JPWO2008139528A1 - Cooling cycle system, natural gas liquefaction facility, cooling cycle system operating method and remodeling method - Google Patents

Abstract

A refrigerant compressor 1 that compresses the refrigerant, a condenser 10 that cools and condenses the refrigerant compressed by the refrigerant compressor 1, a receiver 11 that receives the refrigerant condensed in the condenser 10, and the liquid receiver An expansion mechanism 18 that expands the refrigerant from the container 11, an evaporation mechanism 19 that causes heat exchange with the refrigerant expanded by the expansion mechanism 18 to cool the object to be cooled and evaporate the refrigerant supplied to the refrigerant compressor 1, and an auxiliary Auxiliary for cooling the refrigerant in the liquid receiver 11 before starting the refrigerant compressor 1 by having a pipe 47 for circulating the refrigerant and passing through the pipe 11 and exchanging heat with the auxiliary refrigerant flowing through the pipe 47 A cooling cycle system 61 including a cooling mechanism 62 is configured. Thereby, the required power at the time of starting of a refrigerant compressor can be reduced, and a refrigerant compressor can be started stably even if it uses a drive source with small generated torque at the time of starting.

Description

  The present invention relates to a cooling cycle system, a natural gas liquefaction facility, a method for operating the cooling cycle system, and a modification method.

  In order to convert a gaseous natural gas into a liquefied natural gas suitable for transportation, it is necessary to cool the natural gas to a low temperature of about −150 ° C. in a pressurized state and then expand it to the vicinity of the external pressure. This cooling is realized by a combination of a plurality of cooling cycles such as propane and mixed refrigerant. When starting up these cooling cycles, the temperature and pressure of the refrigerant in the process differs from that during rated operation, so the torque required to start the refrigerant compressor (hereinafter referred to as starting torque) is greater than that during rated operation. Therefore, it has been pointed out that the refrigerant compressor may not be started depending on the torque characteristics of the drive source. To solve this problem, (a) a method of installing a throttle mechanism in the suction line of the refrigerant compressor to reduce the suction flow rate at startup, (b) a method of installing a drive source having a large capacity in advance, and (c) refrigerant compression A method is generally known in which a discharge port is provided on the discharge side of the machine to discharge the refrigerant to the outside. According to Patent Document 1, there is shown a method of starting a compressor by operating an opening degree of a throttle mechanism installed in a suction line of the compressor according to the rotational speed of the compressor.

JP 2002-322996 A

  The temperature and pressure of the refrigerant in the suction section of the refrigerant compressor are equal to the temperature of the refrigerant evaporated by the upstream evaporator and the saturation pressure corresponding to that temperature. Although the method of reducing the suction flow rate of the compressor at the time of start-up including the technique described in Patent Document 1 is effective in reducing the volume flow rate of the intake air, the refrigerant temperature and the saturation pressure at the time of start-up do not restrict the intake air flow rate. Since it is not different from the case, the effect of reducing the mass flow rate of the intake air and reducing the torque required at startup is limited.

  In addition, the method of installing a drive source having a large capacity in advance may reduce the efficiency of the drive source during rated operation, which increases the equipment cost.

  Although the method of discharging the refrigerant discharged from the refrigerant compressor to the outside at the start works well for avoiding the surge of the refrigerant compressor and reducing the starting torque, it takes time to increase the discharge pressure of the refrigerant compressor and the cooling cycle. There is a tendency that the time required for completing the start-up of the system is extended. It is also a disadvantage to waste the refrigerant.

  Accordingly, the present invention provides a cooling cycle system and a natural gas liquefaction facility capable of reducing the required power at the time of starting the refrigerant compressor and stably starting the refrigerant compressor even when using a drive source with a small torque generated at the time of starting. An object is to provide an operation method and a modification method of a cooling cycle system.

  To achieve the above object, a refrigerant compressor for compressing refrigerant, a liquid receiver for receiving the refrigerant condensed after being compressed by the refrigerant compressor, and the refrigerant compression through the auxiliary refrigerant in the liquid receiver An auxiliary cooling mechanism is provided for cooling the refrigerant in the liquid receiver before the machine is started.

  ADVANTAGE OF THE INVENTION According to this invention, the required motive power at the time of starting of a refrigerant compressor can be reduced, and a refrigerant compressor can be started stably even if it uses a drive source with a small generated torque at the time of starting.

1 is an overall configuration diagram of a natural gas liquefaction facility using a cooling cycle system according to a first embodiment of the present invention. The figure showing the rotation speed change at the time of starting of a refrigerant compressor, and the opening change of a suction adjustment mechanism The figure showing the change of the pressure at the time of starting of the refrigerant compressor concerning a 1st embodiment of the present invention, and a flow rate The figure showing the pressure and flow rate change at the time of starting of the refrigerant compressor of a comparative example The figure showing the torque change at the time of starting of the refrigerant compressor concerning a 1st embodiment of the present invention. The figure showing the torque change at the time of starting of the refrigerant compressor of a comparative example Overall configuration diagram of natural gas liquefaction equipment using a cooling cycle system according to a second embodiment of the present invention

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Refrigerant compressor 10 Condenser 11 Receiver 18 Expansion mechanism 19 Evaporation mechanism 21 Heat exchanger 47 Piping 60 2nd cooling cycle system 61 1st cooling cycle system 62 Auxiliary cooling mechanism 62A Auxiliary cooling mechanism 68 Evaporator 80 Evaporator 81 Compressor 82 Condenser 83 Expansion valve 86 Combustor 90 Intake cooling heat exchanger 91 Air compressor 92 Generator 93 Turbine 94 Waste heat recovery boiler 96 Absorption-type cooler 97 Temperature detection means 98 Control device 99 Control valve 101 Bypass piping

Embodiments of the present invention will be described below with reference to the drawings.
The cooling cycle system according to the present invention cools the medium to be cooled by exchanging heat with the refrigerant. The cooled medium after cooling is supplied to a heat utilization facility that uses this medium as a cold heat source. An example of a heat utilization facility is a facility that cools and liquefies natural gas in a gaseous state, for example. When natural gas is liquefied, its volume is reduced and transportation efficiency is improved. Of course, any facility that can use the cooled medium as a heat source is not limited to the facility that liquefies natural gas, and can be used as the cooling medium of the cooling cycle system of the present invention.

<First Embodiment>
(Constitution)
FIG. 1 is an overall configuration diagram of a natural gas liquefaction facility using a cooling cycle system according to a first embodiment of the present invention.

  The main components of the illustrated natural gas liquefaction facility include a main heat exchanger 21 that cools and liquefies natural gas (gas state) introduced from the natural gas introduction pipe 20, and a refrigerant (hereinafter referred to as “first”). A first cooling cycle system 61 that cools a refrigerant that cools natural gas (hereinafter referred to as a “second refrigerant”) and a second refrigerant that is cooled by the first cooling cycle system 61. The 2nd cooling cycle system 60 supplied to the heat exchanger 21 is mentioned. The first cooling cycle system 61 is characterized in that it includes an auxiliary cooling mechanism 62 that cools the first refrigerant with a refrigerant that is a working medium (hereinafter referred to as “auxiliary refrigerant”).

  The first cooling cycle system 61 uses propane as a working fluid, and includes a refrigerant compressor 1, a condenser 10, a liquid receiver 11, an expansion mechanism 18, and an evaporation mechanism 19.

  The refrigerant compressor 1 compresses the first refrigerant from the evaporation mechanism 19, and includes a plurality of stages (three stages in this example) of a low-pressure compressor 2, an intermediate-pressure compressor 3, and a high-pressure compressor 4. ing. The first refrigerant from the evaporation mechanism 19 is guided to the low-pressure compressor 2, the intermediate-pressure compressor 3, and the high-pressure compressor 4 through pipes 56, 55, and 54, respectively. The low-pressure compressor 2, the intermediate-pressure compressor 3, and the high-pressure compressor 4 constituting the refrigerant compressor 1 are coaxially connected to an electric motor 5 that is a drive source, and are rotationally driven by the rotational power of the electric motor 5. The electric motor 5 is driven by electric power supplied from a gas turbine power generator (not shown).

  The condenser 10 is connected to the outlet of the refrigerant compressor 1 via a pipe 50, and cools the first refrigerant compressed by the refrigerant compressor 1 by exchanging heat with a cold heat source such as outside air or seawater. To condense.

  The liquid receiver 11 is connected to the condenser 10 via the pipe 49 and receives the first refrigerant condensed in the condenser 10.

  The expansion mechanism 18 expands the liquid first refrigerant supplied from the liquid receiver 11. In the present embodiment, the expansion mechanism 18 includes a plurality of expansion valves including a high pressure expansion valve 12, an intermediate pressure expansion valve 13, and a low pressure expansion valve 14. By configuring the expansion mechanism 18, the first refrigerant is expanded stepwise to reduce the temperature.

  The evaporating mechanism 19 evaporates the first refrigerant of the first cooling cycle system 61 by taking heat from the second refrigerant passing through the pipe 41 and using the taken heat. In the present embodiment, the evaporation mechanism 19 is configured by a plurality of evaporators of the high-pressure evaporator 15, the intermediate-pressure evaporator 16, and the low-pressure evaporator 17, so that the first refrigerant is expanded and reduced in temperature by the expansion mechanism 18. The second refrigerant flowing through the piping 41 is sequentially cooled and simultaneously the first refrigerant expanded by the expansion mechanism 18 is evaporated.

  The high-pressure evaporator 15 is connected to the liquid receiver 11 via a pipe 51, and the high-pressure expansion valve 12 is provided in the middle of the pipe 51. The intermediate pressure evaporator 16 is connected to the high pressure evaporator 15 via a pipe 52, and the intermediate pressure expansion valve 13 is provided in the middle of the pipe 52. The low pressure evaporator 17 is connected to the intermediate pressure evaporator 16 via a pipe 53, and the low pressure expansion valve 14 is provided in the middle of the pipe 53. The evaporator 15-17 is connected to the high-pressure compressor 4, the intermediate-pressure compressor 3, and the low-pressure compressor 2 of the refrigerant compressor 1 through pipes 54, 55, and 56, respectively.

  A high-pressure suction adjustment mechanism 71, a medium-pressure suction adjustment mechanism 72, and a low-pressure suction adjustment mechanism 73 are provided in the middle of the pipes 54-56 that connect the evaporator 15-17 to the compressors 4, 3, and 2, respectively. The intake air flow rate of the refrigerant compressor 1 can be adjusted according to the operating state. These suction adjusting mechanisms 71-73 can be constituted by valves as shown in FIG. 1, or inlet guide vanes (IGV: inlet guide vanes) can be used for the compressor 2-4.

  The pipe 41 that circulates the second refrigerant of the second cooling cycle system 60 is connected to the gas-liquid separator 27, and the second refrigerant cooled to a predetermined temperature (for example, −35 ° C.) by the first cooling cycle system 61 is Gas-liquid separation is performed by the gas-liquid separator 27, and the liquid phase component is circulated from the gas-liquid separator 27 to the inside of the main heat exchanger 21 via the pipe 43 and the gas phase component via the pipe 44.

  The pipes 43, 44 extending from the gas-liquid separator 27 are taken out of the main heat exchanger 21 after passing through the inside of the main heat exchanger 21, and the nozzles 35, installed inside the main heat exchanger 21. 36 is connected again. Expansion valves 33 and 34 are respectively provided in portions of the pipes 43 and 44 once led to the outside of the main heat exchanger 21, and the second refrigerant flowing through the pipes 43 and 44 is the expansion valves 33 and 34, respectively. The second refrigerant, which is adiabatically expanded to lower the temperature and thus lowered in temperature, is sprayed from the nozzles 35 and 36 into the main heat exchanger 21.

  The main heat exchanger 21 includes heat transfer paths 28, 29, 30, 31, and 32 therein. The transmission path 28 is provided in the middle of the pipe 43, and the liquid phase component of the second refrigerant in the second cooling cycle system 60 separated from the gas phase component by the gas-liquid separator 27 is further fed into the main heat exchanger 21. Heat exchange is performed with a low-temperature second refrigerant. The transmission paths 29 and 30 are provided in the middle of the pipe 44 to exchange heat between the gas phase component of the second refrigerant from the gas-liquid separator 27 and the second refrigerant in the main heat exchanger 21. The transmission paths 31 and 32 are provided in the middle of the natural gas introduction pipe 20 to exchange heat between the natural gas flowing through the pipe 20 and the second refrigerant in the main heat exchanger 21.

  Although not particularly illustrated, natural gas that has been subjected to pretreatment steps necessary for the liquefaction step, such as an acid gas removal step and a water removal step, is led to the natural gas introduction pipe 20 on the upstream side. Further, the natural gas introduction pipe 20 extends to the outside of the main heat exchanger 21 after passing through the main heat exchanger 21. An expansion valve 37 is provided on the downstream side of the main heat exchanger 21 in the natural gas introduction pipe 20.

  The second cooling cycle system 60 uses, for example, a mixed refrigerant composed of methane, ethane, and propane as a working fluid (that is, a second refrigerant). The low-pressure compressor 23, the high-pressure compressor 24, the intermediate cooler 25, and the post-cooler 26 are used. The electric motor 85 is provided. The low-pressure compressor 23 is connected to the main heat exchanger 21 via a pipe 40, and the second refrigerant first stored in the main heat exchanger 21 is connected to the low-pressure compressor 23 via the pipe 40. 23 to the entrance. The low-pressure compressor 23 and the high-pressure compressor 24 are connected via a pipe 48, and the second refrigerant discharged from the low-pressure compressor 23 is guided to the inlet of the high-pressure compressor 24 via the pipe 48. A pipe 41 that passes through the first cooling cycle system 61 is connected to the outlet of the high-pressure compressor 24, and the second refrigerant that has passed through the second cooling cycle system 60 flows through the pipe 41 and is guided to the first cooling cycle system 61. It is burned. The intermediate cooler 25 is provided in the pipe 48, and the post-cooler 26 is provided in the pipe 41. Further, the low-pressure compressor 23 and the high-pressure compressor 24 are coaxially connected to an electric motor 85 that is a driving device thereof. The electric motor 85 is driven by electric power supplied from a gas turbine power generator (not shown).

  The auxiliary cooling mechanism 62 cools the first refrigerant stored in the liquid receiver 11 by exchanging heat with the auxiliary refrigerant flowing in the pipe 47, and cools the first refrigerant of the first cooling cycle system 61. In the present embodiment, propane is also used as the refrigerant in the auxiliary cooling mechanism 62, but the type of auxiliary refrigerant used in the auxiliary cooling mechanism 62 is not limited. The auxiliary cooling mechanism 62 includes an electric motor 84, a compressor 81 driven by the electric motor 84, a condenser 82 that radiates and condenses the auxiliary refrigerant compressed by the compressor 81 to the atmosphere or seawater, and an auxiliary that is condensed by the condenser 82. An expansion valve 83 that expands the refrigerant to generate a low temperature, and an evaporator 80 that cools the first refrigerant in the liquid receiver 11 by exchanging heat with the low-temperature auxiliary refrigerant obtained by expansion by the expansion valve 83. Yes. The pipe 47 is connected to the suction port of the compressor 81 from the discharge port of the compressor 81 through the evaporator 80 in the liquid receiver 11. The condenser 82 and the expansion valve 83 are provided in this order from the upstream side at a portion of the pipe 47 connected from the compressor 81 to the evaporator 80.

  Although the auxiliary refrigerant of the auxiliary cooling mechanism 62 does not necessarily need to be propane, propane is a natural refrigerant and has little influence on global warming, can be shared with the first cooling cycle system 61, and is easily available. There are advantages such as being.

(Operation during steady operation)
First, the operation at the time of steady operation of the natural gas liquefaction facility shown in FIG. 1 will be described. Note that the temperature and pressure at various places in the plant as described below are examples of state quantities assumed during operation, and do not limit the specifications of the plant.

  In the first cooling cycle system 61, the first refrigerant of liquid propane of about 40 ° C. and about 1.5 MPa stored in the receiver 11 is decompressed to about 0.63 MPa by the high-pressure expansion valve 12 while flowing through the pipe 51. By adiabatic expansion, a gas-liquid mixed state of about 9 ° C. corresponding to the saturation temperature of propane of 0.63 MPa is obtained.

  In the high-pressure evaporator 15, the liquid phase portion of the first refrigerant at about 9 ° C. evaporates and the latent heat of evaporation is taken away, so that the second refrigerant of the second cooling cycle system 60 at about 40 ° C. supplied from the pipe 41 is removed. Cooling. Thereafter, the gas-phase first refrigerant is supplied to the high-pressure compressor 4 via the pipe 54 and compressed to about 1.5 MPa. On the other hand, the first refrigerant in the liquid phase is supplied to the intermediate pressure expansion valve 13 via the pipe 52 and is adiabatically expanded to about 0.25 MPa to be in a gas-liquid mixed state at a saturation temperature of about −19 ° C. .

  In the intermediate pressure evaporator 16, the liquid phase portion of the first refrigerant at about −19 ° C. evaporates, and the second refrigerant in the second cooling cycle system 60 is cooled to a lower temperature by taking away the latent heat of evaporation. Thereafter, the gas-phase first refrigerant is supplied to the intermediate pressure compressor 3 via the pipe 55 and compressed to about 0.63 MPa. On the other hand, the first refrigerant in the liquid phase is supplied to the low-pressure expansion valve 14 via the pipe 53 and adiabatically expands to about 0.1 MPa, so that a gas-liquid mixed state of about −41 ° C. which is a saturation temperature is obtained. .

  The low pressure evaporator 17 evaporates all of the first refrigerant at about −41 ° C. to cool the second refrigerant in the second cooling cycle system 60 to about −35 ° C. The evaporated first refrigerant is supplied to the low-pressure compressor 2 via the pipe 56 and compressed.

  On the other hand, since the second refrigerant of the second cooling cycle system 60 cooled to about −35 ° C. is partially liquefied, it is gas-liquid separated by the gas-liquid separator 27. The second refrigerant in the liquid phase separated by the gas-liquid separator 27 is supplied from the pipe 43 to the heat transfer path 28 of the main heat exchanger 21, and further exchanges heat with the second refrigerant having a lower temperature, and is about -100 ° C. Until cooled. The second refrigerant at about −100 ° C. is cooled to about −120 ° C. by adiabatic expansion by the expansion valve 33 and is supplied to the nozzle 35 of the main heat exchanger 21. The second refrigerant sprayed from the nozzle 35 is a liquid phase second refrigerant in the heat transfer path 28, a gas phase second refrigerant in the heat transfer path 29, and a natural heat transfer path 30 inside the main heat exchanger 21. Cool each gas.

  The gas-phase second refrigerant of the gas-liquid separator 27 is supplied from the pipe 44 to the heat transfer path 29 of the main heat exchanger 21, further heat-exchanged with the low-temperature second refrigerant and cooled to about −100 ° C. Further, in the heat transfer path 30 on the downstream side, heat is exchanged with the second refrigerant having a temperature of about −170 ° C. sprayed from the nozzle 35 to be cooled to about −150 ° C., and most of the water is condensed. The second refrigerant at about −150 ° C. is adiabatically expanded by the expansion valve 34 to be cooled to about −170 ° C. and supplied to the nozzle 36 of the main heat exchanger 21. The low-temperature second refrigerant sprayed from the nozzle 36 cools the second refrigerant in the heat transfer path 30 and the natural gas in the heat transfer path 32 to about −150 ° C. inside the main heat exchanger 21.

  The natural gas thus cooled to about −150 ° C. is led to the expansion valve 37 via the pipe 45, where it is adiabatically expanded to near atmospheric pressure and taken out as a liquefied natural gas of about −162 ° C.

  The second refrigerant whose temperature has increased by heat exchange in the heat transfer paths 30 and 32 is reused for cooling the heat transfer paths 28-30 on the downstream side. The second refrigerant after cooling the heat transfer path 28-30 is supplied to the low-pressure compressor 23 of the second cooling cycle system 60 through the pipe 40. Hereinafter, the second refrigerant having reached about 40 ° C. and 5 MPa through the compression by the low-pressure compressor 23 and the cooling by the intermediate cooler 25, the compression by the high-pressure compressor 24, and the cooling by the post-cooler 26 is the first cooling cycle. It is cooled to about −35 ° C. by the system 61 and supplied again to the main heat exchanger 21 to be used for liquefaction of the raw material natural gas.

(Operation at startup)
Next, the operation at the time of activation will be described with reference to FIGS.

  It is desirable that each facility constituting the natural gas liquefaction facility is started in the order of the auxiliary cooling mechanism 62, the first cooling cycle system 61, the second cooling cycle system 60, and the main heat exchanger 21.

  Before the start-up, the temperature of the first refrigerant in the liquid receiver 11 of the first cooling cycle system 61 is warmed by the outside air to about 40 ° C., and the interior of the liquid receiver 11 and the piping 49-56 is The saturation pressure of the first refrigerant (propane) at 40 ° C. is about 1.4 MPa. In the present embodiment, the auxiliary cooling mechanism 62 is activated at this time (before the refrigerant compressor 1 is activated), and the first refrigerant of the liquid receiver 11 of the first cooling cycle system 61 is assumed to have an estimated cooling temperature of about −25. Cool to ° C. By cooling the first refrigerant in the liquid receiver 11 to about −25 ° C., the pressure of the first refrigerant in the first cooling cycle system 61 decreases to about 0.2 MPa, which is the saturation pressure of propane at −25 ° C. .

  At this time, the openings of the high pressure expansion valve 12, the intermediate pressure expansion valve 13, and the low pressure expansion valve 14 are set so that the outlet pressure of each valve during rated operation is about 0.63 MPa, 0.25 MPa, and 0.1 MPa, respectively. It is set in advance.

  FIG. 2A is a diagram illustrating an example of a change in the rotational speed (operation schedule) of the refrigerant compressor 1 of the first cooling cycle system 61 after startup.

  In the operation schedule shown in FIG. 2 (a), it is assumed that the rotational speed reaches the rated value in about 30 seconds if the load of the refrigerant compressor 1 is below the allowable range.

  FIG. 2B is a diagram showing changes in the opening degree of the high pressure suction adjustment mechanism 71, the medium pressure suction adjustment mechanism 72, and the low pressure suction adjustment mechanism 73.

  As shown in FIG. 2B, at the time of start-up, the suction adjusting mechanisms 71-73 are slightly opened (for example, about 30% of opening degree), and the suction of the low pressure compressor 2, the intermediate pressure compressor 3, and the high pressure compressor 4 is performed. Reduce the flow rate and suction pressure, and reduce the starting torque, suction pressure, and discharge pressure. Thereafter, with the passage of time, the opening degree of the suction adjustment mechanism 71-73 is increased, and the suction flow rate / suction pressure is increased to the rated value (fully opened: opening degree 100%).

  FIG. 3 is a diagram showing changes in pressure and flow rate when the refrigerant compressor 1 according to the present embodiment is started.

  Here, the operation of the low-pressure compressor 2, the intermediate-pressure compressor 3, and the high-pressure compressor 4 at the time of starting will be described with reference to FIG.

  In the present embodiment, the refrigerant held in the liquid receiver 11 is cooled to about −25 ° C., and the pressure of the first refrigerant in the first cooling cycle system 61 at the time of start-up is saturated with propane at −25 ° C. The pressure is about 0.2 MPa. Therefore, although the internal pressure of the liquid receiver 11 immediately after the start of the refrigerant compressor 1 is about 0.2 MPa, the suction pipe 56 of the low pressure compressor 2 is caused by the pressure loss action of the low pressure expansion valve 14 and the low pressure suction adjustment mechanism 73. The pressure decreases to 0.1 MPa, which is the rated operating condition (FIG. 3A). The pressure ratio of the low-pressure compressor 2 increases with the rotational speed of the refrigerant compressor 1, and the discharge pressure of the low-pressure compressor 2 increases toward the rated condition of 0.25 MPa (FIG. 3B).

  The suction pressure of the intermediate pressure compressor 3 is determined by the intake flow rate from the pipe 55 and the discharge pressure of the low pressure compressor 2, but toward the rated operating condition of 0.25 MPa as the refrigerant compressor 1 increases in rotation speed. It rises (FIG. 3 (a)). The pressure ratio of the intermediate pressure compressor 3 increases with the rotation speed of the refrigerant compressor 1, and the discharge pressure of the intermediate pressure compressor 3 increases toward the rated condition of 0.63 MPa (FIG. 3B).

  The suction pressure of the high-pressure compressor 4 is determined by the intake flow rate from the pipe 54 and the discharge pressure of the intermediate-pressure compressor 3, but toward the rated operating condition of 0.63 MPa as the rotational speed of the refrigerant compressor 1 increases. It increases (Fig. 3 (a)). The pressure ratio of the high-pressure compressor 4 increases with the rotational speed of the refrigerant compressor 1, and the discharge pressure of the high-pressure compressor 4 increases toward the rated condition of 1.5 MPa (FIG. 3 (b)).

  The time change of the mass flow rate of the first refrigerant flowing in the refrigerant compressor 1 is as shown in FIG. In the figure, the relative value with the suction mass flow rate of the high-pressure compressor 4 during rated operation as 1.0 is shown. The first refrigerant flowing into the intermediate pressure compressor 3 is larger than the flow rate of the low pressure compressor 2 because the first refrigerant sucked from the pipe 55 and the first refrigerant discharged from the low pressure compressor 2 merge. . Further, the first refrigerant flowing into the high-pressure compressor 4 joins the first refrigerant sucked from the pipe 54 and the first refrigerant discharged from the intermediate-pressure compressor 3, so that the flow rate of the intermediate-pressure compressor 3 is further increased. Become more. These flow rate change characteristics are determined by the rotational speed, suction temperature, inlet pressure, and outlet pressure of each refrigerant compressor 1.

  The first refrigerant discharged from the high-pressure compressor 4 becomes a high temperature exceeding 100 ° C., but is cooled to about 40 ° C. by the condenser 10 and flows into the receiver 11 in a liquid or gas-liquid mixed state depending on the pressure condition. To do. In the liquid receiver 11, the temperature is increased to about 40 ° C. by being mixed with the first refrigerant that was initially about −25 ° C. Since the discharge pressure of the high-pressure compressor 4 during rated operation is about 1.5 MPa, which is higher than the saturation pressure, the first refrigerant is stored in the liquid receiver 11 as a liquid.

  The first refrigerant in the liquid receiver 11 is depressurized by the high-pressure expansion valve 12, and finally reaches the rated condition of about 0.63 MPa at the outlet of the high-pressure expansion valve 12, and the temperature is about 9 ° C. which is the saturation temperature.

  In the high-pressure evaporator 15, the liquid phase portion of the first refrigerant evaporates, and the second refrigerant of the second cooling cycle system 60 is cooled by removing latent heat of evaporation. The gas-phase first refrigerant is supplied to the high-pressure compressor 4 through the pipe 54 and compressed. On the other hand, the first refrigerant in the liquid phase is supplied from the pipe 52 to the intermediate pressure expansion valve 13 and adiabatically expands toward the state of 0.25 MPa and −19 ° C., which are rated operating conditions.

  In the intermediate pressure evaporator 16, the liquid first refrigerant adiabatically expanded by the intermediate pressure expansion valve 13 evaporates, and the second refrigerant of the second cooling cycle system 60 is further cooled by taking away latent heat of evaporation. The gas-phase propane refrigerant is supplied from the pipe 55 to the intermediate pressure compressor 3 and compressed. On the other hand, the first refrigerant in the liquid phase is supplied from the pipe 53 to the low-pressure expansion valve 14 and adiabatically expands toward the rated operation conditions of 0.1 MPa and −41 ° C.

  In the low-pressure evaporator 17, the liquid-phase first refrigerant adiabatically expanded by the low-pressure expansion valve 14 evaporates to cool the second refrigerant in the second cooling cycle system 60, and the evaporated first refrigerant is low-pressure compressed from the pipe 56. Supplied to machine 2 and compressed.

(Function and effect)
FIG. 5 is a diagram showing a change in torque when the propane refrigerant compressor is started in the present embodiment.

  The driving torque of the refrigerant compressor 1 is proportional to the required power divided by the rotational speed, and the required power is proportional to the suction mass flow rate of the compressor and the specific enthalpy change. The descriptions “low pressure”, “medium pressure”, and “high pressure” in FIG. 5 are the required torques of the low pressure compressor 2, the intermediate pressure compressor 3, and the high pressure compressor 4, respectively. It is the total value. In addition, the description “drive source” in the figure is a torque curve of a general induction motor. The induction motor has a characteristic of generating a maximum torque at a rotational speed slightly lower than the rated rotational speed, and has a driving torque that is always larger than that during rated operation in the starting process, and the operation of the refrigerant compressor 1 as shown in FIG. It can be seen that it can be started without difficulty by operating the schedule and suction adjustment mechanism.

  On the other hand, FIG. 4 shows changes in the pressure at the inlet / outlet of the compressor and the mass flow rate at the time of startup when the auxiliary cooling mechanism 62 is omitted as a comparative example.

  Also in this comparative example, the temperature of the 1st refrigerant | coolant of the liquid receiver 11 of the 1st cooling cycle system | strain 61 before a start is warmed by external air, and rises to about 40 degreeC. Moreover, the inside of the liquid receiver 11 and the piping 49-56 rises to about 1.4 MPa which is the saturation pressure of propane at 40 ° C.

  After starting the refrigerant compressor 1, the suction pressure of the low-pressure compressor 2, the intermediate-pressure compressor 3, and the high-pressure compressor 4 gradually decreases toward the rated value (FIG. 4 (a)). The discharge pressure of these compressors 2-4 once rises according to the pressure ratio characteristics of the compressor and then gradually decreases toward the rated value (FIG. 4 (b)).

  The mass flow rate of the first refrigerant flowing inside the compressor 2-4 tends to be as shown in FIG. The suction flow rate of the compressor 2-4 has a characteristic that the volume flow rate is substantially constant when the rotation speed is the same, and the mass flow rate increases when the fluid pressure is high. FIG. 4C shows a relative value where the suction mass flow rate of the high-pressure compressor 4 during rated operation is 1.0, but after the start-up, the high-pressure compressor 4 continues until the pressure in the system decreases. It can be seen that the suction mass flow rate reaches 1.5 to twice the rated value.

  FIG. 6 is a diagram showing a change in torque when the refrigerant compressor 1 is started in the comparative example of FIG.

  The content indicated by each line in FIG. 6 corresponds to FIG. In the comparative example, immediately after startup, the suction mass flow rate increases from that at the time of rating, so it can be seen that the torque required at startup is greater than that during rated operation and exceeds the torque generated by the drive source. Therefore, in the comparative example of FIG. 4, the operation schedule shown in FIG. 2 is not established, and the first cooling cycle system 61 cannot be started.

  In contrast, in the present embodiment, as described above, the auxiliary cooling mechanism 62 cools the first refrigerant before the refrigerant compressor 1 is started, thereby greatly reducing the mass flow rate of the intake air of the refrigerant compressor 1. Therefore, the driving torque (required power) required for starting the refrigerant compressor 1 can be reduced, and even if the generated torque of the driving source at the time of starting is small, the first cooling cycle system 61, Specifically, the refrigerant compressor 1 can be started.

  Further, since the refrigerant discharged from the refrigerant compressor 1 is not discharged to the outside at the time of startup, the time required for increasing the discharge pressure of the refrigerant compressor 1 can be reduced, and the time required for completing the startup of the cooling cycle can be reduced. it can. Also, the refrigerant is not wasted.

  Here, as described above, by driving the auxiliary cooling mechanism 62 when the refrigerant compressor 1 is started, the torque at the start of the first cooling cycle system 61 can be reduced. It is not always necessary to operate the mechanism 62, specifically, the electric motor 84 and the compressor 81. However, when the auxiliary cooling mechanism 62 is operated during steady operation, there is an effect of lowering the temperature of the first refrigerant in the liquid receiver 11, and even if the cooling capacity of the condenser 10 is insufficient due to weather conditions or the like, the liquid receiver The first refrigerant of the vessel 11 can be cooled, so that the effect of maintaining the refrigerant flow rate of the first cooling cycle system 61 at the rated flow rate can be expected even when the atmospheric temperature or seawater temperature is high. It can also be expected to contribute to stable production.

<Second Embodiment>
(Constitution)
FIG. 7 is an overall configuration diagram of a natural gas liquefaction facility using a cooling cycle system according to the second embodiment of the present invention.

  The present embodiment relates to a natural gas liquefaction facility having a cooling cycle system as in the first embodiment, but the first cooling cycle system 61, the second cooling cycle system 60, and the main heat exchanger 21. Since this configuration is the same as that of the first embodiment, the illustration is omitted except for a part of the first cooling cycle system 61. For the first cooling cycle system 61, only the refrigerant compressor 1, the condenser 10, the liquid receiver 11, the expansion mechanism 18, and the evaporation mechanism 19 are illustrated in a simplified manner. Also in the present embodiment, the compressors of the first cooling cycle system 61 and the cooling cycle 60 are configured to be driven by electric motors, and the natural gas liquefaction facility of the present embodiment generates electric power supplied to these electric motors. A power generation gas turbine facility 100 and an intake air cooling system 200 for cooling the intake air of the gas turbine facility 100 are provided.

  The power generation gas turbine equipment 100 includes an air compressor 91 that sucks and compresses outside air from an intake duct 87, a combustor 86 that mixes and burns the compressed air and fuel, and generates high-temperature and high-pressure combustion gas. A turbine 93 that is expanded and converted into kinetic energy, and a generator 92 that converts the kinetic energy of the turbine 93 into electric power are provided. The power generation gas turbine facility 100 is connected to a waste heat recovery boiler 94 that recovers exhaust heat from the exhaust of the turbine 93 to generate water vapor, and a stack 95 that releases the exhaust gas that has passed through the exhaust heat recovery boiler 94 to the atmosphere. is there.

  An intake air cooling system 200 that cools the intake air of the power generation gas turbine equipment 100 includes an intake air cooling heat exchanger 90 that cools outside air drawn by the intake air cooling heat exchanger 90, and an intake air cooling heat exchanger 90 that includes an auxiliary cooling mechanism 62A. An evaporator 68 for exchanging heat with the auxiliary refrigerant is provided.

  The intake air cooling heat exchanger 90 is provided on the upstream side of the air compressor 91 in the intake air duct 87, and is connected on the loop via an evaporator 68 and pipes 65 and 66. The refrigerant flowing in the pipes 65 and 66 (hereinafter referred to as “third refrigerant”) is an antifreeze liquid such as ethylene glycol mixed water and circulates in the pipes 65 and 66 by a pump 69 provided in the middle of the pipe 66. Then, when passing through the evaporator 68, it is cooled by exchanging heat with an auxiliary refrigerant of an absorption chiller 96 (described later), returning to the intake air cooling heat exchanger 90 via the pipe 65, and the intake air of the air compressor 92 After being heated by heat exchange, the pump 69 is again sucked into the pump 69 via the pipe 66.

  Further, the intake air cooling system 200 includes a bypass pipe 101 having a control valve 99, a temperature detection means 97 for detecting the temperature of the third refrigerant flowing through the pipe 65, and the control valve 99 according to a detection signal from the temperature detection means 97. Is provided with a control device 98 for controlling the opening degree. The bypass pipe 101 is connected to the discharge side of the pump 69 in the pipe 66 and the pipe 65, and when the adjustment valve 99 is opened, part or all of the third refrigerant discharged from the pump 69 does not pass through the evaporator 68. It merges with the third refrigerant flowing through 65. The temperature detection means 97 is installed in the pipe 65, and the detection signal of the temperature detection means 97 is output to the control device 98. The control device 98 is configured to adjust the opening degree of the control valve 99 according to the detection signal from the temperature detecting means 97 based on the correlation between the preset refrigerant temperature and the opening degree of the control valve 99. Yes.

  The auxiliary cooling mechanism 62A cools the first refrigerant stored in the liquid receiver 11 by exchanging heat with the auxiliary refrigerant flowing in the pipes 57-59, and cools the first refrigerant of the first cooling cycle system 61. To do. In addition to the evaporator 80 in the liquid receiver 11, the auxiliary cooling mechanism 62 </ b> A includes an absorption cooler 96 that drives the water vapor generated from the exhaust heat recovery boiler 94 as a heat source to cool the auxiliary refrigerant. In the present embodiment, ammonia that can be cooled to about −60 ° C. is assumed as an auxiliary refrigerant of the absorption chiller 96, but the type of auxiliary refrigerant used in the auxiliary cooling mechanism 62A is not limited.

  The absorption chiller 96 is connected to the evaporator 80, the evaporator 68, and the exhaust heat recovery boiler 94 in the liquid receiver 11 via pipes 58, 59, and 102, respectively. The pipe 58 connecting the absorption cooler 96 and the evaporator 68 joins the pipe 59 connecting the evaporator 80 in the liquid receiver 11 and the absorption cooler 96 through the evaporator 68. The pipe 57 branched from the pipe 58 is connected to the evaporator 80 in the liquid receiver 11, and is connected to the pipe 59 through a pipe passing through the evaporator 80.

  In the middle of the pipes 57 and 59, shut-off valves 75a and 75b are respectively provided. On the upstream side and the downstream side of the evaporator 68 in the pipe 58, shut-off valves 76a and 76b are respectively provided.

  The supply destination of the low-temperature auxiliary refrigerant generated by the absorption chiller 96 is configured as described above, and by switching the shut-off valves 75a, 75b, 76a, and 76b, the liquid receiving liquid of the cooling cycle 61 is received. It is possible to switch between the evaporator 80 of the vessel 11 and the evaporator 68 of the intake air cooling system of the gas turbine.

(Operation at startup)
When the propane refrigerant cycle 61 is activated, after the power generation gas turbine equipment 100 is activated, the auxiliary cooling mechanism 62 is operated with the cutoff valves 75a and 75b first opened and the cutoff valves 76a and 76b closed. Then, a low-temperature auxiliary refrigerant is supplied to the evaporator 80 from the absorption cooler 96 driven by water vapor from the exhaust heat recovery boiler 94, and the auxiliary refrigerant evaporates in the evaporator 80, whereby the liquid receiver 11 of the cooling cycle 61. The first refrigerant is cooled. The operations of the second cooling cycle system 60 and the main heat exchanger 21 are the same as those in the first embodiment.

  Next, after starting the propane refrigerant cycle 61, the shutoff valves 75a and 75b are closed, the shutoff valves 76a and 76b are opened, and the auxiliary refrigerant is supplied to the evaporator 68 of the intake cooling system of the gas turbine. Then, in the evaporator 68, the third refrigerant having a temperature of about 15 ° C. driven by the pump 69 is cooled to about −30 ° C. along with the evaporation of the auxiliary refrigerant. The third refrigerant cooled to about −30 ° C. is mixed with the third refrigerant having a temperature of about 15 ° C. flowing from the path of the control valve 99 to become a third refrigerant of about 5 ° C. The third refrigerant at about 5 ° C. is supplied to the intake air cooling heat exchanger 90 from the pipe 65, and is heated to about 15 ° C. by exchanging heat with the intake air of the gas turbine. At that time, the intake air of the gas turbine, which was about 30 ° C. at the time of intake from the intake duct 87, is cooled to about 10 ° C. and supplied to the air compressor 91.

  At this time, the degree of solution of the control valve 99 is feedback-controlled by the control device 98 that has received the output signal of the temperature detecting means 97 so that the temperature of the third refrigerant flowing into the pipe 65 becomes about 5 ° C.

(Function and effect)
Also in the present embodiment, the same effect as that of the first embodiment can be obtained by providing the auxiliary cooling mechanism 62A.

  In addition, the gas turbine output may be reduced due to a decrease in intake air density under high temperature conditions such as in summer, but in this embodiment, the gas turbine intake air is cooled to suppress the gas turbine output decrease. can do. In addition, another power generation facility may be prepared for the purpose of compensating for the decrease in the output of the gas turbine. However, in the case of this embodiment, it is difficult to create a separate power generation facility, and the power generation facility is the minimum necessary. It can also be kept to a limit.

  Further, in the intake air cooling system of the gas turbine, by adjusting the temperature of the third refrigerant by the bypass path having the control valve 99, the moisture in the intake air freezes on the surface of the intake air cooling heat exchanger 90 due to excessive cooling. There is also a merit to prevent this.

<How to modify>
The auxiliary cooling facilities 62 and 62A cool the refrigerant of the cooling cycle system having the compressor, thereby reducing the starting torque of the compressor of the cooling cycle system. Therefore, it can be added to the existing cooling cycle system, and the cooling cycle system of the present invention can be configured by additionally installing the auxiliary cooling mechanisms 62 and 62A. Specifically, a refrigerant compressor that compresses the refrigerant, a condenser that cools and condenses the refrigerant compressed by the refrigerant compressor, a receiver that receives the refrigerant condensed by the condenser, and the liquid receiver A cooling cycle system comprising: an expansion mechanism for expanding the refrigerant from the vessel; and an evaporation mechanism for heat-exchanging the refrigerant expanded by the expansion mechanism to cool a cooling target and evaporate the refrigerant supplied to the refrigerant compressor If it already exists, the piping for circulating the auxiliary refrigerant is passed through the receiver for the existing cooling cycle system, and heat exchange is performed with the auxiliary refrigerant flowing through the piping before the refrigerant compressor is started. The cooling cycle system of the present invention can be configured by cooling the refrigerant in the liquid receiver.

<Others>
In the above description, an electric motor is used as a drive source for the refrigerant compressor 1 of the cooling cycle 61 and the compressors 23 and 24 of the cooling cycle 60. However, the gas turbine is connected to the drive shaft of each compressor to drive the gas turbine. It can also be used as a source. The gas turbine may be a single-shaft type or a double-shaft type, but a single-shaft gas turbine that generates a small amount of torque at the time of startup has a greater effect of applying the present invention. In the case where the gas turbine is used as a drive source of the compressor, in the second embodiment described above, the intake of the gas turbine is cooled to prevent the power generation output from being lowered. A configuration may also be adopted in which the intake air of the gas turbine for driving the compressors 23 and 24 in the cycle 60 is cooled. In this case, even when the temperature is high, it is possible to prevent the output of the compressor of each cooling cycle from being lowered, and the effect of stably maintaining the production amount of natural gas liquefaction throughout the year can be obtained.

  Further, although propane or ammonia is used as the auxiliary refrigerant used in the auxiliary cooling mechanisms 62 and 62A, other refrigerant substances may be used as long as the cooling temperature condition is satisfied. Moreover, although the cooling temperature of the 1st refrigerant | coolant of the liquid receiver 11 was assumed to be about -25 degreeC using the auxiliary | assistant cooling mechanisms 62 and 62A, what is necessary is just the assumption temperature which acts on reduction of the starting torque of the refrigerant compressor 1. For example, as an example of the assumed cooling temperature of the first refrigerant in the receiver 11 at the time of starting the refrigerant compressor 1 by the auxiliary cooling mechanisms 62 and 62A, the refrigerant temperature is equal to or lower than the refrigerant temperature at the outlet of the condenser 10 Below the saturation temperature of the refrigerant corresponding to the suction pressure during the rated operation of the machine 1. However, the lower the cooling temperature of the first refrigerant, the smaller the torque at the time of starting, but if the cooling temperature is too low, the difference from the temperature condition at the rated operation becomes large, and the time until the rated operating condition is reached after starting. May become longer. Therefore, the cooling temperature of the first refrigerant in the liquid receiver 11 may be set so as not to overcool in accordance with the torque characteristics of the actual drive source so as not to exceed the torque of the drive source of the refrigerant compressor 1.

  Furthermore, in the present embodiment, the case where the dedicated auxiliary cooling mechanisms 62 and 62A are provided in the single first cooling cycle system 61 has been described as an example, but it is shared by a plurality of cooling cycle systems having refrigerant compressors. It is also possible to provide an auxiliary cooling mechanism. In that case, if an operation schedule is made so as to shift the timing of starting each cooling cycle system, devices other than the evaporator installed in the liquid receiver of each cooling cycle system (the compressor 81, the condenser 82, Since the expansion valve 83, the electric motor 84, the absorption cooler 96 in FIG. 7 and the like do not need to be added, an increase in equipment cost can be suppressed.

Claims (10)

A refrigerant compressor for compressing the refrigerant;
A condenser that cools and condenses the refrigerant compressed by the refrigerant compressor;
A liquid receiver for receiving the refrigerant condensed in the condenser;
An expansion mechanism for expanding the refrigerant from the liquid receiver;
An evaporating mechanism that heat-exchanges with the refrigerant expanded by the expansion mechanism to cool the object to be cooled and evaporate the refrigerant supplied to the refrigerant compressor;
Auxiliary cooling for cooling the refrigerant in the liquid receiver before starting the refrigerant compressor by exchanging heat with the auxiliary refrigerant flowing through the pipe through the auxiliary refrigerant and passing through the pipe A cooling cycle system comprising a mechanism.   2. The cooling cycle system according to claim 1, wherein the auxiliary cooling mechanism cools the refrigerant in the liquid receiver below a refrigerant temperature at an outlet of the condenser before the refrigerant compressor is started. system.   2. The cooling cycle system according to claim 1, wherein the auxiliary cooling mechanism has a refrigerant in the receiver that is equal to or lower than a saturation temperature of a refrigerant corresponding to a suction pressure during a rated operation of the refrigerant compressor before the refrigerant compressor is started. Cooling cycle system characterized by cooling.   2. The cooling cycle system according to claim 1, wherein the auxiliary cooling mechanism expands the compressor that compresses the auxiliary refrigerant, the condenser that condenses the auxiliary refrigerant compressed by the compressor, and the auxiliary refrigerant condensed by the condenser. An expansion valve and an evaporator that is installed in the liquid receiver and expands by the expansion valve and exchanges heat with an auxiliary refrigerant that flows through the pipe to cool the refrigerant in the liquid receiver, Cooling cycle system.   2. The cooling cycle system according to claim 1, wherein the auxiliary cooling mechanism is an absorption chiller that cools the auxiliary refrigerant, and an auxiliary refrigerant that is installed in the receiver and is cooled by the absorption chiller and flows through the pipe. A cooling cycle system comprising an evaporator for heat exchange to cool the refrigerant in the liquid receiver. A first cooling cycle system which is the cooling cycle system according to claim 1;
A heat exchanger that cools and liquefies natural gas using the cooled medium cooled in the first cooling cycle system as a refrigerant;
A natural gas liquefaction facility comprising: a second cooling cycle system that compresses a medium to be cooled that has cooled natural gas by the exchanger and supplies the compressed medium to the first cooling cycle system. In the natural gas liquefaction equipment provided with the cooling cycle system of claim 5,
Gas turbine for power generation, exhaust heat recovery boiler that recovers exhaust heat from the exhaust of the gas turbine to generate water vapor, an intake air cooling system that cools the intake air of the gas turbine, and the auxiliary cooling that is provided in the intake air cooling system An evaporator for exchanging heat between the auxiliary refrigerant of the system and the refrigerant for cooling the intake air;
The auxiliary cooling mechanism is configured such that a pipe is configured so that an auxiliary refrigerant supply destination can be switched between an evaporator installed in the receiver and an evaporator installed in the intake air cooling system. Natural gas liquefaction equipment. A refrigerant compressor that compresses the refrigerant, a condenser that cools and condenses the refrigerant compressed by the refrigerant compressor, a liquid receiver that receives the refrigerant condensed by the condenser, and a refrigerant from the liquid receiver In an operation method of a cooling cycle system comprising: an expansion mechanism that expands; and an evaporation mechanism that causes heat exchange with the refrigerant expanded by the expansion mechanism to cool a cooling target and evaporate the refrigerant supplied to the refrigerant compressor.
An operation method of a cooling cycle system, wherein the refrigerant in the liquid receiver is cooled before the refrigerant compressor is started by passing an auxiliary refrigerant through the liquid receiver and exchanging heat with the auxiliary refrigerant.   9. The operation method of the cooling cycle system according to claim 8, wherein after starting the refrigerant compressor, the auxiliary refrigerant supplied to the evaporator installed in the liquid receiver is used as a refrigerant for cooling the intake air of the gas turbine. An operation method of a cooling cycle system characterized by the above. A refrigerant compressor that compresses the refrigerant, a condenser that cools and condenses the refrigerant compressed by the refrigerant compressor, a liquid receiver that receives the refrigerant condensed by the condenser, and a refrigerant from the liquid receiver In a remodeling method of a cooling cycle system comprising: an expansion mechanism that expands; and an evaporation mechanism that causes heat exchange with the refrigerant expanded by the expansion mechanism to cool a cooling target and evaporate the refrigerant supplied to the refrigerant compressor.
Passing a pipe through which the auxiliary refrigerant flows through the receiver, and heat exchange with the auxiliary refrigerant flowing through the pipe, thereby cooling the refrigerant in the receiver before starting the refrigerant compressor. Remodeling method of cooling cycle system characterized by

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