WO1996036793A1

WO1996036793A1 - Steam injected gas turbine system with steam compressor

Info

Publication number: WO1996036793A1
Application number: PCT/US1996/003008
Authority: WO
Inventors: Michael S. Briesch
Original assignee: Westinghouse Electric Corporation
Priority date: 1995-05-18
Filing date: 1996-03-06
Publication date: 1996-11-21

Abstract

A gas turbine power plant having a heat recovery steam generator that generates steam at high and low pressures. The high pressure steam, which is superheated, is expanded in a steam turbine, thereby producing shaft power. The low pressure steam is pressurized in a steam compressor to an intermediate pressure, which is sufficiently high to allow the steam to be injected into the gas turbine, and then superheated. The gas turbine produces hot compressed gas and then is expanded in a turbine section. The superheated intermediate pressure steam is injected into the combustor of the gas turbine and then expanded in its turbine portion along with the hot gas, thereby increasing the power output and efficiency of the gas turbine.

Description

STEAM INJECTED GAS TURBINE SYSTEM WITH STEAM COMPRESSOR

BACKGROUND OF THE INVENTION

The present invention relates to a gas turbine power plant utilizing steam injection in conjunction with a steam compressor. More specifically, the present invention relates to such a gas turbine power plant in which steam is generated in a heat recovery steam generator at low pressure and then pressurized prior to introduction into the gas turbine.

The low capital cost, short lead times and flexibility of gas turbine-based power plants make them particularly attractive to electrical utilities as a means for generating electrical power. Unfortunately, the inefficiency of a gas turbine standing alone, referred to as a simple cycle system, is relatively low compared to conventional fired boiler steam turbine systems.

The major source of this inefficiency is inherent in the Brayton cycle on which the gas turbine operates. The ideal Brayton cycle operates in three phases - first, work is performed on the fluid (air in the case of a gas turbine) by isentropic compression in a compressor; second, heat is added to the fluid isobarically in a combustor; and, third, the hot compressed fluid is isentropically expanded back down to its initial pressure in the turbine. During the expansion phase much of the energy imparted to the fluid as a result of the compression and heating is recovered in the form of useful work. However, a significant portion of the energy remains in a relatively high-temperature, low-pressure form which, as a practical matter, cannot be recovered by further expansion in the turbine. In a simple cycle system this energy is lost from the cycle when the gas exhausting from the gas turbine is vented to atmosphere.

Consequently, substantial effort has been expended in developing methods for recovering the energy available in the gas exhausting from a gas turbine. One of the most successful methods involves the transfer of sensible heat from the hot exhaust gas to pressurized feed water in a heat recovery steam generator (hereinafter HRSG) . The HRSG generates steam that is expanded in a steam turbine, thereby producing additional rotating shaft power. Since steam turbines operate on the Rankine cycle, rather than the Brayton cycle, power plants employing such a heat recovery method are termed combined cycle power plants .

Typically, a HRSG is comprised of a large duct through which the exhaust gas flows . The duct encloses banks of tubes through which the water/steam flows and over which the gas turbine exhaust gas flows. The surfaces of the tubes provide heat transfer surfaces. There are three basic components in which heat is transferred in a typical HRSG, each comprised of a bundle of tubes: an economizer in which the feed water is heated to near-saturation temperature; an evaporator in which the water heated in the economizer is converted to steam; and a superheater in which the temperature of the saturated steam from the evaporator is raised into the superheat region. In order to obtain maximum efficiency of the steam turbine, it is desirable to generate steam at a high temperature and pressure. However, the maximum pressure at which the steam can be generated is limited by the temperature of the exhaust gas flowing from the gas turbine, since the saturation temperature of water increases with its pressure and only the portion of the heat in the exhaust gas which is above the saturation temperature of the water in the evaporator can be used to generate steam. Hence, although increasing steam pressure increases steam turbine efficiency, it also reduces the quantity of the steam generated and, therefore, the power output.

One approach to maximizing heat recovery by steam generation involves the use of a HRSG that generates steam at multiple pressure levels by employing a separate evaporator at each pressure level. The steam generated at each pressure level is then inducted into the appropriate stage of the steam turbine. In this approach, the gas turbine exhaust gas is directed to the highest pressure evaporator first, then each successive lower pressure level evaporator. Thus, although the temperature of the gas entering the evaporator decreases at each successive pressure level, the saturation pressure (and, hence, saturation temperature of the pressurized water) in each successive evaporator is also reduced, so that additional steam may be produced at each pressure level. Injecting steam into the combustor of a gas turbine has sometimes been used to reduce the NOx generated as a result of the combustion of fuel, or to augment the power output of the gas turbine. In a combined cycle power plant, steam injection for the gas turbine has been accomplished by generating high pressure steam for the steam turbine and then extracting a portion of the steam, at an intermediate pressure, mid-way through the steam turbine and injecting the intermediate pressure steam into the gas turbine combustor. Alternatively, intermediate pressure steam may be generated in the HRSG and then injected into the gas turbine directly.

However, neither of these approaches is optimal since neither facilitate the recovery of low temperature heat from the exhaust gas -- that is, the heat remaining in the exhaust gas after the high temperature heat has been recovered in the generation of high pressure steam. Low temperature heat can only be recovered by generating low pressure steam. Although low pressure steam can be inducted into an intermediate stage of the steam turbine, the power produced by such expansion is relatively small. Also, this approach requires increasing the size of the low pressure steam turbine. Low pressure steam cannot be injected into the gas turbine combustor since modern gas turbines frequently operate at pressures of approximately 1400 to 2000 kPa (200 to 300 psia) . Thus, neither of the approaches heretofore taught by the art -- directing low pressure steam from the HRSG to an intermediate stage of the steam turbine for steam induction or injecting intermediate pressure steam directly from the HRSG into the gas turbine -- facilitates optimum recovery of exhaust heat from the gas turbine exhaust. It is therefore desirable to provide a method and apparatus for recovering heat from the exhaust of a gas turbine by generating steam at low pressure and then injecting that steam into the gas turbine.

SUMMARY OF THE INVENTION Accordingly, it is the general object of the current invention to provide a method and apparatus for recovering heat from the exhaust of a gas turbine by generating steam at low pressure and then injecting that steam into the gas turbine. Briefly, this object, as well as other objects of the current invention, is accomplished in a method of generating power, comprising the steps of (i) producing power in a first rotating shaft by introducing a hot compressed gas at a first pressure into a first turbine for flow therethrough, the hot compressed gas expanding in the turbine so as to produce an expanded gas, (ii) generating a first flow of steam at a second pressure by transferring heat from the expanded gas, the second pressure being less than the first pressure, (iii) pressurizing the first flow of steam to a third pressure, the third pressure being greater than the first pressure, and (iv) introducing the first flow of steam after the pressurizing thereof into the first turbine along with the hot compressed gas, thereby increasing the power produced in the first rotating shaft .

In a preferred embodiment, the second pressure is less than approximately 700 kPa (100 psia) and the first pressure is at least approximately 1380 kPa (200 psia) . The current invention also encompasses an apparatus for generating power, comprising (i) a steam generator having means for generating a first flow of steam at a first pressure by absorbing heat from a flow of expanded gas, and means for generating a second flow of steam at a second pressure by further absorbing heat from the flow of expanded gas, the first pressure being higher than the second pressure, (ii) means for pressurizing the second flow of steam to a third pressure, the third pressure being less than the first pressure and greater than the second pressure, (iii) means for producing compressed air, (iv) a combustor for heating a mixture of the compressed air and the second flow of steam, thereby producing a moisture laden hot gas, and (v) first turbine means for expanding the moisture laden hot gas.

BRIEF DESCRIPTION OF THE DRAWINGS Figure 1 is a schematic diagram of a gas turbine power plant according to the current invention.

Figures 2 (a) and (b) are idealized temperature versus entropy diagrams for the cycles associated with the compressed air and injected steam in the power plant shown in Figure 1.

DESCRIPTION OF THE PREFERRED EMBODIMENT Referring to the drawings, there is shown in Figure 1 a schematic diagram of a gas turbine power plant according to the current invention. The major components of the power plant include a gas turbine 1, a HRSG 2, a steam turbine 3, a condenser 4, a steam compressor 5, and electrical generators 6 and 6' . The gas turbine 1 is comprised of an air compressor 7, a combustor 10, and a turbine 11 that is connected to the compressor by means of a rotating shaft 9 ' . As is conventional, the compressor 7 may include a plurality of alternating rows of rotating blades and stationary vanes. The rotating blades are affixed to discs mounted on the portion of the rotor shaft that extends through the compressor 7 and the stationary vanes are affixed to a casing enclosing the compressor components. Similarly, the turbine 11 may include a plurality of alternating rows of rotating blades and stationary vanes. The rotating blades are affixed to discs that form the portion of the rotor shaft that extends through the turbine 11 and the stationary vanes are affixed to a casing that encloses the turbine components. The combustor 10 may be comprised of a plurality of combustor baskets, each of which forms a combustion chamber, and associated fuel nozzles.

In the gas turbine 1, ambient air 26 is inducted into the compressor 7. In the preferred embodiment, the compressor 7 increases the pressure of the air 26 into approximately the 1380-1720 kPa (200-250 psia) range. The compressed air 27 from the compressor 6 is then heated in a combustor 10 by burning a fuel 28. The fuel 28 may be in a liquid or gaseous form, and is typically No. 2 distillate oil or natural gas. In addition to the fuel 28, superheated steam 50, generated as discussed below, is also injected into the combustor 10. The steam injection may be accomplished by mixing the steam 50 into the compressed air 27 prior to its introduction into the combustor 10 -- for example, by introducing it into the fuel nozzle. Alternatively, the steam 50 may be injected directly into the combustion chamber of the combustor 10 so that it mixes with the products of combustion of the fuel and compressed air. In the preferred embodiment, sufficient fuel 28 is burned in the combustor 10 to heat the hot gas/steam mixture 30 discharged from the combustor into approximately the 1310-1370°C (2400-2500°F) temperature range.

The hot moisture laden gas 30 discharged from the combustor 10, which in the preferred embodiment is in the approximately the 1380-1720 kPa (200-250 psia) range, is then introduced into the turbine section 11 of the gas turbine 1 and expanded therein to essentially atmospheric pressure. In so doing, the temperature of the hot gas 30 is considerably reduced. In the preferred embodiment, the temperature of the expanded gas 31 exhausted from the turbine 11 is in the range of approximately 540-590°C (1000-1100°F) . This expansion produces power in the turbine shaft 9' that drives the compressor 7. The excess power not consumed by the compressor 7 drives the electric generator 6, via shaft 9, thereby generating electrical power.

The expanded gas 31 from the turbine 11 is ducted to the HRSG 2. After leaving the HRSG 2, the cooled exhaust gas 37 is ultimately vented to atmosphere. In the preferred embodiment, the heat transfer in the HRSG 2 reduces the temperature of the exhaust gas 37 into approximately the 90-150°C (200-300°F) temperature range. According to the current invention, the HRSG 2 receives feed water 64 and, by transferring heat to it from the expanded gas 31, converts the feed water into steam at . two pressure levels. The steam generated at high pressure is expanded in the steam turbine 3. The steam generated at low pressure is further pressurized in the steam compressor 5 and then injected into the combustor 10 of the gas turbine 1, as previously discussed.

The method and apparatus for recovering heat from the expanded gas 31 in the HRSG 2, according to the preferred embodiment of the current invention, will now be explained in detail. The HRSG 2 is comprised of a duct 12, which encloses various heat transfer sections (i.e., superheaters, evaporators and economizers) , and an exhaust stack 23. As is conventional, the heat transfer sections may be comprised of multiple rows of finned heat transfer tubes, the number of rows being determined by the amount of heat transfer surface area desired. The water/steam flows within the tubes and the expanded gas 31 flows over the outside surfaces of the tubes.

The heat transfer sections include intermediate and high pressure superheaters 13 and 14, respectively, high and low pressure evaporators 17 and 18, respectively, which may be of the forced or natural circulation type, and high and low pressure economizers 19 and 20, respectively.

The heat transfer sections are arranged to optimize the recovery of heat from the expanded gas 31. Thus, the expanded gas 31 flows first over the intermediate pressure superheater 13, then over the high pressure superheater 14, then over the high pressure evaporator 17, then over the high pressure economizer 19, then over the low pressure evaporator 18, and, finally, over the low pressure economizer 20.

In operation, water 63 from a water supply 38 is combined with condensate 62 from the condenser 4 to form feed water 64 for the HRSG 2. The feed water 64 is pressurized to a relatively low pressure (i.e.,less than approximately 700 kPa (100 psia)) by a pump 25. The pressurized feed water 65 is then directed to the low pressure economizer 20.

The low pressure economizer 20 has sufficient heat transfer surface area to heat the feed water 65 to close to its saturation temperature by the transfer of heat from the expanded gas 36 flowing over the economizer. In order to maintain maximum heat recovery, it is desirable to transfer as much heat as possible in the economizer. However, the temperature of the water must remain below its saturation temperature to avoid steam formation, which impedes the flow of water through the economizer. In the preferred embodiment the water in the low pressure economizer 20 is heated to approximately 5°C (10°F) below its saturation temperature. The heated feed water discharged from the low pressure economizer 20 is then split into first and second streams 66 and 68, respectively. As discussed below, the first feed water stream 66 is used to generate high pressure steam 56 for expansion in the steam turbine 3, while the second feed water stream 68 is used to generate low pressure steam 54 that, after pressurization, is injection into the gas turbine 1. In the preferred embodiment, the ratio of low pressure steam 54 to high pressure steam 56 is at least approximately 0.05.

The first stream of heated feed water 66 from the low pressure economizer 20 is further pressurized by pump 24 to a pressure in excess of 6900 kPa (1000 psia) , and preferably at least 13,800 kPa (2000 psia) . The further pressurized feed water 70 is then directed to the high pressure economizer 19. The high pressure economizer 19 has sufficient heat transfer surface area to heat the feed water 70 to close to its saturation temperature by the transfer of heat from the expanded gas 34 flowing over the economizer. In the preferred embodiment, the water in the high pressure economizer 19 is heated to approximately 11°C (20°F) below the saturation temperature of the steam in the high pressure drum 21.

The heated feed water 72 from the high pressure economizer is then directed to the high pressure steam drum 21, from which it is circulated through the high pressure evaporator 17 and converted to high pressure saturated steam 56 by the transfer of heat from the expanded gas 33 flowing over the evaporator. In the preferred embodiment, the pressure in the high pressure evaporator 17 is maintained above 6890 kPa (1000 psia) , and preferably in the range of approximately 9650-11,000 kPa (1400-1600 psia) . Generating steam at such high pressures optimizes the performance of the steam turbine 3.

From the steam drum 21, the high pressure saturated steam 56 is superheated in the high pressure superheater 14 by the transfer of heat from the expanded gas 32 flowing over this superheater. In the preferred embodiment, the high pressure superheater 14 has sufficient heat transfer surface area to superheat the steam 60 into the approximately 480-570°C (900-1050°F) temperature range. The steam 60 from the high pressure superheater 14 is then directed to the steam turbine 3.

In the steam turbine 3, the superheated high pressure steam 60 is expanded down to the pressure of the condenser 4, which preferably operates at a slight vacuum. In so doing, the steam turbine 3 produces power that drives the electric generator 6' so as to increase the electrical power output from the power plant. The low pressure steam 61 discharged from the steam turbine 3 is then condensed in the condenser 4 and the condensate 62 is returned to the HRSG 2, as previously discussed.

The second feed water stream 68 from the low pressure economizer 20 is directed to the low pressure steam drum 22, from which it is circulated through the low pressure evaporator 18 and converted to low pressure saturated steam 54 by the transfer of heat from the expanded gas 35 flowing over the evaporator. In the preferred embodiment, the pressure in the low pressure evaporator 18 is maintained as low as possible so as to maximize the amount of heat in the expanded gas 35 that can be recovered. The lower the pressure in the evaporator 18, the lower the saturation temperature of the water in the evaporator. The lower the saturation temperature, the greater the amount of heat in the expanded gas that is above the saturation temperature and, therefore, capable of being recovered by the generation of steam.

Therefore, preferably, the pressure in the low pressure evaporator 18 is maintained at less than approximately 700 kPa (100 psia) , and more preferably only approximately 280 kPa (40 psia) . Maintaining such low pressure in the evaporator 18 allows the maximum amount of heat to be recovered from the exhaust gas 35 since the saturation temperature at about 280 kPa (40 psia) is only about 130°C (270°F) . Consequently, if sufficient heat transfer surface area is provided in the low pressure evaporator 18, almost all of the remaining energy in the exhaust gas 35 above 130°C (270°F) can recovered in the low pressure evaporator. In the low pressure economizer 20 even lower temperature heat can be recovered from the exhaust gas 36 discharging from the low pressure evaporator 18.

Thus, according to the current invention, the ordering of the various components of the HRSG 2 with respect to the flow of the expanded gas has been chosen so that heat may be extracted by each component, even though the temperature of the expanded gas 31 is decreasing as it flows through the HRSG.

According to an important aspect of the current invention, the low pressure saturated steam 54 from the steam drum 22 is directed to the steam compressor 5, where its pressure is raised to a level somewhat above that of the compressed air 27 directed to the combustor 10 -- that is, to approximately 1700-2100 kPa (250-300 psia) in the preferred embodiment. This further pressurization is required to raise the pressure of the steam 50 sufficiently to allow it to be injected into the combustor 10 taking into account the small pressure drop associated with the regulation of the steam flow into the combustor, for example, by means of a flow control valve (not shown) . In the embodiment shown in Fig. 1, the steam compressor 5 is driven by a shaft 8 coupled to the gas turbine shaft 9 through the electric generator 6. However, the steam compressor could also be driven by an electric motor receiving current from either of the electric generators 6 and 6' , or by a shaft coupled to the steam turbine 3.

From the steam compressor 5, the saturated intermediate pressure steam 52 is then superheated in the intermediate pressure superheater 13 by the transfer of heat from the expanded gas 31 flowing over this superheater (preferably, the pressure drop experienced by the steam flowing through the intermediate pressure superheater 13 is only about 100 kPa (15 psi) or less) . In the preferred embodiment, the intermediate pressure superheater 13 has sufficient heat transfer surface area to superheat the steam 50 to within approximately 50°F of the temperature of the exhaust gas 31. A temperature versus entropy diagram for the air/gas cycle shown in Fig. 1 is shown in Fig. 2(a), with the temperature being denoted T and the entropy being denoted S. As can be seen, the air 26 inducted into the compressor at point A in the cycle is at pressure ?_{l r} which is essentially ambient pressure. Point B reflects the pressurization of the air in the compressor 7 to the operating pressure P₃ of the combustor 10, which is essentially the maximum pressure for the air/gas cycle. Heat is added to the compressed air 27 by the combustion of the fuel 28 in the combustor 10, thereby isobarically raising the temperature of the hot gas produced thereby to point C, which represents the temperature of the hot gas 31 entering the turbine section 11 of the gas turbine. In the turbine section 11, the hot gas 31 is expanded down to essentially atmospheric pressure again, denoted by point D. The power consumed in the compression portion of this cycle -- that is, from A to B -- is a function of the compression ratio P*-_.:P₃. Similarly, the power produced in the expansion portion of the cycle from C to D is a function of the expansion ratio P₃: _ι- The net power output from the air/gas cycle is the difference between the power consumption and the power production.

The temperature entropy diagram for the injection steam 50 is shown in Figure 2(b) . The saturated steam 54 enters the steam compressor 5 at point E at pressure P₂, which is approximately the operating pressure of the low pressure evaporator 18. The steam 52 is then pressurized to pressure P₃, at point F, in the steam compressor 5. Its temperature is raised from point F to point G by the intermediate pressure super heater 13 and by the combustion of the fuel 28 so that it is at the temperature of the hot gas 30 entering the turbine section 11. In the turbine 11, the steam 50 is expanded, along with the hot gas 30, down to essentially ambient pressure P_x at point H. The power consumed in the compression portion of this cycle from E to F is a function of the compression ratio P₂:P₃. Similarly, the power produced in the expansion portion of the cycle from G to H is a function of the expansion ratio P₃:P_X. The net power output of the steam injection cycle is the difference between the power consumption and the power production. Based on the foregoing, it can be seen that compared to the air/gas cycle, the steam injection cycle will achieve greater efficiency since the compression ratio (which determines the power consumption) is lower, whereas the expansion ratio (which determines the power production) is the same as for the air/gas cycle. Thus, the use of steam injection according to the current invention will not only result in the recovery of exhaust heat, it will improve the efficiency of the gas turbine -- the greater the amount of steam injection, the greater the improvement in efficiency.

It should be understood that the diagrams shown in Figures 2(a) and (b) are idealized. As will readily be appreciated by those skilled in the art, in reality, the compression and expansion are not isentropic, the heating is not isobaric, and there are various pressure losses in the system -- for example, the pressure of the hot gas 30 introduced into the turbine 11 is less than the pressure of the compressed air 27 discharged from the air compressor 7, the operating pressure of the low pressure evaporator 18 is greater than the pressure of the steam 54 introduced into the steam compressor, the pressure of the superheated steam 50 introduced into the gas turbine 1 will be less than the saturated steam discharged from steam compressor 5, etc. Nevertheless, proper design of the system components will result in these pressure drops being relatively small so that the principles of the current invention may be readily achieved. It should be pointed out that the pressure of the feed water 64 must be raised from essentially ambient pressure P_x to the pressure P₂ of the low pressure evaporator by the feed pump 25 so that the net power output from the steam injection cycle will be reduced by the power consumed by the pump 25. However, the pressurization of a liquid, such as the feed water 64, requires much less power consumption than the pressurization of a gas, such as the steam 54. This can be readily seen observing that the work of compression is a function of the integral of the term vdp, where v is the specific volume and dp is the differential pressure. At 40 psia, for example, the specific volume of steam is 0.6542 m³/kg (10.48 ft³/lb) , whereas the specific volume of water is 0.00107 m³/kg (0.0171 ft³/lb) . Thus, for a given pressure rise, the work required to pressurize the feed water 64 from P_x to P₂ is much less than would be required to similarly pressurize the steam 54.

Thus, arranging the steam cycle so that the pressurization from P-*_. to P₂ occurs in the liquid phase, leaving only the pressurization from P₂ to P₃ in the gas phase, results in an increase in efficiency. Of course, all of the pressurization to P₃ could be accomplished in the liquid phase by increasing the operating pressure of the low pressure evaporator 18 from P₂ to P₃, thereby further increasing the efficiency of the steam injection cycle. However, as previously discussed, this would increase the saturation pressure of the water in the low pressure evaporator 18 and dramatically decrease the steam generation rate, and therefore the steam injection rate, as well as the amount of heat recovered from the exhaust gas 31.

Consequently, according to the current invention, the operating pressure P₂ of the low pressure evaporator 18 is selected to optimize the overall efficiency of the power plant based on these two competing factors -- that is, (i) the higher the operating pressure P₂ of the low pressure evaporator, the larger the amount of compression achieved in the liquid phase and, therefore, the lower the work of compression associated with the steam injection cycle and the greater the increase in efficiency per unit of steam mass flow that results from steam injection, and (ii) the lower the operating pressure of the low pressure evaporator, the greater the mass flow of the steam 50 injected into the gas turbine 1 and, therefore, the higher the power output of the gas turbine and the greater the heat recovery by the HRSG 2. As previously discussed, the inventor has found that, depending on the specific cycle parameters, the optimum operating pressure of the low pressure evaporator 18 is less than about 700 kPa (100 psia) , and preferably about 280 kPa (40 psia) . The present invention may be embodied in other specific forms without departing from the spirit or essential attributes thereof and, accordingly, reference should be made to the appended claims, rather than to the foregoing specification, as indicating the scope of the invention.

Claims

CLAIMS:

1. A method of generating power, comprising the steps of: a) producing power in a first rotating shaft by introducing a hot compressed gas at a first pressure into a first turbine for flow therethrough, said hot compressed gas expanding in said turbine so as to produce an expanded gas; b) generating a first flow of steam at a second pressure by transferring heat from said expanded gas, said second pressure being less than said first pressure; c) pressurizing said first flow of steam to a third pressure, said third pressure being greater than said first pressure; and d) introducing said first flow of steam after said pressurizing thereof into said first turbine along with said hot compressed gas, thereby increasing the power produced in said first rotating shaft.

2. The method according to claim 1, wherein said second pressure is less than approximately 700 kPa and said first pressure is at least approximately 1380 kPa.

3. The method according to claim 1, further comprising the step of generating a second flow of steam at a fourth pressure by transferring heat from said expanded gas, said fourth pressure being greater than said second pressure.

4. The method according to claim 3, further comprising the step of expanding said second flow of steam in a second turbine having a second rotating shaft so as to produce additional shaft power.

5. The method according to claim 4, wherein said first turbine is a portion of a gas turbine, and said second turbine is a steam turbine.

6. The method according to claim 1, wherein the step of pressurizing said first flow of steam comprises flowing said first flow of steam through a compressor.

7. The method according to claim 6, wherein said compressor is driven by power produced in said first rotating shaft.

8. The method according to claim 1, further comprising the steps of producing said hot compressed gas to be expanded in said first turbine by compressing air and then burning a fuel in said compressed air in a combustor.

9. The method according to claim 8, wherein the step of introducing said first flow of steam into said first turbine comprises injecting said first flow of steam into said combustor.

10. The method according to claim 8, wherein the step of introducing said first flow of steam into said first turbine comprises injecting said first flow of steam into said compressed air prior to said burning of said fuel therein.

11. The method according to claim 1, further comprising the step of superheating said first flow of steam by transferring heat thereto prior to said introduction of said first flow of steam into said first turbine.

12. The method according to claim 1, wherein said second pressure is less than approximately 700 kPa.

13. The method according to claim 12, wherein said second pressure is approximately 280 kPa.

14. The method according to claim 12, wherein said third pressure is at least approximately 1700 kPa.

15. A method of generating power, comprising the steps of a) generating a first flow of steam at a pressure less than approximately 700 kPa by transferring heat from a flow of expanded hot gas to a flow of feed water; b) pressurizing said first flow of steam to at least approximately 1400 kPa; and c) introducing said first flow of steam after said pressurizing thereof into a flow of hot pressurized gas so as to produce a mixture thereof and expanding said mixture in a first turbine so as to produce shaft power and said expanded gas.

16. The method according to claim 15, further comprising the steps of : a) compressing air so to produce compressed air; and b) heating said compressed air by burning a fuel therein so as to produce said flow of hot pressurized gas.

17. An apparatus for generating power, comprising: a) a steam generator including (i) means for generating a first flow of steam at a first pressure by absorbing heat from a flow of expanded gas, (ii) means for generating a second flow of steam at a second pressure by further absorbing heat from said flow of expanded gas, said first pressure being higher than said second pressure; b) means for pressurizing said second flow of steam to a third pressure, said third pressure being less than said first pressure and greater than said second pressure; c) means for producing compressed air; d) combustor means for heating a mixture said compressed air and said pressurized second flow of steam, thereby producing a moisture laden hot gas; and e) first turbine means for expanding said moisture laden hot gas.

18. The apparatus according to claim 17, further comprising second turbine means for expanding said first flow of steam.