CA1083835A - Regenerative parallel compound dual-fluid heat engine - Google Patents

Abstract

APPLICATION FOR UNITED STATES PATENT

REGENERATIVE PARALLEL COMPOUND
DUAL FLUID HEAT ENGINE

Abstract of the Invention A regenerative, parallel-compound, dual-fluid heat engine is set forth wherein important engine parameters are specified and linked to each other in a manner which maximizes engine efficiency and throughput for an engine of this type.

Description

~ackyround o~ the Invention The present invention relat~s to heat engine and, more particularly, to a dual working fluid en~ine with improved thermal efficiency and throughput.
U.S. patent application Serial No. 534,479 entitled "Parallel-Compound Dual Fluid Heat Engine", issued as patent No.3~y7~ 6~ / on September 7, 1976 to the present assignee, describes a heat engine, which is referred to herein as a dual fluid or Cheng cycle engine, which makes use of two 10 separate working fluids. Each fluid is compressed separately, but they are combined in a single mixture for expansion and heat regeneration. This cycle essentially combines a Brayton cycle and regenerative Rankine cycle system in parallel such that operational limitations of compression ratio in the Brayton cycle, upper temperature in the Rankine cycle, and waste heat rejection in both cycles are removed. Regeneration using the Rankine cycle working fluid is another, very important, feature o this cycle.
The dual fluid engine at first glance has similarities with water injected gas turbine cycles. The following briefly describes these cycles and makes clear the distinction between them and the dual fluid cycle engine.

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8 It has long ~een ~cco~ni~ 1 tl)~t irljccti-)n o~ wate~

9 'into qas turbine powe~plants is an errective me~ns of com~ustor cooling. In ad(lition, this is a telatively !;imple ll means of powec or thcust augmentation. In ~atticull~, this 12 scheme was used in eacly ai~ccaft tu~hojct enqin~s fot tl~tust 13 augmentation ducin(3 takeoF. Tl~e similarit~ of the wat~t- .
l4 injected gas turbine cycles with dual flui~1 cycle is only in ,lS the sense,that two wocking fluids a~e used in the tu~bine 16 together.
17 Although both use the same t~o wo~king fluids, 18 water and ai~, the ope~ation and design o~ water injecte(3 gas l9 turbines and the dual fluid cycle powe~plant is co~pletely different. In water injected gas tucbines, wate~ can he 21 injected either at the front or at the exit end o~ the 22 compressor or directly into the combustoe fo~ cooling with no 23 regenerat~on of the waste,heat fcom the cycle into the water.
24 Water is particularly effective as a coolant because of its lacge latent heat of evaporation. 1]owever, since there is no regeneration, the process has a negative o~ little effect on 27 thermal efficiency.
23 An added purpose of water injection is to provide`

2 thrust oc power augmentation foc sho~t periods only. This is

3 accomplished by the increased mass ~low th~ough the tu~ine 31 or thrust nozzle. Since the engine is not designe(3 fo~
3 continuous operation with water, the amount of water that can I l - ~083835 ':

l be addcd to the cycle is l~mited by the stall cha~acte~istics 2 of the compressor.
3 In di~ect cont~ast, the dual Eluid cycle powe~plant

4 is designed for continuous ope~ation with steam c~eated by S `the ~egene~ative oE heat that would othe~wise be wasted E~om 6 the cycIe. It is impoctant to ~ecognize that the Rankine 7 cycle fluid in the dual-fluid cycle is a wo~king fluid and 8 not a coolant. As will be seen, the p~ope~ combination o~
9 cycle parameters to achieve high thelmal eEficiency in the dual-fluid cycle engine of the present invention ~esults in an increased water-to-air ratio as the design point tu~bine 12 inlet temperature is increased. In pcio~ wate~ ingested gas 13 turbine designs, increased turbine inlet tempe~atu~e always .
14 results in a reduced water-air ratio.
- Gas Turbine with Water Injection 16 and Air-Watec Heat Regenetation 17 A more recent application of water injection in gas 18 turbines is for the purpose of air pollution control. Wate~
19 is injected into the air stream after the compressot to the point of saturation. If a regenerator is used, water is 2i injected before the entrance of the heat exchange~ in the 22 proper amount ~less than 8~i~ so that the wate~ completely evaporates. The air-steam mixtu~e then ~ecove~s the exhaust 2~ heat before it enters She combustion chamber. The e~fect of the steam is to dilute the air so that the flame temperatu~e 26 in the combustion chamber is lowered. The NOX fo~mation of a 27 gas turbine is a strong function of the local flame 28 tempe~atu~e within the combustion zone, and thus, the tesul~
29 of NOX water injection is a ~eduction in the NOX level.
~0 It is known that the specific heat capacity o 31 steam is about twice that of ait. Also, the specific heat 'i ' ' ' . , .

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` 1083835 capacity of wate~ is about twice that o~ steam. Fo~ this 2 ~eason, heat recove~y in watel without ail is fa~ mo~e 3 ef~ective than heat ~ecove~y ln a steam-air mixture. In 4 addition, the same pressu~e ~atio limitation imposed on a ~egene~ative gas tu~bine applies to the ai~-steam 6 regenerative system. The optimum pressure ~atio is usually 71 about 6 to 1. Although the cycle can inclease thloughput and 8 also be slightly bette~ in eEficiency, it is fa~ less than 9 the Cheng cycle engine in both th~oughput and efEiciency.
Steam Injected Tu~bine ~leat Engine 11 Turbine heat engines designed to inject steam with 12 some degree of heat regene~ation have been attempted in the .
13 past, but with failu~e or disappointing ~esults in tecms of 14 efficiency. In fact, efficiencies have been sufficiently low that the series combined cycle engines have been more 16 attcactive and have found comme~cial utilization.
17 Several attempts have been made to improve the 18 steam injected gas tu~bine efficiency with some deg~ee of 19 heat recovery from the ~ngine exhaust. No one, howeve~, has recognized this engine system as the linking of two 21 independent thermodynamic cycles and that the checks and 22 balances of cycle pa~ameters for such an engine a~e 23 inte~locked. Thus, the combination oE the two cycles and 24 engine design parameters are unique as with any other thermodynamic cycle. It has not been ~ecognized that the 26 cycle parameters are limited to a na~row ope~ating range and 27 only in that nar~ow range can high efficiency be ~calized.
28 For example, too much steam results in a poo~ steam 2 cycle because it lacks the high pressu~e ratio oE a pu~e steam cycle. Too litt}e steam results in an engine little 3 diffe~ent from a regenerative gas turbine.

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:` ' 1 1()83835 1 Generally, in any thermodynamic heat engine cycle 2 or system, the engine operating parametets are interdependent 3 and are locked into a narrow operating range Eo~ maximum 4 throughput or efficiency. An analysis oE the pliOr alt in S gas turbines with regeneration and steam injection ihdicates 6 a failure to recognize this interdependence. If the 7 interdependence was recognized, then the nar~ow range of allowable and independent engine parameters that maximize engine efficiency was not found.
Summarv of the Invention 11 In accordance with the present invention, a 12 regenerative, parallel-compound, dual-fluid heat engine is 13¦ provided wherein important engine cycle paramete~s are 14 identified and linked together to maximize efficiency andtor throughput. These parameters include the turbine inlet 16¦ temperature, the overall cycle pressure ~atio, specific heat 17 input cate i.e., 8tu/lb of gas flow (or air-fuel ratio), and 18 the ratio of the Rankine cycle working fluid to the 8rayton 19¦ cycle working fluid.
It wili be shown, with examples of ai~ (for simplicity, 21¦ humidity of ambient air and combustion products are neglected) 22 and water as the working fluids, that the prope~ choice of 23 these pairameters foc the regenerative parallel-compound, 2,4 dual-fluid heat engine of the present invention results in a 2~ powerplant far superior in terms of thermal efficiency ~and 26 thus fuel consumption) compared to any state of the art 27 stationary powerpiant. Thermal efficiency greatcr than 52~ .
28 can be achieved using state of the a~t gas turbine component 29 technology, and efficiencies of 60~ can be ~ealized using advanced high pressure ratio and high temperature 31 technology.
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It should be emphasized that the ope~ation of an 2 engine using two wolking fluids simultaneously is not being 3 claimed as unique to this invention. Rathec, the p~opec 4 choice of cycle paramete~s o~ the unique matching o~
components' sizes to attain high efficiency ot thcoughput

6 and the operational limits of the ~egenetation pa~allel

7 compound dual-fluid cycle a~e the unique teachings oE this

8 invention to the state of'the a~t in heat engines.

9 The p~oper combination of cycle pa~amete~s fo~ the regenerative pa~allel-compound dual fluid cycle ~esults, surprisingly, in an increased propo~tion of liquid (such as 12 water) relative to gas (such as air) as tu~bine inlet .
13 temperature is increased. This is a majo~ distinction of 14 ¦ this cycle from the prio~ art in water injected gas turbine powerplants. In the past, the critical ~elationship between 16 ¦ cycle parameters to achieve high e.ficiency was not 17 ¦ recognized, and increased turbine inlet tempetature resulted 18 in reduced water-air ratios because the ai~-fuel ~atio was 19 not set in the proper proportion.
It has also been found that efficiency is ~elated 21 both to the degree of superheated temperatu~e o~ quality of 22 the regenerated steam, facts not heretofo~e known. It has been fouhd that efficiency is maximized when the steam 2~ entering the combustion chamber is superheated and~is at a maximum superheat temperature and maximum waste heat 2 recovery. This maximum temperature is limited by the 2 turbine exhaust temperature. "Deg~ees of supe~hcat" is 2 defined as the temperatu~e above the boiling tempe~atu~e of 2 ' a liquid at a given pressure. The "quality of steam" is defined as the percentage of vapo~ by mass ve~sus liquid in 3 a wet steam as they are mixed at a constant boiling ~ ,, . ' r.;..

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temperature. Thus, the compression ratio directly influences the degree of superheat or "quality". Increasing the pressure ratio reduces the degree of superheat but too high a pressure ratio places an unwarranted burden on the compression work of the Brayton cycle. This is a fine example of how the proper choice of cycle parameters influences the efficiency of this new cycle.
It has also been found that maximum efficiencies only occur when the engine parameters other than the compression ratio are within a narrow region of permissible values. Maximum efficiency must always be balanced against engine throughput considerations, and hence a practical engine may operate slightly away from the maximum superheated regenerated steam to gain some throughput. The quality of the steam does define a lower boundary for the engine operation for maximum throughput, if the cycle receives heat externally, but is also limited by stoichiometric ratio if a fuel is burnt internally.
Thus, one aspect of the present invention is defined as a dual fluid heat engine comprising: a chamber; compressor means for introducing a first gaseous working fluid into the chamber, the compressor means having a predetermined pressure ratio (CPR); means for introducing a second liquid-vapor working fluid in the form of a vapor within the chamber at a defined second/first working fluid ratio (XMIX); means for heating the first gaseous working fluid and the second working fluid in the vapor form in the chamber at a defined specific heat input rate (SMIR); turbine means responsive to the mixture of the first and ~ second working fluids for converting the engery associated with ;. - ' '- . ' :
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the mixture to mechanical energy, the temperature of the mixture entering the turbine means defining the turbine inlet temperature (TIT); counterflow heat exchanger means for trans-ferring residual thermal energy from the exhausted mixture of first and second working fluids to the incoming second working fluid to thereby preheat the same to a superheated vapor state prior to its introduction within the chamber, and wherein XMIX
and SHIR are selected so that for a given value of TIT, XMIX
is substantially equal to or is greater than XMIXpeak, where XMIXpeak occurs when the following conditions are both met simultaneously: the temperature of the superheated second ~; working fluid vapor is substantially maximized; and, the effec-tive temperature of the exhausted mixture of the first and second working fluids is substantially minimized.
; Another aspect of the present invention involves the method of operating the above defined apparatus at maximum ` efficiency and/or throughput for a given turbine inlet temperature including the steps of pre-heating the second worklng fluid in the heat exchanger to a superheated vapor state prior to its introduction within the chamber; and selecting XMIX an~ SHIR
so that for a given value of TIT, XMIX is substantially equal to or is greater than XMIXpeak, where XMIXpeak, occurs by ;; both: maximizing the temperature of the superheated second working fluid vapor; and minimizing the effective temperature of the exhausted mixture of the first and second working fluids.
Brief Description of the Drawings Fig. 1 is a block diagram of a dual-fluid or 28 Cheng cycle heating engine;

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Fig. 2(a) is a graphical illustration of the temperature-entropy (T-S) diagram of the two working fluids of the dual-fluid, heat engine of Fig. l; Fig. 2(b) is a graphical illustration of the parameter of effective tempera-ture in a dual-fluid engine;
Fig. 3 is a block diagram showing the relative temperature levels on both sides of the heat exchanger 8 illustrated in Fig. l;

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los3s35 l Fig. 4 illustrates the engine cycle ef~iciency of 2 a dual-fluid heat engine plotted as a function of the 3 turbine inlet tempecatures at constant comp~ession catios 4 for peak efficiency operation;
S Figs. 5 - 8 illustrate gcaphically the efficiency 61 of a dual-fluid engine plotted against the specific heat 7 input rate for 1500F - 3000F at constant comp~ession 8 ratios of 10, 20, 30, and 40, respectively;
91 . Fig. 9 is a graphical illustration depicting the interdependency of the turbine inlet temperatu~e and the ll compression ratio in a Cheng cycle heat engine;
12 Figs. 10 - 13 graphically plot the specific heat .
13¦ input rate versus the degree oE supe~heat at diffe~ent 14 values of turbine inlet tempe~atures of the regenelated steam for compres~ion tatio values of 10, 20, 30, and 40, 16 respectively in a dual-fluid heat engine;
17¦ Fig. 14 defines the ~ange of XMIX fo~ a dual-fluid 18 heat engine for opetation from maximum efficiency to high 19 throughput;
201 Figs. 15 - 18 illustrate the dependence of XMIX as 21 a function of the turbine inlet temperature, the compression 22¦ ratio, and specific heat input rate; for compression ~atios -~ 23 of 10, 20 30, and 40, respectively;
24 Fig. 19 is a graphical illustration showing the 251 useful XMIX region as a function of specific heat input rate 26I for the design and operation of a dual-fluid heat engine fo~
271 maximum efficiency to maximum throughput; .
28¦ Figs. 20 - 23 graphically illust~ate engine 291 throughput as a function of the specific heat input rate fo~
301 compression ratios of 10, 20, 30, and 40, ~espectiveIy;
31 Fig. 24 is a g~aph illustrating the na~row ~ange 32 of throughput operating regions as a function oe the ~u~bine , - 8 -.' .
';~':, , .'`' ' ' ,i , ' ~` , 108~835 ., ~ inlet tempe~atuce for a dual-fluid engine designed foc 2 opecation between maximum efficiency and maximum thcoughput;
3 Fig. 25 illustcates the influence of comp~esso 4 efficiency on overall engine ef~iciency foc a dual-fluid cycle engine:
6 Fig. 26 illustrates the influence of the heat 7 exchangee back p~essure on the overall engine efficiency of 8 a dual-fluid cycle heat engine;
Fig. 27~a) g~aphically illustrates a ~ange of compression ~atios ve~sus tucbine inlet tempecatule fo~
11 ¦ p~actical opecation of a dual-fluid cycle engine; Fig. 27tb) 12 illustcates a ~ange of ai~-fuel ~atios (and specific heat 13 input tates) as a function of the tucbine inlet temFe~atuce 14 fo~ practical ope~ation of a dual-fluid cycle heat engine;
Fig. 28 illust~ates a cange for XMIX as a function 16 of tu~bine inlet tempe~atu~e which coccesponds to the ai~-17 fuel ~atio (and specific heat input ~atios) and compcession 18 ~atios illust~ated in Figs. 27(a) and 27(b);
19 ~ Fig. 29 illust~ates the limitatlons of a conventional heat exchange~ integcated into a ~egeneLative 21 pa~allel compound dual-fluid cycle heat engine by 22 illust~ating the engine the~mal efficiency ve~sus ai~-fuel .
23 ratio; ~, 24, Fig. 30 illust~ates gcaphically the effect of the temperatures on both sides of the heat exchange~ of Fig. 1 26 as the air-fuel ratio is incceased; and 27 Fig.,31 illustrates the effect of thcee diffecent .
~28 ai~-fuel ~atios on the heat exchange~.

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::~ ' 1 Desctiption ol Ihe Plete~cd ~mbodiment 2 Figu~e 1 is a block diaglam of one emhodiment oE a 3 dual-fluid o~ Cheng heat engine 10 in acco~dance with the ¦ present invention. The engine typically uses aic as the S first working fluid. Fuel combustion with the ai~ is a 6 typical source of energy, and water is a ty,oical second 7 working fluid. Air enters a throttle 12 to cegulate the air B ptessure prior to entering a comp~esso~ 14 whe~e it is 91 adiabatically compressed. If the comp~ession ~atio of the compressor is below 12:1, the throttle 12 can also se~ve as 11 a carburetor with some of the fuel being int~oduced into the throttle as indicated by 18'. If the comp~ession ~atio oE
~31 compressor 14 is greater than 12:1 without special cooling, .
14¦ spontaneous combustion would result within the compresso~ if 15¦ an air/fuel mixture were compressed. Fo~ highe~ comp~ession 16 ~atios, the fuel must be introduced after compression at 18.
17 Compressor 14 can be of any type, but for a high 18 volume flow ~ate machine a standard axial Elow or 19 -centrifugal flow air comp~essor, eguivalent to those used in conventional gas turbine engines, is desirable.
21 From the compressor 14, the air or ai~/fuel 22 mixture enters the combustion chamber 16. Where Euel has 231 not been introduced into the ai~ flow th~ough the compresso~
24 14 or where additional fuel is desi~ed, it is int~oduced 251 directly into the combustion chamber at 18. Through 261 combustion, heat is added to the air; the combustion 271 products thus heated constitute the Ei~st wo~kinq fluid of 28¦ heat engine 10.
29 The fi~st working fluid can be heated in other ways besides combustion; for example, by sola~ ene~gy o~
31 nuclear energy in combination with a heat exchanger in place : . . , . .

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l oE the comtusto~. For tl-~ remaindet oE this desc~iption, it 2 is assumed that the ~i~st wo~king ~lui~ is heated by 3 combustion. Regardless of the sou~ce o~ heat, the amount o~
4 heat added is re~erred to herein as the SpeciEic llcat Input ~ate (SI~IR~ which is heat input per pound o~ filst wotking 6 fluid flow.
7 The combustor 16 design can be oE the conventional annular or co-annular type used in gas turbine engines 9 ¦ today. However, the region downstream oE the combustion zone must be modified to inject high p~essu~e superheated

11¦ steam in a way to promote good mixing with the combustion

12 products of air. It may be possible to employ more novel

13 I combustor designs which would utilize the steam as an

14 ¦ ejector for minimizing pressure losses. Mixing would take

15 ¦ place in a way similar to that by which diluent air mixès

16 ¦ with the primary zone combustion pcoducts in a conventional

17 combustor.

18 Water, the second working fluid, is compressed to

19 ¦ a high pLessure by pump 22. The high pressure wate~ enters heat regenerator 24 where waste exhaust heat is abs~rbed 21 ¦ from the steam/combustion product mixtu~e exhausted f~om the 22 expander 28. As will be described in greater detail 2 subsequen'tly, the water is heated to vapor. In most cases 24 the steam is superheated, however, it is possible for wet 2 steam to discharge from the heat regenerator. Because of 2 the latent heat of evaporation of wate~, much of the heat 2 absorbed by any water conve~ted to steam is absorbed at ¦
28 essentially constant temperature, i.e. boiling temperature.
2 The heated steam or steam/water mixture from the 3~1 regeneratoc then entecs the combostion chambec 16. ~o ~V ~ -11- I
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l help cool the combustion ch.~mbe~ walls, the steam can first 2 pass through wate~ jackets in the wall of the combustion 3 chamber. ~ny wct steam into o~ just aEte~ the combustion 4 chamber is rapidly evapo~ated into supe~heated steam.
Transfer of thermal energy from the heated combustion 61 products to the steam is accomplished th~ough turbulent 7 ¦ mixing of the two working fluids. The water vapor is mixed 8 with the combustion products only a~te~ combustion is 9 ¦ completed so that the steam does not quench the combustion process. The steam, however, is used to control the ll temperature of the combustion products to ~each a designed l21 turbine inlet tempe~ature, as will be desc~ibed in greate~
l3¦ detail subsequently. .
l4¦ The mixture oef the two working fluid;s then enters lS an expander or coee tuebine 26, which drives the compresso~
l6¦ 14, then entees another expander or wo~k tu~bine 2B. Tilese ~71 expandees convert the thermal energy of the two wo~king l8¦ fluids into mechanical work, to drive the comp~esso~ 14 and l9¦ to produce net work output.
Both the cor~ tu~bine 26 and the work turbine 28 2ll are conventional in the sense that they a~e typical ~eaction 22¦ tu~bine designs. However, they must be specifically 231 designedlfoe the gas mixtu~e of combustion products of air 2~1 and steam because the specific heat ratio and the ave~age 25 1 density of the gas mixtute will change depending upon the 26 ¦ mixtuee ratio. This presents no p~oblem in the ae~odynamic 27 1 design of the turbine in te~ms of flow path areas and blade .

28 ¦ profiles as long as the compression ratio, maximum inlet 29 ¦ temperature, and the Specific Heat Input ~ate a~e known.
30 ¦- There must be a careful mate~ials selection to withstand 31 high temperature; howevee, it is entirely feasible to use a ~/
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1 portion of the steam fo~ film cooling in the tu~bine to 2 replace the comp~essor bleed ai~ as used in conventional 3 high temperatu~e gas turbines.
4 ~leat exchange~ 24 used to ~egene~ate the waste heat from the cycle is a countec flow heat exchange~ The 6 gas side of the heat exchanger contains the gas mixtu~e 7 ¦ which drops in temperatu~e from the powe~ tu~bine 28 8 discharge to a temperature at oc above the satulation 9 ¦ temperature of the water in the gas mixtu~e. This 10¦ saturation temperature is a function of the pa~tial p~essùte 11¦ of the steam in the gas mixture. On the liguid side of the 12 heat exchanger, water under pressure is heated from 13 approximately ambient temperatu~e to boiling tempe~atu~e 14 ¦ where it is evaporated. Wet steam then fo~ms in the water-15 ¦ steam mixture region, and if sufficient heat transfer from 16 the gas mixtuce exists super-heated steam results at the 17 ¦ heat exchanger discharge.
18 ¦ From the heat exchanger 24 the gas mixtu~e is 19 discharged into the condenser 30. The wacer vapor in the gas mixture is at or somewhat highe~ than saturation 21 ¦ temperature. The condenser 30 is a typical wate~-vapor 22 ¦ design such as that now in use at some geothermal powe~
23 installa~ions to condense steam to water. The gas mixture 24 ¦ is ducted to a closed vessel which has water being injected from the top from shower heads. The water droplets absorb the heat from the gas mixture and the water in the mixture 2 condenses and drops to the bottom of the vessel with the 2 cooling water. The remaining gas is vented from the top of 30 ¦ the vessel the atmosphele.

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1 ~Eter some pucirication at the wate~ system 20, 2 the pcoper amount of watec is mete~ed and pumped to the 3 liquid side of the heat exchanger Eo~ tegcne~ation ahead of 4 the combustor. The temaining wate~ is passed th~ough a cooling tower or other cooling means and then ceused in the 6 condenser.
7 The two working fluids, watel and air products, thus follow parallel cycles with the two Eluids being mixed 9 p~ior to the expansion part of the cycle. Since the two fluids are mixed, the output of each is added togethe~, i.e.
11 compounded.
12 The heat enecgy soucce used by the dual fluid 13 engine as disclosed by this invention is not limited as to .
14 fuel type or means of heat input. Hydcocacbons~ fuel gases lS produced by conversion of coal, oc alcohols can be used. In 16 addition, as desccibed above, concent~ated sola~ ene~gy o~ a 17 nuclear reaction also could be the source of heat. However, 18 each fuel will have its own "best" set oE engine ope~ating 19 conditions and cycle parametecs. To simplify the descriptions and explanations offe~ed he~e, all cycle 21 descriptions and ope~ating pa~amete~s of the dual fluid 22 engine are glven in terms of a typical hyd~oca~bon fuel with 23 air and ~ater as the two working fluids. Extension to othe~
24 working fluids such as helium, F~eon, etc., a~nd using nuclear heat sources can be accomplished by established engineeclng p~inciples, well known to those skilled in the 28 The following subpa~agcaphs summa~ize the the~modynamic 2 cycle of the ~egenerative, dual-fluid heat engine oE the 3 pcesent invention.

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, ' , : ., ' ` ' '` 1083835 1 1. Compression oE the two fluids takes place 2 separately. Air is compressed f~om atmosphecic pressure up 3 to the maximum cycle pressure by compressor 14. Wate~ is 4 pumped at ambient temperature to a p~essure somewhat g~eater 5 than the compressor discharge air pressure.
6 2. Combustion takes place in a mixture of air and a suitable fuel in combustor 16. For these examples and illustrative figures, a hydrocarbon fuel is assumed. Water, 9 in the form of superheated steam, is then mixed with the combustion products of air. This steam is the result of 11 water being preheated by the regenecative heat exchanger 24 i2 and is at a somewhat higher pressuce than the combustion gas 13 to promote proper mixing.
14 3. The resultant mixture of combustion products of air and steam, hereafter called the gas mixture, is at a 16 specified maximum turbine inlet tempe~ature and specific 17 heat input rate (SUIR) whose units are Btu/lb of ai~/sec., 18 which dictate the combination of water-air ratio (XMIX).
19 ~Note that XMIX denotes generally the ratio of the liquid-vapor to gaseous working fluids which in this case are water and - 21 air, respectively.) SHIR can be used to determine the air-22 fuel ratio ~AFR) by the fuel types. Expansion of this gas 23 mixture~takes place in tu~bines 26 and 2a. The first or ~- 2 high temperature or core turbine 26 drives the air compressor through a connecting shaft. The second or power 26 turbine is a free turbine 28 which p~ovides the useful 27 output work.
28 4. The gas mixture discharging from the powe~
2 turbine is then passed through a counter flow regenerative heat exchanger 24. This heat exchanger in most cases uses ~-~ the otherwise rejected heat from the cycle to preheat the 3 water to steam at superheat temperatures which is then injected into the combustor 16. Thus, the heat in the `~ - 15 -'"".. : , - , ' -:.. : . , .

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iO~3~35 l cycle is "~egene~ated.~ Fo~ sta~t-up and some special applica-2 tions, after bu~ning can also be provided in the exhau~t gas 3 between the turbine exit and the heat exchangel. The gas 4 ¦ mixture on the hot side drops from the turbine dischatge tempe~atuce to an exhaust tempeeatuce with the satu~ation 6 ¦ temperature of the steam in the gas mixtu~e as a lowe~
7 ¦ limit. The steam is raised to supe~heated tempe~atures, and 8 ¦ depending upon the specified cycle pressu~es and 91 temperatuce, is at or near the maximum supe~heat tempe~ature lO¦ point.
ll Two thermodynamic limits a~e placed on the heat l2¦ exchanger: first, the maximum tempe~atu~e of the wate~
l3¦ aftet waste heat recovery cannot exceed the gas mixtu~e l4¦ temperature at the power turbine discharge. Second, the gas mixture temperature a' the location in the heat exchanget l6¦ where the water boils (saturation temperatu~e) cannot be 17 less than this water saturation tempe~ature. This is called l81 the heat exchanger "neck" and will be discussed fu~the~
l91 subsequently.
201 5. The gas mixture leaves the heat exchange~ 24 2l¦ at or above the saturation temperatu~e of the steam in the 22¦ gas mixture as determined by the partial pressure of the 231 steam, ahd it then passes through condense~ 30. In gene~al, 24 1 no condensation is desirable in the heat exchanger 24.
-25 ¦ Rather, steam condenses to water and is separated from the 26 ¦ mixture in the condenser 30. The remaining p~oducts o~ the 27 ¦ combustion of ak are exhausted to the atmosphe~e. The .
28 1 condensed water is purified, pumped to high p~essu~e and 29 ¦ tecycled to the regenerative heat exchanger.

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~C983835 1 To better illustrate this dual-~luid, pa~allel-2 compound ~egenerative p~inciple, ~eference is madc to thc 3 thermodynamic heat cycle oE Figure 2a illust~ating Eo~ each 4 of the two wo~king Eluids the tcmperatu~e-ent~opy (T S) diagram, which, as will be shown, ace coupled in pa~allel 6 during certain parts of the cycles. This chart is idealized 7 in that minor eEEiciency losses a~e not consideted. In 8 addition the two fluids are treated sepa~ately in their 9 ~rayton and Rankine cycles respective1y fo~ illustrative purposes. Although the two wo~king fluids shown in Figure 11 2a have their ohn separate cycle T-S diagrams, they are ve~y 12 much interdependent.
13 The gaseous working fluid sta~ts at state 1 and is 14 compressed to reach 2. Combustion and steam mixing takes -place to enable the t~ermodynamic state to reach 3.
Expansion together with steam brings the thermodynamic state 17 to 4. Exhaust heat is transferred to the othe~ working 18 fluid and some cooling thus, theoretically, ~etu~ning the 19 thermodynamic state back to 1.
The liquid at 5 is pumped to a pressu~e slightly 21 higher than 2, with essentially little change of tempe~ature 22 and entropy. The high pressure liquid receiving heat ene~gy ?3 from the turbine exhaust mixture is heated to boiling temperature T6. A limitation exists that at any time T6 has to be lower than T6* Tnis will be explained in a late~
2 section. Since T6 is a function of p~essu~e, the p~essure 2 and temperature relationship of the two wo~king flùids 2 begins to become explicit. The liquid is continuously 2 heated either before saturation at 7 in order to allow more -3 Rankine working fluid in the cycle or to be heated to 31 superheat temperature T7' just below T4. The steam is mixed `- 17 -V'- . .

~ - ........... . .

1(~8:3835 I with the combustion p~oducts of ai~ and Euel to ~each T8.
2 The steam tempe~ature T8 is equal to T3 of the gaseous 3 working Eluid. The expansion flom 8 to 9 takes placc with 4 the gaseous wo~king fluid, whe~e T9 equals T4.
The fo~egoing would not be possible if the steam 6 and the gaseous wo~king fluid would expand separately due to 7 thei~ diffe~ence in specific heat ~atio ~k = Cp!Cv;
8 ¦ specific heat at constant pressure/specific heat at constant 9 ¦ volume). If the gaseous working Eluid is ai~, Cpai~ is lû ¦ approximately half that of the steam, but kai~ is usually 11¦ larger than kSteam, hence air helps the steam to convert 12 mo~e heat energy to mechanical work with a sac~ifice oE
13¦ wo~k accomplished by the a k itself. Hence, the fact that 14¦ the two working fluids a~e mixed together becomes ccitical.
The exhaust heat in the steam is also t~ansEe~ed 16¦ to the incoming liquid-vapor, along the path of 9-~ lO, thus 17 making it a unigue regenerative steam cycle. The steam is 18¦ condensed out of the mixture to return to the the~modynamic 19¦ state 5. As describeo above, the two fluids a~e actually physically mixed during the expansion and heat exchange 21¦ processes. It is also important to ~ecognize that waste 221 heat from both the Brayton cycle (a~ea a) and the Rankine 231 cycle (a~ea c) is used in the regeneration process to 241 preheat the incoming water to steam only in the Rankine cycle (area b) before it is mixed with the combustion 26 ¦ products of air.
27 Prom Fig. 2a, it is obvious that the ~atio oE the .

29 two working fluids are not expressed by the T-S diag~am.
The temperature in the T-S diagram is known as the sensible temperature. This means that the tempe~atu~e can be ~, :~. ', , ~easured by a th~rmom~ter. W~th ~hl~ new cycle, an ef~ective temp-erature is defined as:
T ff = c TS + XMIX [hl (TS ) + hfg ( Sl S V Sl S
-Cp (l+XMIX) gas XMIX - Rankine fluid/Brayton fluid Ts~ sensible temperature Ts ~ Liquid boiling temperature - at partial pressure Cp - Specific heat at constant pressure hl - Liquid phase enthalpy, Btu/lb hfg - Latent heat of evaporation; Btu/lb hv ~ Superheat enthalpy; Btu/lb Ps ~ partial pressure of the va~our.
g P ~ effective g g Y
efficiency. If there is no Rankine working fluid, Teff equals T5. Then the cycle becomes essentially a simple Brayton cycle.
At a fixed compression ratio the exhaust temperature measured in terms of Ts is high, so the heat rejection rate is high. When Rankine fluid is introduced into the cycle, the effectiveness of heat re-covery makes the exhaust temperature measured interms of sensible te~mperature low. This can be seen in Figure 2b. If the upper sensible temperature (turbine inlet temperature, T3 and T8) is fixed, then the more Rankine fluid (XMIX) that is introduced, the - higher will be the upper effective temperature and at the same time, the lower sensible temperature after heat recovery is de-creased. But the effective temperature at the lower side (exhaust side) reaches a minimum (Teff i ) at a certain XMIX depends on turbine inlet sensible temperature and cycle pressure ratio.

-I~ 83835 .

l The eEfective temperature beyond Tef~ is highe~
2 due to the large amount oE latent heat o~ evapota~lon being 3 carried away by the ~ankine wotking fluid. Thelefore, at 4 that Teff . the Rankine wo~king [luid to Btayton wocking fluid ratio (XMIXpeak) is uniquely defined. Any other mixtu~e 6 ratio means that more heat than necessa~y is rejected, 7 ~esulting in a lower cycle efficiency. (XMIXpeak) is ¦ influenced by the cycle upper sensible tempe~àtuce and compression ratio. Without the ~ecognition of the 10¦ significance of Teff, the steam injection ~ate (and steam 11 p~operty) and the conditions on the heat exchanger could 12¦ only be arbitrary at best. In other words, the high 13 efficiency potential of this heat engine over the combined .
14 ¦ cycle results only with the recognition of the fo~egoing.
The uniqueness of XMIXpeak is an important element in 16 ¦ defining this new thermodynamic cycle, with all the engine 17 ¦ operating parameters related to each other.
18 Figure 3 is a diagram which shows the relative 19 temperature levels on both sides of the heat exchange~ 24.

20 The following list identifies the subsc~ipts of the va~ious -~
?1 1 temperature notations used there:
22 Mixtu~e Temperatures 2 TMIN Turbine Discharge Temperature TM ~ Neck TemPe 2 TM - l~eat Exchange~ Exit 2 OUT (Minimum value TMSAT) `` 2 TM ~ Satu~ation Temperatu~e of Steam 3 SAT in the mixture ~ - 20 ~ , - ._ . . . .
.. ?
.
... . . . . . ... - ..

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: : ;~ ' :: . .. .

~083835 . ..
1 Water/Steam Tempe~atu~es 2 TLIN Water Inlet Temperatu~e 4 TLSAT Satu~ation (Boiling) l'empe~atu~e 6 TLOUT Steam Discharge Temperatute 81 All the temperatures ~efetred to above a~e sensible 91 temperatuces.
10 `.
TemPerature Constraints TNECK ~ Minimum Temperatu~e Diffe~ential at Neck l3¦ a THoT ~ Minimum Temperatuce Differential 14 at Hot End .
16¦ This sketch points up two basic thermodynamic 17¦ limitations in the heat exchange~: first, the temperatu~e 18 ¦ of the super-heated steam on the water side cannot exceed 19 the temperature of the gas mixture at the power turbine discharge. Second, at the point in the heat exchanger 24

21 1 where the water has reached boiling temperature, the gas

22 ¦ mixture temperature cannot be less than the wate~ boiling

23 ¦ temperatyre~ This is called the "neck" of the heat

24 I exchanger. The neck can be reduced by after burning to increase

25 ¦ thoughput with some sacrifice of efficiency.

26 The two-basic independent parameters that specify

27 ¦ the cycle operating point are the turbine inlet tempe~atu~e

28 ¦ (TIT) and the compressor pressure ~atio tCPR) - also called

29 ¦ the cycle pressure ratio. In fact, once TIT is established, a range of CPR's is permitted, and selection is made based on economic consideration primarily. The remaining two ~ . . .

,~' .. ' .
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: 1 pa~ameters that must be sp~cified a~e ~he allowed speci~ic ¦ 21 heat input rate (SHIR) Btuilb of ai~/sec) or ai~-fuel ~atio ,l 3 ¦ and the steam-air tatio. These two paramete~s a~e directly 4 ¦ coupled to each other and to both CPR and TIT; the:y cannot ¦ 5 be specified independently. Only the engine components J 61 selected acco~ding to a critical choice oE these parameters 71 can produce the best efficiency. -In the conventional sraytOn cycle using a single 9 working fluid, the specific heat input rate, SHIR, ~air-fuel 10¦ ratio in the case of heating by combustion) is uniquely 11 defined once CPR and TIT are specified. However, in the 12¦ dual fluid cycle with a given TIT, the addition o~ steam in 13 the combustor requires an inc~eased SHIR (reduced air-fuel .
14 ratio) as the water rate is increased. The increased heat 15 input per pound of air at the lower air-fuel ratio is offset i 16¦ by the increased heat necessary to achieve super-heated 17¦ steam at the specified TIT. Thus, there is a wide range oE
; 18¦ combinations of SHIR and water-air ratio (X~IX) that can be 191 specified for a given CPR and TIT. These are the four key 20 cycle parameters used to describe the dual fluid cycle.
21 The selection of the design operating point foL
22 the cycle from the wide range of possible combinations of 231 these fo~r parameters is based on requirements fo~ high 241 thermal efficiency and/or high power throughput. Thermal 2sl eEficiency is a direct measure of the powerplant fuel 26 ¦ consumption for a given output of power, and thus has a 27 ¦ major effect on the operating costs oE the powe~plant.
28 ¦ Throughput is the power output measured against the size of 29 the powerplant. This size is most often related to the ai~-31 flow pumped through the compressor. Thus, throughput can be :,, . ' .
..
,, - ,...
.
, ~;, "
., .;, ~(i8~835 : I measured as power pec unit airElow. The initial cost oE the 2 powerplant is roughly inversely p~opo~tional to its 3 throughput.
4 An engine cycle Cannot be designed to achieve both maximum efficiency and maximum throughput, a fact which is 6 common to virtually all heat engine cycles. Fo~ this 7 reason, a na~row region of cycle design paramete~s is 8 described and claimed for this invention which encompasses 9 cycles having maximum efficiency and those which are a lû compromise between high efficiency and high th~oughput.
11 However, the maximum throughput design point can be chosen 12 without too much sacrifice of the efficiency. Prefe~a~ly, - 13¦ turbine inlet temperature can be maximized by steam film .
14 cooling or other methods so that throughput and efficiency can both be increased.
16¦ Ordinarily, the two independent parametets in any 17¦ cycle design are turbine inlet temperatuee ~TIT) and 18¦ compressor pressure tatio (CPR). In the heat engine of the 191 p~esent invention, the temaining patameters of inte~est a~e .
the specific heat input tate (SHIR), the ability oE the 21¦ system to absotb heat and the steam-air ratio (XMIX). Many ~- 22¦ combinations of these two patameters are theoretical1y 231 possible~ but they cannot be specified independently once - 241 TIT has been set and once the specified condition fo~
251 maximum efficiency or high thtoughput is detetmined. Small 26¦ deviations from the examples given later a~e allowed due to 27 compressor and turbine adiabatic efEiciencies and also due .
281 to the specific design limitations of the heat exchange~.
29 The specific limits and the reasons fo~ the limits are now 301 discus~ed.

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l Pe~fo~mance cu~ves of a Cheng cycle engine a~e 2 calculated based on plesent day rea1istic engine componcnts.
3 The comp~essor efEiciency is assumed to be .84, the tutbine 4 efficiency is .90, the combustion eEficiency is .99, and the pressu~e d~op in the combustion chamber is 4~. The steam 6 temperature is only allowed to ceach a level not ovec SO~F
7 below the exhaust temperatu~e of the tu~bine. This 8 assumption is made fo~ -p~actical enginee~ing ~easons tathe~
9 than the thermodynamic limit. The limitation oE the low temperature end at the exit oE the heat exchanger is made so 11 that the gas-steam mixture does not reach the dew point oE
12 the mixtu~e to avoid cocrosion in the heat exchange~. This 13 again is a practical reason. As component efficiency 14 improves in the future, the peak eficiency point will shift towards higher steam to air ratio as will be seen.
16 It should also be realized that with equipment 17 having different efficiencies the pe~fo~mance cu~ves which 18 follow will vary somewhat. ~ut, with the invention of this 19 new cycle, peak efficiency always occu~s at minimum ~eff and XMIXpeak. Hence it is not intended that the following 21 performance curves be interpreted as exact or ironclad.
22 ~ariations will exist depending upon the hardwa~e used.
23 SIn Figure 4, the cycle efficiency is plotted as a 24 function of TIT at constant CPR's for peak XMIX and efficiency. It is obvious that the ove~lapping of the 26 constant CPR curves defines the operating ~elationship - 27 between TIT and CPR. Even with the peak XMIX found, one can .
28 still see that a high CPR is needed to achieve high cycle 29 efficiencies, but their relationships a~e confined to a ve~y

30 ~ =~rlow ban In the useful regions. For exaFple, at F TIT

:.J, : ' ' ' '' ~ '"' . ' ' ~
- , ' .

I oE 1500F, no higher than crR = 10 should be uscd. ~t TIT
2 of 2000F, CPR = 20 or better is preferred. If cPn is lower 3 than 10, this woul~ ~ake the engine suEEec an unnecessa~y 4 loss of efficiency, unless teasons othec than high efEiciency ~such as the requirement o~ lighte~ weight Eoc 6 vehicular oe ai~craft uses) would be desirable. As will be 71 shown later in connection with Figu~e 9, a comp~ession ratio 8 of less than about one-third of the compre~sion ratio to 91 achieve peak efficiency ~esults in a substantial 1055 oE
10¦ efficiency, beyond that for acceptable engine operation.
11 In Figures 5-8, engine efficiency is plotted 12 against SHIR for 1500F-3000F and at constant CPR's of 10, 13 20, 30 and 40, ~espectively. It is quite evident that the 14 efficiency peaks,at certain SHIR for a given TIT and CPR due to a minimum Teff at this point. The interdependency of TIT
16 and CPR is summarized in Figure 9. Note how the thermal 17 efficiency peaks for a given TIT with CPR. The CPR and TIT
}8 optimization can be understood from the t~ade-offs between 19 higher degrees of superheat at the exit of the heat exchanger 24 and more steam. High CPR increases the boiling 21 temperature which tends to lower the turhine exhaust 22 temperature, thus lowering the deg~ee of superheat at the 23 exit of heat exchanger 24. Too low a CPR increases the 24 degree of superheat but lowers the water-to-ai~ ratio, XMIX
at peak. This can be clearly seen in Figures 10-13 which is 26 a plot of SHIR vs. the degcee of supe~heat of the 27 regenerated steam for constant values of CPR. .
28 From Figure 10, at TIT = 1500F, CPR = 10, the -29 degree of superheat of the regenerated steam at the peak is approximately 300F, In the meantime, at TIT = 2000F, CPR

31

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~ 38;~5 1 = 10, the degree oE supe~heat is approximately 650F. When 2 compared with data in Figure 11, at TIr = 20nOF, CPR - 20, 3 the deg~ee of superheat at this peak is 340F and the 4 efficiency and throughput ace both imp~oved ove~ the CPR 5 10 case. When CPR = 40; TIT = 2nO0F, in Figu~e 13, the 6 degree of superheat at the peak is 120F. The efficiency 7 again becomes less than that for the CPR = 20 case as shown 8 in Figure 11. Thus the choice of the deg~ee of superheat 9 is critical to cycle efficiency.
The parameters of CPR and TIT are related uniquely as 11 exhibited in Figure 9. Of particular inte~est is the Eact 12 that the peak efficiencies shown in Figu~es 5-8 occur when 13 the regenerated steam is at the maximum degree of supe~heat, 14 fo~ a given CPR. CPR is related to the maximum possible degree of superheat at the peak such that the degiee of 16 superheat at the peak is in the range of 250F to 650F.
17 Choice oE CPR resulting in deg~ee of supe~heat at the peak, 18 higher than 650F or lower than 250F results in poorer 19 cycle efficiency. Note that efficiency drops oEf fLom a peak efficiency at a given TIT as CPR drops off in value.
21 As an approximate boundary for ~easonably efEicient engine 22 operation, CPR should not go below one-thi~d of the value of 23 CPR at peak efficiency for a given TIT.
24 For maximum throughput consideration, the quality (percent of vapor in the wet steam) of steam leyels off ve~y 26 sharply with SHIR as shown in Figures 10, 11, 12 and 13.
27 This means that at high specific heat input rates and 28 leveling steam quality, the cycle essentially app~oaches a 29 regenerative Rankine cycle engine. At low S~IIR and high superheat, the amount of steam that can be injected becomes - 26 ~
, .
r~
.

I1~383835 -. , l so small, the cycle approdches a ~egene~ative gas turbine.
2 Only in the neighbo~hood of the peak deg~ee of supetheat Is 3 the inte~action oE the two cycles mutually beneEicial to 4 each other. This is another example oE the uniqueness oE
the Rankine to Brayton cycle working fluid ~atio of the 6 present invention.
7 For throughput conside~ation, one can always tolerate higher SHIR but only to the point whe~e va~iation 9 ¦ of steam quality versus SHIR becomes small. As a p~actical matter, once SHIR exceeds approximately twice the value oE
11 ¦ SHIR at peak efficiency, engine eEEiciency is too low fol 12 ¦ normal engine applications. This is evident E~om Figs. 10-13 13, as the superheated steam drops ~apidly as SHIR is 14 ¦ increased.
With the combination of the tesults of Figu~es 10-16 ¦ 13, the XMIX or the engine is bound within a very na~low 17 region as shown in Figure i4. This ~egion can be described ¦ by the equation along W-W:
19 XMIX = 0.0623 + 0.-'217(TIT/1500F)1-65 f~om TIT = 1500F to 3500F
21 ¦ with a +50% width.
22 ¦ The dependence of XMIX as a function of TIT, CPR
23 ¦ and SHIR can be seen in Figu~es 15, 16, 17 and 1~. The ~p~ 24 lower boundary is the limit of the highest degree of superheat using a conventional heat exchanger and 50DF below 26 the exhaust mixture tempe~atu~e. However,-due to the 27 linking of TIT and CPR tFig. 9), the useful XMIX ~egion as a :
28 function of SHIR is shown in Figure 19. Note that the lowe 29 boundary is taken as approximately the highest deg~ee of supe~heat with the best CPR fo~ a given TIT;

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1~ 835 1 Figures 20, 21, '2 and 23 describe the behavio~ of 2 the Cheng cycle engine parameters on thloughput (l~p/lb of 31 air/sec), which is inversely ptopo~tional to the cngine 41 size. Of particular interest is the c~ossover oE the throughputs. Thls indicates that one cannot a~bitrarily 61 increase SHIR o~ XMIX to gain throughput at a given TIT, 7 rather it is better to inc~ease TIT when s~In o~ XMIX is 81 increased, so that both efficiency and throughput are 9 improved. From the crossover behavior of the th~oughput, 10¦ the upper bound of XMIX for maximum througput can be 11¦ approximately defined as the maximum throughput line f~om 12¦ Figures 14 to 19.
13¦ To summarize, all the througllput ope~ating ~egions .
14¦ fall into a very narrow range as shown in Figu~e 2i. The 15¦ interconnection of al~ the engine operating parameters have 16¦ been described. For a given TIT, only a best CPR can 17 produce the peak engine efficiency with a conventional heat 18¦ exchanger. At that peak efficiency, the heat input rate and 19¦ Rankine/Brayton working fluid ratio is unique. Variations 201 are possible only because of different component 21 efficiencies used to const~uct the engine. Inc~easing XMIX
22¦ from (XMIX)peak can improve throughput with a sac~ifice of 23 efficiency, but even that is limited from the crossover 241 properties on TIT. Therefore, for a given TIT, XMIX is bounded from XMIX at peak efficiency to a larger but finite 261 value for maximum throughput. Ovec that value, the engine 271 cannot gain either throughput o~ efficicncy. It is best to 28¦ increase TIT from thereon. XMIX lower than the XMIX at the 29~ p~a~ also 1 5e5 both efE~ciency and thlooghput.

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.
.- .-. . -: . -,. . : . -:
.

` 10~ 335 1 Co~elation of a typical XMIX and a typical SIIIR
2 as a function of TIT with the best CPR operating varIatlons 3 with component efficiency choscn above can be shown in 4 Figure l9. The preferred operating region can also be represented by an equation along R-R with a width of +0.1 6 for XMIX:
7 (XMIX) = 0.178 + 0.026~ ~SHIR~400Btu/lb air)2 05 8 The R-R XMIX = (XMIX)R R - 0.1 boundary reD~esents 9 the best efficiency line or XMIXpeak line. The XMIX = XMIXR
R + 0.1 line represents approximately the lacgest value for 11 XMIX at a given TIT. That line is a recommended line for 12 engine performance and cost tradeoffs. ~Ience, it is not as 13 exact or definite as the maximum efficiency line.
14 From the operating regions shown above of the Cheng cycle engine, it is obvious that a higher TIT produces 16 better engine performance. The choice of TIT depends only 17 on the manufacture's economic conside~ations and turbine 18 cooling methods.
19 It is understood that since steam is available, it can be used as a film cooling media for the turbine and 21 nozzle rather than compressed bleed air. This further 22 reduces the work required by the compressor and also it 23 allows lower combustion chamber pressure drop. The steam 24 used can be at a low temperature, thus reducing the coolant mass flow. This steam is counted as part of the working 26 fluid.
27 This cycle engine is not vcry scnsitivc to 28 component's efficiency, as opposed, for example, to a gas 29 turbine, where the compressor efficiency is the key to a 3l good engine. The Cheng cycle engine traps and recirculates .` , ~' .
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.
.
- . . ..

.

` ~083835 I the waste heat generated ~lue to the inhecent ineEficiency in 2 the cycle. When compressor effi~iency dcops Ecom 90~ to 3 84~, for example, the ovecall thc~mal eEficiency 108s ig 4 ¦ only 0.25~. In a gas turbine such a dcop in th~tmal efficiency could mean more than a 6~ loss. The pecfocmance 6 curves of two different eEficiency compressors are compaced 7 in Figure 25 for a hydrocarbon fuel.
8 Due to the ability to use a high compcession ratio 9 (CPR~, back pressure at the exhaust does not cause afi sevece a loss as in the case of a conventional gas turbine. Figure 11 26 depicts the fact that a 30" watec (1 PSI) increase in 121 exhaust duct back pressure for the Cheng cycle engine only 13 causes about a 1% loss in thermal efficiency at the peak for .
14 CPR = 24, but a greater loss at lower CPR such as CPR = 16.
15¦ Sample data given in the following section 16¦ provides typical engine performances using typical state-of-171 the-art component efficiencies and the design limits of a 18 conventional counter-flow heat exchanger.
19 The data shown in Figures 27a and 27b is the 201 result of numerous parametric combinations of TIT, CPR and 21¦ SHIR (AFR) using a kerosene based fuel. TIT is selected as 22 the independent parameter and regions are shown which cover 231 a range ~f CPR and SHIR (or AFR). These cegions constitute 24 the combination of cycle parameters covered by this Cheng cycle heat engine with an ideal heat exchanger and a 26¦ reasonable component efficiency, and they can be desccibed mathematically in the range oE TIT fcom 1500F to 3000P as .
- 8 I follows:
29 The mean value of CPR as a function of TIT is 30 ¦ labeled as line A-A in Pigure 27a and is expressed as:

~` I . ' ~ :, ' . . : ~

-:-:. - : :: ; - ' , : : .. . .
, . , . .: . : : -,~:. , ., ;

:: . . .

~08~835 1 ~CPR)mean - -21.25 + 21.l4 (TIT/1000) + 3~TIT/1000)2 2 - 1.667(TIT/1000)3 3 for 150nF '-TIT ~. 30nnF

The uppe~ bound of this ~egion is:
6 (CpR)uppe~ CpR)mean x 1.5 for 1500F ~ TIT ~1600F

8 ( PR)Uppe~2 - (CPR)mean x 1.4 fot 1600F ~ TIT ~ 2200F

( )uppe~3 (CPR)mean x 1.3 fo~ 2200P ~ TIT ~ 3000F
11 The lower bound of this region is:
12 ( )lowerl ~ 4.0 fol 1500F TIT ~ 2000F
13 ( )lower2 (CPR)mean/1.4 fo~ 2000F ~ TIT ~ 3000F .

16 The mean value of AFR based on ke~osene type fuel 17 as a function of TIT is labeled as line B-B in Figu~e 27b 18 and is expressed as:
19 (AFR)mean - 209.96 - 170.90 (TIT/1000) + 52.93(TIT~1000) - 5.81(TIT/1000)3 fo~ 1500F ~ TIT < 3000F
21 The upper bound of this region is:
2 ( R)Upperl - (AFR)mean x 1.4 fo~ 1500F c TIT ~ 3000F

2 The lower bound of this ~egion is:
2 ( )lowerl (AFR)mean/1.4 fo~ 1500F ' TIT c 2000F

2 ( )1owe~2 ~ ~AFR)mean/1.5 fo~ 2000F ~ TI~r ~ 2500F .

29 ¦ ( rF ower3 - 15.0 foc 2500-F ~ TI'r c 3000-r 3zl - 31 -V

1..~~....
: . ' .

.

1 For genecality, this cegion can he convclted to a 2 specific heat input in te~ms Or ~tu/lb oE ai~Elow by the 3 following equation:
4 SpeciEic Heat Input (Btu/lb ai~Elow) = 18600./AFR tkerosene based ~uel) 7 ¦ if hydrocarbon liquid Euels are used. Otherwise the 8 app~opriate lower heat value should be used in place oE
9 18600. A second scale is shown on Figuce 27b Eo~ the 10 ¦ specific heat input in terms o Btu/lb oE ai~flow fo~ liquid 11 hydrocarbon fuel.
1~ The specific input rate is mo~e precise than 13 air/Euel ratio, not only because the lowe~ heating value of 14 ¦ fuels vary according to fuel types, but also the combustion 15 ¦ p~oducts can change thermodynamic properties of the wo~king 16 ¦ fluid. SHIR, on the right hand scale of Fig. 27b, takes 17 into account fuel flow rate in the cycle based on kerosene 18 ¦ type fuel. An error ls introduced ELom the above conve~sion 19 formula that SHIR = 18~00/AFR but it is approximately correct for other fuels. If gaseous fuels a~e used, the 21 ¦ compression work on the fuel is neglected by assuming the 22 ¦ fuels are pre-compressed by the supplie~ beEo~e delive~y. A
23 correcti~n factor should be made if the gaseous fuel is not 24 ¦ first compressed, but with the cycle desc~ibed above, such-a 25 ¦ correction is obvious to one skilled in the a~t. ~ith the 26 above conditions in mind, the operating ~egion in te~ms of 27 SHIR can be described as follows: -28 The mean value of SHIR as a function oE TIT is 29 also labelled by B-B on Fig. 27b and is expressed as:

. --, - . - . - - , '. - ' ~083835 103Btu/lh oE ai~
i (SI~IR)mean ~ _ __ ____.__ 2 ~ I o o o ) + 2 - 8 4 6 ( 10 0 0 ) - o 3 l 2 ~ T o o o ) 3 4 ~oc 1500F ~ Tlr ~ 3000F.
The lower bound of SilIR in this cegion is 6 Rlower ~ (SHIR)mean/l.4 foz 1500F~T~rc30oooF

8 The upper bound of SIIIR in this region is:
9 RUppe~l- 1.4 x (SHIR)mean fo~ lS00F ~ TIT S, 2000p 'I
11l uppe~2~ 1.5 x (SHIR)mean fo~ 2000F ~ TIT C 2500F
1~ .,................. .
13 uppe~3 ~ 1240 Btu/lb fo~ TIT ~ 2500F

It is also a peculiar feature of the Cheng cycle ,61 that, at lower TIT, the advantage of the cycle disappeats ~71 and is wo~se in te~ms oE efficiency than the Rankine cycle 18 I below 1100F. Thus, fot the Cheng cycle, TIT highel than 19¦ 1100F is more efficient.
Figs. 27a and 27b identiEy the composite 21 boundaries of SHIR (and AFR) and CPR, as a function of TIT
22 ¦ for the dual fluid cycle heat engine of the ptesent 231 inventior~. Note that the lines labeled E-E ate the loci Oe 241 maximum efficiency, and those labeled P-P rep~esent a 25 ¦ compromise between high efficiency and high th~oughput. The 26 ¦ regions beyond E-E and P-P included within the scope of the 27 ¦ Cheng cycle engine are made to allow for highc~ efficiency 28 I components developed in the future and smallet tcmpetatute 29 limits on future heat exchanger designs.

~ I _33_ r-~ . , - .
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,, 1 The region shown in Figure 23 shows the cange oE
values oE XMIX that will approximately cesult fo~ the ceqions in Figure 27. ~gain this cegion ean be dese~ibed 4 mathematieally.
The mean value Oe XMIX as a funetion of TIT in the middle of the ope~ating cegion of Figuce 28 is expcessed as:
7 (XMIX)mean - 0.20 + 0.0643~TIT/1500) foc 8 150ûF~ TIT ~ 3500F
The uppec bound foc this ~egion is:
(XMIx)upper ~ 0 3 + ~ (TIT/150ûF) 11 The lower bound of this ~egion is:

;3 ( X)lowe~l ~ 0-1 foc 1500 F ~ TIT ~ 3500F .
14 Again, the line E-E which app~oximates maximum thecmal : 15 effieieney for eompressor effieieney of .84 and turbine - 16 effieieney of .90 is eloser to the lowe~ bound, and that 17 whieh is the eompromise between high effieiency and high 18 thcoughput P-P is app~oximated by the upper bound. The 19 upper and lower bound~ shown in Figu~es 14 and 28 cepcesent a eonsistent set oE eyele pacamete~ data foL the Cheng eyele ~- 21 engine with ceasonable eomponent's effieieneies. The cegion 22 beyond E-E and P-P whieh falls within the seope of the Cheng ;~ 23 eyele en~ine of the present invention is antieipated fo~
~future improved engine eomponent effieieneies.
The set of parametrie data given in Figu~es 29, 30 26 and 31 serve to illust~ate the limitations of the 27 eonventional heat exehanger integrated into a regenecative .
28 pacallel eompound dual-fluid eyele powerplant. Tucbine 29 inlet temperatuce is set at 2150F, eompcessoc pressute 3 ratio is set at 27:1 and the ai~-fuel catio in the eombustoc 31 is the independent variable. Component assumptions ace 32 identieal to those given earlie~.
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~1~8;~835 I As the air-fuel ratio increases f~om 2 stoichiometric tl5) to 41.3, efficiency inc~eases flom 46.7 3 I to 52.0% at the peak but at highe~ aie-fuel ratios (or 4 lower SHIR) efficiency deops rapidly [Figure 29). To maintain the turbine inlet temperature at 2150F while 6 inceeasing the aie-fuel ratio ~or decreasing SHIR), the 7 steam-aie ratio drops markedly feom 0.62 to 0.12. As a eesult, power throughput is eeduced by a factor oE more than 91 two - from 760 to 330 hp/lb/sec.
Figure 30 shows how temperatures change on both 11 sides of the heat exchanger as aie-fuel ratio is increased.
'2 The steam quality at the heat exchanger discharge is also 13¦ shown on this figure. .
14 Figure 31 shows the effect of three diffeeent conditions of combustor air-fuel ratio - 20, 35 and 45 on 16 the heat exchange~. Note that two limitations have been ~71 stipulated for the heat exchangee: the neck tempeeature on 18¦ the gas mixture side is assumed to be at least 50F geeater 19I than the water boilin~ temperatuee, and the steam dischaege temperatuce is assumed to be at least 50F less than the gas 21 mixture temperature entering the heat exchangee.
22 Referring to both Figuees 30 and 31, consider the 231 conditio~s at air-fuel ratio = 20. Here the gas mixture 24 dischaege tempeeature is at the satueation temperatuee of 251 the steam in the mixtuee. This is the situation assumed for 2~ all the data qiven in Figuee 27. Wet steam leaves the heat 271 exchanger having a quality of 0.64 and the gas mixture neck 28 temperatuee is set at the limit 50F g~eatee than the water 291 b~iling te perature.

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~1 _ 1 1083ti35 I At an ait-fuel r~tio of 35, the gas mixtuce mus~
2 dischacge at a tempecature of 300F, which is 155F g~ea,tec 8 than the satu~ation tempecature o~ steam in the mixtuce.
4 This elevated dischacge tempecatute is necessacy to balance S the heat in the low tempecature end oE the heat exchangec.
6 The gas mixtuce neck temperature is unchanged but 7 superheated steam is discharged E~om the heat exchangec at B 480F. Even at an elevated temperature at the heat 9 exchanger exit, the rejected heat is less per pound o~
10¦ working fluid than other engine cycles.
ll At an air-fuel ratio of 45, the hot end 12 temperature limit has been reaclled and the discharge ~-13 superheated steam is exactly 50F below the gas mixtuce l4 inlet temperature. To balance the heat on the high , lS temperature end of the heat exchanger, at the neck the , l6¦ temperature increases to 610F, and this forces the gas l7¦ mixture discharge tempeeature of the heat exchanger exit up ,. . .
18 rapidly to 515F to balance the heat'on the low temperature 201 end. It is obvious that waste heat is not lecovered ' --~
sufficiently in this case. Hence, one rejects too much'heat , 2l again. The combination of limits at both the neck and hot 22 ends have degraded the cycle performance significantly as 23 shown in~the thermal efficiency curve on Figure 29. Note 241 that the break point in efficiency on this curve occurs 251 precisely at the point where the steam dischacge tempecature 26 ¦ is first~limited (air-fuel ratio = 41.3); and the steam 28 temperature or the degree of superi~eat of steam out oc the .
i heat exchanger ceaches a peak. It becomes lower on both - 29 ¦ sides of that point. This again points out the dependent ' 31 natuee of cycle parameters~

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:' I Cycle condition!: ~or this example wete 2 intentionally selected to cotrespond to a point on the 3 maximum efficiency line for the ideal data shown on FiguLe 4 27b (curve E-E). Note that the neck limitation of a conventional heat exchanger has ceduced the efficiency ccom 6 55~ to 52~.
¦ In summary, the dual 1uid heat engine oE the 8 present invertion provides vecy high thecmal efficiency 9 while throughput remains remarkably high. The result fot a powerplant installation is reduced fuel consumption and 11¦ therefore reduced opçrating cost coupled with reduced 12¦ powerplant size for a giv~n power output and hecefoce lowec 13¦ initial cost compared with, for example, a combined cycle '41 engine.
l5¦ An Example of Dual Pluid 16¦ CYcle_ ngine Design__ 17¦ The Cheng cycle engine design is complicated 18¦ because the mass flow through the aic compcessor is vecy 19 different fLom the mass flow through the tucbine. Because there are a number of variables in a Cheng cycle en5ine, 2~1 some freedom for the engine design does exist. The 221 following example sets focth typical steps to allow an 23 engine designer to obtain, quickly, engine pacametecs hefore 24 a detailed component map matching and final computation of such an engine is made.
26¦ (l) Enqine size: The first order of importance 27 is to know what powe~ output range is cequiced foc the 28 engine. Unfortunately, the component size cannot be picked 29 until the end of this proceduce.
~2) Selection of the Tucbine Inlet Tempecatuce 31 ¦ (TIT): The Cheng cycle engine generally performs bettec at 32 high TIT, but the cooling methods have to be detecmined first. If satucated steam is used to cool the flrst nozzle .

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I bank and tu~bine blades, li~.~n the mass fl~ ate Eo~ cooling 2 is genecally much smalle~ than that of using the bleed ai~
3 from the compresso~. This efficient cooling method allows a 4 highe~ TIT than a gas turbine. Since some ai~c~a~t gas tu~bines are capable of operating at TIT 2450F with al~
6 bleed cooling, for illust~ative purposes, let us use TIT
7 2500F.
8 (3) Compression ratio ~CPR): Comp~ession ~atio is related to the cost of the comp~esso~. Refe~ing to 101 Figure 9, one can see that if CPR = 40, the efficiency is Il only 1~ better than for CPR = 30. 8ut the cost of the 12 eompresso~ at CPR = 40 is more than 20~ higher than fo~ a 13 compressor with a CPR = 30, so let us assume CPR = 30 would .
141 fit the desired need. The compressor efficiency can 15¦ influence the engine pecformance. However, this influence 16¦ is minimized on the Cheng cycle engine, so compressor 17¦ efficiency of 84~ is assumed acceptable.
18 (4) Degree of Superheat of Steam and Ope~ative 19 Points: It is a standard engineering p~actice that the design point is always slightly away from the peak point.
21¦ In this case, one would pick SHI~ = 640 Btu/lb ai~ (Pigure 22¦ 9) or approximately 460~ of superheat (Figu~e 12). This 23 allows a larger temperature diffe~ence in the heat exchange~
241 relativ.e to the hot end which gains a little better 251 throughput (hp/lb. air/see.).
26 A cross check with Figu~e 7 shows that the ~27 ef.ficiency is only a quarter pe~cent lower than the peak 28¦ effieieney. Thus, SHIR = 640 has been dete~mined. If the 291 heating source is a nuelear reacto~, one can design the reactor surface to supply 5HIR = 640 BTU/lb. air. If it is 32 ~

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1.~8~835 l burning oil, then the fue~ Elow catc is convcrtihle as Wf 2 ¦ 640/18600 = SHlR/low heating value o~ ~uel = 0.0344 lb.
3 ~ucl/lb. ai~.
4 ¦ (5) steam to Air Ratio: ~ith TIT = 2500F, CP~ =
30 and 513IR = 640 ~tu/lb. ai~, one can consult Figu~e 17 and 6 see that the XMIX would be .175 lb. steam/lb. air.
71 Therefore, for l lb./sec. airflow through the air 8 compressor, the tu~bine would have to pass l + Wf + XMIX
91 lb/sec. mixture flow. For this case, it is equal to l +
lO¦ 0.0344 + 0.175 = 1.2094 lb./sec. mixtu~e. The steam to fuel ll¦ ratio is 5.0872 lb. steam/lb. fuel.
l2¦ (6) Selection of Component Sizes: If the engine ~31 is designed to p~oduce lO,000 13p at peak output, f~om Figu~e .
14¦ 22, the throughput, T.P., is 490 Hp/lb. ak /sec. So the l5¦ compressor size is the ratio of engine output, lO,000 11p, '61 divided by T.P. 490 Hp/lb. ai~/sec. o~ 20.41 lb./sec. at CPR
17¦ = 30. The turbine ~low rate would be 24.68 lb./sec. mixture l8¦ flow. With the major components selected, one can p~oceed 19 to pick available components having perfo~mance close to 201 this requirement or design components to f it this 2l¦ requirement. With a refined component mapping, one can then 22¦ begin to predict off peak load characteristics.
231 : This engine would have the~mal e~iciency in the 241 neighborhood of 55% and a throughput of 490 13p/lb. air/sec.
25 ¦ This can be compared with a practical ult~a high tempe~ature 26¦ (TIT = 2800F) combined cycle system of 50.4~ the~mal 2281 efficiency with a throughput of 325 13p/lb. air/sec. Hence; .
1 this engine becomes far superior in efficiency and 29 throughput to the combined cycle.

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1083~i35 , I The physical conrigu~ation of a heat engine ! 2 inco~po~ating the principles of the Cheng cycle which has been described herein reptesents a p~eEe~led embodiment but 4 by no means the only configucation which can be used. To S 1 those skilled in the art, it will be apparent that othe~
6 configurations of, additions to, or substitution of engine 71 components can be used. Also, if efficiency is unimpo~tant 8 to the designer, as where economic conside~ations a~e 91 pa~amount, the design of the heat engine can be expected to va~y considerably from the configu~ation described he~ein.
1l¦ In other words, the Cheng cycle, as desclibed he~ein, sets l2¦ forth a ~elationship between engine pa~amete~s of a dual-l3 fluid engine for maximum eEficiency and/o~ th~oughput. It .
l4¦ is possible, within the teaching of the Cheng cycle, to ~51 build other physical configurations for effectuating these l6¦ relationships and it is also possible to design and build l7 an engine which operates away from peak eEficiency operation 31 whlch iq de c~lbed here.

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Claims (34)

THE EMBODIMENTS OF THE INVENTION IN WHICH AN EXCLUSIVE
PROPERTY OR PRIVILEGE IS CLAIMED ARE DEFINED AS FOLLOWS:

1. A dual fluid heat engine comprising:
a) a chamber;

b) compressor means for introducing a first gaseous working fluid into said chamber, said compressor means having a predetermined pressure ratio (CPR);

c) means for introducing a second liquid-vapor working fluid in the form of a vapor within said chamber at a defined second/first working fluid ratio (XMIX);

d) means for heating said first gaseous working fluid and said second working fluid in the vapor form in said chamber at a defined specific heat input rate (SHIR);

e) turbine means responsive to the mixture of said first and second working fluids for converting the energy associated with the mixture to mechanical energy, the temperature of said mixture entering said turbine means defining the turbine inlet temperature (TIT);

f) counterflow heat exchanger means for transfering residual thermal energy from an exhausted mixture of first and second working fluids to said incoming second working fluid to thereby preheat the same to a superheated vapor state prior to its introduction within said chamber, and wherein XMIX and SHIR are selected so that for a given value of TIT, XMIX is substantially equal to o? is greater than XMIXpeak, where XMIXpeak occurs when the following conditions are both met simultaneously:
i) the temperature of the superheated second working fluid vapor is substantially maximized; and, ii) the effective temperature of said exhausted mixture of the first and second working fluids is substantially minimized.

2. The heat engine of Claim 1 wherein, for a given value of TIT, the CPR is chosen to generally maximize the transfer of the residual thermal energy to said second working fluid.

3. The heat engine of Claim 1 wherein, for a given value of TIT, CPR is chosen to fall within a range bounded by the value of CPR for which maximum transfer of residual thermal energy to said second working fluid occurs, and a value which is not less than one-third of this value.

4. A heat engine as in Claim 3 wherein SHIR falls within a range bounded by SHIR at peak efficiency and 2 X SHIR at peak efficiency, and XMIX is increased above XMIXpeak to maintain a constant TIT.

5. A heat engine as in Claim 1 wherein SHIR falls within a range bounded by SHIR at peak efficiency and 2 X SHIR at peak efficiency, and XMIX is increased above XMIXpeak to maintain a constant TIT.

6. A heat engine as in Claim 1, 2 or 3 wherein said second working fluid comprises water.

7. A heat engine as in Claim 1, 2 or 3 wherein said second working fluid comprises water and said first working fluid comprises air and combustion products.

8. A heat engine as in Claim 4 or 5 wherein said second working fluid comprises water.

9. A heat engine as in Claim 4 or 5 wherein said second working fluid comprises water and said first working fluid comprises air and combustion products.

10. A dual fluid heat engine comprising:
a) a chamber;

b) compressor means for introducing a first gaseous working fluid into said chamber, said compressor means having a predetermined pressure ratio (CPR);

c) means for introducing a second liquid-vapor working fluid in the form of a vapor within said chamber at a defined second/first working fluid ratio (XMIX);

d) means for heating said first gaseous working fluid and said second working fluid in the vapor form in said chamber at a defined specific heat input rate (SHIR);

e) turbine means responsive to the mixture, of said first and second working fluids for converting the energy associated with the mixture to mechanical energy, the temperature of said mixture entering said turbine means defining the turbine inlet temperature (TIT);

(f) counterflow heat exchanger means characterized by having a neck at that point where the second working fluid first begins to change state from a liquid to a vapor, for transfering residual thermal energy from an exhausted mixture of first and second working fluid to said incoming second working fluid thereby preheat the same to a superheated vapor state prior to its introduction within said chamber, and wherein XMIX and SHIR are selected so that for a given--value of TIT, XMIX is substantially equal to of is greater than XMIXpeak, where XMIXpeak occurs when the following conditions are both met simultaneously:
i) the temperature of the superheated second working fluid vapor is substantially maximized at its discharge from said heat exchanger, and ii) wherein the temperature difference at the heat exchange neck between the exhausted mixture of first and second working fluids and the second working fluid is substantially minimized.

11 The heat engine of Claim 10 wherein, for a given value of TIT, the CPR is chosen to generally maximize the transfer of the residual thermal energy to said second working fluid.

12. The heat engine of Claim 10 wherein, for a given value of TIT, CPR is chosen to fall within a range bounded by the value of CPR for which maximum transfer of residual thermal energy to said second working fluid occurs, and a value which is not less than one-third of this value.

13. A heat engine as in Claim 12 wherein SHIR falls within a range-bounded by SHIR at peak efficiency and 2 X SHIR at peak efficiency, and XMIX is increased above XMIXpeak to maintain a constant TIT. 45

14. A heat engine as in Claim 13 wherein SHIR falls within a range bounded by SHIR at peak efficiency and 2 X SHIR at peak efficiency, and XMIX is increased above XMIXpeak to maintain a constant TIT.

15. A heat engine as in Claim 11, 12 or 13 wherein said second working fluid comprises water.

16. A heat engine as in Claim 11, 12 or 13 wherein said second working fluid comprises water and said first working fluid comprises air and combustion products.

17. A heat engine as in Claim 14 wherein said second working fluid comprises water.

18. A heat engine as in claim 14 wherein said second working fluid comprises water and said first working fluid comprises air and combustion products.

19. The method of operating a dual fluid heat engine at maximum efficiency and/or throughput, for a given turbine inlet temperature, which engine comprises:
a chamber;

compressor means for introducing a first gaseous working fluid into said chamber, said compressor means having a predetermined pressure ratio (CPR);

means for introducing a second liquid-vapor working fluid in the form of a vapor within said chamber at a defined second/first working fluid ratio (XMIX);

means for heating said first gaseous working fluid and said second working fluid in the vapor form in said chamber at a defined specific-heat input rate (SHIR);
turbine means responsive to the mixture of said first and second working fluids for converting the energy associated with the mixture to mechanical energy, the temperature of said mixture entering said turbine means defining the turbine inlet temperature (TIT);

counterflow heat exchanger means for transfering residual thermal energy from an exhausted mixture of first and second working fluids to said incoming second working fluid, said method comprising the steps of:

pre-heating the second working fluid in the heat exchanger to a superheated vapor state prior to its introduction within the chamber; and selecting XMIX and SHIR so that for a given value of TIT, XMIX is substantially equal to or is greater than XMIXpeak, where XMIXpeak occurs by both:
i) maximizing the temperature of the superheated second working fluid vapor; and ii) minimizing the effective temperature of the exhausted mixture of the first and second working fluids.

20. The method of Claim 19 including the step of choosing CPR to generally maximize the transfer of the residual thermal energy, for a given value of TIT, to said second working fluid.

21. The method of Claim 19 including the step of choosing SHIR within a range bounded by SHIR at peak efficiency and 2 X
SHIR at peak efficiency, and increasing XMIX above XMIXpeak to maintain a constant TIT.

22. The method of Claim 20 including the step of choosing SHIR within a range bounded by SHIR at peak efficiency and 2 X
SHIR at peak efficiency, and increasing XMIX above XMIXpeak to maintain a constant TIT.

23. A heat engine as in Claim 19, 20 or 21 wherein said second working fluid comprises water.

24.A heat engine as in Claim 19, 20,or 21 wherein.
said second working fluid comprises water and said first working fluid comprises air and combustion products.

25. A heat engine as in Claim 21 or 22 wherein said second working fluid comprises water.

26. A heat engine as in Claim 21 or 22 wherein said second working fluid comprises water and said first working fluid comprises air and combustion products.

27. The method of operating a dual fluid heat engine at maximum efficiency and/or throughput, for a given turbine inlet temperature, which engine comprises:
a) a chamber;
b) compressor means for introducing a first gaseous working fluid into said chamber, said compressor means having a predetermined pressure ratio (CPR);
c) means for introducing a second liquid-vapor working fluid in the form of a vapor within said chamber at a defined second/first working fluid ratio (XMIX);
d) means for heating said first gaseous working fluid and said second working fluid in the vapor form in said chamber at a defined specific heat input rate (SHIR);

e) turbine means responsive to the mixture of said first and second working fluids for converting the energy associated with the mixture to mechanical energy, the temperature of said mixture entering said turbine means defining the turbine inlet temperature (TIT);
f) counterflow heat exchanger means characterized by having a neck at that point where the second working fluid first begins to change state from a liquid to a vapor for transfering residual thermal energy from an exhausted mixture of first and second working fluids to said incoming second working fluid, said method comprising the steps of:
pre-heating the second working fluid in the heat exchanger to a superheated vapor state prior to its introduction within the chamber; and selecting XMIX and SHIR so that for a given value of TIT, XMIX is substantially equal to or is greater than XMIXpeak, where XMIXpeak occurs by both:
(i) generally maximizing the temperature of the superheated second working fluid vapor at its discharge from the heat exchanger; and (ii) generally minimizing the temperature difference at the heat exchanger neck between the exhausted mixture of first and second working fluids and the second working fluid.

28. The method of Claim 27 including the step of choosing CPR to generally maximize the transfer of the residual thermal energy, for given value of TIT, to said second working fluid.

29. The method of Claim 27 including the step of choosing SHIR within a range bounded by SHIR at peak efficiency and 2 x SHIR at peak efficiency, and increasing XMIX above XMIXpeak to maintain a constant TIT.

30. The method of Claim 28 including the step of choosing SHIR within a range bounded by SHIR at peak efficiency and 2 x SHIR at peak efficiency, and increasing XMIX above XMIXpeak to maintain a constant TIT.

31. A heat engine as in Claim 27, 28 or 29 wherein said second working fluid comprises water.

32. A heat engine as in Claim 27, 28 or 29 wherein said second working fluid comprises water and said first working fluid comprises air and combustion products.

33. A heat engine as in Claim 29 or 30 wherein said second working fluid comprises water.

34. A heat engine as in Claim 29 or 30 wherein said second working fluid comprises water and said first working fluid compriese air and combustion products.