WO1996036793A1 - Systeme de turbine a gaz a injection de vapeur equipe d'un dispositif de compression de vapeur - Google Patents

Systeme de turbine a gaz a injection de vapeur equipe d'un dispositif de compression de vapeur Download PDF

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Publication number
WO1996036793A1
WO1996036793A1 PCT/US1996/003008 US9603008W WO9636793A1 WO 1996036793 A1 WO1996036793 A1 WO 1996036793A1 US 9603008 W US9603008 W US 9603008W WO 9636793 A1 WO9636793 A1 WO 9636793A1
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WO
WIPO (PCT)
Prior art keywords
steam
pressure
flow
turbine
gas
Prior art date
Application number
PCT/US1996/003008
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English (en)
Inventor
Michael S. Briesch
Original Assignee
Westinghouse Electric Corporation
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Westinghouse Electric Corporation filed Critical Westinghouse Electric Corporation
Publication of WO1996036793A1 publication Critical patent/WO1996036793A1/fr

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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01K—STEAM ENGINE PLANTS; STEAM ACCUMULATORS; ENGINE PLANTS NOT OTHERWISE PROVIDED FOR; ENGINES USING SPECIAL WORKING FLUIDS OR CYCLES
    • F01K21/00—Steam engine plants not otherwise provided for
    • F01K21/04—Steam engine plants not otherwise provided for using mixtures of steam and gas; Plants generating or heating steam by bringing water or steam into direct contact with hot gas
    • F01K21/042—Steam engine plants not otherwise provided for using mixtures of steam and gas; Plants generating or heating steam by bringing water or steam into direct contact with hot gas pure steam being expanded in a motor somewhere in the plant
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01K—STEAM ENGINE PLANTS; STEAM ACCUMULATORS; ENGINE PLANTS NOT OTHERWISE PROVIDED FOR; ENGINES USING SPECIAL WORKING FLUIDS OR CYCLES
    • F01K21/00—Steam engine plants not otherwise provided for
    • F01K21/04—Steam engine plants not otherwise provided for using mixtures of steam and gas; Plants generating or heating steam by bringing water or steam into direct contact with hot gas
    • F01K21/047—Steam engine plants not otherwise provided for using mixtures of steam and gas; Plants generating or heating steam by bringing water or steam into direct contact with hot gas having at least one combustion gas turbine
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01K—STEAM ENGINE PLANTS; STEAM ACCUMULATORS; ENGINE PLANTS NOT OTHERWISE PROVIDED FOR; ENGINES USING SPECIAL WORKING FLUIDS OR CYCLES
    • F01K23/00—Plants characterised by more than one engine delivering power external to the plant, the engines being driven by different fluids
    • F01K23/02—Plants characterised by more than one engine delivering power external to the plant, the engines being driven by different fluids the engine cycles being thermally coupled
    • F01K23/06—Plants characterised by more than one engine delivering power external to the plant, the engines being driven by different fluids the engine cycles being thermally coupled combustion heat from one cycle heating the fluid in another cycle
    • F01K23/10—Plants characterised by more than one engine delivering power external to the plant, the engines being driven by different fluids the engine cycles being thermally coupled combustion heat from one cycle heating the fluid in another cycle with exhaust fluid of one cycle heating the fluid in another cycle
    • F01K23/106—Plants characterised by more than one engine delivering power external to the plant, the engines being driven by different fluids the engine cycles being thermally coupled combustion heat from one cycle heating the fluid in another cycle with exhaust fluid of one cycle heating the fluid in another cycle with water evaporated or preheated at different pressures in exhaust boiler
    • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02E—REDUCTION OF GREENHOUSE GAS [GHG] EMISSIONS, RELATED TO ENERGY GENERATION, TRANSMISSION OR DISTRIBUTION
    • Y02E20/00—Combustion technologies with mitigation potential
    • Y02E20/16—Combined cycle power plant [CCPP], or combined cycle gas turbine [CCGT]

Definitions

  • 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 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.
  • 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.
  • 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.
  • HRSG heat recovery steam generator
  • 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.
  • 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.
  • 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.
  • the gas turbine exhaust gas is directed to the highest pressure evaporator first, then each successive lower pressure level evaporator.
  • 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.
  • 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.
  • intermediate pressure steam may be generated in the HRSG and then injected into the gas turbine directly.
  • this object 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 .
  • 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.
  • 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.
  • 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 ' .
  • 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.
  • 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.
  • ambient air 26 is inducted into the compressor 7.
  • 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.
  • 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.
  • 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.
  • 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 expanded gas 31 from the turbine 11 is ducted to the HRSG 2.
  • the cooled exhaust gas 37 is ultimately vented to atmosphere.
  • 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.
  • 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 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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 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.
  • 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.
  • 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) .
  • the steam compressor 5 is driven by a shaft 8 coupled to the gas turbine shaft 9 through the electric generator 6.
  • 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.
  • 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) .
  • 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.
  • 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 P 3 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.
  • point D 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*- . :P 3 .
  • the power produced in the expansion portion of the cycle from C to D is a function of the expansion ratio P 3 : ⁇ -
  • 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 P 2 , which is approximately the operating pressure of the low pressure evaporator 18.
  • the steam 52 is then pressurized to pressure P 3 , 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.
  • 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 P 2 :P 3 .
  • the power produced in the expansion portion of the cycle from G to H is a function of the expansion ratio P 3 :P X .
  • the net power output of the steam injection cycle is the difference between the power consumption and the power production.
  • the pressure of the feed water 64 must be raised from essentially ambient pressure P x to the pressure P 2 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.
  • 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.
  • the specific volume of steam is 0.6542 m 3 /kg (10.48 ft 3 /lb)
  • the specific volume of water is 0.00107 m 3 /kg (0.0171 ft 3 /lb) .
  • the operating pressure P 2 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 P 2 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.
  • 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.

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  • Engineering & Computer Science (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Engine Equipment That Uses Special Cycles (AREA)

Abstract

L'invention concerne une centrale électrique à turbine à gaz équipée d'un générateur de vapeur à récupération de chaleur qui fournit de la vapeur à haute et basse pression. La vapeur à haute pression, que l'on surchauffe, est expansée dans une turbine à vapeur, ce qui permet d'obtenir une puissance sur arbre. La vapeur à basse pression est comprimée dans un dispositif de compression de vapeur jusqu'à une pression intermédiaire suffisamment élevée pour permettre une injection de vapeur de turbine à gaz, puis à les surchauffer à son tour. La turbine à gaz fournit un gaz comprimé qui subit ensuite une expansion dans une section turbine. La vapeur comprimée à une pression intermédiaire et surchauffée est injectée dans la chambre de combustion de la turbine à gaz et subit ensuite une expansion dans la section turbine, au même titre que le gaz chaud, ce qui permet d'augmenter la puissance de sortie et l'efficacité de la turbine à gaz.
PCT/US1996/003008 1995-05-18 1996-03-06 Systeme de turbine a gaz a injection de vapeur equipe d'un dispositif de compression de vapeur WO1996036793A1 (fr)

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Application Number Priority Date Filing Date Title
US44434395A 1995-05-18 1995-05-18
US08/444,343 1995-05-18

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Cited By (6)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
RU2463467C1 (ru) * 2011-08-02 2012-10-10 Николай Павлович Иванников Газотурбогидравлическая силовая установка замкнутого цикла для водного транспорта
RU2463468C1 (ru) * 2011-08-02 2012-10-10 Николай Павлович Иванников Газотурбогидравлическая установка замкнутого цикла
EP2957733A1 (fr) * 2014-06-18 2015-12-23 Alstom Technology Ltd Procédé permettant d'augmenter la puissance d'une centrale électrique à cycle combiné et centrale électrique à cycle combiné pour la réalisation de ce procédé
RU167924U1 (ru) * 2016-10-03 2017-01-12 Федеральное государственное бюджетное образовательное учреждение высшего образования "Кубанский государственный технологический университет" (ФГБОУ ВО "КубГТУ") Бинарная парогазовая установка
RU168003U1 (ru) * 2016-10-03 2017-01-16 Федеральное государственное бюджетное образовательное учреждение высшего образования "Кубанский государственный технологический университет" (ФГБОУ ВО "КубГТУ") Бинарная парогазовая установка
CN108361086A (zh) * 2018-02-08 2018-08-03 西安交通大学 一种节能热电解耦系统及运行方法

Citations (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
GB2097476A (en) * 1981-04-27 1982-11-03 Exxon Production Research Co A method for using residue gas in gas turbines
DE3331153A1 (de) * 1983-08-30 1985-03-14 Brown, Boveri & Cie Ag, 6800 Mannheim Gasturbinenanlage fuer offenen prozess
EP0444913A1 (fr) * 1990-02-27 1991-09-04 Turbine Developments Aktiengesellschaft Turbine à gaz
DE4321081A1 (de) * 1993-06-24 1995-01-05 Siemens Ag Verfahren zum Betreiben einer Gas- und Dampfturbinenanlage sowie danach arbeitende GuD-Anlage
EP0676532A1 (fr) * 1994-04-08 1995-10-11 Westinghouse Electric Corporation Système de turbine à gaz avec injection de vapeur et avec une turbine à vapeur à haute pression

Patent Citations (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
GB2097476A (en) * 1981-04-27 1982-11-03 Exxon Production Research Co A method for using residue gas in gas turbines
DE3331153A1 (de) * 1983-08-30 1985-03-14 Brown, Boveri & Cie Ag, 6800 Mannheim Gasturbinenanlage fuer offenen prozess
EP0444913A1 (fr) * 1990-02-27 1991-09-04 Turbine Developments Aktiengesellschaft Turbine à gaz
DE4321081A1 (de) * 1993-06-24 1995-01-05 Siemens Ag Verfahren zum Betreiben einer Gas- und Dampfturbinenanlage sowie danach arbeitende GuD-Anlage
EP0676532A1 (fr) * 1994-04-08 1995-10-11 Westinghouse Electric Corporation Système de turbine à gaz avec injection de vapeur et avec une turbine à vapeur à haute pression

Cited By (8)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
RU2463467C1 (ru) * 2011-08-02 2012-10-10 Николай Павлович Иванников Газотурбогидравлическая силовая установка замкнутого цикла для водного транспорта
RU2463468C1 (ru) * 2011-08-02 2012-10-10 Николай Павлович Иванников Газотурбогидравлическая установка замкнутого цикла
EP2957733A1 (fr) * 2014-06-18 2015-12-23 Alstom Technology Ltd Procédé permettant d'augmenter la puissance d'une centrale électrique à cycle combiné et centrale électrique à cycle combiné pour la réalisation de ce procédé
EP2957731A1 (fr) * 2014-06-18 2015-12-23 Alstom Technology Ltd Procédé pour augmenter la puissance d'une centrale électrique à cycle combiné et centrale électrique à cycle combiné pour l'exécution et la réalisation de ce procédé
CN105201575A (zh) * 2014-06-18 2015-12-30 阿尔斯通技术有限公司 提高联合循环功率装置的功率的方法和实行该方法的装置
RU167924U1 (ru) * 2016-10-03 2017-01-12 Федеральное государственное бюджетное образовательное учреждение высшего образования "Кубанский государственный технологический университет" (ФГБОУ ВО "КубГТУ") Бинарная парогазовая установка
RU168003U1 (ru) * 2016-10-03 2017-01-16 Федеральное государственное бюджетное образовательное учреждение высшего образования "Кубанский государственный технологический университет" (ФГБОУ ВО "КубГТУ") Бинарная парогазовая установка
CN108361086A (zh) * 2018-02-08 2018-08-03 西安交通大学 一种节能热电解耦系统及运行方法

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