JP2954456B2 - Exhaust recirculation combined plant - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to an exhaust heat recovery type combined cycle plant for recovering heat of a gas turbine exhaust gas, and more particularly to an exhaust gas recirculation type combined cycle for circulating exhaust gas from a gas turbine to its air inlet side. Regarding the plant.

[0002]

2. Description of the Related Art An exhaust heat recovery type combined cycle plant includes a gas turbine, an exhaust heat recovery boiler for recovering the exhaust heat thereof, and a steam turbine driven by steam generated by the exhaust heat recovery boiler. The generator is driven by the gas turbine and the steam turbine to obtain an electric output.

There are several types of exhaust heat recovery type combined cycle plants. One example is an exhaust gas recirculation type combined cycle plant that circulates gas turbine exhaust to its air inlet side. 64−
Something like 45924 is known.

[0004]

SUMMARY OF THE INVENTION Exhaust heat recovery type combined cycle plants have been used rapidly in recent years because of their rapid load fluctuation and high efficiency as compared with ordinary thermal power plants. However, there are some problems during partial load operation.

One of the problems is that the thermal efficiency is greatly reduced at the time of partial load. For example, if the thermal efficiency at the time of rating is 1, it is reduced to about 0.8 at 50% load and about 0.6 at 30% load. I will. Combined cycle plants have the advantage of higher load following performance than thermal power plants using ordinary boilers, and are often operated with load changes frequently.In this case, the thermal efficiency of the plant during partial load operation is reduced. We have to sacrifice the decline.

Further, there is a problem that the output of the exhaust heat recovery boiler at the time of partial load greatly fluctuates, and it is difficult to operate the plant. For example, assuming that the output power of the steam turbine generator at the time of rated load is 1, the output is reduced to about 0.43 at 50% load and about 0.12 at 30% load. In addition, the response time of the gas turbine generator (the time from when the amount of fuel supplied to the gas turbine combustor is increased or decreased until the power output of the gas turbine generator changes) is on the order of several seconds, whereas the response time of the steam turbine generator is The response time (the time from when the amount of fuel supplied to the gas turbine combustor is increased or decreased until the power generation output of the steam turbine generator changes) is on the order of several minutes. For this reason, it is very difficult to perform control for obtaining a desired output at the time of load change and immediately thereafter.

[0007] In this regard, the above-mentioned known example discloses that the gas turbine exhaust gas is circulated to the air inlet side. However, there is no description about the existence of the above-mentioned problem at the time of partial load operation and a specific solution thereto. Not shown.

In view of the above, it is an object of the present invention to provide an exhaust gas recirculation type combined plant which can solve the above-mentioned problems at the time of partial load.

[0009]

According to the present invention, an exhaust gas recirculation type combined plant for circulating exhaust gas from a gas turbine to a gas turbine compressor is provided. Then,
Increase the amount of recirculated gas as the load decreases. As a result, it is desirable to adjust the exhaust gas recirculation amount so as to keep the combustion temperature of the gas turbine combustor substantially constant.

[0010]

The combustion temperature can be maintained substantially constant by increasing the intake air temperature by circulating the exhaust gas of the gas turbine to the compressor and increasing the recirculation amount with the decrease in load. In the exhaust heat recovery boiler, since the exhaust temperature of the gas turbine does not change, it is possible to minimize the decrease in the amount of evaporation when the load is reduced.

[0011]

FIG. 1 shows an embodiment of the present invention. The combined cycle plant is roughly composed of a gas turbine device, an exhaust heat recovery boiler device, and a steam turbine device.

Among them, the gas turbine device is driven by a compressor 1 that draws in air and compresses it, a combustor 2 that burns using compressed air and fuel, and a high-temperature and high-pressure combustion gas from the combustor 2. And a gas turbine 3. In many cases, the compressor 1 and the gas turbine 3 are installed on the same axis, and the compressor 1 is driven by the gas turbine 3. A generator 6 is provided on the rotating shaft. The generator 6 is a synchronous machine, and is operated at a constant rotation speed, so that the intake air amount is usually kept constant.

The exhaust gas of the gas turbine 3 has a temperature of 500 degrees or more. In a combined cycle plant, this heat is recovered by a waste heat recovery boiler (HRSG) 4. Specifically, steam is generated by heat exchange between the exhaust gas and the feed water, and the steam is guided to the steam turbine 5 and rotated, thereby driving a generator coupled to the steam turbine 5. In the illustrated example, the steam turbine 5 is connected coaxially with the gas turbine 3, but this may be such that each turbine drives each generator.

Although the combined cycle plant is constructed as described above, in the present invention, a part of the exhaust gas of the gas turbine 3 is supplied to the air inlet of the compressor 1 through the pipe 9 and the recirculation amount control means 10. And reuse. The power generation output of the combined cycle plant of the present invention is determined by adjusting the opening of the fuel amount control valve 7 for controlling the amount of fuel supplied to the combustor 2 and the recirculation amount control means 10 as operating terminals. The position where a part of the exhaust gas is recirculated is set at the outlet of the gas turbine 3 in the figure, but it may be recirculated further downstream, for example, from a part of the exhaust heat recovery boiler. It is possible.

These operation terminals are controlled by operation signals from the general control device 8, and the general control device 8 receives the load request signal Ld from the central power supply command center 20 for the combined cycle plant and controls the entire plant. Although there is another operation target by the central control device 8, only the part related to the combustion control related to the present invention is described here. In short, the load request signal Ld from the central power supply command center 20 is input. Basically, the air amount and the fuel amount are appropriately controlled.

In order to control the fuel amount, first, a difference between the load request signal Ld and the actual load L is obtained by a subtractor AD1, and a fuel amount target signal Fd is obtained by an adjuster PI1. The difference between the fuel amount target signal Fd and the actual fuel amount F is subtracted by a subtractor A.
D2, the fuel amount control valve 7 is adjusted by the adjuster PI2 to determine the amount of fuel to be injected into the combustor. According to this control, as the load increases, the amount of fuel injected into the combustor 2 increases. The control up to this point is not different from the concept of the normally performed control.

The characteristic control of the present invention will be described in the following part, which mainly relates to the control of the recirculated air amount and the correction control of the fuel amount accordingly. In the present invention, for the recirculation air amount control, in the function generator FG1 to which the load request signal Ld is input, an output signal S1 which becomes larger as the load becomes lower is obtained. This signal S1 is provided to the controller PI3 to control the recirculation control means 10.

From the above description, it can be understood that the lower the load, the larger the amount of exhaust gas returned to the air inlet of the compressor 1 is. It has such technical meaning.

First, the gas turbine is rotating at a constant speed, so that the intake air amount (volume flow rate) may be considered to be constant irrespective of the load unless special air amount control is performed. On the other hand, since the fuel amount increases in proportion to the load, the lower the load, the more the amount of air becomes, and the combustion temperature and, consequently, the temperature of the gas turbine exhaust gas tend to decrease.

In the case of the present invention, since the outside air at the atmospheric temperature and the high-temperature gas turbine exhaust gas are mixed to form intake air, and the lower the load, the greater the amount of gas turbine exhaust gas recirculated, the higher the combustion temperature. As a result, it is possible to prevent the gas turbine exhaust gas temperature from decreasing as the load decreases. More desirably, the combustion temperature (gas turbine exhaust gas temperature) can be made substantially constant regardless of the load. From this point of view, the function generator FG1 of FIG. 1 determines the recirculation rate of the exhaust gas amount. Therefore, the output signal S1 of the function generator FG1 changes the gas turbine exhaust gas temperature almost regardless of the load in the illustrated example. It means a signal for keeping it constant.

In the present invention, the combustion temperature is constant in accordance with the output of the function generator 1. However, in actual operation, the combustion temperature may change, so the combustion temperature is determined from the gas turbine exhaust gas temperature and the compressor outlet pressure. The function generator 2 estimates and applies the output of the function generator 1 to the subtractor AD3 to correct the output of the function generator 1. When performing the correction control, the correction control is also performed on the fuel amount side in order to balance the fuel amount and the air amount.

Next, a description will be given of how a desired purpose can be achieved by the process configured and operated as shown in FIG.

FIG. 2 shows the relationship between the volume flow rate and the weight flow rate of the air-fuel mixture sucked into the compressor 1, and the relationship between the recirculation ratio and the load. The volume flow rate is constant regardless of the load. , The weight flow decreases with increasing hot air. The amount of recirculation when high-temperature air is recirculated from the gas turbine outlet is about 20% at 50% load and about 30% at 30% load.

The pressure ratio between the inlet and the outlet of the compressor 1 is uniquely determined by the blade shape of the compressor 1 and the axial flow velocity of the intake air. In the case of the present invention, the temperature of the air-fuel mixture sucked into the compressor 1 changes with the load (the higher the load, the higher the temperature), but the volume flow rate flowing into the compressor 1 does not change, and therefore the axial flow rate does not change. The pressure ratio does not change during the partial load operation. This process is close to an adiabatic change, where the temperature rises with increasing pressure. The ratio of the rise in the outlet temperature to the rise in the mixed air temperature from the outside air temperature is, for example, about 2 when the pressure ratio is 15.

The compressed air is heated in the combustor 2 to a combustion temperature under an isobaric process. According to the present invention, a decrease in the combustion temperature at the time of a load decrease is suppressed, and according to the embodiment of FIG. 1, the combustion temperature is kept constant even at a partial load. Next, the combustion gas works in the process of adiabatic expansion in the gas turbine 3, and a part thereof is consumed for driving the compressor 1 and the generator 6, so that the net output corresponds to the difference.

A part of the exhaust gas of the gas turbine 3 is recirculated as a part of the intake air of the compressor 1 via the pipe 9 and the control means 10. The remaining exhaust gas is recovered into enthalpy by the exhaust heat recovery boiler 4 and then released to the atmosphere. High-pressure steam is generated in the exhaust heat recovery boiler 4 and drives the steam turbine 5 and the generator 6 to generate power.

Next, a description will be given of the fact that a desired effect can be achieved by the present invention. First, FIG. 3 schematically shows the efficiency in a combined cycle plant without exhaust gas recirculation. _GT ,
Η _{ST for} steam turbine efficiency and η _B for waste heat recovery boiler efficiency
The total efficiency η can be expressed as the following (Equation 1).

[0028]

Η = η _GT + (1−η _GT ) η _ST η _B (Equation 1) On the other hand, FIG. 4 schematically shows the efficiency of the combined cycle plant of the present invention. Indicates the temperature of each part, and m indicates the ratio of the amount of recirculated gas.

[0029]

(Equation 2)

In Equation 2, when m = 0 without recirculation, T ₁ = T ₀ , and therefore ν = 1, and Equation 2 matches Equation 1. The second term in [] of Equation 2 indicates the contribution of the enthalpy recovery of the exhaust gas in the exhaust heat recovery boiler. In the present invention, since the heat recovered in the Brayton cycle is transferred to the Rankine cycle to perform work, the efficiency is improved accordingly.

FIG. 5 is a diagram showing the heat cycle of the present invention in the form of an enthalpy i-entropy S diagram, in which the dotted line is a characteristic diagram of the Carnot cycle, the gas turbine portion GT surrounded by a dashed line, and the exhaust heat recovery. The characteristic diagram when the sum of the portions HRSG and ST by the boiler (steam turbine) is a conventional system configuration in which recirculation is not performed, and the portion surrounded by oblique lines are the characteristic diagrams according to the present invention.

As is clear from this characteristic diagram, a portion m due to recirculation is added in the present invention, as compared with the conventional system configuration without recirculation. Thus, the process of the present invention improves the thermal efficiency by making the combustion process of the gas turbine close to the isothermal change characteristic of the Carnot cycle, and partially recovering the enthalpy of the gas turbine exhaust. In this figure, T2 indicates the combustor inlet temperature, T3 indicates the combustor outlet temperature, and T4 indicates the gas turbine outlet temperature.

As described above, the gas turbine exhaust gas of the present invention can be recirculated at any position.
Ratio of the exhaust gas temperature Tα of the exhaust gas recirculating m, showing the relationship between the compressor inlet air temperature T _1. However, the characteristics are as follows: atmospheric temperature T ₀ = 0 ° C., combustor outlet temperature T ₃ = 1200
This is a calculation of the compressor inlet intake air temperature T ₁ at a temperature of ° C. and an oxygen consumption rate of 5%. As is clear from this figure, the selection range of the recirculated gas temperature Tα and the bleed rate m is determined by the condition where the oxygen concentration in the gas turbine exhaust becomes zero, the zero output temperature of the gas turbine, and the exhaust temperature.

FIG. 7 shows the volumetric capacity of the intake air (dot-dash line), the weight flow rate (dotted line), the combustion temperature,
FIG. 5 is a diagram showing a change in gas turbine exhaust temperature, in which the combustion temperature and the gas turbine exhaust temperature are constant irrespective of the output, so that two control parameters are reduced and the operation becomes easy. If the recirculation flow is controlled to keep the compressor inlet volumetric flow constant, the air weight flow of the gas turbine will vary with load as shown in the figure. The weight flow rate at the minimum load is about 1/2 of the rated load.

From the above, an example of an output (load demand) -intake air temperature table for keeping the gas turbine exhaust gas temperature constant is shown. However, this characteristic is a characteristic when the atmospheric temperature is 0 ° C. Therefore, when the recirculation ratio m is 0 at a load of 100%, the intake air temperature is 0 °. And 50%
At the time of load, the intake air temperature may be set to 100 ° C., and from this, the recirculation rate of the gas turbine exhaust gas is determined. Similar characteristics can be similarly set at an arbitrary atmospheric temperature, and the characteristics of the function generator FG1 shown in FIG. 1 are defined by the relationship between the load request Ld and the recirculation ratio m using the atmospheric temperature as a parameter. It was done.

FIG. 9 shows the decrease in efficiency during partial load operation in comparison with the efficiency of a heat cycle (dotted line) without recirculation of gas turbine exhaust gas. In the present invention (solid line), the exhaust gas recirculation flow rate increases at the time of partial load, so that the enthalpy recovery rate of the exhaust gas increases, and therefore the efficiency decrease at the partial load is small. Further, since the axial flow velocity of the air in the compressor 1 does not change, no surge or critical flow occurs at the inlet, and the gas turbine 3 can theoretically be operated to zero output. Therefore, there is a feature that the plant can be operated stably even at a lower load than the conventional method.

FIG. 10 shows the output distribution of the topping cycle (gas turbine system) and the bottoming cycle (exhaust heat recovery boiler and steam turbine system). In the conventional method shown by the dotted line, the gas turbine exhaust gas temperature greatly decreases as the load decreases, and the amount of evaporation of steam generated in the exhaust heat recovery boiler fluctuates greatly. There is a problem that it is lowered. Further, since the bottoming cycle is composed of a complicated steam circuit, the settling time for disturbance is 5 to 10 minutes.
Minutes, and in order to obtain a plant with good follow-up to load requirements, it was necessary to reduce the output fluctuation of the bottoming cycle.

In the present invention, since the enthalpy of the exhaust gas from the gas turbine is recovered, the output of the bottoming cycle becomes slow as the plant output decreases, and the operation of the bottoming cycle having a large time constant with respect to the topping cycle is easy. In addition, there is an effect that the load followability of the plant is good. The effect of a small decrease in the bottoming cycle output is caused by the fact that the exhaust gas temperature of the gas turbine is constant irrespective of load fluctuation, and the slight decrease in efficiency at partial load is due to the decrease in exhaust flow rate. is there. Further, in the conventional method, the exhaust gas temperature fluctuates at the time of partial load. As a result, when the partial load operation is frequently performed, there is a problem that the components of the gas turbine and the HRSG are damaged due to thermal fatigue and creep deformation. According to the present invention, this problem can be prevented.

FIG. 11 is a diagram for explaining the effect of reducing NOx according to the present invention. First, when performing combustion using LNG as fuel in the combustor 2, 5% by weight of steam is generated. This water vapor is recirculated by the recirculation and is introduced again into the combustor.
x generation can be suppressed. This effect has a feature that the NOx reduction effect is greater because the partial load having a larger recirculation amount has a higher water vapor concentration in the intake air. FIG. 12 shows an output change with respect to relative humidity when the atmospheric temperature is used as a parameter (FIG. 12A), and also shows a thermal efficiency (FIG. 12B). As shown in FIG. The higher the ratio, the higher the cycle efficiency and the higher the output.

As shown in FIG. 13, the second embodiment is similar to the first embodiment.
Although the outlet of the pipe 9 is the outlet of the gas turbine 3 in the embodiment, the outlet is provided at the outlet of the HRSG 4 in this embodiment. In this case, there is an effect that the exhaust flow rate of the HRSG 4 and the output of the bottoming cycle are maintained at the values at the rated output even at the time of the partial load. That is, since the change in the cycle heat output is equal to the change in the output of the gas turbine 3, the operation control of the partial load is further facilitated. The cycle thermal efficiency in this case can be expressed by Equation 3. However, since the exhaust temperature of the extraction unit is low (see FIG. 5), the load range to which this embodiment can be applied is limited.

[0041]

(Equation 3)

FIG. 14 shows a third embodiment. In this embodiment, the heat transfer area of the HRSG 4 is increased, and the enthalpy of the recirculated exhaust gas is recovered by the HRSG 4 before being introduced into the compressor 1. Conventionally, the exhaust temperature at the outlet of the HRSG 4 cannot be reduced due to measures against white smoke. However, waste heat can be effectively recovered by the above-described method. The thermal cycle efficiency in this case can be expressed by Equation 4.

[0043]

Η = η _GT (1−η _GT ) ｛η _B + m (1−η _B )｝ η _ST (Equation 4) According to the first to third embodiments, the seasonal variation of the plant output is There is a common effect that it can be prevented. Although the output of the gas turbine 3 fluctuates due to a change in air temperature, in the case of the present invention, the above-described phenomenon can be avoided because the intake air temperature can be fixed. FIG. 15 shows a combination of the method shown in the third embodiment and the conventional example. A fourth embodiment is shown. Load adjustment is linked to the increase and decrease in air flow rate using the inlet guide vane (IGV) 11 provided at the compressor inlet.
It is kept constant regardless of the load. All of the above-described embodiments can be used for all types of combined plants using the gas turbine 3.

FIG. 16 shows a fifth embodiment in which a gas turbine is provided with a variable blade 12 for narrowing the flow path. In the above-described embodiment, when the mass flow rate of the compressor 1 is reduced while maintaining the pressure ratio, a surge or a stall phenomenon may occur. In order to avoid this, the opening degree of the variable blade is adjusted by the control system 8, and the outlet air pressure of the compressor 1 is controlled.

FIG. 17 shows a sixth embodiment having an intercooler 13. At a low partial load, the outlet temperature of the compressor 1 becomes high, and there is a possibility that the air extracted from the compressor 1 cannot effectively cool the blades of the gas turbine 3. In order to avoid this, the air extracted from the compressor is cooled by an intercooler. Further, if the combustion temperature is controlled by the control system 8, the blades of the gas turbine can be cooled more effectively. The fifth and sixth above
According to the embodiment, there is an effect of expanding the partial load zone.

Although FIG. 1 shows a single-shaft structure, the present invention is more suitable for a multi-shaft structure. The reason is that since the exhaust temperature of the gas turbine 3 does not change when the load changes, there is little change in the steam condition between the shafts. FIG. 18 shows an example of application to a multi-shaft combined plant. According to the present invention, even if each gas turbine is in a different load state, the load control of the steam turbine is easy because there is no large difference in the exhaust gas temperature from the gas turbine 3.

FIG. 19 shows a seventh embodiment. Compressor 1,
The gas turbine 3 and the steam turbine 5 are coaxial with each other, and a speed control mechanism 14 is provided at a connection portion with a shaft of the generator 6. Governing mechanism is gear, fluid coupling, thyristor, GTO, etc.
Any generator can be used as long as the generator can maintain a constant rotation speed with respect to a change in the rotation speed of the turbine system.
In the present embodiment, the number of revolutions of the turbine and the volume flow rate of the compressor are controlled in proportion to the square root of the temperature of the mixed gas at the inlet of the compressor. Thus, there is an advantage that the corrected flow rate and the corrected rotation speed of the compressor can be kept constant even at a partial load.

[0048]

According to the present invention, the thermal efficiency at the time of partial load can be increased. In addition, since the combustion temperature and the exhaust gas temperature from the gas turbine can be kept constant irrespective of the load, the operation and controllability is excellent, and the heat damage to the component materials is small. Operation is possible up to gas turbine zero-output operation (up to 30% load per shaft)
And the operating range is wide. Low NOx combustion is possible,
The NOx generation rate in the combustor can be reduced to 1/4 of the conventional rate. By setting the exhaust extraction point downstream of the HRSG, the output of the bottoming cycle can be kept constant during partial load,
Driving and controllability can be further improved. By providing an economizer in the recirculation pipe, the thermal efficiency of the bottoming cycle can be improved.

[Brief description of the drawings]

FIG. 1 is a diagram showing one embodiment of the present invention.

FIG. 2 is a diagram showing load dependence of a combustion temperature and an air flow rate.

FIG. 3 is a diagram showing a cycle configuration of a conventional combined plant.

FIG. 4 is a diagram showing a cycle configuration of the present invention having a recirculation system.

FIG. 5 is a diagram showing a TS diagram of a thermal cycle.

FIG. 6 is a diagram illustrating a selectable range of an extraction air temperature and an extraction amount of exhaust gas.

FIG. 7 is a diagram illustrating a relationship between an air weight flow rate and a load.

FIG. 8 is a diagram showing a relationship between a temperature of an air-fuel mixture and a plant output.

FIG. 9 is a diagram showing that the partial load efficiency of the present invention is good.

FIG. 10 is a diagram showing output distribution of topping and bottoming.

FIG. 11 is a diagram showing that significant reduction of NOx can be achieved by the present invention.

FIG. 12 is a diagram showing a relationship between relative humidity, output, and efficiency.

FIG. 13 is a diagram showing another embodiment of the present invention.

FIG. 14 is a diagram showing another embodiment of the present invention.

FIG. 15 is a diagram showing another embodiment of the present invention.

FIG. 16 is a diagram showing another embodiment of the present invention.

FIG. 17 is a diagram showing another embodiment of the present invention.

FIG. 18 is a diagram showing another embodiment of the present invention.

FIG. 19 is a diagram showing another embodiment of the present invention.

[Explanation of symbols]

DESCRIPTION OF SYMBOLS 1 ... Compressor, 2 ... Combustor, 3 ... Gas turbine, 4
... Exhaust heat recovery boiler (HRSG), 5 ... Steam turbine, 6 ...
Generator 7 Fuel supply system 8 Control system 9 Piping 10
... Valve, 11 ... IGV, 12 ... Variable wing, 13 ... Interk
La, 14 ... governing mechanism.

──────────────────────────────────────────────────続 き Continuing from the front page (72) Inventor Toshihiko Sasaki 3-1-1, Kochi-cho, Hitachi-shi, Ibaraki Pref. Hitachi, Ltd. Hitachi Plant (72) Inventor Hideaki Komatsu 3-1-1, Kochi-cho, Hitachi-shi, Ibaraki No. 1 Hitachi, Ltd., Hitachi Plant (72) Inventor Seiichi Kirikami 3-1-1, Sachimachi, Hitachi City, Ibaraki Prefecture Inside Hitachi, Ltd. Hitachi Plant (72) Inventor Takeshi Suzumura, Sachimachi, Hitachi City, Ibaraki Hitachi 1-1, Hitachi, Ltd. 3-chome Co., Ltd. (72) Inventor Tetsuo Sasada 3-1-1 Sayukicho, Hitachi-shi, Ibaraki Pref. Hitachi Plant, Hitachi, Ltd. (72) Takashi Ikeguchi Tsuchiura-shi, Ibaraki 502 Kandatecho Hitachi Machinery Co., Ltd. (72) Inventor Naruhisa Sugita 502 Kandachicho, Tsuchiura-shi, Ibaraki Pref. Hitachi, Ltd. The laboratory (56) Reference Patent Sho 53-13014 (JP, A) JP Akira 53-143812 (JP, A) (58 ) investigated the field (Int.Cl. ^6, DB name) F02C 3/30 F02C 6/18 F02C 7/08 F02C 9/18

Claims (5)

(57) [Claims] 1. A gas turbine is driven by supplying air compressed by a compressor to a combustor to burn fuel and generate steam by a waste heat recovery boiler using gas turbine exhaust gas as a heat source to drive a steam turbine. In an exhaust gas recirculation plant in which a part of gas turbine exhaust gas is returned to a compressor inlet, a gas turbine associated with a load fluctuation of the combined plant is used.
The compressor so as to suppress fluctuations in the exhaust gas temperature of the compressor.
Adjusting the amount of gas turbine exhaust gas returned to the inlet
Exhaust recirculation plant. 2. A gas turbine is driven by supplying air compressed by a compressor to a combustor to burn fuel, and a steam is generated by an exhaust heat recovery boiler using gas turbine exhaust gas as a heat source to drive a steam turbine. In an exhaust gas recirculation plant in which part of the gas turbine exhaust gas is returned to the compressor inlet, the amount of gas turbine exhaust gas returned to the compressor inlet is adjusted to keep the combustion temperature in the gas turbine combustor constant. Exhaust recirculation type combined plant. 3. A gas turbine is driven by supplying air compressed by a compressor to a combustor to burn fuel, and a steam is generated by an exhaust heat recovery boiler using gas turbine exhaust gas as a heat source to drive a steam turbine. In an exhaust gas recirculation plant adapted to return a part of the gas turbine exhaust gas to the compressor inlet, the exhaust gas recirculation is characterized by increasing the temperature of the gas turbine exhaust gas returned to the compressor inlet as the load on the combined plant decreases. Type combined plant. 4. A compressor having a compressor for sucking air therein and compressing the compressed air, a combustor for burning compressed air and fuel compressed by the compressor, and a gas turbine driven by combustion exhaust gas from the combustor. In an exhaust gas recirculation type gas turbine device configured to return a part of the gas turbine exhaust gas to the compressor inlet, the exhaust gas of the gas turbine accompanying the load fluctuation of the gas turbine is reduced.
At the compressor inlet so as to suppress temperature fluctuations.
An exhaust gas recirculation type gas turbine device, wherein the amount of gas exhaust gas to be returned is adjusted . 5. A compressor having a compressor for sucking air therein and compressing it, a combustor for burning compressed air and fuel compressed by the compressor, and a gas turbine driven by combustion exhaust gas from the combustor. In an exhaust gas recirculation type gas turbine device in which part of the gas turbine exhaust gas is returned to the compressor inlet, adjusting the amount of the gas turbine exhaust gas returned to the compressor inlet to keep the combustion temperature of the gas turbine combustor constant. Exhaust recirculation type gas turbine device.