EP0723626B1 - Cheng-zweistoffwärmekraftmaschine und verfahren zum ihren betrieb - Google Patents

Cheng-zweistoffwärmekraftmaschine und verfahren zum ihren betrieb Download PDF

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EP0723626B1
EP0723626B1 EP92921791A EP92921791A EP0723626B1 EP 0723626 B1 EP0723626 B1 EP 0723626B1 EP 92921791 A EP92921791 A EP 92921791A EP 92921791 A EP92921791 A EP 92921791A EP 0723626 B1 EP0723626 B1 EP 0723626B1
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turbine
working fluid
heat recovery
temperature
engine
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EP0723626A1 (de
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Dah Yu Cheng
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  • This invention relates to a Cheng cycle dual fluid heat engine and a method of operation thereof.
  • the present invention relates to a heat engine, and, more particularly, to the improvement in efficiency and output over and above the parallel regeneration Dual Fluid Cycle or Cheng Cycle in a number of prior United States patents (4128994, 3978661, 4248039, 4297841 and 4417438).
  • the Cheng Cycle is a heat regenerative gas turbine utilizing water-steam only in a heat recovery steam generator (HRSG) in a particular way that the steam is heated to as high a superheated temperature as possible (limited by the exhaust temperature of the turbine), and total heat recovery is maximized, limited only by heat transfer physics at the transition from the water to the evaporative phase of the boiler and the available gas stream temperature differences. This is known as the neck temperature in the HRSG industry. Patented at that condition, the Cheng Dual Fluid Cycle will reach a peak efficiency mixture ratio of liquid to gaseous working fluid, called XMIX peak. The anticipated efficiency gain and its simplicity of operation were cited in aforementioned patents, in literature, and through real operational data.
  • the Cheng Cycle was first commercialized using an Allison Model 501KH gas turbine, then a Kawasaki M1-ACC. Since 1984, ten Cheng Cycle 501KHs have been put into operation, and two Kawasaki Cheng Cycles are in operation in Japan. Numerous operational patents and variations to the turbine have been incorporated.
  • the Cheng Cycle system has enjoyed all the benefits cited, it also suffers from an oversight of the inventor, Dr. Dah Yu Cheng (the current inventor of the Advanced Cheng Cycle. ⁇ ACC ⁇ ).
  • the same gas turbine with a complicated steam turbine can be configured as a combined cycle.
  • it has a reputation of having a potentially slightly higher efficiency than the Cheng Cycle. This is due to the fact that when the Cheng Cycle was conceived in the 1970s, the majority of gas turbines were operating at a metallurgically tolerable turbine inlet temperature for stress and corrosion protection.
  • the Cheng Cycle was conceived at that time without consideration to blade cooling. The trend at that time was going towards ceramic turbines, even internally water cooled turbine blades.
  • Bleeding compressor air certainly is a loss to the gas turbine system, but the. ability to. increase the operating temperature of the turbine more than made up for that loss. This resulted in an increased compressor pressure ratio and also allowed the simple gas turbine cycle efficiency to increase. Up to 10% bleed air is used in advanced fighter gas turbines, with a turbine inlet temperature over 2750°F (1510°C). For industry, an efficiency increase of 1% is considered a major achievement.
  • the new G.E. Frame 7-F has a turbine inlet temperature of 2350°F (1288°C), but is still designed with a relative low pressure ratio so that the exhaust temperature can be high enough to improve the steam cycle part of the combined cycle. A combined cycle efficiency of over 50% have been claimed.
  • Document EP 319699 describes a different thermal cycle than the Advanced Cheng Cycle as the system requires a very high pressure boiler and a low pressure boiler to generate both the very high pressure steam and some low pressure steam.
  • the high pressure steam is cooled down by passing it through a steam turbine 32, which changes the enthalpy of the steam, then subsequently injected 34 into the combustor of the gas turbine.
  • the low pressure steam is used to be injected between the first stage turbine to compress air and the second stage turbine to create a back pressure for the first stage turbine to reduce its ability to produce power by the added steam into the combustion chamber and increase the ability of the second stage turbine to produce more work while cooling the hot working fluid exiting the first turbine to create a balance with the steam injection.
  • the high pressure scheme described in EP 319699 increases the boiling temperature of the heat recovery steam generator, therefore, restricting its ability to generate maximum heat recovery.
  • the low pressure steam addition described in EP 319699 does not enter into the combustion chamber and therefore does not participate in increasing the steam temperature by the combusted gas. As a result, turbine system performance efficiency is degraded.
  • This invention aims to overcome the limitations by an advancement to the Cheng Cycle, hereincalled the Advanced Cheng Cycle (ACC).
  • ACC Advanced Cheng Cycle
  • the hardware configuration of the ACC follows, to a great extent, the old design in valve location, control systems, startup and shutdown processes. A number of difficulties which were encountered with the original Cheng Cycle will also be improved upon by this disclosure.
  • a. Cheng cycle, dual-fluid heat engine comprising: a compressor for compressing a first working fluid, having a compressor outlet; a combustion chamber in fluid communication with the compressor outlet; a turbine having an inlet in fluid communication with the combustion chamber for performing work by expansion of working fluid, and a turbine exhaust; a heat recovery exchanger coupled to the turbine exhaust for heating thereby, having a heat recovery exchanger inlet and an outlet for heating a second working fluid; an injector, for introducing heated second working fluid from the heat recovery exchanger into the combustion chamber; a coolant inlet port for introducing coolant to at least one of turbine nozzles and blades in the turbine; control valve apparatus connected upstream of the injector for selectively throttling flow rate of second working fluid into the turbine, and control means for controlling the control valve apparatus such that: second working fluid injected into the turbine has a temperature less than or substantially equal to temperature of first working fluid at the compressor outlet; and heat recovery exchanger pinch temperature is minimized by maximizing heat recovery from the turbine
  • a method of operation of a Cheng cycle, dual fluid heat engine of the type having: a gas turbine engine including a compressor for compressing a first working fluid, having a compressor outlet, a combustion chamber in fluid communication with the compressor outlet, a turbine unit having an inlet in fluid communication with a combustion chamber for performing work by expansion of working fluid, and a turbine exhaust; a heat recovery steam generator coupled to the turbine exhaust for heating thereby, having a heat recovery inlet and a heat recovery outlet for heating a second working fluid; an injector for introducing heated second working fluid from the heat recovery steam generator into the combustion chamber; a coolant inlet port for introducing coolant to at least one of turbine nozzles and blades in the gas turbine; and control valve apparatus connected upstream of the injector for selectively throttling flow rate of second working fluid into the gas turbine, the method comprising: throttling second working fluid flow rate with the control valve apparatus in accordance with the following operating parameters: second working fluid temperature as injected into the turbine less than or substantially
  • the invention may be embodied in a method or an apparatus which offer several advantages including:
  • Figure 1 illustrates the Cheng Cycle as configured in a series of patents filed in the 1970's by the inventor.
  • Figure 2 illustrates the peak efficiency aspect of steam to air mixture ratio of the Cheng Cycle in the prior art.
  • Figure 3 illustrates the heat transfer limitation to reach maximum heat recovery and maximum superheat as claimed in the prior art.
  • Figure 4 is a temperature and entropy diagram describing cycle efficiencies of the Cheng Cycle as demonstrated in the prior art.
  • Figure 5 is a combined cycle temperature and entropy diagram using a gas turbine and steam turbine cycle for heat regeneration.
  • Figure 6 illustrates the new Advanced Cheng Cycle configuration, including new control valve locations.
  • Figure 7 illustrates a typical bleed air cooled turbine blade.
  • Figure 8 illustrates the Advanced Cheng Cycle with increased pressure ratio.
  • Figure 9 illustrates a comparison of efficiency and the location of XMIX for peak efficiency between the prior art (Cheng Cycle) and the Advanced Cheng Cycle.
  • Figure 10 illustrates the heat transfer differences in the heat recovery boiler and temperature profiles between the prior art (Cheng Cycle) and the Advanced Cheng Cycle.
  • Figure 11 illustrates the Advanced Cheng Cycle optimum description with increased turbine inlet temperature and increased pressure ratio.
  • Figure 12 illustrates the ability for the Advanced Cheng Cycle to increase combustion ratio through turbine matching and compressor map.
  • Figure 1 illustrates the embodiment of the Cheng Cycle according to prior art patents.
  • the configuration indicates the gas turbine has a compressor 10 linked to a turbine 13 by shaft and output to a load.
  • the air intake to compressed air through 1 is being compressed and discharged at 2.
  • the compressed air enters a combustion chamber 12.
  • Fuel is entering the combustion chamber through 11, and steam comes from the heat recovery steam generator (HRSG) to 3.
  • HRSG heat recovery steam generator
  • the mixing of the combusted air and steam reaches a predetermined turbine inlet temperature, then discharges at 4 through turbine 13, exiting at the turbine 5.
  • Exhaust gas then passes through the heat recovery steam generator, which is divided into two parts, a superheater 14, and a water to steam generator 15.
  • a duct burning capability is not depicted here, normally located at 6.
  • the remainder of the heat is recovered by the unit evaporator 15 and exits at 17.
  • the exhaust gas has the option of going through a cleanup or condensing unit 20, then to the atmosphere.
  • Water can be recovered through 20 or can be totally used as a makeup entering or mixing with 19.
  • Water is compressed to a high pressure through a pump 18.
  • the pump exit goes into the steam generator 9 and the evaporator is controlling the steam flow by two valves 16 into the superheater, or to a steam user as a cogeneration unit 17. If used for power generation only, 17 is no longer needed.
  • Figure 2 is an efficiency versus steam to air ratio defined as XMIX such that for a given turbine inlet temperature 30, the efficiency will reach peak, then drop off with even more steam injection. This peak efficiency is the nature of the heat recovery as described in the Cheng Cycle patents and the most important claims, and can be seen in Figure 3.
  • FIG 3 is the exhaust temperature profile and the heat transfer to the HRSG by converting water into high temperature steam.
  • the length indicated is the part length of the heat exchanger in HRSG.
  • the exhaust gas entering the heat exchanger as described in Figure 1 is depicted here as temperature T5.
  • the temperature is dropped to T6 by superheating the steam from T8 to T3.
  • the temperature continuously drops on the exhaust side from T6 to T7 and continuous down to the exhaust T7.
  • the water is entering at T9, reaches the evaporation point for given pressure at T7, and the evaporation is taking place according to the profile 40.
  • the exhaust profile is 41.
  • the temperature difference between the boiler in temperature and exhaust temperature at the smallest point indicated by (T 7 , - T8) is referring to a Delta T pinch or neck.
  • the temperature difference between T5 and T3 is referred to as the Delta T maximum .
  • the maximum enthalpy and the maximum heat recovery in the prior art requires that the upper pinch point Delta T max and T7' minus T8 to be minimized. At that point the peak XMIX is reached.
  • FIG 4 is a temperature entropy diagram in the industry referred to as the T-S diagram.
  • S represents the entropy.
  • the box 50 is referred to as a Carnot Cycle box, and the area indicated inside of the box is the gas turbine cycle diagram such that it consists of four sides.
  • the compression side is indicated by the temperature rise with increased entropy S from T1 to T2.
  • the combustion is taking place to raise the temperature from T2 to T4.
  • the expansion through the turbine is dropping the temperature from T4 to T5.
  • the temperature is dropping through a profile on the bottom to complete the cycle.
  • the square box 50 represents the temperature T4 and the temperature T1 bound by the turbine inlet temperature and ambient temperature of the gas turbine, which is defined as the Carnot Cycle of the gas turbine efficiency.
  • the Carnot Cycle efficiency is usually defined as efficiency equal to 1 minus T1 divided by T4 such that the higher the T4, the higher the Carnot Cycle efficiency.
  • the Carnot Cycle has no indication of the participation by the entropy, therefore the width of the Carnot Cycle box does not enter into play to define the cycle efficiency.
  • the water side of the cycle starts from T9, reaches temperature T8, then the superheater temperature T3 is trying to fill the corner of the Carnot Cycle box. Steam is then further heated from T3 to T4, then expanded synergistically with the air to the same temperature T5. The heat from the steam is also recovered by the additional steam, indicated by the boundary of the box B'. The B' goes through the same cycle and will recover additional steam, B''. Therefore, the Cheng Cycle heat regeneration is a series of heat recovery cycles by steam with maximum heat recovery and maximum entropy to fill the Carnot Cycle box as tightly as possible as a means to increase its efficiency. The area increase between A plus B then B' and B" and so on is the potential of the power output increase of the Cheng Cycle. As was described in the prior art patent, the competing cycle usually is described as a combined cycle such that the combined cycles do not inject steam into the gas turbine, rather it goes through a separate steam cycle.
  • Such a cycle is depicted in Figure 5.
  • the carnot box 50 is the gas turbine part of the Carnot box; however, a steam cycle will occupy the bottom corner of the empty part of the carnot box below the gas turbine cycle A.
  • the steam cycle is indicated by area D.
  • the steam cycle D normally is a high pressure steam, compared to the Cheng Cycle, which is at a lower pressure. Therefore, the potential for power increase is not as great as with the Cheng Cycle.
  • FIG. 6 illustrates the Advanced Cheng Cycle system.
  • the Advanced Cheng Cycle system has a compressor 10 and a turbine 13 linked by shaft and output to the load. Air also comes in at 1, is discharged after the compression at 2 into the combustion chamber 12. Fuel is entering at 11. Steam from the heat recovery generator comes at 3, and the premix to a predetermined turbine inlet temperature is limited by the metallurgy temperature allowed at T4. The mixture gas then expands through the turbine 13 and exhaust at turbine 5. The exhaust gas goes into a superheater 14 and then goes to an evaporator 15 and out of the evaporator at 7. The location between the superheater and the evaporator indicated by 6 can also incorporate additional duct burning for steam generation in the boiler.
  • Path one is a small amount of bleed through the boiler to the power turbine 13 for the primary purpose of cooling turbine blades and nozzles.
  • the steam cooling inlet port 28 for the turbine 13 blades and nozzles is connected through a control steam manifold system 26 and bleeds steam from the steam drum 27.
  • Path two is entering the superheater and discharged. The superheater again has two paths, one through the valve 23 and one through the control valve 25 which enters the steam into the combustion chamber 3.
  • valve 25 control of the steam is different than the prior art patent for the reason that is was discovered that during the cogeneration operation sometimes the superheater with only a control valve between the superheater and the evaporator (indicated in Figure 1), even using chromolly alloy, which has high nickel and chrome content, will still present a rust substance on the surface. While steam is then reintroduced due to load demands, this substance enters into the turbine combustion chamber and lodges into the passageway of the cooling blades as a red substance, which will cause blockage of the cooling air and burn up all the turbines.
  • control valve 23 By optionally putting in the control valve at 25, which alone does not work, further steps are taken to make sure that when steam is not being used in the gas turbine and is used for cogeneration through valve 17 only, that a minimum amount of steam is going through the superheater such that a control by the control valve 23 mixed with the saturated steam at mixing chamber 22 will provide the additional heat for cogeneration purposes. This minimizes the exposure of the wall area of the superheater where steam was not injected into the gas turbines.
  • Additional valves are provided as an option at 24 such that a pressurized gas which will not cause corrosion, such as nitrogen or other excess neutralized gas, is used for the startup to control the pressure of the boiler system such that when the Advanced Cheng Cycle is started, the boiler is going to be boiling at a high pressure instead of having to evaporate at room temperature, then gradually build up the pressure in the boiler as a convention means. This way we speed up the startup process by five-fold. This eliminates the need for linking the gas turbine part of the operation and the steam generation part of the operation, which the ordinary combined cycle has to do.
  • a pressurized gas which will not cause corrosion such as nitrogen or other excess neutralized gas
  • the advantage of this optional configuration is that it retains the pressurized boiler startup operation and independent steam to gas turbine operation so that the gas turbine can, in the case of an electrical generation, start to operate by the turbine to the limit of the inlet temperature as a simple gas turbine cycle, which normally only takes a few minutes.
  • the high exhaust temperature at T5 was limited for the heat recovery boiler operation, because when boiling is happening at low pressure, the steam bubble occupies a large volume. That phenomena is called water swell in the boiler. So one has to bleed the water down at the high-high water mark until the boiler settles down.
  • the high temperature gas is going to the boiler, but the boiler is pressurized by, for example, the nitrogen bottle 29, which adjusts the boiler operating temperature and pressure (normally for this operation around 250 psi).
  • the boiling temperature of the water will then be around 380°F (194°C) instead of 212°F (100°C). Therefore, the boiler can absorb the temperature quickly without concern with water swells.
  • the boiler finally reaches the 380°F (193°C) temperature, it starts to produce steam.
  • the drum pressure will be higher than the allowed nitrogen pressurized pressure.
  • the steam valve 25 will then be opened to first bleed the nitrogen out to maintain the boiler pressure until the steam production rate is high enough to completely empty the nitrogen.
  • the steam temperature at 25 is limited to the gas turbine compressor discharge temperatures such that the cooling air is now provided internally to the turbine blades of the gas turbine, normally through the bypass through internal pass of the gas turbine.
  • the bypass air is mixed with incoming steam to maintain the exhaust exit temperature of the compressor; therefore the coolant to the turbine blade will not exceed previously designed compressor discharge temperature limitations to maintain the turbine blade cooling.
  • Figure 7 depicts typical air bleed cooling for the turbine blades such that the coolant enters from the bottom of the blade through blade 60.
  • the internal passage 61 cools the whole blade, but a certain amount escapes through the leading edge holes 63. After the coolant picks up the heat, it then exhausts at the trailing edge to protect further the heat transfer of the trailing edge. The cooled air then releases out of the holes 62.
  • Figure 8 illustrates a different way to change the parameters in the Advanced Cheng Cycle than indicated in the prior art (Cheng Cycle) such that even with a given fixed turbine inlet temperature (T.I.T), the compressor bleed air temperature normally will not be changed.
  • T.I.T fixed turbine inlet temperature
  • the compressor can be back pressured. Recovering energy from upper gas turbine cycle in the Carnot Cycle box was not previously considered.
  • the increased area due to increase pressure ratio is indicated by A' for a given T.I.T., and the entropy increase is reduced; therefore, the shrinking of the Carnot Cycle box and a much fuller filling of the box is achieved and is due to the increased pressure ratio.
  • the entropy never comes into play as far as efficiency is concerned.
  • the Advanced Cheng Cycle increases the pressure ratio through the back pressure of the compressor with limits of a surge margin to recover energy at the upper side of the cycle diagram inside the carnot box.
  • FIG 9 illustrates the efficiency versus XMIX condition.
  • 30 is the prior art steam injection rate for efficiency curve similar to the one depicted in Figure 2.
  • 31 assumes the same turbine inlet temperature.
  • the steam recovery upper temperature is limited to the compressor air discharge temperature, which allows us to recover more steam before reaching peak efficiency, with, however, a slight drop in efficiency, as indicated in the prior art examples.
  • 30 is penalized by the steam injection, the actual operation of the Advanced Cheng Cycle should operate at a turbine inlet temperature 32, which has a peak efficiency even higher than the peak efficiency of 30.
  • Even the temperature of the superheated steam is less than maximum superheat temperature in the prior art.. Therefore, the efficiency is recovered or gained by the ability to recover the turbine inlet temperature limitation due to steam injection.
  • Figure 10 illustrates the heat transfer temperature profiles when we choose the steam temperature to be limited by the compressor exit temperature 42.
  • the penalized turbine inlet temperature would have an exhaust temperature 41.
  • the turbine blade temperature can be kept constant as if no steam is being injected, one can operate at the higher inlet temperature 43. That indicates that the steam recovery is generated by 44 and more steam energy will still be recovered, except the upper temperature is bound by the compressor air discharge temperature in this case such that the limitation of the turbine inlet temperature to a lower value is removed.
  • Figure 11 illustrates the incorporation of all of the Advanced Cheng Cycle features, changing the TS diagram of the prior art (Cheng Cycle).
  • Carnot Cycle box would have a higher temperature 50', which means that inherent Carnot Cycle efficiency is improved. It is further improved by the higher pressure ratio indicated by the additional area A'. It is also further improved through the additional area which can be produced by the steam in the steam cycle as C, C', C'', etc.
  • Figure 12 illustrates the limitation of the compressor surge margin and pressure ratio increase.
  • the flow rate is indicated by W square root theta divided by delta, which is the compensation for the turbine inlet temperature and ambient pressure conditions.
  • W square root theta divided by delta
  • delta the compensation for the turbine inlet temperature and ambient pressure conditions.
  • On the vertical axis is the pressure ratio of the compressor discharge, and the lines N 1 , N 2 and N 3 are constant compressor RPM lines.
  • the dotted line represents the compressor stalling line, which is never to exceed pressure ratio lines.
  • the startup of an ordinary gas turbine is matched through a startup line such that when high pressure is reached, it is normally obtained through a higher turbine inlet temperature and the RPM is increased. However, when it is used for a power generation case, then the turbine is running at a constant RPM.
  • the RPM is N 3 .
  • the increase of power output is not due to the increase of RPM. Therefore, the increase of the turbine inlet temperature results into higher inlet flow volume required to go through the turbine blade areas, which will require a higher compression ratio. Therefore, for a simple cycle single shaft power generation operation, the surge margin of a compressor map is usually large to accommodate the idle condition of the turbine inlet temperature condition all the way to the fully loaded operations.
  • the startup procedure of the Advanced Cheng Cycle can be seen in Figure 6.
  • the valve 25 is closed until demand for the steam injection.
  • the boiler drum 27 is pressurized either by the heat remaining in the boiler to a designated pressure, or optionally pressurized by the nitrogen bottle 29 through the control regulator valve 24.
  • the gas turbine will be starting up assuming it has no steam injection and to reach the designated turbine inlet temperature for its gas simple cycle operations first.
  • the very high exhaust temperature T5 will quickly generate steam through the steam generating units 14 and 15, regardless of the water level conditions in the boiler, which is totally different than any of the boiler operations we have today.
  • the steam valve 25 When the boiler pressure exceeds the designated pressure, then the steam valve 25 is gradually opened to maintain the drum 27 pressure such that steam is injected into the gas turbine. When a full head of steam is generated, the drum pressure 8 should be higher than the pressure set for the regulator by the valve 24.
  • the steam rate will be synchronized through a control system with the turbine fuel flow. During load fluctuations, the steam will be injected according to a forward flow control scheme as anticipatory steam conditions, because we do have a higher pressure in the drum 27 and necessarily have to wait for steam to be generated all the time.
  • the coordination between the fuel flow and the steam flow will behave like a carburetor equivalent of an internal combustion engine such that turbine inlet temperature T4 will always be under control.
  • the Advanced Cheng Cycle limits the upper temperature not by the exhaust temperature of the gas turbine, but by the discharge temperature of the compressor, which would allow additional heat to be generated in the form of additional steam.
  • the cascading effect of more steam to generate even more steam is the unique aspect of the cycle which will make the whole Advanced Cheng Cycle to perform much better than the prior art (Cheng Cycle) without the limitation on the metallurgic temperatures of turbine blades, as was used in the prior art.
  • the ability to change the turbine area to match the compressor characteristics in the past has been made to match the original compressor discharge pressure ratio by opening up the turbine flow area to accommodate additional steam.
  • This by no means is true any more in Advanced Cheng Cycle in that the matching of the compressor discharge with additional steam will allow it to increase its pressure ratio only limited by the reasonable reserve surge margin for the compressor for a given design turbine inlet temperature. Therefore, when that happens, the compressor air discharge temperature will actually be higher than the original gas turbine operation only temperatures. Therefore, temperature below the actual compressor discharge temperature would be recommended to match the original discharge temperature for better turbine cooling applications. For that reason, the matching of the opening of the turbine areas will be as small as possible in order to back pressure the compressor to have a higher pressure ratio.
  • the steam temp. is 200°F (93.3°C) cooler and 45°F (7.2°C) cooler at maximum output compressor discharge temperature for ISO conditions.
  • the temperature should average at 700°F (371°C).
  • the pinch (or neck) temperature is approximately 100°F (37.7°C).
  • Efficiency is 3% better and output is 1,627 hp (1213 kw) more.

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Claims (23)

  1. Zweifluid-Wärmekraftmaschine mit Cheng-Zyklus mit
    (a) einem einen Kompressorauslaß (2) aufweisenden Kompressor (10) zum Verdichten eines ersten Arbeitsfluids,
    (b) einer mit dem Kompressorauslaß (2) in Strömungsverbindung stehenden Verbrennungskammer (12),
    (c) einer Turbine (13) mit einem mit der Verbrennungskammer (12) in Fluidverbindung stehenden Einlaß (4) zur Leistung von Arbeit durch Expandieren des Arbeitsfluids sowie mit einem Turbinenauslaß,
    (d) einem an den Turbinenauslaß angeschlossenen Wärmewiedergewinnungstauscher (14, 15) zum Erwärmen, mit einem Wärmewiedergewinnungstauscher-Einlaß und einem Auslaß zum Erwärmen eines zweiten Arbeitsfluids,
    (e) einer Einspritzeinrichtung (3) zum Einleiten von erwärmtem zweiten Arbeitsfluid von dem Wärmewiedergewinnungstauscher (14, 15) in die Verbrennungskammer (12),
    (f) einer Kühlmittel-Einlaßöffnung (28) zum Zuführen von Kühlmittel an Turbinendüsen und/oder Schaufeln (60) in der Turbine (13),
    (g) einer stromaufwärts von der Einspritzeinrichtung angeschlossenen Steuerventileinrichtung (25) zum selektiven Drosseln der Strömungsrate von zweitem Arbeitsfluid in die Turbine (13) und
    (h) einer Steuereinrichtung zum Steuern der Steuerventileinrichtung derart, daß
    (i) in die Turbine (13) eingespritztes zweites Arbeitsfluid eine Temperatur aufweist, die kleiner oder im wesentlichen gleich der Temperatur von erstem Arbeitsfluids am Kompressorauslaß (2) ist, und
    (ii) die Klemmtemperatur des Wärmewiedergewinnungstauschers minimiert wird, indem die Wärmewiedergewinnung vom Turbinenauslaß bei einem Mischungsverhältnis von zweitem zu erstem Arbeitsfluid in der Turbine (13) bei gegebenem Spitzenwirkungsgrad maximiert wird.
  2. Wärmekraftmaschine nach Anspruch 1, wobei die Steuereinrichtung die Steuerventileinrichtung derart steuert, daß Grenzwerte für die Turbineneinlaßtemperatur und Flammenbildungsbedingung nicht überschritten werden.
  3. Wärmekraftmaschine nach Anspruch 1 oder 2, mit ferner einer Kühlmittelquelle, die mit der Kühlmittel-Einlaßöffnung (28) derart in Fluidverbindung steht, daß das an der Kühlmittel-Einlaßöffnung (28) austretende Kühlmittel eine Temperatur aufweist, die geringer ist als die Temperatur des ersten Arbeitsfluids am Kompressorauslaß (2).
  4. Wärmekraftmaschine nach einem der Ansprüche 1 bis 3 mit ferner einem ersten Ventil (23), das mit dem Fluidauslaß des Wärmewiedergewinnungstauschers (14, 15) gekoppelt ist, um Arbeitsfluid aus überhitztem Dampf selektiv einer Kombinations-Generatoreinrichtung zuzuführen.
  5. Wärmekraftmaschine nach einem der Ansprüche 1 bis 4 mit ferner
    (a) einem mit dem Einlaß des Wärmewiedergewinnungstauschers in Fluidverbindung stehenden Wärmewiedergewinnungskessel mit einer mit Druck beaufschlagbaren Trommel (27), und
    (b) einer Einrichtung, die die Trommel während eines Kaltstarts der Kraftmaschine mit Druck beaufschlagt.
  6. Wärmekraftmaschine nach Anspruch 5, wobei die Einrichtung zur Druckbeaufschlagung der Trommel (27) eine externe Druckgasquelle ist, die selektiv mit der Trommel und einem Sensor zum Koordinieren anfänglicher Druckbeaufschlagung der Trommel in Verbindung steht.
  7. Wärmekraftmaschine nach Anspruch 6, wobei die Druckgasquelle im wesentlichen keinen Sauerstoff enthält.
  8. Wärmekraftmaschine nach einem der Ansprüche 5 bis 7, wobei die Einrichtung zur Druckbeaufschlagung der Trommel (27) der Kompressor (10) ist, der selektiv mit der Trommel in Verbindung steht.
  9. Wärmekraftmaschine nach einem der Ansprüche 1 bis 8, wobei das in die Kühlmittel-Einlaßöffnung (28) eingeleitete Kühlmittel ein Gas und ein Dampfgemisch aus von dem Wärmewiedergewinnungstauscher (14, 15) zugeführtem zweiten Arbeitsfluids ist.
  10. Wärmekraftmaschine nach einem der Ansprüche 1 bis 8, wobei das in die Kühlmittel-Einlaßöffnung (28) eingeleitete Kühlmittel gesättigter Dampf aus von dem Wärmewiedergewinnungstauscher (14, 15) zugeführtem zweiten Arbeitsfluids ist.
  11. Wärmekraftmaschine nach einem der Ansprüche 1 bis 10, wobei der Wärmewiedergewinnungstauscher (14, 15) ferner aufweist:
    (i) einen Überhitzer (14) mit einem mit dem Auslaß des Wärmewiedergewinnungstauschers gekoppelten Auslaß und einem Einlaß,
    (ii) einem Verdampfer (15) mit einem mit dem Einlaß des Überhitzers gekoppelten Auslaß und einem mit dem Einlaß des Wärmewiedergewinnungstauschers gekoppelten Einlaß, und
    (iii) einem zwischen dem Einlaß und dem Auslaß des Verdampfers angeordneten Wärmewiedergewinnungskessel.
  12. Wärmekraftmaschine nach Anspruch 11, wobei die Kühlmittel-Einlaßöffnung (28) zwischen dem Verdampfer (15) und dem Überhitzer (14) abgezogenes zweites Arbeitsfluid zuführt, um die Turbinendüsen und/oder Schaufeln (60) in der Turbine (13) zu kühlen.
  13. Wärmekraftmaschine nach Anspruch 11 oder 12 mit ferner einer Kombinations-Generatoreinrichtung, die über ein Kombinations-Generatorventil (17) zwischen dem Kessel (27) und dem Überhitzer (14) selektiv abgezogenes zweites Fluid sowie über ein zweites Ventil (23) von dem Überhitzerauslaß selektiv abgezogenes zweites Fluid als Wärmequelle benutzt.
  14. Wärmekraftmaschine nach einem der Ansprüche 5 bis 13, wobei das Volumen-Fassungsvermögen der Kesseltrommel ausreicht, um einen Druckanstieg infolge Schließens der Steuerventileinrichtung (25) bei Abfall des Leistungsbedarfs zu minimieren und zu ermöglichen, daß die Trommel (27) als Energiespeichereinrichtung funktioniert, um bei einem plötzlichen Energiebedarf eine rasche Dampfentnahme als ergänzende Wärmeenergiequelle für den Wärmewiedergewinnungskessel zu ermöglichen.
  15. Verfahren zum Betrieb einer Zweifluid-Wärmekraftmaschine mit Cheng-Zyklus mit (i) einem einen Kompressorauslaß (2) aufweisenden Kompressor (10) zum Verdichten eines ersten Arbeitsfluids, einer mit dem Kompressorauslaß (2) in Strömungsverbindung stehenden Verbrennungskammer (12), einer Turbine (13) mit einem mit der Verbrennungskammer (12) in Fluidverbindung stehenden Einlaß (4) zur Leistung von Arbeit durch Expandieren des Arbeitsfluids sowie mit einem Turbinenauslaß, (ii) einem an den Turbinenauslaß angeschlossenen Wärmewiedergewinnungstauscher (14, 15) zum Erwärmen, mit einem Wärmewiedergewinnungstauscher-Einlaß und einem Auslaß zum Erwärmen eines zweiten Arbeitsfluids, (iii) einer Einspritzeinrichtung (3) zum Einleiten von erwärmtem zweiten Arbeitsfluid von dem Wärmewiedergewinnungstauscher (14, 15) in die Verbrennungskammer (12), (iv) einer Kühlmittel-Einlaßöffnung (28) zum Zuführen von Kühlmittel an Turbinendüsen und/oder Schaufeln (60) in der Turbine (13), und (v) einer stromaufwärts von der Einspritzeinrichtung angeschlossenen Steuerventileinrichtung (25) zum selektiven Drosseln der Strömungsrate von zweitem Arbeitsfluid in die Turbine (13),
       wobei in dem Verfahren die Strömungsrate des zweiten Arbeitsfluids mit der Steuerventileinrichtung nach folgenden Arbeitsparametem gedrosselt wird:
    (i) eine Temperatur von in die Turbine eingespritztem zweiten Arbeitsfluid, die kleiner oder im wesentlichen gleich der Temperatur von erstem Arbeitsfluid am Kompressorauslaß ist, und
    (ii) Maximierung der Wärmewiedergewinnung vom Turbinenauslaß bei einem Mischungsverhältnis von zweitem zu erstem Arbeitsfluid in der Turbine (13) bei gegebenem Spitzenwirkungsgrad, so daß die Klemmtemperatur des Wärmewiedergewinnungstauschers minimiert wird.
  16. Verfahren nach Anspruch 15, wobei die Wärmekraftmaschine ferner eine Druckgasquelle (29) und einen mit der Trommel (27) selektiv in Verbindung stehenden Druckregler, ein mit der Gasturbinenmaschine und dem Wärmewiedergewinnungs-Dampfgenerator gekoppeltes Sensorsystem zur Erfassung von Temperatur und Druck, und ein Steuersystem für den Betrieb von Kraftstoffströmung an die Gasturbine aufweist, und wobei der Wärmewiedergewinnungs-Dampfgenerator einen Überhitzer (14) mit einem Überhitzereinlaß, einen Verdampfer (15) mit einem mit dem Überhitzereinlaß gekoppelten Verdampferauslaß sowie einen zwischen dem Wärmewiedergewinnungseinlaß und dem Verdampferauslaß angeordneten Wärmewiedergewinnungskessel mit einer Trommel aufweist, wobei ferner
    (a) Startbedingungen in der Gasturbinenmaschine und dem Wärmewiedergewinnungs-Dampfgenerator mittels des Steuersystems eingeleitet werden,
    (b) die Steuerventileinrichtung für Leerlauf-Strömungsbedingung der Wärmekraftmaschine eingestellt wird,
    (c) die Trommel mittels der Druckgasquelle auf ungefähr den gewünschten Arbeitsdruck des Wärmewiedergewinnungs-Dampfgenerators mit Druck beaufschlagt wird,
    (d) die Gasturbinenmaschine von Leerlauf auf Vollast, begrenzt durch die maximal zulässige Turbineneinlaßtemperatur, hochgefahren wird, und
       wobei zum Drosseln die Druckgasquelle abgeschaltet wird, wenn das Sensorsystem anzeigt, daß der Druck des Wärmewiedergewinnungs-Dampfgenerators höher ist als der Kompressor-Auslaßdruck.
  17. Verfahren nach Anspruch 16, wobei ferner während des Abschaltens der Maschine
    (a) der Auslaß der Gasturbinenmaschine von der Last genommen wird,
    (b) die Steuerventileinrichtung (25) geschlossen wird, um die Strömung von zweitem Arbeitsfluid in die Gasturbinenmaschine anzuhalten,
    (c) die Kraftstoffströmung an die Gasturbinenmaschine mittels des Steuersystems abgeschaltet wird, und
    (d) der Kessel heruntergekühlt wird.
  18. Verfahren nach Anspruch 16 oder 17, wobei während der Druckbeaufschlagung die zur Beaufschlagung der Trommel dienende Druckgasquelle (29) eine unabhängige Druckgasquelle ist und deren Abgabedruck an die Trommel mittels des Druckreglers auf ungefähr 5% unter dem Arbeitsdruck des Wärmewiedergewinnungs-Dampfgenerators eingestellt wird, und wobei der Druckregler die Rückströmung von erstem Arbeitsfluid an die Druckgasquelle gestattet, wenn der Trommeldruck den Druckgas-Abgabedruck überschreitet.
  19. Verfahren nach Anspruch 18, wobei die unabhängige Druckgasquelle (19) Flaschenstickstoff ist.
  20. Verfahren nach einem der Ansprüche 16 bis 19, wobei die Reihenfolge der Druckbeaufschlagung (c) und des Hochfahrens (d) folgendermaßen umgekehrt wird:
    die Gasturbinenmaschine wird von Leerlauf auf Vollast, begrenzt durch die maximal zulässige Turbineneinlaßtemperatur, hochgefahren und
    die Trommel (27) wird mit von dem Kompressor abgegebenen (10) ersten Arbeitsfluid auf ungefähr den gewünschten Arbeitsdruck des Wärmewiedergewinnungs-Dampfgenerators mit Druck beaufschlagt, so daß die Trommel mit Abgas der Gasturbinenmaschine als Druckgasquelle mit Druck beaufschlagt und erwärmt wird.
  21. Verfahren nach einem der Ansprüche 16 bis 20, wobei die Wärmekraftmaschine ferner ein zwischen dem Wärmewiedergewinnungs-Dampfgenerator (14, 15) und der Steuerventileinrichtung (25) mit der Maschine gekoppeltes erstes Ventil (23), ein an den Wärmewiedergewinnungs-Dampfgenerator angeschlossenes zweites Ventil (17), eine mit dem zweiten und dem ersten Ventil gekoppelte Mischkammer (22) und eine stromabwärts von dem ersten und dem zweiten Ventil (23; 17) mit der Mischkammer (22) gekoppelte Kombinations-Generatoreinrichtung aufweist,
       wobei ferner zum Betrieb der Wärmekraftmaschine im Kombinations-Generatormodus
    (a) die Steuerventileinrichtung (25) geschlossen wird, um eine Einleitung von zweitem Arbeitsfluid in die Turbine zu verhindern;
    (b) das zweite Ventil (17) geöffnet wird, um zweites Arbeitsfluid der Mischkammer (22) zuzuführen, und
    (c) das erste Ventil (23) ausreichend geöffnet wird, um zweites Arbeitsfluid durch den Überhitzer (14) abzulassen und damit der Möglichkeit einer Oxidation desselben entgegenzuwirken und der Mischkammer (22) zusätzliche Wärme für den Kombinations-Generatorbetrieb zuzuführen.
  22. Verfahren nach einem der Ansprüche 16 bis 21, wobei der Wärmewiedergewinnungs-Dampfgenerator (14, 15) eine Leitungsverbrennungseinrichtung (6) aufweist, um dem ersten Arbeitsfluid zwischen dem Überhitzer und dem Verdampfer zusätzliche Wärme zu entziehen und diese zweitem Arbeitsfluid in dem Wärmewiedergewinnungskessel zuzuführen,
       wobei ferner erstes Arbeitsfluid durch die Leitungsverbrennungseinrichtung geleitet und zweites Wärmefluid in dem Kessel erwärmt wird.
  23. Verfahren nach Anspruch 14 oder 15, wobei ferner Kühlmittel mit einer Temperatur, die unter der Temperatur des ersten Arbeitsfluids am Kompressoreinlaß liegt, in die Kühlmittel-Einlaßöffnung (28) eingeleitet wird
EP92921791A 1992-09-30 1992-09-30 Cheng-zweistoffwärmekraftmaschine und verfahren zum ihren betrieb Expired - Lifetime EP0723626B1 (de)

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PCT/US1992/008351 WO1994008128A1 (en) 1991-04-02 1992-09-30 Cheng dual fluid regenerative cycle

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EP0723626B1 true EP0723626B1 (de) 2002-03-27

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