JP5630977B2 - Intake air cooling apparatus for gas turbine, gas turbine and gas turbine combined cycle power plant equipped with the same, and output increasing method - Google Patents

Description

  The present invention relates to an intake air cooling device for a gas turbine, a gas turbine and a gas turbine combined cycle power plant including the same, and an output increasing method.

  2. Description of the Related Art Conventionally, in a power plant equipped with a gas turbine mainly composed of a compressor, a combustor, and a turbine, a gas turbine combined cycle (hereinafter referred to as “GTCC”) having exhaust heat utilization equipment utilizing exhaust heat exhausted from the turbine. The power plant is known. As a specific example, there is one that includes a boiler that uses exhaust heat of a gas turbine and a steam turbine that uses steam generated by the boiler. In this GTCC power generation plant, power generation can be performed by a steam turbine that uses exhaust heat of the turbine in addition to the gas turbine, so that the power generation efficiency can be improved as a whole.

In such a GTCC power plant, the output is affected by the temperature of the intake air sucked into the compressor. That is, especially in the summer, the atmospheric temperature rises, so the density of the intake air decreases, the mass flow rate decreases, and the output decreases. In order to suppress such a decrease in output, some have a cooling device for cooling the intake air. By the way, when the intake air is cooled, when the temperature of the intake air becomes lower than the dew point temperature, water vapor in the intake air is condensed and mist is generated.
When this mist flows into the gas turbine, the rotor blades and the stationary blades of the compressor are damaged such as erosion, and the durability is adversely affected.

  In the following Patent Document 1, a gas turbine having a gas turbine and a refrigeration facility for cooling the intake air, a mist separator for removing mist condensed by cooling of the refrigeration facility is provided between the refrigeration facility and the gas turbine. The installed gas turbine equipment is described.

JP 2000-145477 A

However, in the prior art, there is a problem that the efficiency of the gas turbine is reduced because pressure loss occurs when the intake air passes through the mist separator.
Moreover, in the prior art, since another apparatus such as a mist separator is newly added to the plant facility, there is a problem that the cost and maintenance performance of the entire facility are deteriorated.

  The present invention has been made in view of such circumstances, and the object thereof is to prevent the gas turbine from sucking mist without adding another device on the flow path of the cooled intake air. An object of the present invention is to provide an intake air cooling device for a gas turbine, a gas turbine and a gas turbine combined cycle power plant including the same, and a method for increasing output.

In order to achieve the above object, the present invention employs the following means.
That is, an intake air cooling device for a gas turbine according to the present invention is an intake air cooling device for a gas turbine that is used in a gas turbine including a compressor, a combustor, and a turbine, and is an intake air that is sucked into the compressor from the outside. A heat exchanger capable of cooling the heat exchanger, the heat exchanger extending over an upper side and a lower side in the direction of gravity, and a plurality of heat transfer tubes intersecting the air flow direction of the intake air, and the plurality of heat exchanger tubes of and a cooling medium flowing inside the heat transfer tubes, the plurality of heat transfer tubes, the so that Do greater cooling capacity than the heat transfer tube heat exchanger tube located on the airflow upstream side is positioned on the airflow downstream side, the The heat transfer tube on the upstream side of the airflow has a larger orifice diameter than the heat transfer tube on the downstream side of the airflow .
According to this configuration, the heat exchanger capable of cooling the intake air includes a plurality of heat transfer tubes, and the plurality of heat transfer tubes are cooled more than the heat transfer tubes located on the airflow upstream side than the heat transfer tubes located on the airflow downstream side. Since the capacity is arranged so as to increase, when the intake air is cooled below the dew point temperature, the heat transfer tube located on the upstream side of the airflow relatively generates mist and is located on the downstream side of the airflow. Mist generation is relatively suppressed in the heat transfer tube. Then, the mist generated by being cooled by the heat transfer tube located on the upstream side of the air flow adheres to the heat transfer tube on the downstream side of the air flow, falls over the surface, and falls downward in the gravity direction.
As a result, even if mist is generated along with the cooling of the intake air, the mist is captured by the heat transfer tube itself that cools the intake air as compared with the case where the cooling capacity is made constant from the upstream side to the downstream side of the airflow. Since the ratio increases, it is possible to suppress the mist from flowing to the compressor inlet of the gas turbine located on the downstream side of the airflow.
Therefore, it is possible to effectively cool the intake air and prevent the gas turbine from sucking mist without adding another device on the flow path.

Furthermore, the heat transfer tube on the upstream side of the airflow has a larger orifice diameter than the heat transfer tube on the downstream side of the airflow, so the flow rate of the cooling medium in the heat transfer tube on the upstream side of the airflow can be reduced with a relatively simple configuration. The flow rate of the cooling medium in the heat transfer tube on the side can be increased, and the cooling effect of the heat transfer tube on the upstream side of the airflow can be relatively increased. In addition, it is possible to relatively easily adjust the cooling effect.

An intake air cooling device for a gas turbine according to the present invention is an intake air cooling device for a gas turbine that is used in a gas turbine including a compressor, a combustor, and a turbine. The intake air is sucked into the compressor from the outside. A heat exchanger capable of cooling the heat exchanger, the heat exchanger extending over an upper side and a lower side in the direction of gravity, and a plurality of heat transfer tubes intersecting the air flow direction of the intake air, and the plurality of heat exchanger tubes A plurality of heat transfer tubes, and the plurality of heat transfer tubes have a cooling capacity greater than that of the heat transfer tubes located on the upstream side of the airflow than the heat transfer tubes located on the downstream side of the airflow. The upstream heat transfer tube has at least one of an inner surface and an outer surface rougher than the heat transfer tube on the downstream side of the airflow.
According to this configuration, since at least one of the inner surface and the outer surface of the heat transfer tube on the upstream side of the airflow is rough, the flow of the cooling medium or the suction air can be disturbed, and the cooling effect of the heat transfer tube on the upstream side of the airflow Can be made relatively large.

An intake air cooling device for a gas turbine according to the present invention is an intake air cooling device for a gas turbine that is used in a gas turbine including a compressor, a combustor, and a turbine. The intake air is sucked into the compressor from the outside. A heat exchanger capable of cooling the heat exchanger, the heat exchanger extending over an upper side and a lower side in the direction of gravity, and a plurality of heat transfer tubes intersecting the air flow direction of the intake air, and the plurality of heat exchanger tubes A plurality of heat transfer tubes, and the plurality of heat transfer tubes have a cooling capacity greater than that of the heat transfer tubes located on the upstream side of the airflow than the heat transfer tubes located on the downstream side of the airflow. The upstream side heat transfer tube has a larger flow area than the downstream side heat transfer tube.
According to this configuration, the flow path area of the heat transfer tube on the upstream side of the airflow is larger than that of the heat transfer tube on the downstream side of the airflow, so the cooling medium flow rate of the heat transfer tube on the upstream side of the airflow is relatively simple. Becomes relatively large, the temperature rise of the cooling medium is reduced, and the cooling effect of the heat transfer tube on the upstream side of the airflow can be relatively increased.

The heat transfer tube on the upstream side of the airflow is characterized in that the temperature of the cooling medium is lower than that of the heat transfer tube on the downstream side of the airflow.
According to this configuration, since the temperature of the cooling medium of the heat transfer tube on the upstream side of the airflow is lower than that of the heat transfer tube on the downstream side of the airflow, the cooling effect of the heat transfer tube on the upstream side of the airflow can be relatively increased. In addition, the cooling effect can be adjusted relatively easily.

Further, the two heat transfer tubes adjacent to each other in the airflow direction are set so that the flow directions of the cooling medium are opposite to each other.
According to this configuration, since the two heat transfer tubes adjacent to each other in the airflow direction are set so that the flow direction of the cooling medium is opposite, it is possible to reduce the temperature difference above and below the gravity direction of the intake air. it can. That is, in each heat transfer tube, since the cooling medium heated by exchanging heat with the suction air on the upstream side in the cooling medium flow flows toward the downstream side in the cooling medium flow, the downstream side in the cooling medium flow Compared with the above, the cooling effect on the upstream side becomes higher. However, since the two heat transfer tubes adjacent to each other in the airflow direction are set so that the flow direction of the cooling medium is opposite, in these two heat transfer tubes, the cooling effect is equalized across the top and bottom in the direction of gravity. Is done. For this reason, if the suction air continuously flows along these two heat transfer tubes, the suction air will be cooled uniformly over the gravity direction of the suction air, so the temperature difference between the suction air in the gravity direction is reduced. can do.

Moreover, the 1st connection heat exchanger tube unit formed by each upper end of the two heat exchanger tubes adjacent to each other in the said airflow direction is provided.
According to this configuration, since the first connection heat transfer tube unit is formed by connecting the upper ends of the two heat transfer tubes adjacent to each other in the airflow direction, the flow direction of the cooling medium is opposite with a simple configuration. Can be set to

Moreover, the 2nd connection heat exchanger tube unit formed by each connecting the lower end of two heat exchanger tubes mutually adjacent | abutted in the said airflow direction is characterized by the above-mentioned.
According to this configuration, since the second connection heat transfer tube unit in which the lower ends of the two heat transfer tubes adjacent to each other in the airflow direction are connected is provided, the flow direction of the cooling medium is opposite with a simple configuration. Can be set to

In addition, at least a part of the plurality of heat transfer tubes includes a fin member extending radially outward from the outer peripheral surface.
According to this configuration, since the fin member extending radially outward from the outer periphery is provided, the contact area with respect to the intake air is increased. That is, the cooling effect is improved by increasing the heat transfer area, and when the heat transfer tube is provided on the downstream side, the mist adhesion amount is increased. Thereby, while cooling of suction air is accelerated | stimulated, the mist capture | acquisition effect can be improved.

The fin member may have a spiral shape.
According to this configuration, since the fin member has a spiral shape, the captured mist falls continuously over the fin member. Thereby, it is possible to quickly guide the captured mist downward and to prevent the captured mist from being blown to the downstream side of the airflow by the intake air.

The fin member has a funnel shape.
According to this configuration, since the fin member has a funnel shape, the captured mist is collected on the outer peripheral side of the heat transfer tube, and continuously falls over the surface of the heat transfer tube. Thereby, it is possible to quickly guide the captured mist downward and to prevent the captured mist from being blown to the downstream side of the airflow by the intake air.

Further, the plurality of heat transfer tubes are arranged in a staggered manner when viewed from above, and a part of the plurality of heat transfer tubes is in a flow path width direction intersecting the air flow direction and the gravity direction, A plurality of heat transfer tube stages are arranged at intervals, the heat transfer tube stages are stacked in a plurality in the airflow direction, and the heat transfer tube stages adjacent to each other in the airflow direction are It is characterized by being arranged with a gap in the flow path width direction.
According to this configuration, since the plurality of heat transfer tubes are arranged in a staggered manner, the air flow direction of the heat transfer tube stage on the upstream side of the air flow and the gravity direction intersect among the heat transfer tube stages adjacent to each other in the air flow direction. The suction air that has passed through the gap in the flow path width direction can be sufficiently brought into contact with each heat transfer tube in the heat transfer tube stage on the downstream side of the airflow. Thereby, the cooling effect and the mist capturing effect can be enhanced.

Further, the center-to-center distance between two heat transfer tubes adjacent to each other in the flow path width direction is set to be twice or less the outer diameter of the heat transfer tube.
According to this configuration, since the center-to-center distance between two heat transfer tubes adjacent to each other in the flow path width direction is not more than twice the outer diameter of the heat transfer tube, the suction that has passed through the heat transfer tube stage on the upstream side of the airflow Air can be reliably brought into contact with each heat transfer tube in the heat transfer tube stage on the downstream side of the airflow.

Further, in one of the two heat transfer tube stages adjacent to each other in the air flow direction, a wall surface interval from the outer peripheral surface of the heat transfer tube located at one end in the flow channel width direction to the flow channel wall surface of the suction air On the other hand, at least a part of the outer peripheral surface of the heat transfer tube located at one end in the flow path width direction is located within the wall surface interval in the one when viewed from the air flow direction. It is characterized by that.
According to this configuration, the suction air passing between the heat transfer tube located at one end of the heat transfer tube stage on the upstream side of the airflow and the flow path wall surface becomes the flow width direction of the heat transfer tube stage on the downstream side of the airflow. It can be made to contact the heat exchanger tube located at one end. Thereby, it is possible to prevent the intake air that passes through the plurality of heat transfer tubes without coming into contact with the heat transfer tubes along the vicinity of the wall surface at one end side in the flow path width direction, thereby enhancing the cooling effect and the mist capturing effect. it can.

Further, in one of the two heat transfer tube stages adjacent to each other in the air flow direction, the wall surface distance from the outer peripheral surface of the heat transfer tube located at the other end in the flow channel width direction to the flow channel wall surface of the suction air is The outer diameter of the heat transfer tube is equal to or smaller than the outer diameter of the heat transfer tube. It is characterized by being located.
According to this configuration, the suction air passing between the heat transfer tube located at the other end in the flow path width direction of the heat transfer tube stage on the upstream side of the air flow and the flow path wall surface is the flow path width of the heat transfer tube stage on the downstream side of the air flow. It can be made to contact the heat transfer tube located at the other end in the direction. As a result, the intake air that passes through the plurality of heat transfer tubes without contacting the heat transfer tubes along the vicinity of the wall surface at the other end side in the flow path width direction can be prevented, and the cooling effect and the mist trapping effect are enhanced. Can do.

In addition, a gas turbine according to the present invention includes any one of the above-described gas turbine intake air cooling devices.
According to this configuration, since any one of the above-described gas turbine intake air cooling devices is provided, it is possible to prevent the mist from being sucked into the gas turbine without adding another device on the flow path of the cooled intake air. it can.

A gas turbine combined cycle power plant according to the present invention uses any one of the above-described gas turbine intake air cooling devices, a gas turbine including a compressor, a combustor, and a turbine, and exhaust heat from the gas turbine. And a waste heat utilization means.
According to this configuration, since any one of the above-described gas turbine intake air cooling device and the gas turbine is provided, the mist is sucked into the gas turbine without adding another device on the flow path of the cooled intake air. Can be deterred.

The output increasing method according to the present invention is characterized in that any of the above-described gas turbine intake air cooling devices is additionally installed in an existing gas turbine or gas turbine combined cycle power plant.
According to this configuration, since any one of the above-described gas turbine intake air cooling devices is provided, it is possible to prevent the mist from being sucked into the gas turbine without adding another device on the flow path of the cooled intake air. it can.

  The gas turbine intake air cooling device according to the present invention can prevent the gas turbine from sucking mist without adding another device on the flow path of the cooled intake air.

  Further, according to the gas turbine and the gas turbine combined cycle power plant according to the present invention, it is possible to prevent the mist from being sucked into the gas turbine without adding another device on the flow path of the cooled intake air. .

It is a schematic block diagram of GTCC power plant G1 concerning the embodiment of the present invention. It is a horizontal sectional view showing a schematic structure of the first heat exchanger 21 concerning the embodiment of the present invention. It is schematic structure sectional drawing of the 1st heat exchanger 21 which concerns on embodiment of this invention, Comprising: It is the II sectional view taken on the line in FIG. It is a block diagram which shows schematic structure of the cooling control part 24 of the intake-air-cooling apparatus 2 for gas turbines concerning embodiment of this invention. It is a figure which shows the criterion of the determination means 43 of the cooling control part 24 of the intake-air-cooling apparatus 2 for gas turbines which concerns on embodiment of this invention. It is a block diagram which shows schematic structure of the dispersion | distribution control part 60 of the intake-air humidity control apparatus 3 for gas turbines which concerns on embodiment of this invention. It is operation | movement explanatory drawing of the 1st heat exchanger 21 which concerns on embodiment of this invention. FIG. 6 is a schematic cross-sectional view showing a modified example 21A of the first heat exchanger 21 according to the embodiment of the present invention, which corresponds to FIG. 3. It is a schematic block diagram of the fin member 32c which concerns on embodiment of this invention. It is a schematic block diagram of the fin member 32d which concerns on embodiment of this invention. It is a principal part expanded sectional view which shows the modification 21B of the 1st heat exchanger 21 which concerns on embodiment of this invention, Comprising: It is a figure which shows the heat exchanger tube 32 which varied flow path area (alpha). It is a principal part expanded sectional view which shows the modification 21C of the 1st heat exchanger 21 which concerns on embodiment of this invention, Comprising: It is a figure which shows the heat exchanger tube 32 provided with the orifice 32f. It is a figure of the plant composition which omitted the intake air humidity control device for gas turbines concerning the embodiment of the present invention.

Embodiments of the present invention will be described below with reference to the drawings.
FIG. 1 is a schematic configuration diagram of a GTCC power plant G1 according to an embodiment of the present invention. As shown in FIG. 1, the GTCC power plant G1 includes a power generation gas turbine 1, a gas turbine intake air cooling device 2, a gas turbine intake air humidity control device 3, a generator 4, a boiler, and a steam turbine. And a waste heat utilization means (not shown) schematically configured.

  The power generation gas turbine 1 compresses the intake air A sucked from the atmosphere (external air) into the compressor 1a, supplies the compressed air to the combustor 1b, and mixes it with the fuel F in the combustor 1b for combustion. Thus, by generating combustion gas and supplying it into the turbine 1c, it is possible to rotate the rotor 1d with a blade structure (not shown) and generate power with the generator 4. Further, the combustion gas flowing through the turbine 1c is supplied to exhaust heat utilization means (boiler, etc.) (not shown) as exhaust gas g1, and steam is generated using the exhaust heat.

The gas turbine intake air cooling device 2 includes a first heat exchanger (heat exchanger) 21, a refrigerator 22, a second heat exchanger 23, cold water C1, cooling water C2, a cooling control unit 24, A storage unit 25 (see FIG. 4), an atmospheric temperature measurement unit 26, and an atmospheric humidity measurement unit 27 are provided.
The 1st heat exchanger 21 is provided in the suction flow path from the exterior to the compressor 1a inlet, and heat-exchanges between the cold water C1 and the suction air A which flow through the inside. The heat received by the cold water C1 from the intake air A is released to the outside through the refrigerator 22, the cooling water C2, and the second heat exchanger 23.
Note that means other than the refrigerator 22 may be used as long as the intake air A can be cooled by the first heat exchanger 21.

2 is a horizontal sectional view showing a schematic configuration of the first heat exchanger 21, and FIG. 3 is a sectional view taken along the line II in FIG.
As shown in FIGS. 2 and 3, the first heat exchanger 21 includes a pipe line 31 that also serves as a suction flow path, a plurality of heat transfer pipes 32, and a drain 33.

  The pipe line 31 intersects the gravity direction with the gravity direction, and constitutes a suction flow path for the suction air A along a substantially horizontal direction. As shown in FIG. 1, the intake air A that has passed through the pipe line 31 is sucked into the compressor 1 a via the gas turbine intake humidity control device 3.

  As shown in FIG. 3, the heat transfer tube 32 is erected so as to penetrate the pipe line 31 vertically in the gravitational direction, and each axis is directed in the substantially gravitational direction.

As shown in FIG. 2, the plurality of heat transfer tubes 32 are arranged in a staggered manner in the horizontal direction in the flow path.
Specifically, some of the plurality of heat transfer tubes 32 are spaced apart in the flow channel width direction (direction intersecting the gravity direction and the airflow direction) (the gravity direction and the airflow direction). As shown in FIG. 3, four (35A to 35D) are superposed on the heat transfer tube stages 35 arranged in a row at intervals in the airflow direction. In each of the heat transfer tube stages 35A to 35D, as shown in FIG. 2, the distance L between the centers of the heat transfer tubes 32 adjacent to each other in the flow path width direction is not more than twice the outer diameter D of the heat transfer tube 32 (L ≦ 2D). ) Is set.
As shown in FIG. 2, the heat transfer tube stages 35 (35A to 35D) adjacent to each other in the airflow direction are arranged with a gap in the flow path width direction being shifted. More specifically, of the two heat transfer tube stages 35 (for example, 35A and 35B) adjacent to each other in the airflow direction, the center of the heat transfer tube 32 of the heat transfer tube stage 35 (for example, 35B) on the downstream side of the airflow is When viewed from the air flow direction, the heat transfer tube stage 35 (for example, 35A) on the upstream side of the air flow is located at the center of the interval between the heat transfer tubes 32 adjacent to each other in the flow path width direction (in other words, in the middle of the center-to-center distance L). It is arranged to do.

Further, the arrangement of the heat transfer tube stages 35A to 35D at each end in the flow path width direction is as shown in FIG. 2, where the wall surface distance from the heat transfer tube 32 to the flow wall 31a at each end is S. It is set as follows. First, the wall surface spacing S in the heat transfer tube stages 35A and 35C is set to be equal to or smaller than the outer diameter D of the heat transfer tube 32 (S ≦ D).
The wall surface spacing S between the heat transfer tube stages 35B and 35D is set to be much smaller than the outer diameter D (S << D).
Thus, in the two heat transfer tube stages 35 (35A to 35D) adjacent to each other in the airflow direction (for example, 35A and 35B), when viewed from the airflow direction, one of the heat transfer tube stages 35 (for example, 35B) A part of the outer peripheral surface of the heat transfer tube 32 at one end in the channel width direction is set so as to be positioned within the wall surface interval S of the other heat transfer tube stage 35 (for example, 35A).

  Each of the heat transfer tube stages 35A to 35D is supplied with cold water C1 having a different temperature from the refrigerator 22, and the temperature of the cold water C1 supplied to each heat transfer tube 32 of the heat transfer tube stage 35A is as follows. It is the lowest, and the temperature of the cold water C1 supplied in the order of the heat transfer tube stages 35B, 35C, 35D is increased. Note that the flow rate of the cold water C1 is substantially constant in each heat transfer tube 32, and flows from below to above.

  The drain 33 is provided in the lower part of the pipe line 31 on the downstream side of the heat transfer pipe stage 35D. A suction pump (not shown) is connected to the drain 33. In addition, the lower part of the pipe line 31 is inclined so as to gradually go downward from the upstream side of the heat transfer pipe stage 35 </ b> A toward the drain 33.

Returning to FIG. 1, the cooling control unit 24 of the gas turbine intake air cooling apparatus 2, a suction air A is cooled by the first heat exchanger 21 in accordance with the required output W _PR of power generation gas turbine 1. The cooling control unit 24 compares the dew point output W _PDT and the required output W _PR when the compressor inlet temperature T _IN of the intake air A at the inlet of the compressor 1a is operated at the dew point temperature T _d and the required output W _PR. in is large, balance associated with power generation on the basis of the set electricity price P _E is the condition that satisfies a predetermined criterion, the compressor inlet temperature T _iN to below the dew point temperature T _d.

FIG. 4 is a block diagram showing a schematic configuration of the cooling control unit 24 of the gas turbine intake air cooling device 2.
Cooling control unit 24, a required output _{W PR} and differential revenue calculation means 41 for calculating a difference income INC based on the difference output ΔW and electricity price _{P E} is the difference between the dew point output _{W PDT,} saturation dew point temperature _{T d} difference cost calculating means for calculating a fuel difference cost C _F is the fuel cost increment power generation gas turbine 1 when air is cooled to the required output W _PR is the compressor inlet temperature T _iN corresponding to the required inlet temperature T _PR 42, by comparing the difference revenue INC and the difference cost C _C, and a determination means 43 whether or not to below the dew point temperature T _d of the compressor inlet temperature T _iN.
In the present embodiment, of the difference cost C _C consisting of consumable costs and fuel differential cost C _F etc. of the blade or the like, and fiction fuel differential cost C _F occupying most as the difference cost C _C.

  The differential income calculation means 41 includes a dew point temperature calculation unit 41a, a dew point output calculation unit 41b, a differential output calculation unit 41c, a differential income calculation unit 41d, and a differential refrigeration power calculation unit 41e.

The dew point temperature calculation unit 41 a calculates the dew point temperature T _d from the atmospheric humidity φ input from the atmospheric humidity measurement unit 27 and the atmospheric temperature T _amb input from the atmospheric temperature measurement unit 26. The dew point temperature _Td may be obtained from the above-described NC diagram (for example, known from “How to read and use the thorough master air diagram”, edited by the Society for Air Conditioning and Sanitation Engineering, 1998, pp16)). And it is good also as a structure calculated | required directly from a dew point thermometer instead of the dew point temperature calculating part 41a.

Dew point output calculation unit 41b calculates the dew point output W _PDT is output when the compressor inlet temperature T _IN from dew point temperature T _d calculated in dew point temperature calculating section 41a becomes dew point temperature T _d. Incidentally, correspondence between the dew point temperature _{T d} and dew point output _{W PDT,} for example, "Gas Turbine Theory 5th Edition" , Sarabanamuttoo, HIH, et al. , 2001 ", etc. Alternatively, it may be obtained from an experiment or a trial run.

Difference output calculation unit 41c, the operation and the dew point output _{W PDT} that is calculated on the dew point output calculation unit 41b, based on the required output _{W PR,} the difference output ΔW which is the difference between the dew point output _{W PDT} and the required output _{W PR} To do.

The difference income calculation unit 41d is a power price stored in advance in the storage unit 25 as a difference between the difference output ΔW calculated by the difference output calculation unit 41c and the power increment ΔW _IN calculated by the difference refrigeration power calculation means 41e. multiplied by the P _E, it calculates the difference between revenue INC.
The power price P _E, the more recent ones are preferred.

  The differential refrigeration power calculation means 41e includes a required inlet temperature calculation unit 41e1, a specific enthalpy difference calculation unit 41e2, a differential refrigeration capacity calculation unit 41e3, and a differential refrigeration power calculation unit 41e4.

Request inlet temperature calculation section 41e1 includes a required output W _PR which is input from the outside based on the operating time of the power demand, is stored in advance in the storage unit 25, the required output W compressor inlet temperature required to obtain a _PR T predetermined relationship with a _iN request inlet temperature _{T PR} and the required output _{W PR} (in Fig. 4 (1)) from, and calculates the required inlet temperature _{T PR.}
Here, the electric power demand includes electric power that can be sold. For example, when electric power can be sold by private power generation, the electric power that can be sold or the rest of the electric power sold to the customer by the electric power company In the case where power can be sold to other electric power companies, the electric power that can be sold to other electric power companies is included.
The predetermined relationship between the required output W _PR and the required inlet temperature T _PR is, for example, “Gas Turbine Theory 5th Edition”, Sarabanamuttoo, HIH, et al. , 2001 "etc. can be used.

The specific enthalpy difference calculation unit 41e2 calculates the dew point temperature T _d calculated by the dew point temperature calculation unit 41a from the atmospheric temperature T _amb input from the atmospheric temperature measurement unit 26 and the atmospheric humidity φ input from the atmospheric humidity measurement unit 27. and request the inlet temperature calculation unit NC diagram stored in advance in the storage unit 25 from the computed required inlet temperature T _PR to 41e1 (in FIG. 4 (2), as known above) based on the dew-point temperature T calculates the specific enthalpy difference Δh in the case of cooling _d of the saturated air to the required inlet temperature T _PR.

Differential cooling capacity calculating portion 41e3 calculates the difference refrigerating capacity ΔH obtained by the product of the flow rate G _IN of the computed specific enthalpy difference Δh and the suction air A to a specific enthalpy difference calculation section 41E2. As illustrated in FIG. 4, the flow rate G _IN of the intake air A may be obtained from the gas turbine control device 1 e, or a separate gas turbine intake air flow rate calculation unit may be provided so that the intake air from the compressor inlet temperature T _IN it may be configured for calculating the flow rate _{G iN} of a. In the case of providing a gas turbine intake flow rate calculation unit, for example, “Gas Turbine Theory 5th Edition”, Sarabanamuttoo, HIH, et al. , It can be calculated the flow rate _{G IN} of the suction air A using the method depicted in 2001 "or the like.

Difference refrigeration power calculating portion 41e4, based on the difference refrigerating capacity ΔH that is calculated on the difference refrigerating capacity calculating portion 41e3 and the COP of the refrigerator 22, and calculates the freezing mobility increment [Delta] W _IN. As shown in FIG. 4, the coefficient of performance COP of the refrigerator 22 may be provided from the refrigerator control device 22 a or may be provided from the storage unit 25.

The differential cost calculation means 42 includes a fuel differential cost calculation unit 42a.
Fuel difference cost calculation unit 42a includes a differential output ΔW that is calculated on the difference output calculation unit 41c, per unit calorific value stored in advance in the storage unit 25 the fuel price P _F and the power generation efficiency E _{G (in} FIG. 4 (3 )) And the fuel differential cost _CF is calculated. The power generation efficiency E _G, as illustrated in FIG. 4, may be configured to provide a gas turbine control device 1e.
Incidentally, the power generation efficiency E _G is a value obtained by dividing the fuel heating value to inject the output (W) of GTCC power plant G1 not deducted frozen mobility (Q).

FIG. 5 is a diagram illustrating the determination criteria of the determination unit 43.
The determination unit 43 determines whether or not the required output W _PR is larger than the dew point output W _PDT and the differential income INC is larger than the differential cost C _C (fuel differential cost C _F ).

Based on the determination result of the determination unit 43, the cooling control unit 24, when W _PR ≦ W _PDT , is (difference income INC−difference cost C _C ) ≦ 0 (balance is deficit or 0), and, (differential revenue INC- difference cost _C C)> 0 ... (balance black) at both time, without the suction air a to below the dew point temperature _{T d,} is operated at a dew point temperature _{T d} above.

Further, in the case of W _PR > W _PDT , when (difference income INC−difference cost C _C ) ≦ 0 (the balance is in red or 0), the intake air A is not made less than the dew point temperature T _d , and the dew point temperature T Drive at _d .
On the other hand, in the case of W _PR > W _PDT , when (difference income INC−difference cost C _C )> 0... (Balance is in black), the intake air A is cooled below the dew point temperature T _d .

Returning to FIG. 1, the gas turbine intake air conditioning device 3 is arranged between the gas turbine intake air cooling device 2 and the compressor 1 a in the suction flow path, and is sucked into the compressor 1 a. A moisture absorption means 50 capable of absorbing moisture in the air A and releasing the contained moisture to humidify the intake air A is provided.
The hygroscopic means 50 includes a hygroscopic agent 51, a low moisture content hygroscopic agent storage unit 52, a high moisture content hygroscopic agent storage unit 53, a spraying unit 54, and a spraying control unit 60.

As the hygroscopic agent 51, a liquid material such as a LiBr aqueous solution or a powder material such as a powdered silica gel can be used.
In this hygroscopic agent 51, a low moisture content hygroscopic agent 51a having a relatively small moisture content and a high moisture content hygroscopic agent 51b having a relatively large moisture content are separately stored.

  The low moisture content hygroscopic storage part 52 stores the low moisture content hygroscopic agent 51a. This low moisture content hygroscopic storage part 52 has a valve body 52a and a pipe line 52b for a funnel-shaped recovery hole (not shown) formed in the lower part of the suction flow path, and a valve body 52c and a pipe for a spray part 54. The paths 52d are connected to each other.

  The high moisture content hygroscopic storage part 53 stores the high moisture content hygroscopic agent 51 b having a relatively large moisture content in the hygroscopic agent 51. This high moisture content hygroscopic storage part 53 has a valve body 53a and a pipe line 53b for a funnel-shaped recovery hole (not shown) formed in the lower part of the suction flow path, and a valve body 53c and a pipe for a spray part 54. The paths 53d are connected to each other.

  The valve body 53a, the pipe line 53b, the valve body 52a, and the pipe line 52b described above constitute a hygroscopic agent recovery mechanism 55.

  The spraying part 54 is provided in the upper part of the suction flow path, and when the valve bodies 52c and 53c are selectively switched by the spraying control part 60, the low moisture content hygroscopic agent 51a and the high moisture content hygroscopic agent 51b are provided. Alternatively, select and spray toward the suction air A.

FIG. 6 is a block diagram showing a schematic configuration of the spray control unit 60 of the gas turbine intake humidity control apparatus 3.
The spraying control unit 60 switches and controls the spraying unit 54. Moreover, the dispersion | distribution control part 60 absorbs the water | moisture content of the suction air A, when the low moisture content hygroscopic agent 51a is sprayed, the high moisture content hygroscopic agent 51b which contained the increased moisture content is stored in the high moisture content hygroscopic agent storage part. 53. More specifically, the valve body 52a is closed, the valve body 53a is opened, and the high-moisture-containing moisture absorbent 51b accumulated in the lower part of the suction flow path is introduced into the pipe line 53b from a funnel-shaped recovery hole (not shown), The high moisture content hygroscopic agent storage unit 53 is introduced.

  On the other hand, when the high moisture content hygroscopic agent 51b is sprayed, the low moisture content hygroscopic agent 51a whose moisture content is reduced by releasing moisture into the suction air A is introduced into the low moisture content hygroscopic agent storage unit 52. More specifically, the valve body 52a is opened, the valve body 53a is closed, and the low moisture content hygroscopic agent 51a accumulated in the lower part of the suction flow path is introduced into the pipe line 52b from a funnel-shaped recovery hole (not shown), It introduces into the low moisture content hygroscopic storage part 52.

  The spray control unit 60 includes a required inlet temperature calculation unit 61, a dew point temperature calculation unit 62, a saturated water vapor pressure calculation unit 63, an atmospheric water vapor partial pressure calculation unit 64, a solution surface water vapor partial pressure calculation unit 65, And an operation mode determination unit 66.

Request inlet temperature calculation unit 61 requests the inlet temperature is the same configuration as the arithmetic unit 41e1, the required output W _PR inputted from the outside, pre stored in the storage unit 69 and the required output W _PR required inlet temperature T _{PR Is} calculated from the predetermined relationship ((1) in FIG. 6).
The required inlet temperature calculation unit (41e1, 61) is provided in each of the spraying control unit 60 and the cooling control unit 24. However, the required inlet temperature calculation unit (41e1, 61) is provided only in one side, the other is omitted, and the other is omitted. It is good also as a structure.

The dew point temperature calculation unit 62 includes an atmospheric temperature T _amb input from the atmospheric temperature measurement unit 56, an atmospheric humidity φ input from the atmospheric humidity measurement unit 57, and an NC diagram previously stored in the storage unit 69 (FIG. 6). The dew point temperature _Td of the intake air A is calculated from (2)).
Note that the same NC diagram as that stored in the storage unit 25 described above can be used.
In addition, the dew point temperature calculation unit (41a, 62) is configured to be provided in each of the spray control unit 60 and the cooling control unit 24. It is good.

Saturated vapor pressure calculating unit 63, NC diagram stored in the storage unit 69 on the basis of the ((2) in FIG. 6), the suction air corresponding to the operation on the request inlet temperature calculation unit 61 requests the inlet temperature T _PR A saturated water vapor pressure E of A is calculated.

The atmospheric water vapor partial pressure calculation unit 64 is an NC diagram (in FIG. 6) stored in advance in the storage unit 69 and the atmospheric temperature T _amb and the atmospheric humidity φ input from the atmospheric temperature measurement unit 56 and the atmospheric humidity measurement unit 57. Based on (2)), the water vapor partial pressure PS _amb of the intake air A is calculated.

The on-solution water vapor partial pressure calculating unit (humectant water vapor partial pressure calculating unit) 65 receives the hygroscopic agent temperature _{TL of} the high moisture-containing hygroscopic agent 51b input from the hygroscopic agent temperature measuring unit 68 and the hygroscopic agent concentration measuring unit 67. high and moisture concentration _{X L} of water containing desiccant 51b, prestored Duhring diagram in the storage unit 69 that is input from (((3) in FIG. 6)) (for example, "absorption chiller and heat pump", Takada fall one Author, 1989, Xing English Kasha, p.10 ") and on the basis, calculates the water vapor partial pressure PS _L on the surface of the high moisture content moisture absorbent 51b.

The operation mode determination unit 66 compares the required inlet temperature _TPR with the dew point temperature _Td, and in the operation mode in which the low moisture content hygroscopic agent 51a is sprayed when the required inlet temperature _TPR is lower than the dew point temperature _Td. Judge that there is.
Further, the operation mode determination unit 66 compares the required inlet temperature _TPR with the dew point temperature _Td, and when the required inlet temperature _TPR is higher than the dew point temperature _Td , the operation mode determination unit 66 is input from the atmospheric temperature measurement unit 56. when the atmospheric temperature _{T amb} ((6) in FIG. 6) is higher than the required inlet temperature _{T PR,} water vapor partial pressure PS _L on the calculated high water content desiccant 51b surface, the saturated vapor pressure calculating section 63 An operation mode in which the high moisture content moisture absorbent 51b is sprayed when it is smaller than the calculated saturated water vapor pressure E and higher than the calculated water vapor partial pressure PS _amb of the intake air A (E> PS _L > PS _amb ). It is determined that

Next, the operation of the GTCC power plant G1 having the above configuration will be described.
In the following description, the case where the suction air A is cooled below the dew point will be described. More specifically, the case where the intake air A is cooled below the dew point temperature _Td based on the determination by the determination means 43 will be described.

The cooling control unit 24 sets the cooling amount of the entire heat transfer tube 32 according to the enthalpy to be reduced from the suction air A in order to cool the suction air A to less than the dew point temperature _Td . Within this cooling amount range, each temperature of the cold water C1 is set so that the temperature of the cold water C1 increases in the order of the heat transfer tube stages 35A <35B <35C <35D. And the cold water C1 set to these each temperature is supplied to each heat exchanger tube 32 of the heat exchanger tube stages 35A-35D by the refrigerator 22. As shown in FIG. 3, the cold water C <b> 1 that has flowed into the heat transfer tubes 32 of the heat transfer tube stages 35 </ b> A to 35 </ b> D flows from the lower side to the upper side and returns to the refrigerator 22.

On the other hand, the air sucked from the outside flows through the suction flow path as the suction air A and flows toward the inlet of the compressor 1a. Then, after flowing into the pipe line 31 of the first heat exchanger 21, it reaches the heat transfer pipe stage 35 </ b> A and sequentially passes through the gaps between the heat transfer pipes 32.
As shown in FIG. 7, the intake air A is cooled by exchanging heat with cold water C1 when passing through the heat transfer tube stage 35A, and the moisture contained in the intake air A when the intake air A becomes less than the dew point temperature _Td. Condenses into mist M.

As shown in FIG. 7, the suction air A (arrow Y1) that has passed through the gap in the heat transfer tube stage 35A collides with the heat transfer tube 32 of the heat transfer tube stage 35B, and then flows downstream along the outer peripheral surface of the heat transfer tube 32. It flows toward the side. Further, the suction air A (the other end (arrow Y2)) passing through the wall surface spacing S between the one end and the other end in the flow path width direction of the heat transfer tube 32 is at one end and the other end of the heat transfer tube stage 35B in the flow path width direction. After colliding with the heat transfer tube 32, it flows along the outer peripheral surface of the heat transfer tube 32 toward the downstream side of the airflow. At this time, the mist M contained in the intake air A adheres to the heat transfer tube 32 of the heat transfer tube stage 35B in a condensed manner and is captured.
The intake air A is also cooled when flowing along the outer peripheral surface of the heat transfer tube stage 35B, and the moisture contained in the intake air A is newly condensed by this cooling to become mist M. At this time, since the cooling effect of the heat transfer tube stage 35B is smaller than that of the heat transfer tube stage 35A, the cooling amount and the amount of mist M generated are smaller than those of the heat transfer tube stage 35A.

Similarly, the suction air A that has passed through the heat transfer tube stage 35B flows along the outer peripheral surface of the heat transfer tube 32 after colliding with the heat transfer tube 32 in the heat transfer tube stage 35C, and adheres mist M and is cooled. At this time, since the cooling effect in the heat transfer tube stage 35C is smaller than that in the heat transfer tube stage 35A, the cooling amount and the amount of mist M generated are smaller than those in the heat transfer tube stage 35B.
The suction air A that has passed through the heat transfer tube stage 35C slightly adheres mist M when it passes through the heat transfer tube stage 35D.

In this way, when the intake air A passes through each of the heat transfer tube stages 35A to 35D, it sequentially exchanges heat with the cold water C1 to reduce the enthalpy of the intake air A. Then, the temperature of the intake air A that has passed through the heat transfer tube stage 35D becomes a temperature lower than a predetermined dew point temperature _Td . The intake air A that has passed through the heat transfer tube stage 35D flows toward the compressor 1a.

  The mist M adhering to each heat transfer tube 32 flows down in the direction of gravity along the heat transfer tube 32 and reaches the lower portion of the pipe line 31. And it flows toward the drain 33 by the inclination of the lower part of the pipe line 31 and is stored in the drain 33. In addition, the water stored in the drain 33 is discharged | emitted outside via a suction pump.

As described above, according to the present embodiment, the first heat exchanger 21 capable of cooling the intake air A includes the plurality of heat transfer tubes 32, and the plurality of heat transfer tubes 32 are located on the upstream side of the airflow. Since the heat pipe 32 is disposed so as to have a cooling effect greater than that of the heat transfer pipe 32 located on the downstream side of the airflow, when the suction air A is cooled to a temperature lower than the dew point temperature _Td , the heat transfer located on the upstream side of the airflow is arranged. Generation of mist M is relatively active in the heat pipe 32, and generation of mist M is relatively suppressed in the heat transfer pipe 32 positioned on the downstream side of the airflow. Then, the mist M generated by being cooled by the heat transfer tube 32 located on the upstream side of the airflow adheres to the heat transfer tube 32 on the downstream side of the airflow and falls down in the gravity direction over the outer peripheral surface.
Thereby, even if mist M is generated along with the cooling of the suction air A, the ratio of capturing the mist M by the heat transfer tube 32 itself that cools the suction air A as compared with the case where the cooling effect is made constant in the airflow direction. Therefore, it is possible to prevent the mist M from flowing downstream of the airflow.
Therefore, it is possible to prevent the gas turbine from sucking the mist M without adding another device on the flow path of the cooled intake air A.

  Further, since the temperature of the cold water C1 is lower in the heat transfer tube 32 on the upstream side of the airflow than the heat transfer tube 32 on the downstream side of the airflow, the cooling effect of the heat transfer tube 32 on the upstream side of the airflow can be relatively increased. In addition, it is possible to relatively easily adjust the cooling effect.

  Further, since the plurality of heat transfer tubes 32 are arranged in a zigzag shape, the suction air A that has passed through the gap in the flow path width direction of the heat transfer tube stage 35 on the upstream side of the airflow is transferred to the heat transfer tube stage 35 on the downstream side of the airflow. The heat transfer tube 32 can be sufficiently brought into contact. Thereby, the cooling effect and the mist capturing effect can be enhanced.

  In addition, since the center-to-center distance L between two heat transfer tubes 32 adjacent to each other in the flow path width direction is less than twice the outer diameter of the heat transfer tube 32, the suction that has passed through the heat transfer tube stage 35 on the upstream side of the airflow The air A can be reliably brought into contact with each heat transfer tube 32 in the heat transfer tube stage 35 on the downstream side of the airflow.

  Further, the suction air A passing through the wall surface spacing S between the two heat transfer tubes 32 positioned at both ends of the heat transfer tube stage 35 on the upstream side of the airflow and the flow passage wall surface 31a is transferred to the downstream side of the airflow. The heat transfer tube 32 can be brought into contact with the heat transfer tubes 32 located at both ends in the flow path width direction. Thereby, the suction air A which passes without contacting the heat exchanger tube 32 along the flow-path wall surface 31a vicinity can be prevented, and a cooling effect and a mist capture | acquisition effect can be improved.

Further, according to the GTCC power plant G1, the compressor output temperature T _IN is set on the condition that the required output W _PR is compared with the output at the dew point temperature T _d and that the balance due to power generation satisfies a predetermined standard. because it below the dew point temperature T _d, it is possible to prevent the compressor inlet temperature T _iN is the balance and independent below the dew point temperature T _d due to the power generation. As a result, the suction air A is limited to a temperature _lower than the dew point temperature _Td based on the balance of power generation, so that the suction air A can be efficiently cooled. In addition, when the balance due to power generation satisfies a predetermined standard, the intake air A is cooled to a temperature _lower than the dew point temperature _Td. It becomes possible.

  Furthermore, since the moisture absorption means 50 capable of absorbing moisture of the intake air A and releasing the contained moisture to humidify the intake air A is provided, when the mist M is generated, the mist M can be absorbed. When the humidity is low, the intake air A can be cooled by releasing the contained water. Thereby, the mist M which goes to the compressor 1a inlet of a gas turbine can be removed, and the damage of the components of the gas turbine 1 for electric power generation can be suppressed. Further, when the humidity of the intake air A is low, the intake air A can be further cooled, and the output of the power generation gas turbine 1 can be improved.

  In the above-described embodiment, the cold water C1 that has flowed into the heat transfer tubes 32 of the heat transfer tube stages 35A to 35D is configured to flow from below to above, but may flow from above to below. In the two heat transfer tubes 32 adjacent in the airflow direction, the flow direction of the cold water C1 may be opposite.

FIG. 8 is a schematic cross-sectional view of a modification 21A of the first heat exchanger 21 and corresponds to FIG.
As shown in FIG. 8, in the modification 21A, the upper end 32a of one heat transfer tube 32 in the heat transfer tube stage 35A and the upper end 32a of one heat transfer tube 32 in the heat transfer tube stage 35B are connected by a curved tube 36a. The first connected heat transfer tube unit 37A is configured. Similarly, the upper end 32a of one heat transfer tube 32 in the heat transfer tube stage 35C and the upper end 32a of one heat transfer tube 32 in the heat transfer tube stage 35D are connected by a curved tube 36a to form a first connected heat transfer tube unit 37B. doing.
Further, the lower end 32b of one heat transfer tube 32 in the heat transfer tube stage 35B and the lower end 32b of one heat transfer tube 32 in the heat transfer tube stage 35C are connected by a curved tube 36b to form a second connected heat transfer tube unit 38. So that they are connected.

  With such a configuration, when the cold water C1 is sent from the refrigerator 22 and the cold water C1 flows from the lower end 32b of the heat transfer tube 32 of the heat transfer tube stage 35A, the heat transfer tube stages 35B to 35D are passed through the curved tubes 36a and 36b. The cold water C1 flows through the heat transfer tubes 32 in order while changing the flow direction in the opposite direction.

At this time, in the heat transfer tube 32 of the heat transfer tube stage 35A, the chilled water C1 flows upward while exchanging heat with the suction air A in the lower part of the heat transfer tube 32 in the gravity direction (while raising the temperature). Compared with the lower cooling effect, the lower side of the intake air A is cooler than the upper side. However, since the heat transfer tube 32 of the heat transfer tube stage 35B constitutes the first connected heat transfer tube unit 37A in which the flow direction of the cold water C1 is opposite to that of the heat transfer tube 32 of the heat transfer tube stage 35A, the heat transfer tube stage 35A The heat transfer tube stage 35B cools the upper part more strongly than the lower part with respect to the suction air A whose lower part is lower than the upper part, and equalizes the temperature difference in the vertical direction.
Even when the intake air A passes through the second connected heat transfer tube unit 38 and the first connected heat transfer tube unit 37B, the temperature difference in the vertical direction is made uniform. Moreover, since the cold water C1 flows in the order of the heat transfer tube stages 35A to 35D while exchanging heat with the intake air A (while raising the temperature), the temperature rises in the order of the heat transfer tube stages 35A <35B <35C <35D.

  According to this configuration, since the two heat transfer tubes 32 adjacent to each other in the airflow direction are set so that the flow direction of the cold water C1 is opposite, the temperature difference between the intake air A and the gravity direction is reduced. be able to. That is, in each heat transfer tube 32, the cold water C1 heated and heated with the suction air A on the upstream side in the cold water C1 flow flows toward the downstream side in the cold water C1 flow. The cooling effect on the upstream side is higher than that on the downstream side. However, since the two heat transfer tubes 32 adjacent to each other in the airflow direction are set so that the flow direction of the cold water C1 is opposite, the cooling effect is exerted on the two heat transfer tubes 32 in the gravity direction. Equalized. For this reason, when the suction air A continuously flows along these two heat transfer tubes 32, the suction air A is cooled uniformly in the gravity direction of the suction air A. A temperature difference can be reduced. Further, the temperature of the cold water C1 can be easily increased in the order of the heat transfer tube stages 35A <35B <35C <35D.

Note that the operation procedure shown in the above-described embodiment, various shapes and combinations of the constituent members, and the like are examples, and various modifications can be made based on design requirements and the like without departing from the gist of the present invention.
For example, in the embodiment described above, the heat transfer tubes 32 are arranged in a staggered manner, but may be arranged in a matrix, for example.

  Moreover, as shown in FIG. 8, as a method of changing the temperature of the cold water C1 for each heat transfer tube 32 as in the above-described embodiment, a method of flowing the heat transfer tube stages 35A to 35D in series as shown in FIG. Even if the method of suppressing the temperature rise by enlarging the heat capacity and increasing the flow rate of the cold water C1 and increasing the flow area α of the upstream heat transfer tube 32 as in the modified example 21B shown in FIG. Good.

Further, in the embodiment described above, suction air A in Dennetsukandan 35A was cooled to be less than the dew point temperature T _d, not necessarily be not suction air A in the heat transfer tube stage 35A is less than the dew point temperature T _d Good. For example, the suction air A is cooled to less than the dew point temperature _{Td in the} process of flowing through the heat transfer tube stages 35A and 35B on the upstream side of the airflow, and the mist M of the suction air A in the process of flowing through the heat transfer tube stages 35C and 35D on the downstream side of the airflow It is good also as a structure to capture. That is, if the cooling effect of the heat transfer tube stages 35A and 35B on the upstream side of the airflow is greater than the cooling effect of the heat transfer tube stages 35C and 35D on the downstream side of the airflow, the effect of the present invention can be obtained.
Further, in the embodiment described above, the heat transfer tube stage 35 is configured in four stages, but any number of stages may be used as long as it is a plurality of stages.

Moreover, in embodiment mentioned above, although the temperature of the cold water C1 which flows through the heat exchanger tube stage 35A-35D was varied, it comprised so that the cooling effect of the airflow upstream side might become larger than the airflow downstream side, The flow rate of the cold water C1 flowing through 35D may be varied so that the cooling effect on the upstream side of the airflow is greater than that on the downstream side of the airflow. For example, by increasing the size of the flow area of each heat transfer tube 32 of Dennetsukandan 35A~35D in this order, it is possible to increase the cooling effect in the order of Dennetsukandan 35 A to 35 D. For example, as in Modification 21C shown in FIG. 12, the cooling water C1 is cooled by decreasing the size of the orifice diameter β of each of the heat transfer tubes 32 of the heat transfer tube stages 35A to 35D in the order of 35A>35B>35C> 35D. The effect can be reduced in the order of the heat transfer tube stages 35A>35B>35C> 35D. The orifice 32f may be provided on either the inlet side or the outlet side of the heat transfer tube 32.

  Moreover, in embodiment mentioned above, although it did not mention in particular about the surface roughness of the inner surface and outer surface of the heat exchanger tube 32, the surface roughness of either one of the inner surface and outer surface of the heat exchanger tube 32 of an airflow upstream side was mentioned. May be made coarser than the heat transfer tube 32 on the downstream side of the airflow. According to this configuration, the flow of the cold water C1 can be disturbed by increasing the surface roughness of the inner surface, and the flow of the intake air A can be disturbed by increasing the surface roughness of the outer surface. The cooling effect of the heat pipe 32 can be relatively increased.

In the above-described embodiment, the first connection heat transfer tube units 37A and 37B and the second connection heat transfer tube unit 38 are connected. However, the first connection heat transfer tube units 37A and 37B and the second connection heat transfer tube unit 38 are connected. 38 is not necessarily continuous.
Moreover, you may provide 1st connection heat exchanger tube unit 37A, 37B and the 2nd connection heat exchanger tube unit 38 separately, without continuing. Further, only one of the first connected heat transfer tube unit 37 and the second connected heat transfer tube unit 38 may be provided.

In the embodiment described above, nothing is provided on the outer peripheral surface of the heat transfer tube 32. However, as shown in FIG. 9, a helical fin member 32c extending radially outward from the outer peripheral surface. May be provided.
According to this configuration, since the contact area on the heat transfer tube 32 side with respect to the intake air A is increased, the cooling effect is improved by increasing the heat transfer area, and when the heat transfer tube 32 is provided on the downstream side of the airflow, , Mist M adhesion amount increases. Thereby, while cooling of the suction air A is accelerated | stimulated, the mist capture | acquisition effect can be improved.
Moreover, since the fin member 32c is spiral, the captured mist M continues to fall down along the fin member 32c. Thereby, the captured mist M can be quickly guided downward to prevent the suction air A from being blown to the downstream side of the airflow.

  Further, as shown in FIG. 10, the fin member 32d is configured in a funnel shape, the vertical through hole 32e is provided on the inner peripheral side of the fin member 32d, and a plurality of fin members 32d are provided. As the cooling effect is increased, the trapped mist M is collected on the outer peripheral surface side of the heat transfer tube 32, and continuously falls over the outer peripheral surface of the heat transfer tube 32. Thereby, the captured mist M can be quickly guided downward to prevent the suction air A from being blown to the downstream side of the airflow.

  Moreover, as shown in FIG. 13, the gas turbine intake humidity control device 3 may be omitted, and the plant configuration may be further simplified.

Further, all the data of the NC diagram may be stored, or only the relationship between the temperature and the saturation pressure may be stored, and other data such as the dew point temperature _Td may be obtained by calculation each time. .

DESCRIPTION OF SYMBOLS 1 ... Gas turbine 1a for electric power generation ... Compressor 1b ... Combustor 1c ... Turbine 2 ... Intake cooling device 21 for gas turbines ... First heat exchanger (heat exchanger)
31 ... Pipe line 31a ... Channel wall surface 32 ... Heat transfer tube 32a ... Upper end 32b ... Lower end 32c, 32d ... Fin member 32e ... Vertical through hole 33 ... Drain 35 (35A-35D) ... Heat transfer tube stages 36a, 36b ... Curved tube 37A , 37B ... first connected heat transfer tube unit 38 ... second connected heat transfer tube unit 64 ... atmospheric water vapor component calculating unit 65 ... hygroscopic agent water vapor partial pressure calculating unit 66 ... operation mode determining unit 67 ... hygroscopic agent concentration measuring unit 68 ... hygroscopic agent Temperature measurement unit 69 ... storage unit A ... suction air C1 ... cold water (cooling medium)
L ... Center distance M ... Mist S ... Wall spacing G1 ... GTCC power plant

Claims (17)

An intake air cooling device for a gas turbine used in a gas turbine including a compressor, a combustor, and a turbine,
A heat exchanger capable of cooling the intake air sucked into the compressor from the outside,
The heat exchanger extends across an upper side and a lower side in the direction of gravity, and a plurality of heat transfer tubes intersecting the air flow direction of the suction air;
A cooling medium flowing inside the plurality of heat transfer tubes,
The heat transfer tubes on the upstream side of the airflow are heat transfer tubes on the downstream side of the airflow so that the heat transfer tubes located on the upstream side of the airflow have a higher cooling capacity than the heat transfer tubes located on the downstream side of the airflow. An intake air cooling device for a gas turbine characterized in that the orifice diameter is larger than that of the gas turbine. An intake air cooling device for a gas turbine used in a gas turbine including a compressor, a combustor, and a turbine,
A heat exchanger capable of cooling the intake air sucked into the compressor from the outside,
The heat exchanger extends across an upper side and a lower side in the direction of gravity, and a plurality of heat transfer tubes intersecting the air flow direction of the suction air;
A cooling medium flowing inside the plurality of heat transfer tubes,
The heat transfer tubes on the upstream side of the airflow are heat transfer tubes on the downstream side of the airflow so that the heat transfer tubes located on the upstream side of the airflow have a higher cooling capacity than the heat transfer tubes located on the downstream side of the airflow. An intake air cooling device for a gas turbine, wherein at least one of the inner surface and the outer surface is rougher than the inner surface. An intake air cooling device for a gas turbine used in a gas turbine including a compressor, a combustor, and a turbine,
A heat exchanger capable of cooling the intake air sucked into the compressor from the outside,
The heat exchanger extends across an upper side and a lower side in the direction of gravity, and a plurality of heat transfer tubes intersecting the air flow direction of the suction air;
A cooling medium flowing inside the plurality of heat transfer tubes,
The heat transfer tubes on the upstream side of the airflow are heat transfer tubes on the downstream side of the airflow so that the heat transfer tubes located on the upstream side of the airflow have a higher cooling capacity than the heat transfer tubes located on the downstream side of the airflow. An intake air cooling device for a gas turbine characterized in that the flow passage area is larger than that of the gas turbine. The gas turbine according to any one of claims 1 to 3 , wherein the heat transfer tube on the upstream side of the airflow has a lower temperature of the cooling medium than the heat transfer tube on the downstream side of the airflow. Intake cooling system. Two heat transfer tubes that are adjacent to each other in the air flow direction, the gas according to claims 1 in any one of the 4, characterized in that the flow direction of the cooling medium is set to be opposite Turbine intake air cooling system. The intake-air cooling apparatus for a gas turbine according to claim 5 , further comprising a first connected heat transfer tube unit in which upper ends of two heat transfer tubes adjacent to each other in the airflow direction are connected. The intake cooling device for a gas turbine according to claim 5 or 6 , further comprising a second connected heat transfer tube unit in which respective lower ends of two heat transfer tubes adjacent to each other in the air flow direction are connected. . It said plurality of at least a portion of the heat transfer tubes, gas turbine as claimed in any one of claims 1 7, characterized in that it comprises a fin member formed extending from the outer peripheral surface radially outward Intake cooling system. The intake cooling device for a gas turbine according to claim 8 , wherein the fin member has a spiral shape. The intake cooling device for a gas turbine according to claim 8 , wherein the fin member has a funnel shape. The plurality of heat transfer tubes are arranged in a staggered manner when viewed from above,
A part of the plurality of heat transfer tubes constitutes a heat transfer tube stage that is arranged in a plurality at intervals in the flow path width direction intersecting the air flow direction and the gravity direction,
A plurality of the heat transfer tube stages are stacked with an interval in the airflow direction, and the heat transfer tube stages adjacent to each other in the airflow direction are arranged with a gap in the flow path width direction. intake air cooling apparatus for a gas turbine according to any one of claims 1 to 1 0 to. Center distance a gas turbine intake air cooling according to claim 1 1, characterized in that it is more than 2 times the outer diameter of the heat transfer tubes of two heat exchanger tubes adjacent to one another in the flow path width direction apparatus. In one of the two heat transfer tube stages adjacent to each other in the airflow direction, a wall surface interval from the outer peripheral surface of the heat transfer tube located at one end in the flow channel width direction to the flow channel wall surface of the suction air is determined by the heat transfer tube. It is below the outer diameter,
On the other hand, at least a part of the outer peripheral surface of the heat transfer tubes positioned in the flow path width direction one end, viewed from the air flow direction, to claim 1 2, characterized in that located in the inner wall surface interval in the one The intake-air cooling apparatus for gas turbines as described. In one of the two heat transfer tube stages adjacent to each other in the airflow direction, a wall surface distance from the outer peripheral surface of the heat transfer tube located at the other end in the flow channel width direction to the flow channel wall surface of the suction air is the heat transfer tube. Less than the outer diameter of
On the other hand, at least a part of the outer peripheral surface of the heat transfer tubes positioned in the flow path width direction other end, viewed from the air flow direction, according to claim 1 3, characterized in that located in the inner wall surface interval in the other An intake air cooling device for a gas turbine as described in 1. Gas turbines equipped with intake air cooling apparatus for a gas turbine as claimed in any one of claims 1 1 4. An intake cooling apparatus for a gas turbine according to any one of claims 1 to 1 4,
A gas turbine comprising a compressor, a combustor, and a turbine;
A gas turbine combined cycle power plant comprising exhaust heat utilization means for utilizing exhaust heat from the gas turbine. Power augmentation method characterized by additionally provided an intake air cooling apparatus for a gas turbine of any one of the existing gas turbine or gas turbine combined cycle power plant of claims 1 1 4.

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