through the unred boiler 39 having 17,200 square feet of heat exchange surface, the unred water heater 3l having 12,000 square feet of heat ex` change surface, and the exhaust duct to the stack, which they reach at a pressure of 13.9 p. s. i. a. and a temperature of 293 F.

The boiler 30 generates at 3507F. and 134.6 p. s. i. a. the 2334 moles of steam per hour which are added to the circuit at the combustor, using 38.07 million B. t. u. per hour and cooling'the gases from 822 F. to 420 F. The water heater provides the boiler feed, heating 2334 moles of water per hour from F. to 350 F., and using 11.43 million B. t. u. per hour.

Comparing the cycle of Fig. 2 with the plural compressor and intercooler systems of the prior art, the improved cycle produces about 78 per cent more power at aboutV the same air rate, using a second stage compressor which requires about 22 per cent more power because it handles about 5 per cent more moles taken in at about 14 per cent higher absolute temperature, a combustor and a turbine which handle about 32 per cent more moles, a turbine which generates and transmits about 37 per cent more power and an additional spray chamber, but using about 38 per cent less heat exchange surface. The additional cost of making the second compressor, combustor amasar and: turbine slightly larger about balances the saving on` heat exchange equipment, so that the cost per `horsepower should* be decreased more than 40 per cent.

. A further imodication of the invention which is` shown. diagrammatically in Fig. 3 involves the use. of.` multiple stages of compressors and tur bines withl reheatingl of.` the gas` between the two turbine: or expansion stages. 3 `the compression is. divided lntoi` three stages comprising. the compressors 441., 42 and 43, eachy mounted on` the` shaft 44 which` shaftn isvdriven by the turbines 45v and 46?.` Power. is transmitted from` the` turbine to shaft; 641` to. drive the compressors and also the load 4T.. Asoompared with` prior art installations using an equivalent com-y bination of plural compressorsand turbines, but without utilizing thenovel cooling concept of the invention, theinstallation depicted in. Fig. 3 producesl109 per cent` more power with about the same thermall eiiciency.

Thiamodiiication, as indicated in the` drawings. utilizes `intercooling between the sequential4 compression `stages and reheating' between the `expansion stages. The operation willl have been. appreciated from the earlier discussion of` the first two` modifications. Air from air` supply 418 is washed in washer 49 and is passed to the first compressor ML. Compressed airffrom compressor 4l isthen cooled. in thetadiabaticcooler 50and is passed sequentially through the second' :com-

pressorv42 and cooler 51.` to the third compressor 43. The` mixture of compressed air and `Water. vapor from` compressor 43` as` shown is continu*- ously chargedto the iirsthcombustor 52.

As` in.` the previously described drawing, steam isfad'mitte'di to the combustion zone to `effect the desired cooling. of the combustion gases. Such steam is.A formed in a. suitable steam generator such as the unredboiler 53T, the heat for this generator being supplied as shown', by indirect heat exchange withhot gases from the second turbine 4.6'. These exhaust gases are further cooled in the water heater 514` while preheating the water fedv to boiler53'..

In.'` the rst combustor 52 the mixtureV `of air and steam meets theifuell fed fronriuel1 supply 55 `and the products. of', combustion` as shown are expanded inY turbine f 45. From'. the: rst' turbine theexpanded gases pass.` to the. second combustion 561fed1s with `fuel frorrrsupply 55; and` .the products of combustion are'- expanded in the.v secondi turbine 46. Theturbine exhaust-as notedrisxutilized togen'erate. steam'` and. preh'eat` the water `supply. to. the steam4 generator thus most effectively loweringrthe stack temperature. of the` exhaust gases.

This installation employs 72 pounds of. air: per second to `produce 27,710 `horsepowerV with 3.5.6 per .cent1 thermal'. eiciency to produce, as; previa i ouslynoted,. 109 per cent more powerwithabout theV same thermal. eiiiciency as; presently utilized installations of the equivalent 'plural` compression and; turbine stages.

Gompressors 4l and 42 and thei irst` intercool'er .5 0L are'` identical. with.' those previously dezscribed` 2,; Ysimilarly the. inietlair conditions andthe: conditions at the; `discharge of; secondicompressor.AZaraidentical withtthoseo Eig. 2.V The. second adiabatic'. cooler 5I introduces -650 moles of water` vapor per hour into the circuit concomitantly reducing-the.l temperature toi 260 F'. This second intercooler andconnections con.- sumefliapLs. i; aand thezngasesenterfthethird compressor. at` .132 gp. s.. i.V a: The. thirdieompres- As4 shown in 10 sor raises the. pressure to 396 p. s.. i. a. and the temperature to 553 F. absorbing 6430 horse.- power and putting. 21.51 million B. t. u. into the mixture. off air and steam.

The mixture of steam and air passesfrom the compressor' 43 to the. combustor 52 andis admixed with fuel introduced at the rate of 124.25 million B. t.` u. (350 moles of methane) per hour and steam entering` from generator 53 at the. rate or" 4000 moles per hour. The thermal loss in combustor 52. is 3.18 million B. t. u. per hour,xand the mixture` of combustion products and steam enters turbine i5 at l500 F. During` passage through the combustor 52 and its connections. the pressure drop4 is 25 p. s. i. a..,andihence the pressure-attire inletof turbine 45. is` 371 p; s; i. a. The gases: expand in turbine 45 down. to 69.5 p. s. i. a.; the turbine efliciency being 'per cent, thetemperature ofithe gases exhausted from tur.- bine45 is 962 F.

in the. second combustor 55 fuel is added at the rate of 73.75 million. B. t. u. (214. moles oi' methane) per hour.` The loss in the second combustor is` 2.16 million B. t. u. per hour and the gases enterthe second turbine at 1500 F; The pressure` drop` through the second combustor and connections is `5.4. p. s. i. a. and thus: the pressure of the gases at theinlet to the second turbineis 64.1 p. s. i. a.

The gases. expand turbine 46 down to 14A p. s. i. a.; with the turbine efficiency 85 per cent, the: temperature of the exhaust gases from turbine. 46. is 1021 F.

The enthalpy drop inthe turbine; 45 is 67.45 millionB.` t. u. per hour, of which 1.85 million B. t. u. per hour are lost; the enthalpy drop in turbine 45: is: 62.32 million B. t. u. per hour.. of which 1.76 million B. t. u. per hour are lost; thus the two turbines convert into usedmechanical energy 126.16 million B. t. u; perhoulr, which isie9g470 horseposer. The three compressors recycle 21.,760 horsepower to the air,: leaving 27,710 horsepower. for the'load, so that the thermal4 eiii.- ciency is 35.6 per cent.

From turbine `13.6, asf noted', the gases iiow throughftheunred boiler 53;, which has` 26,700 square feet `of heat.. exchange surface, andrgencrates at. p. s.. i. a. and 445 F. the.v 72,000 pounds of' steam perl hour added at therst combustor, and transfers 60.61 million B.. t.."u. per hour` from the gases to the-boiling liquid` While cooling the .gases to 508 F.

From' the unred boiler the gases iiow` through the' unred water heater 541 and exhaust duct to theI stack. The vwater heater` has 32,500 square feet` of heat exchange surface, and it heats from .80 F. to 445 F. the 72,000 poundgof water per hour which` the boiler converts. to steam, putting 27.07l millionB. t. u. per hourinto thewater and coolingthexgases to 260- F.

Inpassing through the boiler; Water heater and exhaust duct the gases` suier a'. pressure drop of` 0.5 p. s. i. a., `and:y reach. the stack'at 13.9` p. s. i. a.

Toproduce 1'09: per cent more. powerthan the comparable prior art multiple: compressor-mul# tiple turbinef units with the same air'owf rate andthe same thermal eiliciency, the. cycle de.-` picted inFig. 3` requires 22 per.` cent more coin'- pressor` horsepower, combustors. and turbines which handle 159 per cent more moles-of gas,.and a.. turbine which `generates and` transmits@ .'59 per centmore powenrand. two" adiabatic intercoolers; but it requires 38 per cent lessaheatexchange surface, yand none of` thefheat. exchange surface 11 has to operate at high temperature in the cycle of Fig. 3, while part of the heat exchange surface in the comparable prior art cycle must operate at temperatures which are high for ordinary steel.

It will be appreciated that various mechanical modifications may be made without departing from the spirit of the invention. For ex-k ample, each turbine may be divided into two parts, each with its own shaft, in series, and these may operate at different speeds; or the stream of gas passing to a combustor may be split into two or more streams, each of which may then pass to a separate combustor and turbine. Also, while, as described, it is preferred to mix the added steam with the combustion products before the turbine inlet, it may be arranged that combustion products are supplied to some of the turbinenozzles and steam to the others, so that the mixing takes place in the turbine.

The usual methods of controlling the speed and load may be used with the apparatus of this invention, but in addition the quantity of steam added may be decreased at part load; when the quantity of steam added is decreased, the pressureand temperature at which thesteam is added. may be increased, thereby improving the f thermal eiciency at part load.

The water used in the adiabatic intercoolers is preferably free from non-volatile matter.; however, in some cases water containing non-volatiles in solution may be used, provided an excess of water is introduced, and the unevaporated excess in allowed to flow to waste carrying the non-volatiles away with it. For marine application the boiler may be fed with sea water, in which case there must be suilcient excess feed to carry away the non-volatiles.

The heatV remaining in the gases leaving the water heater may be used for heating water, or for comfort heating, or `for. preparing distilled water in evaporators, which distilled water may then be used for the adiabatic coolers or the boiler.

The applications of the invention described above have been for power outputs of thousands of horsepower. The invention may also be used for lower power outputs, with proportionately greater benefits. The axial-flow compressor is not so. suitable for low iiow rates, such as would be needed for a power output of less thanone thousand horsepower, as is the centrifugal com.- pressor, but the efficiency of the centrifugal compressor is substantially lower. Since in the improved process of the present invention the power .taken by the compressor is only about half of that generated by the turbine when the efficiency of the compressor is 86 per cent, whereas in the conventional gas turbine the 86 per cent efficient compressor takes about two thirds of the generated power, the present invention makes it practical to operate with a compressor having only-60 Vper cent efficiency. Thus the internal combustion turbine is made practical for poweroutputs of the orderof 200 horsepower, with a thermal efciency of 15 per cent.

The improvements obtained by the present invention require for full exploitation the use of about 10 pounds of water per poundof fuel, `which could be a limiting feature in some cases where water is scarce or for mobile applications where the water must be carried. In such cases partial advantage may be. obtained by using a smaller quantity of water.

It is particularly to be observed that the fundamental concept of the invention is the use ofsteam as a coolant and the generation vof at least part of theV steam with heat which could not otherwise be made available for conversion into mechanical energy anyway. In the earlier pe-l riod of the development of the gas turbine many suggestions were advanced as to driving the turbine by a mixture of gas and steam. These sug` gestions were advanced at a time when the gas turbine then known had difficulty driving its own compressor. In the modern successful gas tur-A bine the criteria are maximum simplicity in design and maximum reliability in operation.

While the conversion of water intosteam under pressure has the advantage that it requires nov mechanical energy, exceptfor the little demanded for the feed water pump, this must be paid for with latent heat and a steam power plant cannot work at high thermal efficiency except by operating at high pressure. On the other hand, the modern successful gas turbine operates at moderate pressure and high temperature. Hence, in the past, substituting steam for air as a cooling medium could only be justified when the thermal ef-Y ficiency of the gas turbine which used the lsteam was less than about 12 per cent.

Using the heat in the exhaust gas from the turbine to generate the steam improves the efciency, to be sure, but the heat in the exhaust gas can also be used to preheat the air. It makes no difference from the fuel consumption standpoint whether fuel is burned to generate steam in the combustion chamber to mix with air preheated by the exhaust gases, or fuel is burned in the combustion chamber to heat cold air which is then mixed with steam generated by heat abstracted from the exhaust gases-provided the. amount of heat recycled from the exhaust gases is the same in both cases. But the amount of heat recycled from the exhaust gases need not be the same in the two cases.

If the cost of the equipment were unimportant, the compression of the air could take place at constant temperature, or at least with substantially no increase in temperature, and then the amount of heat transferred from the exhaust gases to the Vcompressed air could be suiiicient to increase the temperature of the compressed air almost but not quite up to the temperature at which the exhaust gases leave the turbine; in such a case the temperature of the exhaust gases would be reduced almost to atmospheric.

However, the cost of the equipment is important. It is therefore not economic to attempt to obtain very low final compressor discharge temperature, nor to increase the temperature. of the compressed air by transfer from the exhaust gases beyond a temperature still substantially lower than that at which the exhaust leaves lthe turbine. In the examples given of gasturbne processes which are representative of the present state of the art, the lowest temperature at which the exhaust gases leave the regenerator is 475 F. In this case the amount of heat carried to the stack by the exhaust gases is over 40 per cent of that supplied as fuel. This is heatwhich theoretically could be recovered, but practically can not be, in the currently employed process.

By the present invention, however, a substantial fraction of this heat can be recovered and put to use in the power generation cycle. This recovered heat may be used to improve the thermal eiciency, orto make up for theloss which takes place when steam is generated by heat which could have. been used for heating the amasar:

13 air. In the il'rst described embodimentI thelobjective is lower machine cost with the same ther-mal ei'ci'ency, though. the possibility of improved thermal efficiencyis` mentioned. In the second embodiment the objective is higher thermal eiciency, without excessive machine cost.

Wliilethe reduction in cost -perzhorsepower will'l be less if the quantity of'water used is less, there willbepossibleian offsetting gain in thermal eiliciency. rlhi-swgainiin thermal:` efficiency may be obtained by Vone or more ofthe expediente previously mentioned; vthat is, by generating the steam at a `higher' temperature and pressure; or by superheating it in an unflred superheater through which the exhaust gases `now to the boiler: orby passing' the exhaust gases through a regenerator; which heats vthe air frornthe compressor, and then through the boiler.

`Itwill be understood thatthe ratio of water to air which is to be `used in eachl case is determined by the form in which the `benefits of the invention areto beI obtained, provided suficient water is available. With arelatively `small ratio of' water to `air both thermal efficiency and machine cost are improved. As the ratio of water to air is increased, the cost per horsepower decreases more andmore, but the thermal efficiency increases to amaximum, and then decreases. In the examples given the water-air ratio has been chosen `to obtain low cost per horsepower while keeping the thermal efficiency aboutthe same as in the comparable cycles of .the present. stateof `the art. Still lower cost per horsepower may be obtained by employing higher .watervair ratios, but thethermal efficiency will be less.`

Inapplying the invention to the improvement ofz a particular gas turbinecycle, the temperature at...whch` the exhaust gases pass to the stack is an. important guide. Thus in the example, given in connection with one prior art installation the exhaust gases passto the stack at a temperature of 868*F..carrying 76.9 per cent of the heat of the-fuel, and lthe thermal `efficiency is 18.2 `per cent. In the example given in connection with prior methods using preheated air the thermal efficiency is 25.0 per cent, the improvement being due to the regenerator which reduces the temperature of the exhaust gases to 552 F. and returns to the process enough heat to raise the temperature of the air from 470 F. `to 800 F. In the example of the invention given in connection with Fig. 1 the thermal emciency is 24.8 per cent, the improvement over the comparable prior art installation being due to the boiler and water heater which reduce the teinperature of the exhaust gases to 254 F. In these examples, therefore, the use of the present invention to obtain a large decrease in `cost per horsepower with about the same thermal eniciency employs such a Water-air ratio as to decrease the exhaust gas temperature to 254 F. in comparison to the decrease in exhaust gas temperature to 552 F. which takes place when a regenerator is used.

It will be noted that in the representative prior method using multiple compressors with regenerative heating of air the exhaust gases pass to the stack at 475 F. while in the comparable cycle of the invention shown in Fig. 2 the exhaust gases pass tothe stack at `293 F.; also, that in the prior art using multiple compressors and multiple turbines the exhaust gases pass to the stack at 480 F. while in the comparable. cycle of Fig. 3` the exhaust gasespasdl to the stack at` 260 F.v Thus `in `all examples the present invention is applied to bringA the; temperature of the exhaust gases. to a value sub-' stantially lower than regeneratorscould, butstill substantially higher than atmospheric; l

While a larger regenerator would allow more heat to` be recovered, it. is not possiblefin any casefor a regenerator to reduce the temperature of thetexhaust gases below the temperature at which the air leaves the compressor, this: temperature being as high as 470 F. in these examples, while by using a large water-air` ratio the temperature of the exhaust gases. can be brought close to atmospheric, thereby decreasing the sensible heat loss. While the power outputincreases with the water-air ratio, the latent` heat loss increases also. It is therefore an fimportant feature of the invention to employ Tsuch a water-air ratio and so much heat transfer surface in each case as to obtain the: desired balance between thermal efficiency and cost of equipment, controlling the thermal efliciencyby the balance between latent heat loss and sensible heat loss and the equipment cost by the balance between heat transfer surface cost and rotating equipment cost. The heat transfer surface may be distributed among water heater, boiler, superheater, `and regenerator, depending upon the water-air ratio employed and the relativecost. of the various forms.

While preferred embodiments of the invention have been described, it is to be understood that these are given to illustrate the underlying principles of the invention and not as limiting its useful scope except as such limitations are imposed by the appended claims.

I claim:

l. A power production process comprising the steps of generating steam from liquid water; compressing atmospheric air to a temperature substantially superatmospheric; passing the compressed air to a zone where fuel is introduced and combustion takes place continuously, but without increase of pressure; passing the combustion products and excess air in admixture with the steam continuously through an expansion zone where mechanical energy is abstracted, in amount substantially greater than the mechanical energy absorbed in the compression step; and passing the expanded mixture in indirect heat exchange relationship with the water to generate the steam, the pressure at which the steam is generated being not less than the pressure at the entrance to the expansion zone, and the quantity of heat abstracted from the expanded mixture for recycling to the expansion zone being large enough to reduce the temperature of the expanded mixture to a value :not substantially greater than the temperature of the air at the end of the compression step whereby the heat absorbed as latent heat in the steam generation step is supplied by abstraction of sensible heat from said expanded mixture without requiring the combustion of additional fuel to supply such latent heat.

2. The process of claim l wherein the steam is superheated by heat abstracted from the expanded mixture, and the expanded mixture passes in indirect heat exchange relation with the steam through the steam superheating zone before said mixture enters the steam generation zone.

3. The process of claim 1 wherein the water is heated to the saturation temprature from a temperature substantiallylower by means of heat abstracted from the expanded mixture whereby the temperature of the expanded mixture is reduced to a value substantially lower than the temperature of the air at the end of the compression step.

4. The process of claim 1 wherein the pressure at which the steam is generated is made substantially higher than the pressure at the entranceto the expansion zone, whereby the waterair ratio is controlled to increase theV thermal efficiency.

5. The process of claim 1, wherein are comprised the steps of compressing the air in a plurality of stages and expansion takes place in a series of zones through which combustion products pass sequentially of transferring at least a substantial portion of the heat removed from the compressed air during the cooling between stages and at least a substantial portion of the sensible heat of the gases after the abstraction of mechanical energy therefrom to water to generate steam therefrom at substantially superatmospheric pressure, and adding said steam to the mixture of combustion products and excess air before the expansion. Y

'6. The process of claim 5 wherein there are employed a plurality of combustion zones, arranged sequentially, in each of which fuel is introduced, With an expansion zone between two consecutive combustion zones, and an expansion zone after the last combustion zone in the flow path.

7. The process of claim 5 wherein the expansion zone in which the steam is admixed with combustion products is not the first in the series.

8. The process of claim 1 wherein the quantity of steam admixed with the combustion products and excess air is decreased as the power required is decreased.

9. The process of claim 8 wherein the temperature of the steam before admixture is increased as the power required is decreased.

16 10. A power production process comprising the steps of continuously generating saturated steam from liquid water compressing air to a tempera- `ture substantially superatmospheric; continuously passing the compressed air to a zone where fuel is introduced and combustion takes place continuously, but Without increase of pressure, the rate of fuel introduction being so selected with respect to the rate at which compressed air enters said combustion Zone that the ratio of air to fuel in the feed tothe combustion zone is substantially greater than stoichiometric; continuously passing the combustion products and excess air in admixture with the steam through an expansion zone where mechanical energy is abstracted, in amount substantially greater than the mechanical energy absorbed in the compression step; and passing the expanded mixture in indirect heat exchange relationship with the water to generate the steam, the pressure at which the steam is generated being not less than the pressure at the entrance to the expansion zone, whereby the heat absorbed as latent heat in the steam generation step is supplied by abstraction of sensible heat from said expanded mixture without requiring thecombustion of additional fuel to supply such latent heat. Y

References Cited in the le of this patent UNITED STATES PATENTS Number Name Date 884,821 Zoelly Sept. 3, 1907 996,324 deFerranti June 27, 1911 1,400,813 Graemiger Dec. 20, 1921 1,887,001 v Zetterberg Nov. 8, 1932 1,982,664 Holzwarth Dec. 4, 1934 2,078,958 Lysholm May 4, 1937 2,244,467 Lysholrn June 3, 1941 2,469,678 Wyman May 10, 1949 2,549,819 Kane Apr. 24, `1951 FOREIGN PATENTS Number Country Y Date 243,692 Switzerland Jan, 16, 1947

Images