NZ210934A - Using water vapour in air supply of power-producing turbine as thermal diluent - Google Patents

Description

210934 NEW ZEALAND PATENTS ACT, 1953 No.: 210934 pate. 24 January 1985 i COMPLETE SPECIFICATION "PROCESS FOR PRODUCING POWER" ^ We, FLUOR CORPORATION, a corporation organized and existing under the laws of the State of Delaware, USA, of 3333 Michelson Drive, Irvine, California 92730, United States of America, hereby declare the invention for which t / we pray that a patent may be granted to xcutju.6, and the method by which it is to be performed, to be particularly described in and by the following statement: - (followed by page -la-) 210934 PROCESS FOR PRODUCING POWER BACKGROUND OF THE INVENTION 1. Field oE the Invention v This invention relates to a process Cor producing mechanical energy or electric power in which a combustion turbine is used for conversion of the chemical energy in a fuel. 2. Description of Prior Art When a working fluid is used in an engine to produce mechanical energy or electrical power from the chemical energy contained in a fuel, the working fluid is pressurized and, following combustion of the fuel, the energy thus released from the fuel is absorbed into the working fluid as heat. The working fluid with the absorbed energy is then expanded to produce mechanical energy which may in turn be used to drive a generator to produce electrical power. Unconverted energy is rejected in the exhaust in the form of heat, only a portion of which may be recovered and utilized. The efficiency of the engine is at a maximum when the temperature of the working fluid entering the expansion stage is also at a maximum.

In the case of combustion turbines, air compression is used for the pressurization step and direct combustion o£ the —Id- 210934 fuel into the compressed air is the energy addition step. Expansion in the turbine produces the mechanical energy and the unconverted heat is carried off by the turbine exhaust. The efficiency of the combustion turbine is at a maximum when the combustion temperature itself is at a maximum, and this occurs when the fuel is burned in the presence of the pressurized air under:, stoichiometric conditions, i.e., enough air is present for complete combustion, but without any excess. conditions, however, the resulting temperature is approximately 4000® F, which is in excess of the metallurgical limits of the turbine. As a result, it is necessary to utilize a large excess of air in the combustion step, which acts as a thermal diluent and reduces the temperature of the combustion products to approximately 2000° F. The necessity to use a large excess of air under pressure in turn creates a,large parasitic load on the system, because compression of the air requires mechanical energy and thus reduces the net power produced from the system, as well as reducing the overall efficiency of the system. cycles is that the pressurization step requires compression of air. Compression of a gas is very inefficient, since mechanical energy is required, which is the highest form oE energy and degrades into thermal energy. The mechanical energy required for air compression cam.be reduced by utilizing interstage cooling, that is, by cooling the temperature of the compressed air between successive stages of a multiple stage compression process. However, from an overall cycle efficiency standpoint, Interstage cooling can be utilized advantageously if the heat removed from the compressed air in the inter-cooler can be efficiently When fuel oil is burned with air under stoichiometric Another disadvantage of existing combustion turbine 210934 recovered and utilized. If the entire heat is simply rejected to the atmosphere, the overall cycle efficiency is actually decreased, since it results in the consumption of more fuel to compensate for the energy lost through the inter-cooler. Accordingly, rather than simply rejecting the heat, in commercial practice, the high compressor horsepower requirement has been tolerated while containing the heat in the compressed air stream.

"Even in light of the foregoing limitations, it is very desirable to use a combustion turbine engine, because it is able to operate at the highest temperature of engines that use a working fluid to convert chemical energy in a fuel to mechanical energy. However, due to the high exhaust temperature that is inherent to a combustion turbine engine, the efficiency of the cycle is limited, and as a. result, the exhaust from the engine is used as the heat source to operate another engine such as a steam turbine to increase the overall efficiency of utilization of the fuel. Such a system is called a combined cycle system and is widely used in the industry. Another use for the energy contained in the combustion turbine exhaust is to raise superheated steam which is injected back into the combustor of the combustion turbine, see, e.g., CJ.S. Patent Mo. 3,978,661. Yet another method is to preheat the air leaving the compressor against the engine exhaust and simultaneously'use interstage cooling during compression (see Kent's Mechanical Engineers Handbook, 1950). These systems show higher overall efficiencies with respect to the utilization of the chemical energy contained in a fuel, but as will be explained subsequently herein, are inherently less efficient than the process of the present invention.

A combined cycle cannot take full advantage of air-compressor inter-cooling, because the temperature of the 210934 heat rejected in the air compressor inter-cooler is too low to be recovered for efficient use such as steam generation. A small portion of this heat.may be recovered for boiler feed water preheating as described in Agnet, U.S. Patent No. 3,335,565 but this results in more heat being rejected with the stock gases and results in little, if any, net increase in either heat recovery or cycle efficiency. Recently, direct water injection into the air stream as a means of inter-cooling has been proposed.

However, there are two disadvantages with this. One is that the temperature of air leaving the inter-cooling step is limited by the dew point temperature of the saturated air. Also, by the direct injection of water into the air in the intercooler, the added water vapor which serves as a thermal diluent needs to be compressed in the successive stages after the inter-cooler, which precludes realizing the full advantage of water vapor substitution as a means of saving compression power. of water into the compressed air after preheating both the air and the water. However, this means of evaporation requires a higher temperature level to achieve useful moisture loading of the air because the air and water leave the evaporator in equilibrium with each other. This method of water evaporation is less efficient than the present invention which can take advantage of air entering the saturator at low temperatures. since the evaporation of water (steam generation) occurs at a constant temperature, whereas the heat release occurs at varying temperatures. The following diagram shows the heat release curve and the water evaporation line: Foote, in U.S. Patent 2,869,324, describes evaporation The steam cycle has an inherent high irreversibility 210934 w, water evaporation (steam generation) line Cumulative Heat Duty As can be seen from the diagram, with steam generation, a small temperature difference between the heat source and heat absorbing fluid cannot be maintained, and this leads to a high irreversibility in the system and hence a lower efficiency.

A combined cycle plant is also expensive since it requires an additional steam turbogenerator, steam drums, surface condenser for condensing steam turbine exhaust, and cooling towers to reject the heat from the surface condenser to the atmosphere.

A steam injected cycle cannot take full advantage of air-compressor inter-cooling for. the same reasons as a combined cycle. Also this cycle involves the generation of steam and hence has the same irreversibility associated with it as described for a combined cycle, although eliminating the steam turbogenerator, surface condenser and cooling towers, and reducing the parasitic load of air compression by displacing some of the air with steam. This is an improvement over a water injected cycle ~ ~ ~ " 210934 where water is directly injected into the combustor. The injected water displaces some of the diluent,air, but there is a tremendous irreversibility associated with this. The evaporation of the ~r water in the combustor uses energy from the fuel at the highest temperature, which results in an overall reduction of efficiency. Also with the water injected cycle, the heat available from the turbine exhaust still remains to be utilized.

The heat used for generation of steam in a steam injected cycle, is of a much higher quality, i.e., temperature level, than is desirable. For example, typically for a combustion turbine operating at a pressure ratio of 11, the steam pressure required for injection should be at least 200 psla. The corresponding saturation temperature of the steam is 382° F.

This requires that a heat source be available at much higher temperatures and heat down to only 420" F can be utilized without unreasonable temperature pinches.

The inter-cooled regenerative cycle uses inter-cooling during the air-compression step, and compressed air preheated against the turbine exhaust before the air enters the combustor. The optimum pressure ratio for this cycle is about 6 to 7. The heat released in the inter-cooler is all lost to atmosphere. Also the temperature of gas leaving the air pre-heater is around 500° F, and the heat contained in these gases is all wasted. All the thermal diluent is compressed, leading to a large parasitic load, which results in poor overall efficiency for the system.

Martinka 0.3. Patent 2,186,706, describes the replacement portion of the air for combustion with water vapor 210934- derived by directly contacting the compressed air with heated water in a-humidification operation. The heat required for this humidification operation is supplied by inter-coolers in the air compressor. Makeup water for the system picks up additional heat from the gas turbine exhaust. The net effect of such a system is a reduction in the parasitic load of air compression and, thus, an Increase in cycle efficiency.

Nakamura et al., in U.S. Patent 4,537,023, describe a system similar to that of U.S. Patent 2,186,706, in which an after-cooler is used for the air compressor. The after-cooler reduces the temperature of the water leaving the humidifier, which in turn allows recovery of lower-level heat to a greater extent. The decrease in heat-rate resulting from the addition of the after-cooler is approximately 1.4 percent, based on the data presented in the Nakamura et al. patent.

Both the Martinka and the Nakamura et al. systems reject heat from the cycle through the stack gases. Rejection of heat is a consequence of the second law of thermodynamics and any power cycle converting heat to power must reject some heat. To improve the, cycle efficiency, it is not only Important to minimize the quantity of heat being rejected, but also, to minimize the temperature at which the heat is rejected. In both the Martinka and Nakamura et al. systems, the quality of heat being rejected is solely set by the stack temperature which constrains the cycle efficiency. 210934 SUMMARY OF THE INVENTION The present invention provides a process for producing mechanical energy or electric power from a fuel, utilizing a combustion turbine in which some or all of the excess air, which is used as thermal diluent and working fluid, is replaced with water vapor. The water vapor is introduced into the system in a very efficient manner, by pumping as a liquid followed by low temperature evaporation. Pumping a liquid requires very little mechanical energy compared to gas (air) compression. Also, evaporation of the water is accomplished using low level heat, in a counter-current multistage humidification operation.

Humidifying in multistages permits, the temperature of the compressed air to follow closely the temperatures of the heating medium, which minimizes thermodynamic irreversibilities.

The process of the present invention strives to minimize • * simultaneously the quantity and quality of heat rejections, and this results in a significant improvement in the cycle thermal efficiency. Low level heat is rejected from the compressed air during inter-cooling and just prior to humidification. The product of quality and quantity of heat rejected in this manner is lower than that rejected through the stack gases. This may be accomplished via air-coolers, cooling water exchangers or through a refrigeration system.

The process of the present invention also reduces the parasitic load of compressing the diluent air and achieves a more thermally efficient power production cycle. Humidification of the compressed air also leads to a reduction of nitrogen oxide emissions, which, of course, is a major environmental benefit. The invention also provides the means for humidifying the '- ■ r'/ 210934 compressed air in a thermodynamically efficient manner, using direct contact of the compressed air in a saturator, which permits the air to be humidified with relatively cold water and without the requirement of a steam boiler.

Figure 1 is a schematic depiction of the process of the present invention in one preferred embodiment,utilizing a two-stage air compressor, axially coupled to a turbine.

Figure 2 is a schematic depiction of the process of the present invention, utilizing, variations in the mode of low level heat rejection.

Figures 3 and 4 are graphs of the effect of pressure ratio on cycle efficiency and the effect of air temperature to humidifier on cycle efficiency. into the first stage of the dual stage air compressor, 2 and 3, which are coupled together axially at 4. The compressed air exiting the first stage of the compressor 2 through line 5 is at a temperature of approximately 300" to 400"F and passes through heat exchanger 6 where it undergoes heat exchange relationship with water passing through line 7. The temperature of the compressed air is thus reduced to approximately 40® to about 250" F, typically about 70" to 140° F, and thereafter is passed through line 8 to the second stage, 3, of the air compressor.

BRIEF DESCRIPTION OF THE DRAWINGS Description of Preferred Embodiments Referring to Figure 1, air through line 1 is introduced 210934 The compressed air exiting the air compressor through line 10 is at a temperature of approximately 300° to about 400° F and passes through heat exchanger 11 in which it experiences heat exchange with water passing through line 12. The temperature of the compressed air is thus reduced to approximately 40° to about 230° F, typically about 115" to about 200*Fi Water in line 7, following heat exchange in heat exchanger 6, is introduced into the top section of saturator 15 at a temperature of about 300° to about 400°F. Within the saturator, the air and water are contacted counter-currently in multi-stages, which improves the thermodynamic efficiency. The operating pressure of the saturator is about 200 psi to about 600 psi, and the water temperature is approximately 300 to about 400°F. The water remaining after vaporization is removed from the bottom o£ saturator 15 through line 16 and pumped at 17 through exchanger 18 wherein heat is rejected and line 19 to either line 7 and heat exchanger 6 or line 13 and 12 to heat exchanger 11, as desired. Low level heat from the intercooler and _the aftercooler are thus rejected to the atmosphere.

The humidified air exits saturator 15 through line 20 as essentially saturated air at approximately 250°F ta about 350"F and is passed through heat recovery unit 21 in heat exchange, relationship with the exhaust from turbine 22 to preheat the saturated air prior to introduction to combustor 24. The Euel for combustion is introduced through line 25 and the combusted gaseous product exits through line 26 to drive turbine 22. The turbine is coupled axially, at 4, to the air compressor and also to generator 30 for the' production of electrical power. While the compressor, turbine and generator are described and illustrated as coupled on a single axle, it will be appreciated X'vTyi'li' £ Z ' \\ ;£/ 5 °\ ( 2W01//9S6o| 7 210934 that other arrangements may be used, as will be readily understood by those skilled in the art.

Within heat recovery unit 21, the hot exhaust from the gas turbine is passed in heat exchange relationship with water to heat the water to the appropriate temperature for humidification within saturator 15, as illustrated. Thus, water through line 31 may be taken thereby to the heat recovery unit as illustrated. Additionally, of course, makeup water may be added through line 32 by pump 33 as is necessary to maintain the necessary water Inventory in the system.

Variations in the mode of low level heat rejection are, of course, possible, and certain of these are depicted in Figure 2. Thus, in the embodiment here illustrated, heat rejection occurs in exchanger 35, wherein compressed ait from the after-cooler 11 undergoes heat exchange against water to increase the temperature of the water, after which the cooled compressed air is introduced into the lower section of saturator 15. Provision may also be made for heat rejection in exchanger 37 in which compressed air from the inter-cooler is heat exchanged against cooling water or refrigerant prior to introduction into stage II of the multiple stage compressor. In this embodiment, makeup water is heated by heat exchange in exchanger 38 before combining with water in line 7 and passage through inter-cooler 6.

The process of this invention is shown as a stand alone power generation cycle. This process may, if desired, be integrated with other process facilities for further optimization of energy conversion. In a cogeneration configuration, a portion of the hot turbine exhaust would be utilized to produce steam for other purposes. In a reverse manner, the cycle can be integrated with heat recovery in other processes to increase the supply of ■li- Jpv 210934 heated water to the humidification operation. In this manner, the power cycle of the present invention can be used to a greater extent than other cycles in integration with a plant producing large quantities of low temperature level heat such as a coal gasification plant or a geothermal facility, because humidifi-cation can be achieved at such low temperatures while the work producing step of expansion in the turbine occurs at much higher pressure ratios. Also, the cycle may be used with reheat turbines more efficiently because this cycle optimizes at higher pressure ratios. In reheat turbines, the first turbine operates at a high pressure where partial expansion occurs, additional fuel is fired in a second combustor, and the hot gases are expanded to near atmospheric pressure in the second turbine. The results of rejecting heat to cooling water are presented in Figure 3 in the form of a plot of the pressure ratio versus cycle thermal efficiency. For comparison a similar plot for the * Nakamura et al. system is also presented in the same Figure.

These efficiencies were calculated using a consistent set of design criteria established in the Nakamura et al. patent as follows: EXAMPLE 1: NAKAMURA PROCESS (Example 2, Figure 2) (I) Conditions (a) Efficiencies Compressor adiabetic efficiency nC = 0.89 Turbine adiabetic efficiency nT = 0.91 Mechanical efficiency nM = 0.99 Generator efficiency nG = 0.985 Combustion efficiency nB = 0.999 (b) Ambient air conditions at compressor inlet Temperature 15°.c Pressure 1.033 ata.

Relative humidity 60% Flow rate }dry air 1 Kg-mole/sec }H2O 0.0101 .JCg-mole/s 210934 (c) Fuel Kind Temperature High heating value (0° C.) Low heating value (0° C.) (d) Total pressure loss (e) Replenishing water Temperature Flow rate (f) Turbine inlet conditions Pressure Temperature (g) Minimum temperature difference for heat-exchanger High temperature regenerator R^ Low temperature regenerator R2 Fuel preheater R^ Intercooler IC (h) Miscellaneous The compressive forces of the fuel, replenishing water and water at the bottom of the exchanging tower are assumed to be negligible while the total auxiliary power is taken as O.j percent of the generator output. Further, as to the cooling air for the turbine, the availability of low temperature compressed air in the regenerative gas turbine cycle is taken into account to determine its required amount.

III) Result? (a) Waste gas Temperature Flow cate (b) Compressor outlet temperature (AC2) (c) Sending end power output (d) Sending end thermal efficiency (LHV) natural gas 15° C 245,200 Kcal./Kg-mole 221,600 "kcal/Kg-mole . 2% ° C 0.132 Kg-mole/sec. 6 ata. 1,000® C ° C ° C ° C ° C 82.7° C 1.15 Kg-mole/sec. 148° C 8690 KW 50 . 2% EXAMPLE 2: RAO PROCESS (I) Conditions (a) Efficiencies Compressor adiabetic efficiency Turbine adiabetic efficiency Mechanical efficiency Generator efficiency Combustion efficiency nC = 0.89 ^nT = 0.91 nM = 210934 (b) Ambient air conditions at compressor inlet Temperature Pressure Relative humidity Flow rate }dry alt }h2o 15s C 1.033 ata. 60% 1 Kg-mole/sec. 0.0101 Kg-mole sec. (c) Fuel Kind Temperature High heating value (0* c. ) Low heating value (0" C.) (d) Total pressure loss (e) Replenishing Hater Temperature Flaw rate natural gas 15° C 245,200 Keal/Kg-mole 221,600 K^al/Kg-mole . 2% °C 0.144 Kg-mole/sec . (f) (g l Turbine inlet conditions Pressure Temperature Minimum temperature difference for heat exchanger and/or exchanger outlet condition High temperature regenerator R-^ Low temperature regenerator R2 Fuel Preheater R3 Inter Cooler IC Selfheat exchanger (SR) Intercooler Outlet IC2 Rejecting Aftercooler RAC (h) 6 ata. 1,000° C °C 20 "C 30°C 20°C 20°C 35°C 48°C Miscellaneous The compressive forces of the fuel, replenishing water and water at the bottom of the exchanging tower are assumed to be negligible while the total auxiliary power is taken as 0.3 percent of the generator output. Further, as to the cooling air for the turbine, the availability of low temperature compressed air in the regenerative gas turbine cycle is taken into account to determine its required amount. 210934 (II) Results (a) Waste Gas Temperature Flow Rate 75.6°C 1.18 Kg-mole/sec . ' 157° (b) ' Compressor outlet Temperature (AC2) (c) Sending end power output (d) .Sending end thermal efficiency (LHV) 10947KW 51.06% The Nakamura et al. system shows a peak efficiency at a pressure ratio of approximately 6 for a gas turbine firing temperature of 1832 F. The cycle of the present invention, however, shows a peak efficiency at a pressure ratio of approximately 10.5 for the same gas turbine firing temperature of 1832 F. Comparing the peak performance for the two systems, the heat-rate for the process of the present invention is approximately 1.6 percent lower than that for the Nakamura et al. system. This improvement is actually higher than the improvement by Nakamura et al. from utilizing the after-cooler. Also, the process of the present invention makes it possible to take advantage of higher pressure ratios, for example in the range of 6:1 to 34"sj^, and thus increases the engine specific power. as a function of the temperature of the compressed air entering the humidification operation. This plot shows that the cycle efficiency is not necessarily maximized when the inlet air temperature to the humidifier is minimized. The optimum temperature depends on the simultaneous reduction of quality and quantity of heat rejected. compressed air for the humidification operation is done to achieve the lowest possible water temperature from the humidifier. This, however, does not result in peak efficiency Figure 4 is a plot of the cycle thermal efficiency drawn According to U.S. Patent 4,537,023, precooling of the 210934 for the cycle as evidenced by Figure 4 which shows plots of temperature of air entering the humidifier and the resulting water leaving the humidifier versus cycle efficiency. The peak efficiency occurs when the quality and quantity of heat rejection are simultaneously minimized.

Another disadvantage with the system of U.S. Patent 4,537,023, is that the temperature difference between the air entering the humidifier and water leaving the humidifier is set by the temperature difference used in designing the after-cooler. This forces an added constraint on the system and fixes the temperature of the water leaving the saturator at a higher temperature than the corresponding temperature that results from the process of the present invention.

A major advantage of the process of the present Invention is a significant improvement in thermal efficiency. Appreciation for this improvement in thermal efficiency, compared to U.S. Patent 4,537,023 will be realized by the following. In a 500 MH power plant, with the Nakamura et al. process, the fuel required using a gas turbine with a firing temperature of 1832°F = 500 x 1000 KW x 6800 BTU x 24 x 365 hrs KWH yr = 2.98 x 107 MMBTU/year With the process of the present invention, the fuel required = 500 x 1000 KW x 6700 BTU x 24 x 365 hrs KWH yr = 2.93^ 107 MMBTU/year -16 210934 Hence, fuel savings with the Improved Power Cycle = (2,98 i 107 - 2.93 x 107) MMBTU/year' = 0,05 x 107 MMBTU/year This corresponds to an annual saving (with fuel cost at us$4/MMBTU) of 0.05 x 107 MMBTU xlus$£_ =us$2 ,x 106/year year MMBTU The process of the present Invention may also be used to convert low level heat from another plant such as a gasification plant or refinery into mechanical energy or electrical power, at a much higher efficiency than other methods. The fuel used in the combustion engine serves to upgrade the recovered low level heat. Thus, for example, when the low level heat recovered by preheating the humidifier circulating water, in the range of 300° to 140° F from a gasification plant is converted to electric power, the effective efficiency of conversion is as high as approximately 18%. The imported heat may be used to evaporate additional water to provide total water in the range of 0.26 to 0.5 pounds per pound of dry air.

The standard of efficiency of conversion of such low level heat may be calculated for U.S. Patent No. 4,085,591, "Continuous Flow, Evaporative-Type Thermal Energy Recovery Apparatus and Method for Energy Recovery", where a pressurized gas, e.g. air, is humidified in a spray chamber, and expanded through a gas turbine, to take advantage of the higher specific volume of humidified air. The resulting efficiency with this system is, less than 5%. Also there are a number of inherent disadvantages. To produce appreciable amounts of power, very large equipment is required since the system pressure is limiting. This system cannot "upgrade" the recovered low level 210934 energy, since it cannot be used in conjunction with a combustion engine. that, with the process o£ the present invention, chemical energy, or low level heat supplemented with chemical energy, may be converted to mechanical energy or electrical power at a very high efficiency. It will also be appreciated that significant environmental benefits will result from the process of the present invention, including conservation of energy resources and reduction in thermal pollution due to the higher efficiency, a reduction in water consumption, particularly as compared to either the combined cycle or the steam injected cycle, and a reduction in nitrogen oxide emissions. With combined cycle plants, steam must be injected into the combustor to reduce such emissions, which in turn leads to a decrease in efficiency, which is overcame by the present invention. tlons and modifications of the process of the present invention may become apparent to those skilled in the art. Thus, for example, a plurality of inter-coolers may be used as well as more than two stages of air compression. Also, the inlet air to the compressor may be cooled using a refrigeration system to improve both the efficiency and capacity of the system. The air leaving the inter-cooler may also be further cooled using the refrigeration system and the saturator water may also be precooled, using a refrigeration system, before it enters the inter-cooler. Additionally, saturators of designs other than that illustrated may be used, such as a design where the water would be introduced at a plurality of locations. It is accordingly to be understood that all such modifications and variations are to be considered within the scope of the present invention, as defined 3 cla It will be appreciated from the foregoing description In light of the foregoing description, certain varia-

Claims (14)

210934 WHAT WE CLAIH ISs

1. A process for producing power utilizing a combustion turbine, comprising humidifying compressed ai-r in multistage countercurrent flow prior to combustion, to provide water vapor as thermal diluent for combustion in said turbine, water being at a temperature below its boiling point,at the operating pressure when in contact with said compressed air passing said compressed, air in heat exchange relationship with said water prior to humidification, whereby the temperature of said water is increased and the temperature of said compressed 9ir is decreased, and rejecting heat from the power production cycle by praeoollna said ntamareaaed a-jr medium prior to humidification.

2. A process for producing power utilizing a combustion gas turbine, comprising: (a) compressing air to a predetermined pressure; (b) after-cooling the compressed air against water, whereby the temperature of said water is increased and the temperature of said compressed air is decreased; (c) contacting said compressed air with said heated water in a multistage counter-current operation to humidify said compressed air and provide water vapor as thermal diluent for combustion in said turbine; (d) said step of contacting the compressed air with heated water being preceded by heat rejection from the cycle by pre-cooling the compressed air. -19- 210934

3. A process for producing power utilizing a combustion gas turbine; comprising: (a) compressing air to a predetermined pressure; (b) the said compression step being performed with inter-stage cooling against water and using the heated waber in the humidification of the compressed air; (c) contacting the compressed air with said heated water in a multistage counter-current operation for the humidification operation; (d) said step of inter-cooling during the compression including heat rejection from the cycle.

4. A process for producing power utilizing a combustion gas turbine, comprising: (a) compressing air to a pre determined pressure; {b) the said compression step being performed with inter-stage cooling against wdter and using the heated water in the humidification of the compressed air; (c) contacting the compressed air with said ' < heated water in a multistage counter-current operation for the humidification operation; (d) burning fuel together with the humidified compressed air; (e) driving a gas turbine for production of power; (f) said step of inter-cooling during compression including heat rejection from circulating water prior to its use for intercoolihg. -20- 210934

5. A process for producing power utilizing a combustion gas turbine, comprising: (a) comp-assing air to a predetermined pressure; (b) the said compression step being performed with aftercooling against water and using the heated water in the humidification of the compressed air; heated water in a multistage counter-current operation for the humidification operation; (d) burning the fuel together with the humidified "compressed air; ' . (f) said step of after-cooling after compression including heat rejection from circulating water prior to its use for after-cooiing.

6. The processes of claims 1, 2, 3, 4 or 5 in which the humidified compressed medium is preheated against the gas turbine exhaust ahd then mixed with the fuel for combustion.

7. The processes of claims 1, 2, 3, 4, 5 or 6 in which water heated in the gas turbine exhaust is used in the humidification operation.

8. The processes of claims 1, 2, 3, 4, 5, 6 or 7 in which water heated by sources external to the power cycle is used in the humidification operation. (c) contacting the compressed airJwith-said (e) driving a gas turbine for production- of power; -21- 210934

/ 9. The process of claims 7 and 8 in which the overall I compression ratio is between 6:1 and 34:1.

10. The process of claim 8 in which the water vapor content of the humidified compressed air is between 0.26:'and 0.5 pounds per pound of dry gaseous medium.

11. The process of claims 1 through 5 in which the fuel is preheated by the gas turbine exhaust.

12. Tha process of claims 1 through 5 iri which a fuel I gas is humidified using water heated in the intercooler, after icooler and gas turbine exhaust.

13. ,A process as claimed in any one of claims 1 to 12 when performed substantially as hereinbefore described with reference to any one of the accompanying drawings.

14. Power produced by a process as claimed in any one of the pouceding claims. dated thisday of odtb&t A. J. PAKKJi SON per agents for the applicants fC " H0CTO&'.;: ■'y -22-