DE69232521T2 - CHENG DUAL-SUBSTANCE HEAT ENGINE AND METHOD FOR THEIR OPERATION - Google Patents

Description

This invention relates to a two-fluid Cheng cycle heat engine and a method for operating the same.

The present invention relates to a heat engine and, more particularly, to improving efficiency and power output over the two-fluid cycle or Cheng cycle with parallel recovery in a number of previous US patents (US-A-4,128,994, US-A-3,978,661, US-A-4,248,039, US-A-4,297,841 and US-A-4,417,438).

Summarizing the background of the previous patents and publications, the Cheng cycle concerns a heat recovery gas turbine where steam is used only in a heat recovery steam generator (HRSG) in a specific way, where steam is heated to as high a superheat temperature as possible (limited by the turbine exhaust temperature) and where the total heat recovered is maximized, limited only by the physics of heat transfer at the water to steam transition of the boiler and the available temperature differences of the gas stream. This is known in the HRSG industry as the throat temperature. The two-fluid Cheng cycle patented under this condition achieves a peak efficiency mixing ratio between the liquid and gaseous working fluid, referred to as the XMIX peak. The expected efficiency gain and simplicity of operation were cited in the mentioned patents and literature, and they occurred in real operating data.

The Cheng cycle was first commercially used using an Allison 501KH gas turbine and then a Kawasaki M1-ACC engine. Since 1984, ten 501KH engines have been fitted with Cheng Cycle has been put into operation, and two Kawasaki machines with Cheng cycle are in operation in Japan. Numerous operating patents and modifications to the turbine have been added.

All Cheng cycle systems have had all the advantages mentioned and are also supervised by the inventor, Dr. Dah Yu Cheng (the inventor of the present advanced Cheng cycle {ACC}). The same gas turbine that has a complicated steam turbine can also be configured with a combined cycle. Although it is not possible to improve the output power to the same extent, it is believed that it may be possible to achieve a slightly higher efficiency than the Cheng cycle. This is because at the time the Cheng cycle was developed, in the 1970s, most gas turbines operated at a metallurgically acceptable turbine inlet temperature for stress and corrosion protection. The Cheng cycle was developed at that time without considering blade cooling. The trend at that time was towards ceramic turbines and even internally water-cooled turbine blades. The advance in cooling by diverting compressor air through a cooling passage, strategically introduced at a stagnation point and on the pressure side of the trailing edge of the blade, has made it possible to dramatically increase the temperature of the gas stream without increasing the surface temperature of the turbine blade above the previous metallurgical operating temperature limit without cooling.

The diversion of compressor air is certainly a loss for the gas turbine system, but the ability to increase the operating temperature of the turbine more than compensated for this loss. This resulted in an increased compressor pressure ratio and also made it possible to increase the efficiency of the simple gas turbine cycle. Up to 10% of diverted air is used in advanced fighter gas turbines. with the turbine inlet temperature above 2750ºF (1510ºC). In the industry, a 1% increase in efficiency is considered a significant achievement. The new GE 7-F frame has a turbine inlet temperature of 2350ºF (1288ºC), but it is still designed for a relatively low pressure ratio so that the exhaust temperature can be high enough to improve the steam cycle portion of the combined cycle. Combined cycle efficiency of more than 50% has been claimed.

It has become apparent that an uncooled gas turbine converted to a Cheng cycle can be improved by maximum superheating and that difficulties arise in operating without the Cheng cycle steam inlet at this gas turbine inlet temperature.

In Table 1, which lists Allison turbine performance guarantees, the 501KH and the 501KH steam inlet engine are shown in the last two columns. The 501KH engine has a turbine inlet temperature of 1895 #F (1035ºC) and the 501KH steam inlet engine has a turbine inlet temperature of 1800ºF (982ºC). The loss of 95ºF (35ºC) of available temperature represents a potential loss of 3.0% in efficiency and a potential increase of 1000 kW in output power. The root cause of the problem was that when energy is recovered at the highest thermodynamic potential, in other words at the highest (maximum) temperature of the steam (see prior art patents), the steam temperature becomes much higher than the compressor air outlet temperature, knowing that additional fuel must be burned to bring the steam up to the turbine inlet temperature. By mixing the steam and the compressor outlet air, a mixture temperature is obtained that is higher than the compressor outlet temperature.

If bypass air passages were provided to achieve a lower compressor air outlet temperature, the higher mixing temperature resulted in a loss of blade surface cooling ability. The only way to limit blade surface temperature would be to reduce the turbine operating temperature. This causes a possible loss of efficiency and power output, resulting in higher machine costs.

Document EP 319699 describes a temperature cycle different from the advanced Cheng cycle, where the system requires a very high pressure boiler and a low pressure boiler to produce the very high pressure steam and some low pressure steam. The high pressure steam is cooled by passing it through a steam turbine 32 which changes the enthalpy of the steam and it is then introduced into the combustion chamber of the gas turbine at 34. The low pressure steam is admitted between the first stage turbine for compressing air and the second stage turbine for creating a back pressure for the first stage turbine to reduce its ability to generate power from the added steam fed into the combustion chamber and to increase the ability of the second stage turbine to generate more work while cooling the hot working fluid exiting the first turbine to create equilibrium with the steam inlet. The high pressure scheme described in EP 319699 increases the boiling temperature of the heat recovery steam generator, limiting its ability to generate maximum heat recovery. The added low pressure steam described in EP 319699 does not enter the combustion chamber and therefore does not contribute to increasing the steam temperature from the combusted gas. Therefore, the efficiency of operating the turbine system is compromised.

The object of the invention is to overcome the limitations by an evolution of the Cheng cycle, referred to here as an advanced Cheng cycle (ACC). The hardware configuration of the ACC follows the old design to a large extent in terms of valve arrangement, control systems, start-up and shutdown processes. A number of difficulties encountered with the original Cheng cycle are also improved by this disclosure.

According to the present invention, there is provided a Cheng cycle two-fluid heat engine comprising a compressor having a compressor outlet for compressing a first working fluid, 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 expanding the working fluid, and having a turbine outlet, a heat recovery exchanger connected to the turbine outlet for heating, 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 supplying coolant to turbine nozzles and/or blades in the turbine, a control valve means connected upstream of the injector for selectively throttling the flow rate of the second working fluid into the turbine, and a control means for controlling the control valve means such that injected second working fluid has a temperature which is less than or substantially equal to the temperature of the first working fluid at the compressor outlet, and the clamping temperature of the heat recovery exchanger is minimized by recovering heat from the turbine outlet at a temperature between the first and second Working fluid in the turbine is maximized at a given peak efficiency.

According to the present invention there is further provided a method of operating a two-fluid Cheng cycle heat engine having a gas turbine with a compressor having a compressor outlet for compressing a first working fluid, 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 expanding the working fluid and with a turbine outlet, a heat recovery exchanger connected to the turbine outlet for heating, a heat recovery exchanger inlet and an outlet for heating a second working fluid, an injection device for introducing heated second working fluid from the heat recovery exchanger into the combustion chamber, a coolant inlet opening for supplying coolant to turbine nozzles and/or blades in the turbine, and an upstream of the injection device connected control valve means for selectively throttling the flow rate of the second working fluid into the turbine, in which method the flow rate of the second working fluid is throttled with the control valve means according to the following operating parameters: a temperature of second working fluid injected into the turbine which is less than or substantially equal to the temperature of first working fluid at the compressor outlet, and maximizing heat recovery from the turbine outlet at a mixing ratio between the second and first working fluids in the turbine at a given peak efficiency so that the clamping temperature of the heat recovery exchanger is minimized.

The invention can be implemented by a method or device which has several advantages, which are:

(a) the reduction in turbine inlet temperature is reduced when a steam inlet is used in a gas turbine so that the maximum design temperature of the turbine can only be available for gas for the Cheng cycle,

(b) a choice of coolant is provided, which can enhance the cooling of the gas turbine blades relative to the turbine inlet temperature,

(c) efficiency is improved by increasing the turbine inlet temperature beyond the Cheng cycle, thereby maintaining the turbine blade surface temperature,

(d) a new configuration of a steam control valve is introduced to prevent rusting in the superheater,

(e) the surface area of the superheater is reduced,

(f) the efficiency of power generation is increased by increasing the compressor discharge pressure,

(g) a facility is provided for rapid ramp-up of the advanced Cheng cycle; and

(h) the combined cycle generator operation for producing mechanical (electrical) power and steam is maintained in such a way as to follow the varying load independently.

Fig. 1 shows the Cheng cycle according to a series of patents filed by the inventor in the 1970s.

Fig. 2 shows the peak efficiency aspect of the steam-air mixture ratio in the state-of-the-art Cheng cycle.

Fig. 3 shows the heat transfer limitation to achieve maximum heat recovery and maximum superheat as stated in the prior art.

Fig. 4 is a temperature versus entropy plot describing cycle efficiencies of the Cheng cycle as shown in the prior art.

Fig. 5 is a graph of temperature versus entropy for a combined cycle under Use of a gas turbine and a steam turbine for heat recovery.

Fig. 6 shows the configuration of the new advanced Cheng cycle including new control valve locations.

Fig. 7 shows a typical bled air cooled turbine blade.

Fig. 8 shows the advanced Cheng cycle with an increased pressure ratio.

Fig. 9 shows a comparison of the efficiency and location of XMIX for peak efficiency between the state-of-the-art (Cheng cycle) and the advanced Cheng cycle.

Fig. 10 shows the differences in heat transfer in the heat recovery boiler and temperature profiles between the state-of-the-art (Cheng cycle) and the advanced Cheng cycle.

Fig. 11 shows the optimal description of the advanced Cheng cycle with an increased turbine inlet temperature and an increased pressure ratio.

Fig. 12 shows the ability of the advanced Cheng cycle to increase the combustion ratio through turbine tuning and compressor mapping.

In Fig. 1, the embodiment of the Cheng cycle according to prior art patents is shown. The configuration shows that the gas turbine has a compressor 10 connected by a shaft to a turbine 13 and an outlet to a load. The air admitted through 1 is compressed to compressed air and discharged at 2. The compressed air enters a combustion chamber 12. Fuel enters the combustion chamber through 11 and steam comes from the heat recovery steam generator (HRSG) at 3. The mixture of combusted air and steam reaches a predetermined turbine inlet temperature, is then admitted to the turbine 13 at 4 and exits at 5. from the turbine. Exhaust gas then passes through the heat recovery steam generator which is divided into two parts, namely a superheater 14 and a water to steam generator 15. The hot exhaust gas enters the superheater 14 and gives up the heat to superheat the steam entering at 8 and exiting at 3. A line combustor, normally located at 6, is not shown here. The remaining heat is recovered by the evaporation unit 15 and exits at 17. It is possible for the exhaust gas to pass through a purification or condensation unit 20 and it is then exhausted to the atmosphere. Water may be recovered through 20 or used entirely as an additive entering at 19 or mixed in at 19. Water is compressed to a high pressure by a pump 18. The pump output enters the steam generator 9 and the evaporator controls the steam flow through two valves, the steam flow entering the superheater through 16 and being supplied to a steam user through 17, thus forming a combination generator unit. If used only for power generation, 17 is no longer required.

The above description of this component diagram is typical of the Cheng cycle configuration described and patented in the prior art patents.

Fig. 2 shows the relationship between efficiency and the steam-to-air ratio defined as XMIX, such that efficiency reaches a peak for a given turbine inlet temperature 30 and then drops off with even greater steam input. This peak efficiency is the nature of the heat recovery described in the Cheng cycle patents and the main claims, as can be seen in Fig. 3.

Fig. 3 shows the exhaust gas temperature profile and heat transfer to the HRSG by converting water to high temperature steam. The length indicated is the partial length of the heat exchanger in the HRSG. It is stated here that the exhaust gas entering the heat exchanger described in Fig. 1 is at temperature T5. The temperature decreases from T8 to T3 to temperature T6 by superheating the steam. The temperature decreases continuously on the exhaust side from T6 to T7 and then continuously to the exhaust gas temperature T7. The water entering at T9 reaches the evaporation point for a given pressure at T7 and evaporation takes place according to profile 40. The exhaust gas profile is 41. The temperature difference between the boiler inlet temperature and the exhaust gas temperature at the smallest point indicated by (T7'-T8) is called the Delta Tpinch or throat temperature. The temperature difference between T5 and T3 is called the Delta Tmaximum. The maximum enthalpy and maximum heat recovery in the prior art require that the upper pinch point Delta Tmax and T7' minus T8 be minimized. At this point the peak value of XMIX is reached.

Fig. 4 is a temperature-entropy diagram, referred to in the industry as a T-S diagram. S represents the entropy. The box 50 is referred to as a Carnot cycle box, and the area indicated within the box is the gas turbine cycle diagram, so it has four sides. The compression side is indicated by the temperature rise with increased entropy S from T1 to T2. Combustion occurs by increasing the temperature from T2 to T4.

On the gas side, the temperature is reduced from T4 to T5 by expansion through the turbine. Before the gas leaves the power plant, it enters the heat recovery boiler. The temperature drops through a profile at the bottom, completing the cycle. The square box 50 represents the temperature T4 and the temperature T1, which are limited by the turbine inlet temperature and the ambient temperature of the gas turbine, where this is defined as the Carnot cycle of gas turbine efficiency.

To improve the efficiency of this cycle, the area of a circle filled in the Carnot cycle must be maximized. The efficiency of the Carnot cycle is usually defined as efficiency equal to 1 minus T1 divided by T4, so that the higher T4 is, the higher the efficiency of the Carnot cycle. There is no indication of the influence of entropy in the Carnot cycle, so the box of the Carnot cycle plays no role in defining the efficiency of the cycle.

The water side of the cycle starts at T9, reaches temperature T8, and the superheater temperature T3 then tries to fill up the corner of the Carnot cycle box. The steam is then further heated from T3 to T4 and then expanded synergistically with the air and brought to the same temperature T5. The heat from the steam is also recovered by the additional steam as indicated by the box boundary B'. B' goes through the same cycle and recovers additional steam as indicated by B". Therefore, heat recovery according to the Cheng cycle is a series of heat recovery cycles by steam with maximum heat recovery and maximum entropy filling the Carnot cycle box as closely as possible to increase efficiency. The area increase between A plus B and then B' and B" and so on is the potential power output increase of the Cheng cycle. As described in the prior art patent, the competing cycle is usually described as a combined cycle, so that the combined cycles do not introduce steam into the gas turbine, and rather a separate steam cycle is run.

This cycle is shown in Fig. 5. The Carnot box 50 is the gas turbine part of the Carnot box, but a steam cycle occupies the lower corner of the empty part of the Carnot box below the gas turbine cycle A. The Steam cycle is indicated by the area D. The steam cycle D is usually a high pressure flow, which is comparable to the Cheng cycle which occurs at a lower pressure. Therefore, the potential for power increase is not as great as in the Cheng cycle.

In Fig. 6 the advanced Cheng cycle system is shown. The advanced Cheng cycle system has a compressor 10 and turbine 13 connected by a shaft and there is an output to the load. Air also enters at 1 and after compression is discharged into the combustion chamber 12 at 2. Fuel enters at 11. Steam from the heat recovery generator enters at 3 and premixing to a given turbine inlet temperature is limited by the allowable metallurgy temperature at T4. The gas mixture then expands through the turbine 13 and exits the turbine at 5. The exhaust gas enters a superheater 14 and then passes to an evaporator 15 and exits the evaporator at 7. The location indicated by 6 between the superheater and the evaporator may also have an additional line combustion device for generating steam in the boiler.

On the water side, pretreated water enters the pump 18, is discharged into the boiler 9 and evaporates to saturated steam which has three possible paths. The first path has a small amount of leakage which passes through the boiler to the power turbine 13 and has the main purpose of cooling the turbine blades and nozzles. The steam cooling inlet port 28 for the blades and nozzles of the turbine 13 is connected through a control steam distribution system 26 and admits steam from the steam drum 27. The second path admits and discharges steam into and from the superheater. The superheater again has two paths, one passing through the valve 23 and one passing through the control valve 25 whereby the steam enters the combustion chamber 3. As indicated, the control of the steam through valve 25 differs from that of the prior art patent because it has been discovered that during combination generator operation the superheater sometimes forms a rust substance on its surface if only one control valve is present between the superheater and the evaporator (indicated in Fig. 1), even when using a Chromolly alloy having a high nickel and chromium content. As steam is reintroduced in response to load demands, this substance then enters the turbine combustion chamber and deposits as a red substance in the passage of the cooling blades, thereby blocking the cooling air and burning all of the turbines. If the control valve is optionally included at 25, which alone will not work, further steps are taken to ensure that when steam is not being used in the gas turbine and only passes through valve 17 for combination generator operation, a minimum amount of steam passes through the superheater so that the steam mixed with the saturated steam at the mixing chamber 22 under control of control valve 23 provides the additional heat for combination generator operation. This minimizes exposure of the superheater wall surface when steam has not been introduced into the gas turbines. Additional valves are provided as an option at 24 so that a non-corrosive pressurized gas such as nitrogen or other additional neutralized gas is used at start-up to control the pressure of the boiler system so that when the advanced Cheng cycle is initiated, the boiler boils at a high pressure rather than having to evaporate at room temperature, the pressure in the boiler then being gradually built up in the conventional manner. In this way, the start-up process is accelerated by five times. This makes it unnecessary to link the gas turbine operating section with the steam generation operating section, as is the case with the usual combined cycle.

The advantage of this optional configuration is that it maintains the start-up process of the pressure vessel and the independent process of admitting steam to the gas turbine, so that when generating power, the gas turbine can start operating in a simple gas turbine cycle up to the inlet temperature limit, which normally only takes a few minutes. The high exhaust gas temperature at T5 was limiting for the operation of the heat recovery boiler because the steam bubbles occupy a large volume when boiling occurred at a low pressure. This phenomenon is called boiler water swell. Therefore, the water must be drained at the high water mark until the boiler has settled.

In this case, the high temperature gas enters the boiler, but the boiler is pressurized, for example by the nitrogen cylinder 29, which sets the boiler operating temperature and pressure (normally around 250 psi in this operation). The boiling temperature of the water is then around 380ºF (194ºC) instead of 212ºF (100ºC). Therefore, the boiler can absorb the temperature quickly without danger of water swelling. When the boiler finally reaches the temperature of 380ºF (193ºC), it begins to generate steam. The drum temperature is then higher than the allowable nitrogen pressure.

The steam valve 25 is then opened to release the nitrogen slowly at first to maintain the boiler pressure until the steam generation rate is high enough to completely empty the nitrogen. At this moment, the nitrogen cylinder control system will completely shut off the nitrogen flow because the drum pressure is higher than the nitrogen pressure. When this happens, the high temperature water does not absorb nitrogen gas, so that no venting is required. All nitrogen has been vented into the combustion chamber. The additional steam generated then begins to introduce steam into the combustion chamber until the boiler has reached the equivalent of the full amount of steam prescribed by the design parameters, which can be introduced through the control valve 25 to achieve higher loads and higher efficiency.

When steam is first introduced through port 23, the turbine inlet temperature decreases so that the fuel flow rate at 11 increases to maintain the turbine inlet temperature T4. Therefore, the output continues to increase without increasing the turbine inlet temperature until the full amount of steam has been generated. The turbine is now producing its maximum output at maximum efficiency.

If the diverted steam is not used to cool the turbines, the steam temperature is limited to the gas turbine compressor outlet temperatures at 25, so that the cooling air is now supplied internally to the gas turbine turbine blades, normally by bypassing through the gas turbine internal passage. The diverted air is mixed with incoming steam to maintain the compressor exhaust outlet temperature, and therefore the turbine blade coolant does not exceed pre-established compressor outlet temperature limitations to maintain turbine blade cooling.

If the steam temperature at T3 is lower than the compressor discharge temperature, this is certainly an option to achieve flexibility of operation with a slight loss of efficiency, but with an increase in output power. The character is similar to that previously described in the prior art (Cheng cycle), but here the parameter chosen for Achieving peak efficiency is different and is controlled by this new invention.

Figure 7 shows a typical air bypass cooling for the turbine blades, with the coolant passing from the bottom of the blade through the blade 60. The inner passage 61 cools the entire blade, but a certain amount escapes through the holes 63 at the leading edge. After the coolant has absorbed the heat, it exits at the trailing edge to further protect the trailing edge heat transfer. The cooled air is then exhausted from the holes 62.

In Fig. 8 a different way of changing the parameters in the advanced Cheng cycle than in the prior art (Cheng cycle) is given, where even for a given fixed turbine inlet temperature (T.I.T.) the compressor bleed air temperature is normally not changed. However, using an operating map the compressor can be provided with a back pressure. The recovery of energy from the upper gas turbine cycle in the Carnot cycle box has not been considered so far. The increased area due to the increased pressure ratio is given by A' for a given T.I.T. and the entropy increase is reduced, therefore a shrinkage of the Carnot cycle box and a much more complete filling of the box due to the increased pressure ratio are achieved. However, as has been explained, entropy never plays a role in terms of efficiency. The advanced Cheng cycle increases the pressure ratio by the back pressure of the compressor, with limits set by a pressure pulse range, to recover energy at the top of the cycle diagram within the Carnot box.

In Fig. 9 the efficiency is shown against the XMIX condition. 30 is the curve which, similar to that shown in Fig. 2, indicates the efficiency versus the steam inlet rate of the prior art. At 31 the same turbine inlet temperature is assumed. The upper Steam recovery temperature is limited to the compressor air outlet temperature, allowing more steam to be recovered before peak efficiency is reached, but with a small drop in efficiency as indicated in the prior art examples. Because 30 is hindered by the steam inlet, actual operation of the advanced Cheng cycle should occur at a turbine inlet temperature 32, with peak efficiency even higher than the peak efficiency at 30. Even the superheated steam temperature is less than the maximum superheat temperature in the prior art. Therefore, the ability to recover the turbine inlet temperature through the steam inlet restores or increases efficiency.

By preferentially cooling the cooling blades by saturated steam, which is normally at a temperature lower than the compressor outlet temperature, this ability to increase the turbine inlet temperature can again be maintained at a maximum steam temperature.

In Fig. 10, heat transfer temperature profiles are shown when the steam temperature is chosen to be limited by the compressor outlet temperature 42. Usually, the limited turbine inlet temperature would be at the exhaust gas temperature 41. However, because the turbine blade temperature can be held constant as if no steam is being admitted, operation can be carried out at the higher inlet temperature 43. This indicates that steam recovery is generated by 44 and more steam energy continues to be recovered, but the upper temperature in this case is limited by the compressor air outlet temperature, so that the limitation of the turbine inlet temperature to a lower value is removed.

In Fig. 11 the inclusion of all features of the advanced Cheng cycle is shown, which change the TS diagram from the state of the art (Cheng cycle). It First, it is seen that the Carnot cycle box has a higher temperature 50', which means that the natural efficiency of the Carnot cycle is improved. It is further improved by the higher pressure ratio, which is indicated by the additional area A'. It is further improved by the additional area that can be created by the steam in the steam cycle as C, C', C" etc.

Fig. 12 shows the limitation of the pressure pulse range and the pressure ratio increase of the compressor. The flow rate is given by

which is the compensation for the turbine inlet temperature and ambient pressure conditions. On the vertical axis is the pressure ratio of the compressor outlet and the lines N1, N2 and N3 are lines of constant rotational speed of the compressor. The dotted line represents the compressor stall line which must never exceed the pressure ratio lines. The start-up of an ordinary gas turbine is consistent with a start-up line so that when a high pressure is achieved, it is normally maintained by a higher turbine inlet temperature and the rotational speed increases. However, when used for power generation, the turbine runs at a constant rotational speed.

In this case, the rotation speed is N3. The increase in output power is not due to the increase in rotation speed. Therefore, the increase in turbine inlet temperature results in a higher inlet flow volume having to pass through the turbine blade areas, which requires a higher compression ratio. Therefore, in a simple single-shaft cyclic power generation process, the pressure pulse range of a compressor map is usually large to accommodate the idle condition of the turbine inlet temperature over the entire range up to full load operation.

In the prior art (Cheng cycle), opening the turbine sections was considered so that the pressure pulse range does not exceed the pressure pulse range of the compressor. However, allowing steam and higher temperatures in the advanced Cheng cycle is not only useful in closing the turbine flow section to increase the pressure ratio, but also results in higher efficiency and higher output power at work. This is the reason why the advanced Cheng cycle can enable the high pressure ratio of the turbines.

In Fig. 1, prior art operation in terms of fuel flow rates and efficiency (as in the Allison catalog in Table 1) was described as an example based on Cheng cycle parameters, with the steam temperature 50ºF (10ºC) below the turbine exhaust temperature. Because the gas turbine pressure ratio is 9.3 and the compressor discharge temperature is lower, an example of advanced Cheng cycle operation is given with the steam recovery temperature limited by the gas turbine compressor discharge temperature. An example of energy balance and efficiency and output improvement is a comparison of operation of the Cheng cycle at 1800ºF (982ºC), limited by the current Allison warranty values to 1895ºF (1035ºC), with the gas turbine only optional (column 501KH in Table 1), and the advanced Cheng cycle, with the turbine blade temperature now controlled by the superheated steam temperature to a lower selected superheat temperature (see Fig. 2) related to the compressor discharge temperature for a given turbine. For example, comparing the prior art (Cheng cycle) to the advanced Cheng cycle only results in a steam temperature of 700ºF (371ºC).

It is obvious from the operational viewpoint of the figures that the superheater of the unit is smaller and that because of such a large temperature difference in the heat transfer to the evaporator, the surface increase due to the additional steam generation is not sufficiently high, so that the overall cost of the boiler is reduced. On the other hand, the efficiency and power are increased, which leads to the advanced Cheng cycle having a higher efficiency and a lower manufacturing cost, thus giving a commercial advantage for competition in the market.

The start-up procedure for the advanced Cheng cycle is shown in Fig. 6. The valve 25 is closed until there is a demand for steam inlet. The boiler drum 27 is either pressurized to a set pressure by the heat remaining in the boiler or alternatively pressurized by the nitrogen bottle 29 via the controlling regulator valve 24. During start-up, the gas turbine is started on the assumption that it will not experience steam inlet and that the set turbine inlet temperature is reached first for its simple gas cycle operation. Because of the very high exhaust gas temperature T5, steam is generated very rapidly by the steam generating units 14 and 15 regardless of the water level conditions in the boiler, which is completely different from any present-day boiler operation.

When the boiler pressure exceeds the set pressure, the steam valve 25 is gradually opened to maintain the pressure in the drum 27 so that steam is admitted to the gas turbine. When the full amount of steam is being produced, the drum pressure 8 should be higher than the pressure set by the valve 24 for the governor. The steam rate is synchronized with the fuel flow through the turbine by a control system. During load fluctuations, the steam is admitted according to expected steam conditions according to a predictive flow control scheme because a higher pressure is present in the drum 27 and one must always wait for steam to be generated. The coordination between the fuel flow and the steam flow behaves like a carburetor in an internal combustion engine so that the turbine inlet temperature T4 is always under control. Under part load conditions, T4 is reduced and the maximum amount of heat is recovered by the heat recovery generators. In the case of night-time combination generator operation, maximum steam generation by the heat recovery generators is required. In the case of night-time combination generator operation where maximum steam generation is required but a minimum electrical load is required, the steam is completely bypassed through the combination generator valve 17 and the valve 25 is closed and the valve 23 is slightly opened to pass a small amount of saturated steam through the superheater for mixing in the mixing chamber 22 for the combination generator operation. This protects the superheater surface even when no steam is being introduced into the gas turbine. The peak efficiency point for a given steam-air ratio still applies even when the steam temperature is limited to the compressor outlet because efficiency is lost at the lower steam temperature. The advanced Cheng cycle limits the upper temperature not by the gas turbine exhaust temperature but by the compressor outlet temperature, thus allowing additional heat to be generated in the form of additional steam. The cascade effect of more steam generating even more steam is the only aspect of the cycle that makes the entire advanced Cheng cycle work much better than the prior art (Cheng cycle) without the limitation of turbine blade metallurgical temperatures that occurred in the prior art.

In summary, changing the turbine area to adjust compressor characteristics has historically been done by opening the turbine flow area to accept additional steam to adjust the original compressor discharge pressure ratio. This is no longer true for the advanced Cheng cycle because adjusting the compressor discharge with additional steam allows the pressure ratio to be increased only limited by the reasonable reverse pressure pulse range for the compressor for a given fixed turbine inlet temperature. When this occurs, the compressor discharge air temperature is therefore actually higher than the temperatures during original gas turbine operation alone. Therefore, a temperature below the actual compressor discharge temperature is recommended to adjust to the original discharge temperature for better turbine cooling. For this reason, the adjustment of the opening of the turbine surfaces is as small as possible to exert a back pressure on the compressor in order to achieve a higher pressure ratio. Therefore, the parameters of the cycle in terms of pressure ratio, turbine inlet temperature and superheat temperature, which limits the amount of steam, are linked in a similar way to the prior art (Cheng cycle), but it works in a different range in terms of steam-air ratios in order to achieve the high efficiency and high output power.

In order for the example 501KB to become KH with steam inlet, the pressure ratio is increased from 9.3:1 to 11.5:1. This in turn increases the efficiency. The turbine inlet temperature is also now set so that a simple cycle is used at the temperatures of a gas turbine. The cycle is no longer limited by single-point operation because it is clear from the operation of the Cheng cycle from the state of the art that in a combination generator operation, the steam can be switched between a combination generator demand and intake demands for power generation. By adding line combustion, steam can be generated without using only the waste heat from the turbine, so that the electrical power or mechanical output and the steam demand can be completely independent.

TABLE 1

Allison Specifications for Industrial Gas Turbine Engines 501-K Specifications

Operating Conditions: International Standard Day - Sea Level - No Losses 59ºF/15ºC - Nominal Steam Inlet [C]

[A] Based on gaseous fuel

[B] Including one power turbine

[C] Based on 5 lb/s (2.2 kg/s) steam inlet at 90ºF (32.2ºC)

[D] Depending on the selected turbine power

TABLE 2

Comparison of functionality improvements between Cheng Cycle 501KH and Advanced Cheng Cycle 501KH

[A] Initial temperature of Air compressor without increasing the output pressure

[B] Due to the higher thermal conductivity in the hot stream, a slightly lower steam HRSG temperature is recommended to maintain the blade surface temperature.

From Table 2 above, the vapor temperature under ISO conditions is 200ºF (93.3ºC) and 45ºF (7.2ºC) cooler at the discharge temperature at maximum compressor output. In hot summer and cold winter, the temperature should average about 700ºF (371ºC). The clamp temperature (or throat temperature) is about 100ºF (37.7ºC). The efficiency is 3% higher and the output is 1627 hp (1213 kW) higher.

Claims (23)

1. Two-fluid heat engine with Cheng cycle with (a) a compressor (10) having a compressor outlet (2) for compressing a first working fluid, (b) a combustion chamber (12) in flow connection with the compressor outlet (2), (c) a turbine (13) with an inlet (4) in fluid communication with the combustion chamber (12) for performing work by expanding the working fluid and with a turbine outlet, (d) a heat recovery exchanger (14, 15) connected to the turbine outlet for heating, with a heat recovery exchanger inlet and an outlet for heating a second working fluid, (e) an injection device (3) for introducing heated second working fluid from the heat recovery exchanger (14, 15) into the combustion chamber (12), (f) a coolant inlet opening (28) for supplying coolant to turbine nozzles and/or blades (60) in the turbine (13), (g) a control valve device (25) connected upstream of the injection device for selectively throttling the flow rate of second working fluid into the turbine (13) and (h) a control device for controlling the control valve device such that (i) the second working fluid injected into the turbine (13) has a temperature which is less than or substantially equal to the temperature of the first working fluid at the compressor outlet (2), and (ii) the clamping temperature of the heat recovery exchanger is minimized by maximizing the heat recovery from the turbine outlet at a mixture ratio of second to first working fluid in the turbine (13) at a given peak efficiency. 2. Heat engine according to claim 1, wherein the control device controls the control valve device such that limit values for the turbine inlet temperature and flame formation condition are not exceeded. 3. Heat engine according to claim 1 or 2, further comprising a coolant source which is in fluid communication with the coolant inlet opening (28) such that the coolant exiting at the coolant inlet opening (28) has a temperature which is lower than the temperature of the first working fluid at the compressor outlet (2). 4. A heat engine according to any one of claims 1 to 3, further comprising a first valve (23) coupled to the fluid outlet of the heat recovery exchanger (14, 15) for selectively supplying working fluid from superheated steam to a combination generator device. 5. Heat engine according to one of claims 1 to 4, further comprising (a) a heat recovery vessel in fluid communication with the inlet of the heat recovery exchanger, having a pressurizable drum (27), and (b) a device for pressurising the drum during a cold start of the engine. 6. A heat engine according to claim 5, wherein the means for pressurizing the drum (27) is an external source of pressurized gas selectively connected to the drum and a sensor for coordinating initial pressurization of the drum. 7. A heat engine according to claim 6, wherein the pressurized gas source contains substantially no oxygen. 8. Heat engine according to one of claims 5 to 7, wherein the device for pressurizing the drum (27) the compressor (10) which is selectively connected to the drum. 9. Heat engine according to one of claims 1 to 8, wherein the coolant introduced into the coolant inlet opening (28) is a gas and a vapor mixture of second working fluid supplied from the heat recovery exchanger (14, 15). 10. Heat engine according to one of claims 1 to 8, wherein the coolant introduced into the coolant inlet opening (28) is saturated vapor from the second working fluid supplied from the heat recovery exchanger (14, 15). 11. Heat engine according to one of claims 1 to 10, wherein the heat recovery exchanger (14, 15) further comprises: (i) a superheater (14) having an outlet coupled to the outlet of the heat recovery exchanger and an inlet, (ii) an evaporator (15) having an outlet coupled to the inlet of the superheater and an inlet coupled to the inlet of the heat recovery exchanger, and (iii) a heat recovery boiler arranged between the inlet and the outlet of the evaporator. 12. A heat engine according to claim 11, wherein the coolant inlet opening (28) supplies withdrawn second working fluid between the evaporator (15) and the superheater (14) in order to cool the turbine nozzles and/or blades (60) in the turbine (13). 13. A heat engine according to claim 11 or 12, further comprising a combination generator device which uses second fluid selectively withdrawn via a combination generator valve (17) between the boiler (27) and the superheater (14) and second fluid selectively withdrawn from the superheater outlet via a second valve (23) as a heat source. 14. A heat engine according to any one of claims 5 to 13, wherein the volume capacity of the boiler drum is sufficient to minimize a pressure increase due to closure of the control valve device (25) when the power demand drops and to enable the drum (27) to function as an energy storage device to enable rapid steam extraction as a supplementary heat energy source for the heat recovery boiler in the event of a sudden energy demand. 15. Method for operating a two-fluid heat engine with Cheng cycle with (1) a compressor (10) having a compressor outlet (2) for compressing a first working fluid, a combustion chamber (12) in flow connection with the compressor outlet (2), a turbine (13) with an inlet (4) in fluid connection with the combustion chamber (12) for performing work by expanding the working fluid and with a turbine outlet, (ii) a heat recovery exchanger (14, 15) connected to the turbine outlet for heating, with a heat recovery exchanger inlet and an outlet for heating a second working fluid, (iii) an injection device (3) for introducing heated second working fluid from the heat recovery exchanger (14, 15) into the combustion chamber (12), (iv) a coolant inlet port (28) for supplying coolant to turbine nozzles and/or blades (60) in the turbine (13), and (v) a control valve device (25) connected upstream of the injection device for selectively throttling the flow rate of second working fluid into the turbine (13), wherein in the method the flow rate of the second working fluid is throttled with the control valve device according to the following working parameters: (i) a temperature of second working fluid injected into the turbine that is less than or substantially equal to the temperature of first working fluid at the compressor outlet, and (ii) maximizing heat recovery from the turbine outlet at a mixing ratio of second to first working fluid in the turbine (13) at a given peak efficiency such that the clamping temperature of the heat recovery exchanger is minimized. 16. The method of claim 15, wherein the heat engine further comprises a pressurized gas source (29) and a pressure regulator selectively connected to the drum (27), a sensor system coupled to the gas turbine engine and the heat recovery steam generator for sensing temperature and pressure, and a control system for operating fuel flow to the gas turbine, and wherein the heat recovery steam generator comprises a superheater (14) having a superheater inlet, an evaporator (15) having an evaporator outlet coupled to the superheater inlet, and a heat recovery boiler having a drum arranged between the heat recovery inlet and the evaporator outlet, further comprising (a) starting conditions in the gas turbine engine and the heat recovery steam generator are initiated by means of the control system, (b) the control valve device is set for the idle flow condition of the heat engine, (c) the drum is pressurized by the compressed gas source to approximately the desired operating pressure of the heat recovery steam generator, (d) the gas turbine engine is ramped up from idle to full load, limited by the maximum permissible turbine inlet temperature, and whereby the compressed gas source is switched off for throttling if the sensor system indicates that the pressure of the heat recovery steam generator is higher than the compressor discharge pressure. 17. The method of claim 16, further comprising during the shutdown of the machine (a) the outlet of the gas turbine engine is removed from the load, (b) the control valve device (25) is closed to stop the flow of second working fluid into the gas turbine engine, (c) the flow of fuel to the gas turbine engine is shut off by means of the control system, and (d) the boiler is cooled down. 18. A method according to claim 16 or 17, wherein during pressurization the compressed gas source (29) serving to pressurize the drum is an independent compressed gas source and its delivery pressure to the drum is adjusted by means of the pressure regulator to approximately 5% below the working pressure of the heat recovery steam generator, and wherein the pressure regulator allows the return flow of first working fluid to the compressed gas source when the drum pressure exceeds the compressed gas delivery pressure. 19. The method of claim 18, wherein the independent pressurized gas source (19) is bottled nitrogen. 20. Method according to one of claims 16 to 19, wherein the order of pressurization (c) and start-up (d) is reversed as follows: the gas turbine engine is ramped up from idle to full load, limited by the maximum permissible turbine inlet temperature, and the drum (27) is pressurized with first working fluid delivered by the compressor (10) to approximately the desired working pressure of the heat recovery steam generator, so that the drum is pressurized and heated with exhaust gas from the gas turbine engine as a pressurized gas source. 21. A method according to any one of claims 16 to 20, wherein the heat engine further comprises a heat recovery steam generator (14, 15) connected between the heat recovery steam generator (14, 15) and the control valve device (25) a first valve (23) coupled to the machine, a second valve (17) connected to the heat recovery steam generator, a mixing chamber (22) coupled to the second and first valves, and a combination generator device coupled to the mixing chamber (22) downstream of the first and second valves (23; 17), Furthermore, for operating the heat engine in the combination generator mode (a) the control valve device (25) is closed to prevent introduction of second working fluid into the turbine; (b) the second valve (17) is opened to supply second working fluid to the mixing chamber (22), and (c) the first valve (23) is opened sufficiently to discharge second working fluid through the superheater (14) to counteract the possibility of oxidation thereof and to supply additional heat to the mixing chamber (22) for combination generator operation. 22. A method according to any one of claims 16 to 21, wherein the heat recovery steam generator (14, 15) has a line combustion device (6) to extract additional heat from the first working fluid between the superheater and the evaporator and to supply it to second working fluid in the heat recovery boiler, wherein first working fluid is further passed through the line combustion device and second heating fluid is heated in the boiler. 23. The method of claim 14 or 15, further comprising introducing coolant into the coolant inlet port (28) at a temperature below the temperature of the first working fluid at the compressor inlet.