JP2010133605A - Ejector type refrigerating cycle - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To sufficiently obtain both compression efficiency improvement effects of compression means and driving power reduction effects in an ejector type refrigerating cycle. <P>SOLUTION: Heat of a refrigerant discharged from a first compression means 11a of a two-stage compression type compressor 10 is released by a radiator 12, and the refrigerant is decompressed by a nozzle 13a of an ejector 13. A flow of a refrigerant evaporated by a suction side evaporator 19 is branched by a suction side branch part 22. One refrigerant is sucked from a refrigerant suction port 13b of the ejector 13 and the other refrigerant is sucked to a second compression means 21a of the compressor 10 via a flow regulating valve 38. A second compression means 21a discharge refrigerant and a refrigerant made to flow out from the ejector 13 are mixed with each other and sucked to the first compression means 11a. Thus, both of the driving power reduction effects by pressure-rising action by the ejector 13 and the compression efficiency improvement effects of the first and second compression means 11a, 21a are sufficiently obtained, so as to effectively improve a COP. <P>COPYRIGHT: (C)2010,JPO&INPIT

Description

  The present invention relates to an ejector-type refrigeration cycle that is configured to increase the pressure of a refrigerant in multiple stages and include an ejector as refrigerant decompression means.

  Conventionally, an economizer refrigeration cycle, an ejector refrigeration cycle, and the like proposed for improving the coefficient of performance (COP) of a vapor compression refrigeration cycle are known.

  For example, as an economizer-type refrigeration cycle, a high-stage decompression unit that decompresses and expands a high-pressure refrigerant, a gas-liquid separator that separates gas-liquid of an intermediate-pressure refrigerant decompressed by the high-stage decompression unit, and a gas-liquid separation A low-stage decompression unit that decompresses and expands the liquid-phase refrigerant separated in the vessel to form a low-pressure refrigerant, a low-stage compression unit that pressurizes the low-pressure refrigerant until it becomes an intermediate-pressure refrigerant, and an intermediate-pressure refrigerant as the high-pressure refrigerant. A cycle including a high-stage compression unit that boosts the pressure until it reaches is known.

  In this economizer-type refrigeration cycle, the refrigerant that has been pressurized by the low-stage compression means and the gas-phase refrigerant separated by the gas-liquid separator are mixed, and the mixed intermediate-pressure refrigerant is mixed with the high-pressure refrigerant by the latter-stage compression means. Thus, the pressure of the refrigerant is increased in multiple stages. As a result, the pressure difference between the suction pressure and the discharge pressure in each compression means is reduced compared with the case where the pressure of the low-pressure refrigerant is increased to a high-pressure refrigerant by a single compression means, and the compression efficiency of each compression means is reduced. Has improved.

  Further, for example, as an ejector-type refrigeration cycle, an ejector that decompresses and expands a high-pressure refrigerant, a gas-liquid separator that separates gas-liquid of the refrigerant decompressed by the ejector, and a liquid phase separated by the gas-liquid separator A cycle is known that includes decompression means that decompresses and expands the refrigerant and flows out to the evaporator inlet side, and compression means that pressurizes the gas-phase refrigerant separated by the gas-liquid separator until it becomes a high-pressure refrigerant.

  In this ejector-type refrigeration cycle, the refrigerant flowing out of the evaporator from the refrigerant suction port of the ejector is sucked by the pressure drop of the refrigerant injected from the nozzle part of the ejector, thereby reducing the kinetic energy loss during decompression expansion in the nozzle part. Collected. The recovered kinetic energy (hereinafter referred to as recovered energy) is converted into pressure energy by the diffuser portion of the ejector.

  As a result, for the refrigeration cycle that does not employ an ejector as the decompression means, the refrigerant pressure on the suction side of the compression means is increased above the refrigerant evaporation pressure in the evaporator, thereby reducing the driving power of the compression means.

Furthermore, Patent Document 1 discloses an ejector-type refrigeration cycle that employs an ejector as a means corresponding to the high-stage decompression means of the economizer-type refrigeration cycle. As a result, Patent Document 1 attempts to improve COP by simultaneously obtaining the effect of improving the compression efficiency in the above-described economizer refrigeration cycle and the effect of reducing the driving power of the compression means by the ejector.
Japanese Patent No. 4-320762

  By the way, as described above, the effect of improving the compression efficiency in the economizer refrigeration cycle is obtained by reducing the pressure difference between the suction pressure and the discharge pressure in each compression means. Therefore, in the economizer refrigeration cycle, in order to effectively improve COP, it is necessary to control the pressure of the intermediate pressure refrigerant to an appropriate value.

  Here, in the ejector refrigeration cycle of Patent Document 1 shown in the overall configuration diagram of FIG. 34, in order to control the pressure of the intermediate pressure refrigerant to an appropriate value, the suction side evaporator 19 and the refrigerant suction port 13b of the ejector 13 are used. It is necessary to appropriately adjust the opening degree of the flow rate adjusting means (flow rate adjusting valve) 38 disposed between them.

  For example, in order to increase the pressure of the intermediate pressure refrigerant, the flow rate of the flow rate adjustment valve 38 is reduced, and the flow rate of the refrigerant sucked directly into the low-stage compression means 21a from the outlet side of the suction side evaporator 19 is increased. There must be. However, if the opening degree of the flow rate adjustment valve 38 is reduced, the pressure loss in the refrigerant flow path from the suction side evaporator 19 to the refrigerant suction port 13b of the ejector 13 increases.

  As a result, the flow rate of the suction refrigerant sucked into the ejector 13 from the refrigerant suction port 13b is reduced, and the recovery energy of the ejector 13 is also reduced. In other words, in the ejector refrigeration cycle of Patent Document 1, if the pressure of the intermediate pressure refrigerant is controlled to an appropriate value in order to obtain the effect of improving the compression efficiency in the economizer refrigeration cycle, the effect of reducing the driving power of the compression means by the ejector is sufficient. It becomes difficult to obtain.

  In other words, in the ejector type refrigeration cycle of Patent Document 1, it is difficult to control the operation of the cycle so as to simultaneously obtain the effect of improving the compression efficiency in the economizer type refrigeration cycle and the effect of reducing the driving power of the compression means by the ejector. There's a problem.

  In view of the above points, the present invention is to increase the refrigerant pressure in multiple stages and to sufficiently obtain both effects of improving the compression efficiency of the compression means and reducing the driving power in the ejector refrigeration cycle including the ejector. Objective.

  In order to achieve the above object, in the first aspect of the present invention, the first compression means (11a) that compresses and discharges the refrigerant and the heat dissipation that dissipates the high-pressure refrigerant discharged from the first compression means (11a). The refrigerant is sucked from the refrigerant suction port (13b) by the flow of the high-speed jet refrigerant jetted from the nozzle (13a) that decompresses and expands the high-pressure refrigerant that has flowed out of the radiator (12). The ejector (13) that mixes the refrigerant and the suction refrigerant sucked from the refrigerant suction port (13b) to increase the pressure, the suction side evaporator (19) that evaporates the refrigerant, and the suction side evaporator (19) flowed out. The suction side branch (22) that branches the refrigerant flow and causes one of the branched refrigerant to flow out to the refrigerant suction port (13b) side, and the other refrigerant branched at the suction side branch (22) Flow rate adjustment means for adjusting the flow rate ( 8) and second compression means (21a) that compresses and discharges the refrigerant whose flow rate is adjusted by the flow rate adjustment means (38), and the first compression means (11a) is the second compression means (21a). It is characterized by an ejector type refrigeration cycle that compresses and discharges an intermediate pressure refrigerant that is a mixture of the refrigerant discharged from the refrigerant and the refrigerant discharged from the ejector (13).

  According to this, even if the flow rate adjusting means (38) changes the flow rate of the refrigerant sucked into the second compression means (21a), the refrigerant suction port (13b) of the ejector (13) from the suction side evaporator (19). ), The flow rate of the refrigerant flowing from the suction side branch portion (22) to the refrigerant suction port (13b) side can be reliably ensured.

  Therefore, the refrigerant can be sucked from the refrigerant suction port (13b) of the ejector (13), and the pressure elevating action can be exerted on the ejector (13). As a result, the drive power reduction effect of the compression means by the ejector (13) can be reliably obtained.

  Moreover, the intermediate pressure can be controlled so that the pressure difference between the suction pressure and the discharge pressure in the first and second compression means (11a, 21a) is reduced in a state where the ejector (13) exerts a pressure increasing action. it can. Therefore, the same compression efficiency improvement effect as the above-described economizer refrigeration cycle can be obtained.

  As a result, both the compression efficiency improvement effect and the driving power reduction effect of the compression means can be sufficiently obtained, and the COP of the ejector refrigeration cycle can be effectively improved.

  According to a second aspect of the present invention, in the ejector refrigeration cycle according to the first aspect, the gas-liquid refrigerant flowing out from the ejector (13) is separated, and the separated gas-phase refrigerant is sucked into the first compression means 11a. The outflow side gas-liquid separator (17) that flows out to the side and the liquid-phase refrigerant separated by the outflow side gas-liquid separator (17) are expanded under reduced pressure and flow out to the inlet side of the suction side evaporator (19) And suction side pressure reducing means (18).

  Specifically, the refrigerant is supplied from the outflow side gas-liquid separator (17) to the suction side evaporator (19) via the suction side pressure reducing means (18), and the suction side evaporator (19). Can demonstrate the refrigerating capacity.

  Further, since the saturated gas phase refrigerant separated by the outflow side gas-liquid separator (17) can be sucked into the first compression means (11a), the second compression means (21a) can suck only the discharged refrigerant. On the other hand, the COP can be further improved by reducing the compression work when the refrigerant is isentropically compressed in the first compression means (11a).

  According to a third aspect of the present invention, in the ejector-type refrigeration cycle of the second aspect, the ejector is disposed between the outlet side of the ejector (13) and the suction side of the first compression means (11a), and the ejector (13) An outflow side evaporator (16) for evaporating the outflowed refrigerant is provided. According to this, not only the suction side evaporator (19) but also the outflow side evaporator (16) can exhibit the refrigerating capacity.

  Further, in the suction side evaporator (19), the refrigerant evaporating pressure corresponds to the suction action of the injected refrigerant, and in the outflow side evaporator (16), the refrigerant evaporating pressure after being boosted by the boosting action of the ejector (13) Therefore, the refrigerant evaporation temperatures of the outflow side evaporator (16) and the suction side evaporator (19) can be set to different temperatures.

  According to a fourth aspect of the present invention, in the ejector-type refrigeration cycle according to the first aspect, the branch part (42) for branching the flow of the refrigerant flowing out from the ejector (13), and one outlet of the branch part (42) An outlet side evaporator (16) disposed between the suction side and the suction side of the second compression means (21a) and evaporating one refrigerant branched at the branch portion (42); and a branch portion (42) And suction side decompression means (18) for decompressing and expanding the other branched refrigerant to flow out to the suction side evaporator (19) inlet side.

  According to this, specifically, the refrigerant is supplied to both the outflow side evaporator (16) and the suction side evaporator (19), and the refrigerating capacity is exhibited in both the evaporators (16, 19). Can do. Further, similarly to the third aspect of the invention, the refrigerant evaporation temperatures of both evaporators (16, 19) can be set to different temperatures.

  According to a fifth aspect of the present invention, in the ejector refrigeration cycle according to any one of the first to fourth aspects, the internal heat that exchanges heat between the refrigerant flowing out of the radiator (12) and the low-pressure side refrigerant of the cycle. An exchange (30) is provided. According to this, the enthalpy difference (refrigeration capacity) between the enthalpy of the inlet side refrigerant (19) inlet side refrigerant and the enthalpy of outlet side refrigerant can be expanded, and COP can be improved.

  Specifically, the low-pressure side refrigerant of the cycle may be a refrigerant that has flowed out of the suction-side evaporator (19) as in the invention described in claim 6, or as in the invention described in claim 7. The refrigerant may be sucked into the second compression means (21a).

  According to an eighth aspect of the present invention, in the ejector refrigeration cycle according to the first aspect, to the branching portion (32) for branching the flow of the refrigerant flowing out from the radiator (12), and the suction side evaporator (19) A suction side decompression means (18) for decompressing and expanding the inflowing refrigerant, and the nozzle section (13a) decompresses and expands one of the refrigerants branched at the branch section (32), thereby sucking the decompression means (18). Is characterized in that the other refrigerant branched at the branch portion (32) is expanded under reduced pressure.

  According to this, specifically, the refrigerant is supplied to the suction side evaporator (19) via the branch part (32) and the suction side pressure reducing means (18), and the suction side evaporator (19) supplies the refrigerating capacity. Can be demonstrated.

  According to the ninth aspect of the present invention, in the ejector refrigeration cycle according to the eighth aspect, the ejector is disposed between the outlet side of the ejector (13) and the suction side of the first compression means (11a), and from the ejector (13). An outflow side evaporator (16) for evaporating the outflowed refrigerant is provided.

  According to this, not only the suction side evaporator (19) but also the outflow side evaporator (16) can exhibit the refrigerating capacity. Furthermore, similarly to the third aspect of the invention, the refrigerant evaporation temperatures of the outflow side evaporator (16) and the suction side evaporator (19) can be set to different temperatures.

  According to a tenth aspect of the present invention, in the ejector refrigeration cycle according to the eighth or ninth aspect, the other refrigerant branched at the branching portion (32) is dissipated, and the suction side decompression means (18) inlet side An auxiliary radiator (12e) is provided to flow out into the air.

  According to this, since the low enthalpy refrigerant cooled by both the radiator (12) and the auxiliary radiator (12e) can flow into the suction side evaporator (19), the suction side evaporator (19). COP can be improved by expanding the enthalpy difference (refrigeration capacity) between the enthalpy of the inlet side refrigerant and the enthalpy of the outlet side refrigerant.

  Moreover, since the refrigerant flowing into the nozzle part (13a) of the ejector (13) is not cooled by the auxiliary radiator (12e), the enthalpy does not decrease with respect to the refrigerant flowing into the suction side pressure reducing means (18). Thereby, the amount of recovered energy in the nozzle part (13a) can be increased, and the COP can be further improved.

  According to an eleventh aspect of the invention, in the ejector refrigeration cycle according to any one of the eighth to tenth aspects, the other refrigerant branched at the branch portion (32) and the low-pressure side refrigerant of the cycle are heated. An internal heat exchanger (30) to be exchanged is provided. According to this, the enthalpy difference (refrigeration capacity) between the enthalpy of the inlet side refrigerant (19) inlet side refrigerant and the enthalpy of outlet side refrigerant can be expanded, and COP can be improved.

  According to a twelfth aspect of the present invention, in the ejector refrigeration cycle according to any one of the eighth to tenth aspects, the refrigerant in the decompression / expansion process in the suction side decompression means (18) and the low-pressure side refrigerant of the cycle are heated. An internal heat exchanger (31) to be replaced is provided. According to this, the enthalpy difference (refrigeration capacity) between the enthalpy of the inlet side refrigerant (19) inlet side refrigerant and the enthalpy of outlet side refrigerant can be expanded, and COP can be improved.

  Specifically, the low-pressure side refrigerant of the cycle may be a refrigerant that has flowed out of the suction-side evaporator (19) as in the invention described in claim 13, or as in the invention described in claim 14. The refrigerant may be sucked into the second compression means (21a).

  In the invention according to claim 15, in the ejector refrigeration cycle according to claim 1, the first branch portion (32) configured to be able to branch the flow of the refrigerant flowing out from the radiator (12), and an ejector ( 13) between the second branch portion (42) configured to be able to branch the refrigerant flow flowing out from the second branch portion (42) and the second compression means (21a) suction side. An outflow side evaporator (16) that is disposed and evaporates one refrigerant branched at the second branch portion (42), and a first and a second that expand the refrigerant flowing into the suction side evaporator (19) under reduced pressure. 2 suction side decompression means (18, 28), and the nozzle part (13a) decompresses and expands one of the refrigerants branched at the first branch part (32), thereby providing first suction side decompression means (18). Expands the other refrigerant branched at the first branch part (32) under reduced pressure. The second suction-side decompression means (28) decompresses and expands the other refrigerant branched at the second branching section (42), and the suction-side evaporator (19) includes first and second suction-side decompression means ( 18, 28), at least one of the refrigerant expanded under reduced pressure is evaporated.

  According to this, the refrigerant flow is branched only by the first branch part (32), the refrigerant flowing out from the first suction side decompression means (18) is supplied to the suction side evaporator (19), and the ejector ( 13) By supplying the refrigerant flowing out from 13) to the outflow side evaporator (16) through the second branch portion (42), both the outflow side evaporator (16) and the suction side evaporator (19) A cycle configuration that demonstrates the refrigerating capacity can be realized.

  Further, the refrigerant flow is branched only by the second branch section (42), and the refrigerant flowing out from the ejector (13) is supplied to both the outflow side evaporator (16) and the suction side evaporator (19). Thus, it is possible to realize a cycle configuration in which both the outflow side evaporator (16) and the suction side evaporator (19) exhibit refrigeration capacity.

  Furthermore, by branching the flow of the refrigerant in both the first and second branch parts (22, 32) and supplying the refrigerant to both the outflow side evaporator (16) and the suction side evaporator (19), A cycle configuration in which the refrigerating capacity is exhibited in both the outflow side evaporator (16) and the suction side evaporator (19) can be realized. In each cycle configuration, the refrigerant evaporating temperatures of the outflow side evaporator (16) and the suction side evaporator (19) can be set to different temperatures, as in the third aspect of the invention.

  In the invention described in claim 16, in the ejector refrigeration cycle according to claim 15, internal heat for exchanging heat between the other liquid-phase refrigerant branched at the first branch portion (32) and the low-pressure side refrigerant of the cycle. An exchange (30) is provided.

  According to this, the enthalpy difference (refrigeration capacity) between the enthalpy of the inlet side refrigerant and the enthalpy of the outlet side refrigerant of both the evaporators (16, 19) of the outflow side evaporator (16) and the suction side evaporator (19) is obtained. It can be expanded to improve COP.

  Specifically, the low-pressure side refrigerant of the cycle may be a refrigerant that has flowed out of the suction-side evaporator (19) as in the invention described in claim 17, or as in the invention described in claim 18. The refrigerant may be sucked into the second compression means (21a).

  According to a nineteenth aspect of the present invention, in the ejector refrigeration cycle according to the first aspect, the flow of the high-pressure refrigerant discharged from the first compression means (11a) is branched, and one of the branched refrigerants is dissipated in the radiator. (12) The branch part (32) that flows out to the side, the branch refrigerant radiator (12d) that radiates heat of the other refrigerant branched at the branch part (32), and the high pressure that flows out from the branch refrigerant radiator (12d) It is characterized by comprising suction side decompression means (18) for decompressing and expanding the refrigerant to flow out to the inlet side of the suction side evaporator (19).

  According to this, since the heat exchange capability (heat dissipation capability) of the radiator (12) and the branch refrigerant radiator (12d) can be changed independently, the heat exchange capability of the branch refrigerant radiator (12d) and the suction The heat exchange capacity (endothermic performance) of the side evaporator (19) can be easily adapted. Therefore, it is easy to stabilize the operation of the cycle.

  Moreover, the enthalpy of the refrigerant | coolant which flows into the nozzle part (13a) of an ejector (13) is unnecessary because the heat exchanger (12) employ | adopts that whose heat exchange capability is lower than a branch refrigerant | coolant radiator (12d). Can be avoided. Thereby, the amount of recovered energy in the nozzle part (13a) can be increased, and COP can be improved.

  According to the twentieth aspect of the present invention, in the ejector refrigeration cycle according to the eighteenth or nineteenth aspect, the ejector (13) is disposed between the outlet side of the ejector (13) and the suction side of the first compression means (11a). And an outflow side evaporator (16) for evaporating the refrigerant that has flowed out.

  According to this, not only the suction side evaporator (19) but also the outflow side evaporator (16) can exhibit the refrigerating capacity. Furthermore, similarly to the third aspect of the invention, the refrigerant evaporation temperatures of the outflow side evaporator (16) and the suction side evaporator (19) can be set to different temperatures.

  According to a twenty-first aspect of the invention, in the ejector refrigeration cycle according to the nineteenth or twentieth aspect, an internal heat exchanger that exchanges heat between the refrigerant flowing out of the branch refrigerant radiator (12d) and the low-pressure side refrigerant of the cycle. 30). According to this, the enthalpy difference (refrigeration capacity) between the enthalpy of the inlet side refrigerant (19) inlet side refrigerant and the enthalpy of outlet side refrigerant can be expanded, and COP can be improved.

  Specifically, the low-pressure side refrigerant of the cycle may be a refrigerant that has flowed out of the suction-side evaporator (19) as in the invention described in claim 22, or as in the invention described in claim 23. The refrigerant may be sucked into the second compression means (21a).

  According to a twenty-fourth aspect of the present invention, in the ejector refrigeration cycle according to any one of the first to twenty-third aspects, the high-pressure side pressure reducing means (14) for decompressing and expanding the refrigerant flowing out to the inlet side of the nozzle portion (13a). It is characterized by providing.

  According to this, the gas-liquid two-phase refrigerant can be caused to flow into the nozzle portion (13a) of the ejector (13) by the pressure reducing action of the high pressure side pressure reducing means (14), and the liquid phase is supplied to the nozzle portion (13a). In contrast to the case where the refrigerant is introduced, the boiling of the refrigerant in the nozzle portion 13a can be promoted, and the nozzle efficiency can be improved.

  Therefore, the amount of recovered energy can be increased and the boosting capability of the ejector (13) can be increased, so that the COP can be further improved. Furthermore, since the refrigerant passage area of the nozzle part (13a) can be enlarged compared with the case where the liquid phase refrigerant is caused to flow into the nozzle part (13a), the nozzle part (13a) can be easily processed. As a result, the manufacturing cost of the ejector (13) can be reduced.

  According to a twenty-fifth aspect of the present invention, in the ejector refrigeration cycle according to the first to twenty-fourth aspects, the high-pressure side refrigerant of the cycle discharged from the first compression means (11a) is guided to the suction-side evaporator (19). A bypass passage (25) and an opening / closing means (26) for opening and closing the bypass passage (25) are provided.

  According to this, when the suction side evaporator (16) is frosted, the high temperature refrigerant discharged from the first compression means (11a) flows into the suction side evaporator (16) by opening the opening / closing means (26). Can be defrosted.

In the invention according to claim 26, the first compression means (11a) for compressing and discharging the refrigerant, the use side heat exchanger (54) for exchanging heat between the refrigerant and the heat exchange target fluid, the refrigerant, Refrigerant channel switching for switching between an outdoor heat exchanger (53) for exchanging heat with the outside air, a refrigerant channel in a cooling operation mode for cooling the fluid to be heat exchanged, and a refrigerant channel in a heating operation mode for heating the fluid to be heat exchanged High speed at which the refrigerant radiated in one of the means (51, 52) and the outdoor heat exchanger (53) and the use side heat exchanger (54) is injected from the nozzle portion (13a) for decompressing and expanding. An ejector (13) that sucks the refrigerant from the refrigerant suction port (13b) by the flow of the injected refrigerant and mixes the injected refrigerant and the suction refrigerant sucked from the refrigerant suction port (13b) to increase the pressure, and outdoor heat exchange (53) and use A suction side branch section (22) for branching the flow of the refrigerant evaporated in the other heat exchanger of the heat exchanger (54) and allowing the branched one refrigerant to flow out to the refrigerant suction port (13b) side; The second flow rate adjusting means (38) for adjusting the flow rate of the other refrigerant branched by the suction side branch portion (22) and the second flow rate adjusted by the flow rate adjusting means (38) are compressed and discharged. Compression means (21a),
The first compression means (11a) compresses and discharges the intermediate pressure refrigerant obtained by mixing the refrigerant discharged from the second compression means (21a) and the refrigerant flowing out from the ejector (13), and the refrigerant flow switching means (51 , 52) in the cooling operation mode, the refrigerant discharged from the first compression means (11a) is radiated by the outdoor heat exchanger (53), and the refrigerant evaporated by the use side heat exchanger (54) is discharged. In the heating operation mode, the refrigerant discharged from the first compression means (11a) is radiated by the use side heat exchanger (54) and the outdoor heat is switched to the refrigerant flow path leading to the suction side branch (22) side. It is characterized by switching to the refrigerant flow path which guides the refrigerant evaporated in the exchanger (53) to the suction side branch part (22) side.

  According to this, even if the flow rate adjusting means (38) changes the refrigerant flow rate sucked into the second compression means (21a) in any of the cooling operation mode and the heating operation mode, the suction side evaporation is performed. Since the pressure loss of the refrigerant flow path from the container (19) to the refrigerant suction port (13b) of the ejector (13) is not greatly increased, the refrigerant flows from the suction side branch (22) to the refrigerant suction port (13b) side. The refrigerant flow rate can be secured reliably.

  Therefore, the refrigerant can be sucked from the refrigerant suction port (13b) of the ejector (13), and the pressure elevating action can be exerted on the ejector (13). As a result, the driving power reduction effect of the compression means by the ejector (13) can be reliably obtained in any operation mode.

  Moreover, the intermediate pressure can be controlled so that the pressure difference between the suction pressure and the discharge pressure in the first and second compression means (11a, 21a) is reduced in a state where the ejector (13) exerts a pressure increasing action. it can. Therefore, in any operation mode, the same compression efficiency improvement effect as the above-described economizer refrigeration cycle can be obtained.

  As a result, both the compression efficiency improvement effect and the driving power reduction effect of the compression means can be sufficiently obtained, and the COP of the ejector refrigeration cycle can be effectively improved.

  Further, specifically, as in the invention described in claim 27, in the ejector-type refrigeration cycle according to claim 26, the outflow side gas-liquid separation that separates the gas-liquid of the refrigerant flowing out from the ejector (13). The gas-phase refrigerant outlet of the outflow side gas-liquid separator (17) is connected to the suction port side of the second compression means (21a), and the refrigerant flow path switching means (51, 52) In the cooling operation mode, the refrigerant radiated by the outdoor heat exchanger (53) is caused to flow into the nozzle portion (13a), and the liquid-phase refrigerant separated by the outflow side gas-liquid separator (17) is used for heat exchange on the use side. In the heating operation mode, the refrigerant radiated by the use side heat exchanger (54) is caused to flow into the nozzle part (13a) and the outflow side gas-liquid separator (17). ) Liquid phase refrigerant separated in And it switches the refrigerant flow path for flowing into the exchanger (53).

  According to a twenty-eighth aspect of the present invention, in the ejector refrigeration cycle according to any one of the first to twenty-seventh aspects, a first discharge capacity changing means (11b) for changing a refrigerant discharge capacity of the first compression means (11a). ) And second discharge capacity changing means (21b) for changing the refrigerant discharge capacity of the second compression means (21a). The first discharge capacity changing means (11b) and the second discharge capacity changing means (21b) The refrigerant discharge capacities of the first compression means (11a) and the second compression means (21a) can be changed independently of each other.

  According to this, the refrigerant discharge capacity of the first compression means (11a) and the refrigerant discharge capacity of the second compression means (21a) are independently adjusted, and either the first or second compression means (11a, 21a) is adjusted. Can be operated while exhibiting high compression efficiency. Therefore, COP as the whole ejector type refrigerating cycle can be further improved.

  According to a twenty-ninth aspect of the present invention, in the ejector refrigeration cycle according to any one of the first to twenty-eighth aspects, the first compression means (11a) and the second compression means (21a) are in the same housing. It is accommodated and it is comprised integrally. According to this, the first compression means (11a) and the second compression means (21a) can be reduced in size, and the entire ejector refrigeration cycle can be reduced in size.

  As in the invention described in claim 30, in the ejector refrigeration cycle according to any one of claims 1 to 29, the first compression means (11a) increases the pressure of the refrigerant until it reaches a critical pressure or more. It may be.

  In addition, the code | symbol in the bracket | parenthesis of each means described in this column and the claim shows the correspondence with the specific means as described in embodiment mentioned later.

(First embodiment)
An example in which the ejector refrigeration cycle of the present invention is applied to a refrigerator will be described with reference to FIGS. This refrigerator cools the inside of a freezer that is a space to be cooled to an extremely low temperature of about −30 to −10 ° C. FIG. 1 is an overall configuration diagram of an ejector refrigeration cycle 100 of the present embodiment.

  In the ejector-type refrigeration cycle 100, the compressor 10 sucks refrigerant and boosts and discharges the refrigerant in multiple stages. Specifically, the compressor 10 of the present embodiment drives the first and second compression units 11a and 21a and the first and second compression units 11a and 21a in a single housing 10a. 1 is a two-stage booster type electric compressor configured to accommodate the second electric motors 11b and 21b.

  As the first and second compression means 11a and 21a, various compression mechanisms such as scroll-type compression means and vane-type compression means can be employed. The first and second electric motors 11b and 21b are each independently controlled in operation (number of rotations) by a control signal output from a control device to be described later. Any format may be adopted.

  And by this rotation speed control, the refrigerant | coolant discharge capability of the 1st, 2nd compression means 11a and 21a is each changed independently. Accordingly, the first and second electric motors 11b and 21b of the present embodiment constitute first and second discharge capacity changing means for changing the refrigerant discharge capacity of the first and second compression means 11a and 21a, respectively. .

  The housing 10a is provided with a suction port 10b for sucking low-pressure refrigerant, an intermediate-pressure port 10c for flowing intermediate-pressure refrigerant, and a discharge port 10d for discharging high-pressure refrigerant. These ports 10b to 10d are connected to the first and second compression means 11a and 21a inside the housing 10a.

  Specifically, the suction port 10b is connected to the suction port of the second compression means 21a, and the intermediate pressure port 10c is communicated with the discharge port of the second compression means 21a and the suction port of the first compression means 11a. The discharge port 10d is connected to the discharge port of the first compression means 11a. Therefore, the first compression unit 11a sucks, compresses and discharges the intermediate pressure refrigerant obtained by mixing the refrigerant discharged from the second compression unit 21a and the refrigerant flowing in from the intermediate pressure port 10c.

  Further, as shown in FIG. 1, a radiator 12 is connected to the discharge port 10 d of the compressor 10. The radiator 12 is a heat exchanger for heat radiation that radiates and cools the high-pressure refrigerant by exchanging heat between the high-pressure refrigerant discharged from the compressor 10 and outside air (outside air) blown by the cooling fan 12a. . The cooling fan 12a is an electric blower in which the number of rotations (the amount of blown air) is controlled by a control voltage output from the control device.

  In the ejector refrigeration cycle 100 of the present embodiment, a normal chlorofluorocarbon refrigerant is employed as the refrigerant, and a subcritical refrigeration cycle in which the high-pressure side refrigerant pressure does not exceed the critical pressure of the refrigerant is configured. Therefore, the radiator 12 functions as a condenser that condenses the refrigerant.

  An ejector 13 is connected to the outlet side of the radiator 12. The ejector 13 is a refrigerant decompression unit that decompresses and expands the refrigerant, and is also a refrigerant circulation unit that circulates the refrigerant by suction of a refrigerant flow ejected at high speed.

  Specifically, the ejector 13 squeezes the passage area of the liquid-phase refrigerant flowing out of the radiator 12 to a small size, and expands the liquid-phase refrigerant in an isentropic manner under reduced pressure, and the refrigerant injection port of the nozzle portion 13a. The refrigerant suction port 13b is arranged so as to communicate and sucks the refrigerant flowing out from a suction side evaporator 19 described later.

  Further, a diffuser portion 13c that mixes the high-speed refrigerant flow ejected from the nozzle portion 13a and the suction refrigerant from the refrigerant suction port 13b to increase the pressure at the downstream portion of the refrigerant flow of the nozzle portion 13a and the refrigerant suction port 13b. Is provided.

  The diffuser portion 13c is formed in a shape that gradually increases the refrigerant passage area, and functions to decelerate the refrigerant flow to increase the refrigerant pressure, that is, to convert the velocity energy of the refrigerant into pressure energy. Furthermore, the accumulator 17 is connected to the exit side of the diffuser part 13c.

  The accumulator 17 is an outflow-side gas-liquid separator that separates the gas-liquid refrigerant flowing out of the diffuser portion 13c and stores excess liquid-phase refrigerant in the cycle. An intermediate pressure port 10 c of the compressor 10 is connected to the gas-phase refrigerant outlet of the accumulator 17, and a suction-side evaporator 19 is connected to the liquid-phase refrigerant outlet via an electric variable throttle mechanism 18. ing.

  The variable throttle mechanism 18 is a suction-side decompression unit that further decompresses and expands the liquid refrigerant flowing out of the accumulator 17. More specifically, the variable throttle mechanism 18 includes a valve body configured to be able to change the throttle opening, and an electric actuator including a stepping motor that changes the throttle opening of the valve body. Yes. The operation of the variable aperture mechanism 18 is controlled by a control signal output from the control device.

  The suction-side evaporator 19 evaporates the low-pressure refrigerant and exhibits a heat-absorbing effect by exchanging heat between the low-pressure refrigerant decompressed and expanded by the variable throttle mechanism 18 and the freezer air circulated by the blower fan 19a. This is an endothermic heat exchanger. Therefore, in this embodiment, the air in a freezer becomes a heat exchange target fluid. The blower fan 19a is an electric blower in which the rotation speed (the amount of blown air) is controlled by a control voltage output from the control device.

  Connected to the outlet side of the suction side evaporator 19 is a suction side branch portion 22 that branches the flow of the refrigerant flowing out of the suction side evaporator 19. The suction side branch portion 22 is configured by a three-way joint having three inflow / outflow ports, and one of the inflow / outflow ports is a refrigerant inflow port, and two are refrigerant outflow ports. Such a three-way joint may be constituted by joining pipes having different pipe diameters, or may be constituted by providing a plurality of refrigerant passages having different passage diameters in a metal block or a resin block.

  The refrigerant suction port 13b of the ejector 13 is connected to one refrigerant outlet of the suction side branch portion 22, and the flow rate adjusting valve 38 constituting the flow rate adjusting means is connected to the other refrigerant outlet. As is clear from FIG. 1, devices that generate pressure loss such as a variable throttle mechanism are arranged in the refrigerant flow path from one refrigerant outlet of the suction side branch portion 22 to the refrigerant suction port 13b. Absent.

  Accordingly, the pressure loss of the refrigerant flow path from one refrigerant outlet of the suction side branching section 22 to the refrigerant suction port 13b slightly changes depending on the flow rate of the circulating refrigerant, but greatly changes when the ejector refrigeration cycle 100 is operated. There is nothing.

  The basic configuration of the flow rate adjusting valve 38 is the same as that of the variable throttle mechanism 18. More specifically, the pressure loss of the refrigerant passage inside the flow rate adjusting valve 38 is formed smaller than the pressure loss of the refrigerant passage inside the variable throttle mechanism 18. Further, the suction port 10 b of the compressor 10 is connected to the outlet side of the flow rate adjustment valve 38.

  A control device (not shown) includes a known microcomputer including a CPU, a ROM, a RAM, and the like and peripheral circuits thereof. This control device performs various calculations and processes based on the control program stored in the ROM, and controls the operation of the various electric actuators 11b, 12a, 18, 19a, 21b, 38 and the like described above.

  Therefore, this control device functions as a first discharge capacity control means for controlling the operation of the first electric motor 11b as the first discharge capacity change means, and operates as the second electric motor 21b as the second discharge capacity change means. It also has a function as a second discharge capacity control means for controlling the above. Of course, the first discharge capacity control means and the second discharge capacity control means may be configured by different control devices.

  In addition, the control device is provided with a detection value of a sensor group (not shown) such as an outside air sensor for detecting the outside air temperature, an inside temperature sensor for detecting the inside temperature, an operation switch for operating the refrigerator, and the like. Various operation signals from an operation panel (not shown) are input.

  Next, the operation of the present embodiment in the above configuration will be described based on the Mollier diagram of FIG. When the operation switch of the operation panel is turned on, the control device operates the first and second electric motors 11b and 21b, the cooling fan 12a, the variable throttle mechanism 18, the blower fan 19a, and the flow rate adjustment valve 38. Thereby, the compressor 10 sucks the refrigerant, compresses it, and discharges it.

First refrigerant discharged from the compression means 11a state of the compressor 10 is a _2-point in FIG. The high-temperature and high-pressure gas-phase refrigerant discharged from the first compression means 11a flows into the radiator 12, exchanges heat with the blown air (outside air) blown from the cooling fan 12a, dissipates heat and condenses (a ₂ points → b ₂ points).

The refrigerant radiated by the radiator 12 flows into the nozzle portion 13a of the ejector 13 and is decompressed and expanded in an isentropic manner (b ₂ point → c ₂ point). And the pressure energy of a refrigerant | coolant is converted into speed energy at the time of this decompression | expansion expansion, and a refrigerant | coolant is injected at high speed from the refrigerant | coolant injection port of the nozzle part 13a. Due to the refrigerant suction action of the injection refrigerant, one refrigerant branched at the suction side branch portion 22 from the refrigerant suction port 13b is sucked.

Then, the suction refrigerant sucked from the injection refrigerant and the refrigerant suction port 13b that is injected from the nozzle portion 13a is mixed flows into the diffuser portion 13c of the ejector 13 (c ₂ points → d ₂ points and i ₂ points → d ₂ points). In the diffuser portion 13c, the refrigerant speed energy is increased to pressure energy by increasing the passage area, so that the refrigerant pressure increases (d ₂ point → e ₂ point).

The refrigerant flowing out of the diffuser portion 13c flows into the accumulator 17 and is separated into a gas phase refrigerant and a liquid phase refrigerant (e ₂ point → f ₂ point and e ₂ point → g ₂ point). The gas-phase refrigerant flowing out from the gas-phase refrigerant outlet of the accumulator 17 is sucked into the first compression means 11a from the intermediate pressure port of the compressor 10.

On the other hand, the liquid-phase refrigerant that has flowed out from the liquid-phase refrigerant outlet of the accumulator 17 is further decompressed and expanded isoenthalpically by the variable throttle mechanism 18 to reduce its pressure (g ₂ point → h ₂ point). Variable decompressed and expanded refrigerant is in throttle mechanism 18, and flows into the suction side evaporator 19, the air blowing fan 19a is evaporated by absorbing heat from the internal air circulated blown (h ₂ points → i ₂ points).

  Thereby, the air in a warehouse is cooled. At this time, the control device adjusts the throttle opening degree of the variable throttle mechanism 18 so that the refrigerant evaporation temperature in the suction side evaporator 19 becomes the temperature set on the operation panel. The flow of the refrigerant flowing out from the suction side evaporator 19 flows into the suction side branching section 22 and flows into the refrigerant suction port 13b side of the ejector 13 and the refrigerant flow flowing into the flow rate adjusting valve 38 side. Divided.

Here, the flow rate ratio Gα / Gβ of the refrigerant flow rate Gα flowing into the refrigerant suction port 13b side with respect to the refrigerant flow rate Gβ flowing into the flow rate adjustment valve 38 side is determined according to the valve opening degree of the flow rate adjustment valve 38. As described above, the refrigerant that has flowed from the suction side branch portion 22 to the refrigerant suction port 13b is sucked into the ejector 13 from the refrigerant suction port 13b (i ₂ point → d ₂ point).

On the other hand, the refrigerant flowing into the flow rate adjustment valve 38 side from the suction side branch portion 22 drops in pressure due to its pressure loss when passing through the flow rate adjustment valve 38 (i ₂ point → j ₂ point). Further, the refrigerant whose flow rate is adjusted by the flow rate adjusting valve 38 is sucked into the second compression means 21a of the compressor 10 and compressed until it reaches an intermediate pressure (j ₂ point → k ₂ point).

  At this time, the control device controls the operations of the first and second compression units 11a and 21a and the flow rate control valve 38 so that the COP of the ejector refrigeration cycle 100 as a whole approaches the maximum.

  Specifically, the control device controls the operation of the flow rate control valve 38 so that the suction refrigerant flow rate sucked from the refrigerant suction port 13b, that is, the refrigerant flow rate Gα is equal to or higher than a predetermined flow rate. Further, the control device controls the first and second compression means 11a and 21a in order to improve the mechanical efficiency of the first and second compression means 11a and 21a in a state where the operation of the flow control valve 38 is controlled as described above. Are controlled so as to be substantially equal to each other.

  Note that the compression efficiency means that the amount of increase ΔH1 is actually calculated when ΔH1 is the amount of increase in the enthalpy of the refrigerant when the refrigerant is isentropically compressed by the first and second compression means 11a, 21a. 1, the value divided by the enthalpy increase ΔH2 of the refrigerant when the refrigerant is pressurized by the second compression means 11a, 21a.

  For example, when the rotation speed and the pressure increase amount (pressure difference between the discharge pressure and the suction pressure) of the first and second compression means 11a and 21a increase, the temperature of the refrigerant rises due to the frictional heat, and the actual enthalpy increase ΔH2 Increases the compression efficiency.

The intermediate pressure refrigerant whose pressure has been increased by the second compression means 21a is mixed with the refrigerant flowing out from the gas-phase refrigerant outlet of the accumulator 17 in the compressor 10 (f ₂ point → l ₂ point and k ₂ point → l _2. point). The mixed refrigerant is compressed again is sucked into the first compression means 11a (l ₂ points → a ₂ points).

  Since the ejector-type refrigeration cycle 100 of this embodiment operates as described above, the suction side evaporator 19 can cause the refrigerant to exhibit an endothermic effect to cool the inside of the freezer.

  Furthermore, since the valve opening degree of the flow rate adjusting valve 38 is adjusted so that the refrigerant can be reliably sucked from the refrigerant suction port 13b of the ejector 13, the diffuser portion 13c of the ejector 13 can surely exert a pressure increasing action. Therefore, the power reduction effect of the compressor 10 by the boosting action of the ejector 13 described above can be reliably obtained.

  In addition, the intermediate pressure is controlled so as to reduce the pressure difference between the suction pressure and the discharge pressure in the first and second compression means 11a and 21a while reliably obtaining the power reduction effect of the compressor 10 due to the boosting action of the ejector 13. Therefore, it is possible to obtain the same compression efficiency improvement effect as that of the economizer refrigeration cycle described above.

  As a result, both the compression efficiency improvement effect and the driving power reduction effect of the compression means can be sufficiently obtained, and the COP of the ejector refrigeration cycle can be effectively improved.

  Furthermore, since the saturated gas phase refrigerant can be sucked into the first compression means 11a from the gas phase refrigerant outlet of the accumulator 17, the first compression means 11a is different from the case where only the refrigerant discharged from the second compression means 21a is sucked. The COP can be further improved by reducing the compression work when the refrigerant is isentropically compressed.

  Moreover, since the refrigerant that has flowed out of the radiator 12 flows directly into the nozzle portion 13a of the ejector 13, the amount of pressure reduction in the nozzle portion 13a of the ejector 13, that is, the pressure between the nozzle portion 13a inlet side pressure and the nozzle portion 13a outlet side pressure. The pressure difference increases. As a result, the difference between the enthalpy of the refrigerant on the inlet side of the nozzle portion 13a and the enthalpy of the refrigerant on the outlet side of the nozzle portion 13a can be increased to increase the amount of recovered energy, and COP can be further improved.

  The ability to effectively improve the COP as in this embodiment is that the refrigerant evaporation temperature of the suction side evaporator 19 is -30, for example, as in this embodiment, where the pressure difference between the high-pressure refrigerant and the low-pressure refrigerant is large. It is extremely effective in a refrigeration cycle apparatus that lowers to an extremely low temperature of -10 ° C.

(Second Embodiment)
In the present embodiment, as shown in the overall configuration diagram of FIG. 3, the refrigerant flowing out from the radiator 12 outlet side to the nozzle portion 13 a inlet side of the ejector 13 is decompressed with respect to the ejector refrigeration cycle 100 of the first embodiment. An example in which a fixed throttle 14 is added as a high-pressure side pressure reducing means for expansion will be described. As the fixed throttle 14, a capillary tube, an orifice, or the like can be employed. Other configurations are the same as those of the first embodiment.

  Next, the operation of the present embodiment will be described based on the Mollier diagram of FIG. In addition, the code | symbol which shows the state of the refrigerant | coolant in FIG. 4 uses the same code | symbol as the code | symbol which shows the state of the same refrigerant | coolant in FIG. 2, and has changed only the subscript. The same applies to the Mollier diagram described in the following embodiments.

When the ejector refrigeration cycle 100 of this embodiment is operated, the refrigerant flowing out of the radiator 12 is decompressed and expanded at the fixed throttle 14 until it becomes a gas-liquid two-phase state (b ₄ point → b ′ ₄ point). Then, the refrigerant that has flowed out of the fixed throttle 14 flows into the nozzle portion 13 a of the ejector 13. Other operations are the same as those in the first embodiment.

  Therefore, also in the ejector type refrigeration cycle 100 of the present embodiment, the same effect as that of the first embodiment can be obtained. Further, the gas-liquid two-phase refrigerant can be caused to flow into the nozzle portion 13a of the ejector 13 by the pressure reducing action of the fixed throttle 14, and the liquid phase refrigerant is caused to flow into the nozzle portion 13a. The boiling of the refrigerant can be promoted, and the nozzle efficiency can be improved.

  As a result, the amount of energy recovered by the ejector 13 can be increased and the boosting capability of the diffuser portion 15c can be increased, so that the COP can be further improved. Furthermore, since the refrigerant passage area of the nozzle portion 13a can be enlarged as compared with the case where the liquid phase refrigerant is caused to flow into the nozzle portion 13a, the processing of the nozzle portion 13a is facilitated. As a result, the manufacturing cost of the ejector 13 can be reduced.

  In this embodiment, the example in which the fixed throttle 14 is employed as the high-pressure side decompression unit has been described. However, a variable diaphragm mechanism may be employed as the high-pressure side decompression unit. For example, a temperature-type expansion valve, an electric expansion valve, or the like that adjusts the valve opening degree so that the degree of superheat of the suction-side evaporator 19 outlet-side refrigerant falls within a predetermined range may be employed as the high-pressure side pressure reducing means. .

(Third embodiment)
In the present embodiment, an example in which an internal heat exchanger 30 is added to the ejector refrigeration cycle 100 of the first embodiment will be described as shown in the overall configuration diagram of FIG. The internal heat exchanger 30 performs heat exchange between the refrigerant flowing out of the radiator 12 passing through the high-pressure side refrigerant flow path 30a and the low-pressure side refrigerant of the cycle passing through the low-pressure side refrigerant flow path 30b.

  More specifically, the low-pressure side refrigerant of the cycle in the present embodiment is a refrigerant sucked into the second compression means 21 a of the compressor 10. In addition, as a specific configuration of the internal heat exchanger 30, a double-pipe heat exchanger in which an inner pipe that forms a low-pressure side refrigerant flow path 30b is arranged inside an outer pipe that forms the high-pressure side refrigerant flow path 30a. The configuration is adopted. Of course, the high-pressure side refrigerant flow path 30a may be an inner pipe and the low-pressure side refrigerant flow path 30b may be an outer pipe.

  Further, a configuration in which the refrigerant pipes forming the high-pressure side refrigerant flow path 30a and the low-pressure side refrigerant flow path 30b are brazed and joined to exchange heat may be employed. Other configurations are the same as those of the first embodiment.

Next, the operation of the present embodiment will be described based on the Mollier diagram of FIG. When the ejector-type refrigeration cycle 100 of this embodiment is operated, the enthalpy of the low-pressure refrigerant flowing out from the flow regulating valve 38 is increased by the action of the internal heat exchanger 30 with respect to the first embodiment (j ₆ points → j ′ ₆ points), the enthalpy of the high-pressure refrigerant flowing out of the radiator 12 decreases (b ₆ points → b ′ ₆ points). Other operations are the same as those in the first embodiment.

  Thereby, since the dryness of the injection refrigerant injected from the nozzle part 13a can be reduced and the flow rate of the injection refrigerant can be reduced, the refrigerant pressure flowing out from the diffuser part 13c can be reduced. Therefore, the enthalpy of the refrigerant in the accumulator 17 can be reduced, and the enthalpy of the liquid phase refrigerant flowing from the accumulator 17 to the suction side evaporator 19 can also be reduced.

  Therefore, not only the same effects as in the first embodiment can be obtained, but also the enthalpy difference between the enthalpy of the inlet side refrigerant and the enthalpy of the outlet side refrigerant of the suction side evaporator 19 can be enlarged. As a result, the refrigeration capacity that can be exhibited by the ejector refrigeration cycle 100 can be increased, and the COP can be further improved.

(Fourth embodiment)
In the present embodiment, an example in which the configuration of the internal heat exchanger 30 is changed with respect to the ejector refrigeration cycle 100 of the third embodiment as shown in the overall configuration diagram of FIG. 7 will be described.

  More specifically, the low-pressure side refrigerant of the cycle in the present embodiment is a refrigerant that is sucked from the suction-side branch portion 22 to the refrigerant suction port 13 b of the ejector 13 among the refrigerant that has flowed out from the suction-side evaporator 19. Of course, the low-pressure side refrigerant of the cycle may be a refrigerant that has flowed out from the suction side branch portion 22 to the flow rate adjustment valve 38 side. Other configurations are the same as those of the first embodiment.

Next, the operation of the present embodiment will be described based on the Mollier diagram of FIG. When the ejector type refrigeration cycle 100 of this embodiment is operated, the enthalpy of the refrigerant flowing from the suction side branch portion 22 to the refrigerant suction port 13b side is increased by the action of the internal heat exchanger 30 (i ₈ points → i ′ ₈ points), the enthalpy of the liquid refrigerant flowing out of the radiator 12 is reduced (b ₈ points → b ′ ₈ points). Other operations are the same as those in the first embodiment.

  Therefore, when the ejector-type refrigeration cycle 100 of this embodiment is operated, not only the same effect as in the first embodiment can be obtained, but also the inlet-side refrigerant of the suction-side evaporator 19 can be obtained as in the third embodiment. The enthalpy difference between the enthalpy and the enthalpy of the outlet side refrigerant can be increased, and the refrigerating capacity that the ejector refrigeration cycle 100 can exhibit can be increased. As a result, COP can be further improved.

(Fifth embodiment)
In the present embodiment, as shown in the overall configuration diagram of FIG. 9, with respect to the ejector refrigeration cycle 100 of the first embodiment, the ejector 13 outlet side (specifically, the diffuser portion 13 c outlet side) and the compressor 10. The example which has arrange | positioned the outflow side evaporator 16 between these intermediate pressure ports 10c is demonstrated.

  The basic configuration of the outflow side evaporator 16 is the same as that of the suction side evaporator 19. Further, the outflow side evaporator 16 exchanges heat between the refrigerant flowing out from the ejector 13 and the blown air blown from the blower fan 16a, thereby evaporating the refrigerant flowing out of the ejector 13 and exerting an endothermic effect. It is a vessel.

  Accordingly, in the present embodiment, the blown air blown from the blower fan 16a is also the heat exchange target fluid. The blower fan 16a is an electric blower in which the rotation speed (the amount of blown air) is controlled by a control voltage output from the control device. Other configurations are the same as those of the first embodiment.

When the ejector-type refrigeration cycle of this embodiment is operated, not only can the endothermic effect be exerted in the suction side evaporator 19 as in the first embodiment, but also the Mollier diagram of FIG. As shown in FIG. 4, the liquid phase refrigerant in the process from the point e _{10 to the} point e ′ ₁₀ can be evaporated to exert an endothermic effect. Thereby, the blowing air from the blowing fan 16a can also be cooled. Other operations are the same as those in the first embodiment.

  At this time, in the outflow side evaporator 16, the refrigerant evaporates at a temperature higher than the refrigerant evaporation temperature in the suction side evaporator 19. That is, in the suction side evaporator 19 and the outflow side evaporator 16, the refrigerant evaporates at different temperature zones. Thereby, in this embodiment, while being able to acquire the effect similar to 1st Embodiment, the air in the refrigerator in a refrigerator which preserve | saves food, a drink, etc. at the low temperature of 0 to 10 degreeC with the ventilation fan 16a, for example. Can also be cooled.

(Sixth embodiment)
An example in which the ejector refrigeration cycle 200 of the present invention is applied to a refrigeration / refrigeration apparatus will be described with reference to FIGS. This freezing / refrigeration apparatus cools the inside of the refrigerator, which is the cooling target space, to a low temperature of about 0 to 10 ° C, and further cools the inside of the freezer, which is another cooling target space, to an extremely low temperature of about -30 to -10 ° C. To do. FIG. 11 is an overall configuration diagram of the ejector refrigeration cycle 200 of the present embodiment.

  The ejector refrigeration cycle 200 of the present embodiment is obtained by changing the components and the connection mode thereof, that is, changing the cycle configuration, with respect to the ejector refrigeration cycle 100 of the first embodiment which is the premise thereof. is there. Therefore, in FIG. 11, the same reference numerals are given to the same or equivalent parts as in the first embodiment, and the description thereof is omitted. This also applies to the entire configuration diagram described in the following embodiment.

  In the ejector refrigeration cycle 200, a branch portion 32 that branches the flow of the refrigerant that has flowed out of the radiator 12 is connected to the outlet side of the radiator 12. The basic configuration of the branch portion 32 is the same as that of the suction side branch portion 22.

  One refrigerant outlet of the branch part 32 is connected to the nozzle part 13 a side of the ejector 13, and the other refrigerant outlet is connected to the suction side evaporator 19 side of the ejector 13. Further, the outlet side evaporator 16 similar to that of the fifth embodiment is connected to the outlet side of the diffuser portion 13 c of the ejector 13, and the intermediate pressure port 10 c of the compressor 10 is connected to the refrigerant outlet side of the outlet side evaporator 16. Is connected.

  In addition, a suction side evaporator 19 is connected to the other refrigerant outlet of the branch portion 32 via a variable throttle mechanism 18 which is a suction side pressure reducing means. Further, a suction side branching portion 22 is connected to the outlet side of the suction side evaporator 19. The refrigerant suction port 13b of the ejector 13 is connected to one refrigerant outlet of the suction side branch portion 22, and the inlet side of the flow rate adjusting valve 38 is connected to the other refrigerant outlet. Other configurations are the same as those of the first embodiment.

  Next, the operation of the present embodiment in the above configuration will be described based on the Mollier diagram of FIG. When the ejector-type refrigeration cycle 200 of this embodiment is operated, the refrigerant flowing out of the radiator 12 flows through the branch portion 32 into the refrigerant portion flowing into the nozzle portion 13a side of the ejector 13 and the suction side evaporator 19 of the ejector 13. It is divided into the refrigerant flow flowing into the side.

  Here, in this embodiment, when the refrigerant evaporation temperature of the suction side evaporator 19 reaches a predetermined temperature, the refrigerant flow rate Gnoz flowing into the nozzle portion 13a side and the refrigerant flow rate Ge flowing into the refrigerant suction port 13b side The flow rate characteristics (pressure loss characteristics) of the nozzle unit 13a and the variable throttle mechanism 18 are determined so that the flow rate ratio Ge / Gnoz of the nozzles 13a and the variable throttle mechanism 18 is an optimum flow rate ratio at which high COP can be exhibited as a whole cycle.

The intermediate-pressure refrigerant flowing from the branch portion 32 to the nozzle part 13a side, forward flow (b ₁₂ points → c ₁₂ points → d ₁₂ points → e ₁₂ points) of the nozzle portion 13a → the diffuser portion 13c, to the discharge side evaporator 16 Inflow.

Refrigerant flowing into the discharge side evaporator 16 evaporates by absorbing heat from the refrigerator air circulated blown by the blower fan 16a ₍₁₂ points → f ₁₂ points e). Thereby, the air in a refrigerator is cooled. Then, the refrigerant flowing out of the outflow side evaporator 16 is sucked into the first compression means 11a from the intermediate pressure port 10c of the compressor 10.

On the other hand, the intermediate pressure refrigerant that has flowed out from the branch portion 32 to the suction side evaporator 19 side is further decompressed and expanded isoenthalpically by the variable throttle mechanism 18 to reduce its pressure (b ₁₂ point → h ₁₂ point). . Variable decompressed and expanded refrigerant is in throttle mechanism 18, and flows into the suction side evaporator 19, the air blowing fan 19a by such refrigerant is evaporated by absorbing heat from the freezer in air circulated blown (h ₁₂ points → i ₁₂ points). Thereby, the air in a freezer is cooled.

  The flow of the refrigerant flowing out from the suction side evaporator 19 flows into the suction side branching section 22 and flows into the refrigerant suction port 13b side of the ejector 13 and the refrigerant flow flowing into the flow rate adjusting valve 38 side. Divided.

The refrigerant that has flowed out from the suction side branch portion 22 toward the refrigerant suction port 13b is sucked from the refrigerant suction port 13b (i ₁₂ points → d ₁₂ points). On the other hand, the refrigerant flowing into the flow rate adjustment valve 38 side from the suction side branch portion 22 decreases its pressure when the flow rate is adjusted by the flow rate adjustment valve 38 and is sucked into the second compression means 21a to become an intermediate pressure. (I ₁₂ points → j ₁₂ points → k ₁₂ points).

The intermediate pressure refrigerant whose pressure has been increased by the second compression means 21a is mixed with the refrigerant flowing out of the outflow side evaporator 16 in the compressor 10 (f ₁₂ point → l ₁₂ point and k ₁₂ point → l ₁₂ point). . The mixed refrigerant is sucked into the first compression means 11a and compressed again (point l ₂ → point a ₂ ). Other operations are the same as those in the first embodiment.

  Since the ejector refrigeration cycle 200 of this embodiment operates as described above, the inside of the refrigerator and the freezer can be cooled. That is, since the flow of the refrigerant is divided so that the flow rate ratio Ge / Gnoz becomes the optimum flow rate ratio in the branching portion 22, the refrigerant is appropriately supplied to both the outflow side evaporator 16 and the suction side evaporator 19. it can. Therefore, both the outflow side evaporator 16 and the suction side evaporator 19 can exhibit the cooling action simultaneously.

  At this time, the refrigerant evaporation pressure of the suction side evaporator 19 becomes the pressure after the pressure is increased by the diffuser portion 13c and then further reduced by the variable throttle mechanism 18, and the refrigerant evaporation pressure of the outflow side evaporator 16 is the pressure of the suction side evaporator 19. The refrigerant evaporation pressure becomes the pressure after being increased by the second compressor 21 and the diffuser unit 13c.

  Therefore, the refrigerant evaporation pressure (refrigerant evaporation temperature) of the suction side evaporator 19 can be made sufficiently lower than the refrigerant evaporation pressure (refrigerant evaporation temperature) of the outflow side evaporator 16. As a result, the outflow evaporator 16 can be used for cooling in a low-temperature refrigerator, and the suction-side evaporator 19 can be used for cooling in a cryogenic freezer.

  Further, similarly to the first embodiment, both the power reduction effect of the compressor 10 due to the boosting action of the ejector 13 and the effect of improving the compression efficiency similar to the economizer refrigeration cycle can be sufficiently obtained. COP of the refrigeration cycle can be effectively improved.

  Moreover, in the present embodiment, the refrigerant flows in the order of the compressor 10 → the radiator 12 → the branch portion 32 → the ejector 13 → the outflow side evaporator 16 → the compressor 10, and the compressor 10 → the radiator 12 → the branch portion 22 → variable. The refrigerant flows in the order of the throttle mechanism 18 → suction side evaporator 19 → suction side branching part 22 → ejector 13 → outflow side evaporator 16 → compressor 10, and further compressor 10 → heat radiator 12 → branch part 22 → variable throttle. The refrigerant flows in the order of mechanism 18 → suction side evaporator 19 → suction side branching section 22 → flow rate adjusting valve 38 → compressor 10.

  That is, since the flow of the refrigerant passing through the evaporators such as the outflow side evaporator 16 and the suction side evaporator 19 is annular, lubricating oil (refrigeration machine oil) for the first and second compressors 11 and 21 is used as the refrigerant. Even if mixed, this oil can be prevented from staying in the outflow side evaporator 16 and the suction side evaporator 19.

(Seventh embodiment)
In the present embodiment, as shown in the overall configuration diagram of FIG. 13, the outflow-side evaporator 16 and the blower fan 16a are eliminated from the ejector refrigeration cycle 200 of the sixth embodiment, and the accumulator 17 and the internal heat are further eliminated. An example in which the exchanger 30 is added will be described. Note that the ejector refrigeration cycle 200 of the present embodiment is applied to a refrigerator similar to the first embodiment.

  The internal heat exchanger 30 flows out from the high-pressure side refrigerant flowing out from the branch portion 32 passing through the high-pressure side refrigerant flow path 30a to the variable throttle mechanism 18 side and from the suction side evaporator 19 passing through the low-pressure side refrigerant flow path 30b. Heat exchange is performed with the low-pressure side refrigerant of the cycle. More specifically, the low-pressure side refrigerant of the cycle in the present embodiment is a refrigerant that has flowed out of the outflow side evaporator 19.

  Of course, as in the third embodiment, the low-pressure side refrigerant of the cycle may be a refrigerant sucked into the second compression means 21a. Other configurations are the same as those of the sixth embodiment.

Next, the operation of the present embodiment will be described based on the Mollier diagram of FIG. When the ejector refrigeration cycle 200 of this embodiment is operated, the enthalpy of the refrigerant flowing out from the suction-side evaporator 19 is increased by the action of the internal heat exchanger 30 with respect to the sixth embodiment (i ₁₄ points → i ′ ₁₄ points), the enthalpy of the refrigerant flowing out from the branch part 32 to the suction side evaporator 19 side is reduced (b ₁₄ → b ′ ₁₄ points).

Further, refrigerant flowing from the diffuser portion 13c of the ejector 13, since the discharge side evaporator 16 is discontinued, the gas-liquid separation flows into the accumulator 17 (e ₁₄ points → f ₁₄ points). The gas-phase refrigerant separated by the accumulator 17 is sucked into the first compression means 11a from the intermediate pressure port 10c of the compressor 10. Other operations are the same as in the sixth embodiment.

  Therefore, when the ejector refrigeration cycle 200 of the present embodiment is operated, the inside of the freezer can be cooled by the heat absorbing action of the refrigerant in the suction side evaporator 19. Further, similarly to the first embodiment, both the power reduction effect of the compressor 10 due to the boosting action of the ejector 13 and the effect of improving the compression efficiency similar to the economizer refrigeration cycle can be sufficiently obtained. COP of the refrigeration cycle can be effectively improved.

  Furthermore, the enthalpy of the refrigerant flowing into the suction side evaporator 19 can be reduced by the action of the internal heat exchanger 30 as compared with the sixth embodiment. Therefore, since the enthalpy difference between the enthalpy of the inlet side refrigerant and the enthalpy of the outlet side refrigerant of the suction side evaporator 19 can be increased to increase the refrigeration capacity, the COP can be further improved.

  In addition, in the present embodiment, in the internal heat exchanger 30, heat exchange is performed between the high-pressure refrigerant flowing through the refrigerant passage extending from the branch portion 32 outlet side to the variable throttle mechanism 18 inlet side and the low-pressure refrigerant sucked into the refrigerant suction port 13b. Therefore, the enthalpy of the refrigerant flowing from the branch portion 32 to the nozzle portion 13a is not unnecessarily reduced.

  Thereby, the further COP improvement effect can be acquired. The reason is that the amount of recovered energy in the nozzle portion 13a can be increased by unnecessarily reducing the enthalpy of the refrigerant flowing into the nozzle portion 13a.

  This will be explained in more detail. As the enthalpy of the refrigerant flowing into the nozzle portion 13a increases, the slope of the isentropic line becomes gentle (small). Therefore, in the case where the nozzle portion 13a is isentropically expanded by the same pressure, the higher the enthalpy of the nozzle portion 13a inlet-side refrigerant, the higher the enthalpy of the nozzle portion 13a inlet-side refrigerant and the enthalpy of the nozzle portion 13a outlet-side refrigerant. The difference (recovered energy amount) increases.

  Accordingly, as the enthalpy of the refrigerant flowing into the nozzle portion 13a increases, the amount of recovered energy in the nozzle portion 13a increases. As the amount of recovered energy increases, the amount of pressure increase in the diffuser section 13c can be increased, and a further COP improvement effect can be obtained.

(Eighth embodiment)
An example in which the ejector refrigeration cycle 300 of the present invention is applied to a refrigeration / refrigeration apparatus similar to that of the sixth embodiment will be described with reference to FIGS. FIG. 15 is an overall configuration diagram of the ejector refrigeration cycle 300 of the present embodiment.

  In the ejector refrigeration cycle 300, as shown in FIG. 15, the accumulator 17 is abolished with respect to the fifth embodiment, and the flow of the refrigerant flowing out of the diffuser part 13c is branched to the outlet side of the diffuser part 13c of the ejector 13. A branch portion 42 is provided.

  The basic configuration of the branch portion 42 is the same as that of the suction side branch portion 22. This branch part 42 branches the flow of the refrigerant flowing out from the diffuser part 13c, and flows it out to the outflow side evaporator 16 side and the variable throttle mechanism 18 side. In the present embodiment, the inlet side of the suction side evaporator 19 is connected to the outlet side of the variable throttle mechanism 18.

  Further, the branch portion 42 has a flow direction of the refrigerant flowing out from the one refrigerant outlet 42b to the outlet-side evaporator 16 side and a flow direction of the refrigerant flowing out from the other refrigerant outlet 42c to the variable throttle mechanism 18 side. Further, it is formed in a substantially Y shape so as to face the target direction and intersect at an acute angle with respect to the flow direction of the refrigerant flowing into the refrigerant inflow port 42a from the outlet side of the diffuser portion 13c.

  Therefore, the refrigerant that has flowed into the branch portion 42 flows out of the branch portion 42 without unnecessarily reducing the flow velocity when the flow is branched. Thereby, the flow velocity (dynamic pressure) of the refrigerant flowing out from the ejector 13 at the branching portion 42 is maintained. Of course, the branch portion 42 is not limited to this, and may be formed in a substantially T-shape or the like.

  Next, the operation of the present embodiment in the above configuration will be described based on the Mollier diagram of FIG. When the ejector refrigeration cycle 300 according to the present embodiment is operated, the refrigerant flowing out of the diffuser unit 13c causes the refrigerant flowing into the outflow side evaporator 16 side and the refrigerant flowing into the variable throttle mechanism 18 side at the branching unit 42. Divided into flow.

  Here, in the present embodiment, the refrigerant flowing into the outflow evaporator 16 side is set by setting the refrigerant passage area on the refrigerant outlet 42b side of the branch portion 42 larger than the refrigerant passage area on the refrigerant outlet 42c side. The flow rate G1 is set to be larger than the refrigerant flow rate G2 flowing into the variable throttle mechanism 18 side.

Refrigerant flowing from the branch portion 42 to the discharge side evaporator 16 evaporates by absorbing heat from the refrigerator air circulated blown by the blower fan 16a (e ₁₆ points → f ₁₆ points). Thereby, the air in a refrigerator is cooled. Then, the refrigerant flowing out of the outflow side evaporator 16 is sucked into the first compression means 11a from the intermediate pressure port 10c of the compressor 10.

On the other hand, the refrigerant that has flowed into the variable throttle mechanism 18 from the branch portion 42 is further decompressed and expanded in an isenthalpy manner to reduce its pressure (e ₁₆ → h ₁₆ points). Variable decompressed and expanded refrigerant is in throttle mechanism 18, and flows into the suction side evaporator 19, the air blowing fan 19a is evaporated by absorbing heat from the freezer in air circulated blown (h ₁₆ points → i ₁₆ points). Thereby, the air in a freezer is cooled. Other operations are the same as those in the fifth embodiment.

  As in the fifth embodiment, the ejector refrigeration cycle 300 of the present embodiment uses the outflow side evaporator 16 for cooling in a low-temperature refrigerator and the suction-side evaporator 19 for cooling in a cryogenic freezer. Can be used.

  Further, as in the fifth embodiment, both the power reduction effect of the compressor 10 due to the boosting action of the ejector 13 and the effect of improving the compression efficiency similar to the economizer refrigeration cycle can be sufficiently obtained. COP of the refrigeration cycle can be effectively improved.

  Further, since the refrigerant flow rate G1 flowing from the branching portion 42 to the outflow side evaporator 16 side is larger than the refrigerant flow rate G2 flowing from the branching portion 42 to the variable throttle mechanism 18 side, more refrigerant can be supplied. Heat can be dissipated by the radiator 12. Thereby, the endothermic amount of the refrigerant, that is, the refrigerating capacity of the cycle can be expanded as a whole cycle, and the COP can be further improved.

(Ninth embodiment)
An example in which the ejector refrigeration cycle 400 of the present invention is applied to a refrigeration / refrigeration apparatus similar to that of the sixth embodiment will be described with reference to FIGS. FIG. 17 is an overall configuration diagram of the ejector refrigeration cycle 400 of the present embodiment.

  In the ejector refrigeration cycle 400, as shown in FIG. 17, a first branching portion 32 that branches the flow of the refrigerant flowing out of the radiator 12 is connected to the outlet side of the radiator 12. The basic configuration of the first branch portion 32 is the same as that of the branch portion 32 of the sixth embodiment.

  The nozzle portion 13a of the ejector 13 is connected to one refrigerant outlet of the first branch portion 32, and the first variable throttle mechanism 18 side as the first suction side pressure reducing means is connected to the other refrigerant outlet. Yes.

  Further, the refrigerant inlet 42 a of the second branch portion 42 is connected to the outlet side of the diffuser portion 13 c of the ejector 13. The basic configuration of the second branch portion 42 is the same as that of the branch portion 42 of the eighth embodiment. The outflow side evaporator 16 is connected to one refrigerant outlet 32b of the second branch portion 42, and the second variable throttle mechanism 28 as the second suction side pressure reducing means is connected to the other refrigerant outlet 42c. Yes.

  The basic configuration of the first and second variable aperture mechanisms 18 and 28 is the same as that of the variable aperture mechanism 18 of the first embodiment. Furthermore, the first and second variable throttle mechanisms 18 and 28 of the present embodiment can fully close their throttle openings.

  Therefore, when the first variable throttle mechanism 18 fully closes the throttle passage, the flow of the refrigerant that has flowed out of the radiator 12 is not branched by the first branch portion 32, and its total flow rate flows to the ejector 13 side. To do. Further, when the second variable throttle mechanism 28 fully closes the throttle passage, the flow of the refrigerant flowing out from the diffuser portion 13c is not branched by the second branch portion 42, but the total flow rate is reduced to the outflow side evaporator 16. To the side.

  The outlet side of the first and second variable throttle mechanisms 18, 28 is connected to a merging portion 20 that joins the refrigerant flows that have flowed out of the first and second variable throttle mechanisms 18, 28. The basic configuration of the merging unit 20 is the same as that of the second branching unit 42. That is, in the junction part 20, two out of the three inflow / outflow ports 20a to 20c are the refrigerant inflow ports 20b and 20c, and one is the refrigerant outflow port 20a.

  More specifically, in the merging portion 20 of the present embodiment, the flow direction of the refrigerant flowing from the first variable throttle mechanism 18 into the one refrigerant inlet 20b, and the other refrigerant inlet from the second variable throttle mechanism 28. The flow direction of the refrigerant flowing into 20c is formed in a substantially Y shape so as to face the target direction and intersect at an acute angle with respect to the flow direction of the refrigerant flowing out from the refrigerant outlet 20a to the suction side evaporator 19. .

  Therefore, the refrigerant that has flowed into the merging portion 20 flows out of the merging portion 20 without unnecessarily reducing the flow velocity when the flows are merged. Thereby, the flow velocity (dynamic pressure) of the refrigerant that has flowed out of the first and second electric expansion valves 19 and 29 is maintained in the merging portion 20. A suction-side evaporator 19 is connected to the refrigerant outlet 20 a of the junction 20. Other configurations are the same as those of the sixth embodiment.

  Next, the operation of the present embodiment in the above configuration will be described based on the Mollier diagram of FIG. Here, in the ejector-type refrigeration cycle 400 of the present embodiment, the control device controls the first and second variable throttle mechanisms 18 and 28 to the throttle state or the fully closed state, so that the following three types of cycle configurations are configured. Can be realized.

  When the control device sets the first variable throttling mechanism 18 to the fully closed state and the second variable throttling mechanism 28 to the throttling state, it is possible to realize a cycle configuration in which the refrigerant flow is branched only by the second branch portion 42 (hereinafter, referred to as “the second variable throttling mechanism”). The operation mode in this cycle configuration is called the low pressure branch operation mode). The operation in the low-pressure branch operation mode is the same as that in the eighth embodiment.

  When the control device places the first variable throttling mechanism 18 in the throttling state and the second variable throttling mechanism 28 in the fully closed state, it is possible to realize a cycle configuration in which the refrigerant flow is branched only by the first branch portion 32 (hereinafter, referred to as “the first variable throttling mechanism 28”). The operation mode in this cycle configuration is called a high-pressure branch operation mode). The operation in the high-pressure branch operation mode is the same as that in the sixth embodiment.

  When the control device places both the first and second electric expansion valves 19 and 29 in the throttled state, it is possible to realize a cycle configuration in which the refrigerant flow is branched at the same time in the first branch portion 32 and the second branch portion 42 (hereinafter referred to as the following) The operation mode in this cycle configuration is called the simultaneous branch operation mode). The operation of this simultaneous branch operation mode will be described later.

  Moreover, each said operation mode is switched based on the refrigerating capacity or external temperature required for a cycle. In this embodiment, at normal load where normal refrigeration capacity is required, the mode is switched to the low pressure branch operation mode, which requires higher refrigeration capacity than at normal load, and the flow rate of refrigerant circulating in the cycle is higher than at normal load. When the load is high, switch to the high-pressure branch operation mode.

  In addition, the refrigeration capacity is not required more than during normal load, and the flow rate of refrigerant circulating in the cycle is lower than during normal load, or when the outside air temperature is lower than a predetermined reference temperature, the cycle When the difference between the high pressure and the low pressure becomes smaller than a predetermined pressure difference, the simultaneous branching operation mode is switched.

Here, the operation in the simultaneous branching mode will be described based on the Mollier diagram of FIG. Simultaneous branch operation mode, the flow of the refrigerant flowing out of the radiator 12, as in the sixth embodiment, in the first branch portion 32 (b ₁₈ points in FIG. 18), flows into the nozzle portion 13a side of the ejector 13 And a refrigerant flow flowing into the suction side evaporator 19 side.

The refrigerant flowing out from the first branch portion 32 toward the nozzle portion 13a flows in the order of the nozzle portion 13a → the diffuser portion 13c (b ₁₈ point → c ₁₈ point → d ₁₈ point → e ₁₈ point). On the other hand, the refrigerant that has flowed out from the first branch portion 32 to the suction-side evaporator 19 side is further decompressed and expanded isothermally by the variable throttle mechanism 18 to reduce its pressure (b ₁₈ point → kα ₁₈ point). .

  The refrigerant that has flowed out of the diffuser portion 13c is divided into a refrigerant flow that flows into the outflow-side evaporator 16 side and a refrigerant flow that flows into the second variable throttle mechanism 28 side at the second branch portion.

The refrigerant flowing into the outflow side evaporator 16 from the second branch part 42 absorbs heat from the air in the refrigerator circulated by the blower fan 16a and evaporates (e ₁₈ point → f ₁₈ point). Thereby, the air in a refrigerator is cooled. Then, the refrigerant flowing out of the outflow side evaporator 16 is sucked into the first compression means 11a from the intermediate pressure port 10c of the compressor 10.

On the other hand, the refrigerant that has flowed into the second variable throttle mechanism 18 from the second branch portion 42 is further decompressed and expanded in an enthalpy manner to reduce its pressure (e ₁₈ point → hβ ₁₈ point). The refrigerant decompressed and expanded by the second variable throttle mechanism 28 merges with the refrigerant decompressed by the first variable throttle mechanism 18 at the junction 20 (hβ ₁₈ point → hγ ₁₈ point and hα ₁₈ point). → hγ ₁₈ points).

The refrigerant that has flowed out of the merging portion 20 flows into the suction-side evaporator 19, absorbs heat from the freezer air circulated by the blower fan 19a, and evaporates (hγ ₁₈ points → i ₁₈ points). Thereby, the air in a freezer is cooled. Other operations are the same as those in the sixth and eighth embodiments.

  Since the ejector refrigeration cycle 400 of this embodiment operates as described above, the same effects as those of the eighth embodiment can be obtained in the low-pressure branch mode. In the high-pressure branch mode, the same effect as in the sixth embodiment can be obtained.

  Further, also in the simultaneous branching mode, the outflow side evaporator 16 can be used for cooling in the low-temperature refrigerator, and the suction side evaporator 19 can be used for cooling in the cryogenic freezer, as in the first embodiment. In addition, it is possible to sufficiently obtain both the power reduction effect of the compressor 10 due to the boosting action of the ejector 13 and the compression efficiency improvement effect similar to that of the economizer refrigeration cycle, and the COP of the ejector refrigeration cycle can be effectively achieved. Can be improved.

  Further, in the simultaneous branching operation mode, it is possible to realize a cycle configuration in which the refrigerant flowing out from both the first variable throttle mechanism 18 and the second variable throttle mechanism 28 is supplied to the suction side evaporator 19. Accordingly, the refrigerant supplied to the suction-side evaporator 19 with respect to the cycle configuration in which the refrigerant flowing out from one of the first variable throttle mechanism 18 and the second variable throttle mechanism 28 is supplied to the suction-side evaporator 19. It becomes easy to increase the flow rate.

(10th Embodiment)
In the present embodiment, as shown in the overall configuration diagram of FIG. 19, an auxiliary radiator that further dissipates the refrigerant that has flowed out from the branch portion 32 to the variable throttle mechanism 18 side with respect to the ejector refrigeration cycle 200 of the sixth embodiment. An example in which 12e is provided will be described.

  The auxiliary radiator 12 is a heat-dissipating heat exchanger that further heats and cools the high-pressure refrigerant by exchanging heat between the high-pressure refrigerant flowing out of the radiator 12 and the outside air (outside air) blown by the cooling fan 12a. It is. Moreover, the heat radiator 12 of this embodiment is reducing the heat exchange capability by reducing a heat exchange area with respect to 6th Embodiment.

  In FIG. 19, the cooling fan 12a is disposed in the vicinity of the radiator 12 for clarity of illustration, but the cooling fan 12a also blows outside air to the auxiliary radiator 12e. Of course, the outside air may be blown to the radiator 12 and the auxiliary radiator 12e from independent blowing fans. Other configurations are the same as those of the sixth embodiment.

Next, the operation of the present embodiment will be described based on the Mollier diagram of FIG. When the ejector refrigeration cycle 200 of the present embodiment is operated, the refrigerant discharged from the first compression means 11a of the compressor 10 dissipates heat and condenses in the radiator 12 as in the first embodiment, and the gas is condensed. It becomes a liquid two-phase state (a ₂₀ points → b ₂₀ points). This is because the heat exchange capability of the radiator 12 is reduced compared to the sixth embodiment described above.

  The high-pressure refrigerant that has flowed out of the radiator 12 flows into the branch portion 32 and is divided into a refrigerant flow that flows into the nozzle portion 13 a side of the ejector 13 and a refrigerant flow that flows into the suction side evaporator 19 side of the ejector 13.

As in the sixth embodiment, the refrigerant flowing from the branch portion 32 to the nozzle portion 13 side of the ejector 13 flows in the order of the nozzle portion 13a → the diffuser portion 13c → the outflow side evaporator 16 of the ejector 13 and is sucked from the pressure port 10c to the first compression means 11a (b ₂₀ points → c ₂₀ points → d ₂₀ points → e ₂₀ points → f ₂₀ points).

On the other hand, the high-pressure refrigerant that has flowed out from the branch portion 32 to the suction-side evaporator 19 side is further cooled by the auxiliary radiator 12e and becomes a liquid phase state (b ₂₀ points → b ′ ₂₀ points). The refrigerant that has flowed out of the auxiliary radiator 12e flows in the order of the variable throttle mechanism 18 → the suction side evaporator 19 and flows into the suction side branch portion 22 (b ′ ₂₀ points → h ₂₀ points → i ₂₀ points). Other operations are the same as in the sixth embodiment.

  Therefore, when the ejector refrigeration cycle 200 of the present embodiment is operated, the same effect as that of the sixth embodiment can be obtained. Furthermore, since the heat exchange capability of the radiator 12 is reduced, the enthalpy of the refrigerant flowing into the nozzle portion 13a is not unnecessarily reduced. Therefore, the COP improvement effect by not reducing the enthalpy of the refrigerant | coolant which flows in into the nozzle part 13a similarly to 7th Embodiment can be acquired.

(Eleventh embodiment)
In the present embodiment, an example in which an internal heat exchanger 31 is added to the ejector refrigeration cycle 200 of the fifth embodiment will be described as shown in the overall configuration diagram of FIG. The internal heat exchanger 31 includes a refrigerant in a decompression / expansion process passing through a fixed throttle 18a serving as a suction-side decompression unit constituting the high-pressure side refrigerant flow path, and a second compression of the compressor 10 passing through the low-pressure side refrigerant flow path 31b. Heat is exchanged with the refrigerant sucked by the means 21a.

  As a specific configuration of the internal heat exchanger 31, a double-tube heat exchanger configuration in which a fixed throttle 18 made of a capillary tube or the like is disposed inside an outer tube forming the low-pressure side refrigerant flow path 31b. Adopted. Of course, a configuration in which heat is exchanged by brazing the fixed throttle 18a and the refrigerant pipe forming the low-pressure side refrigerant flow path 31b may be employed.

  Further, when the variable throttle mechanism 18 is employed as the suction side pressure reducing means, the heat exchange is performed by brazing the refrigerant pipe forming the low pressure side refrigerant flow path 31b to the outer peripheral surface of the throttle passage of the variable throttle mechanism 18. Can be adopted. Other configurations are the same as those of the sixth embodiment.

Next, the operation of the present embodiment will be described based on the Mollier diagram of FIG. When the ejector-type refrigeration cycle 200 of this embodiment is operated, the enthalpy of the refrigerant flowing into the suction side branch portion 22 is increased by the action of the internal heat exchanger 31 with respect to the fifth embodiment (i ₂₂ points → i ′ ₂₂ points), the enthalpy of the decompression expansion process in the fixed throttle 18a is reduced (b ₂₂ points → h ₂₂ points).

  In other words, the refrigerant passing through the fixed throttle 18a is cooled to the temperature of the refrigerant flowing out of the suction side evaporator 19 while expanding under reduced pressure, and the enthalpy is reduced. Thereby, compared with 6th Embodiment, the enthalpy of the refrigerant | coolant which flows in into the outflow side evaporator 16 and the suction side evaporator 19 can be decreased. Other operations are the same as in the sixth embodiment.

  Therefore, when the ejector refrigeration cycle 200 of the present embodiment is operated, the same effect as that of the sixth embodiment can be obtained. Furthermore, the enthalpy of the refrigerant flowing into the suction side evaporator 19 can be reduced by the action of the internal heat exchanger 31 with respect to the sixth embodiment. Thereby, the enthalpy difference between the enthalpy of the inlet-side refrigerant and the enthalpy of the outlet-side refrigerant of the suction-side evaporator 19 is increased to increase the refrigeration capacity, and COP can be further improved.

(Twelfth embodiment)
An example in which the ejector refrigeration cycle 500 of the present invention is applied to the same refrigeration / refrigeration apparatus as in the sixth embodiment will be described with reference to FIGS. FIG. 23 is an overall configuration diagram of an ejector refrigeration cycle 500 of the present embodiment.

  In the ejector type refrigeration cycle 500, as shown in FIG. 23, a branch portion 32 similar to that of the sixth embodiment is connected to the discharge port side of the compressor 10. The radiator 12 is connected to one refrigerant outlet of the branch portion 32, and the branch refrigerant radiator 12d is connected to the other refrigerant outlet.

  The branching refrigerant radiator 12d performs heat exchange between the high-pressure refrigerant flowing out from the other refrigerant outlet of the branching part 22 and the outside air (outside air) blown by the cooling fan 12f, and dissipates the high-pressure refrigerant for cooling. It is a heat exchanger for heat dissipation.

  Further, in the present embodiment, the heat exchange area (heat radiation performance) of the first radiator 12 is reduced by reducing the heat exchange area of the first radiator 12 with respect to the branch refrigerant radiator 12d. It is lower than the heat exchange capacity. The cooling fan 12f is an electric blower whose number of rotations (amount of blown air) is controlled by a control voltage output from the control device.

  Further, the nozzle portion 13a of the ejector 13 is connected to the outlet side of the first radiator 12, and the variable throttle mechanism 18 is connected to the outlet side of the branch refrigerant radiator 12d. Other configurations are the same as those of the sixth embodiment.

Next, the operation of the present embodiment will be described based on the Mollier diagram of FIG. Operating the refrigerant cycle 500 of the present embodiment, the refrigerant discharged from the first compression means 11a of the compressor 10 (a ₂₄ point in FIG. 24), it flows into the branch portion 32, the first radiator 12 The refrigerant flow is divided into a refrigerant flow flowing into the side and a refrigerant flow flowing into the branch refrigerant radiator 12d side.

  Here, in the present embodiment, the flow rate ratio Gr1 / Gr2 between the refrigerant flow rate Gr1 flowing into the first radiator 12 side and the refrigerant flow rate Gr2 flowing into the branch refrigerant radiator 12d side can exhibit a high COP as a whole cycle. The passage areas (pressure loss characteristics) of the respective refrigerant passages in the branch portion 32 are determined so as to obtain the optimum flow rate ratio.

The refrigerant flowing into the first radiator 12 side exchanges heat with the blown air (outside air) blown from the cooling fan 12a to dissipate heat and condense (a ₂₄ points → b1 ₂₄ points). On the other hand, the refrigerant flowing into the branching refrigerant radiator 12d side exchanges heat with the blown air (outside air) blown from the cooling fan 12f to dissipate heat and condense (a ₂₄ points → b 2 ₂₄ points).

  At this time, since the heat exchange capacity of the first radiator 12 is set lower than the heat exchange capacity of the branch refrigerant radiator 12d, the enthalpy of the refrigerant flowing out from the first radiator 12 is the branch refrigerant radiator 12d. It becomes larger than the enthalpy of the refrigerant that has flowed out.

The refrigerant that has flowed out of the first radiator 12 flows in the order of the nozzle portion 13a of the ejector 13 → the diffuser portion 13c → the outflow side evaporator 16 in the order of the first pressure from the intermediate pressure port 10c of the compressor 10 as in the sixth embodiment. It is sucked into the compression means 11a (b1 ₂₄ points → c ₂₄ points → d ₂₄ points → e ₂₄ points → f ₂₄ points).

On the other hand, the refrigerant flowing out from the branch refrigerant radiator 12d flows in the order of the variable throttle mechanism 18 → the suction side evaporator 19 and flows into the suction side branch portion 22 (b2 ₂₄ points → h ₂₄ points → i ₂₄ points). Other operations are the same as in the sixth embodiment.

  Therefore, when the ejector refrigeration cycle 200 of the present embodiment is operated, the same effect as that of the sixth embodiment can be obtained.

  Furthermore, since the heat dissipation performance of the first radiator 12 and the branch refrigerant radiator 12d can be changed independently, for example, the heat dissipation performance of the branch refrigerant radiator 12d and the heat absorption performance of the suction side evaporator 19 are easily adapted. And the heat dissipation performance of the first and second radiators 12 and 12d and the heat absorption performance of the outflow evaporator 16 can be easily matched. Therefore, it is easy to stabilize the operation of the cycle.

  In addition, since the heat exchange capacity of the first radiator 12 is lower than the heat exchange capacity of the branch refrigerant radiator 12d, the enthalpy of the refrigerant flowing into the nozzle portion 13a of the ejector 13 is unnecessarily reduced. You can avoid that. Therefore, the COP improvement effect by not reducing the enthalpy of the refrigerant | coolant which flows in into the nozzle part 13a similarly to 7th Embodiment can be acquired.

(13th Embodiment)
In the present embodiment, an example in which a bypass passage 25 and an on-off valve 26 are added to the ejector refrigeration cycle 100 of the first embodiment will be described as shown in the overall configuration diagram of FIG.

  The bypass passage 25 is a refrigerant flow path that guides the high-pressure refrigerant discharged from the first compression means 11 a of the compressor 10 to the suction-side evaporator 19 by bypassing the radiator 12. Specifically, the bypass passage 25 is constituted by a refrigerant pipe that connects between the compressor 10 and the radiator 12 and between the variable throttle mechanism 18 and the suction side evaporator 19.

  The opening / closing valve 26 is an opening / closing means for opening / closing the bypass passage 25 and is an electromagnetic valve whose opening / closing operation is controlled by a control signal output from the control device. Furthermore, the refrigerant passage area when the on-off valve 26 is opened is smaller than the refrigerant passage area of the bypass passage 25. Accordingly, the refrigerant flowing through the bypass passage 25 is decompressed when passing through the on-off valve 26.

  The reason why the on-off valve 26 having a pressure reducing function is adopted as the on / off valve 26 is that not only the pressure difference between the pressure of the inlet side refrigerant and the pressure of the outlet side refrigerant is secured, but also the compressor This is because if the high-pressure refrigerant discharged from 10 is directly flowed into the suction-side evaporator 19, there is a concern that the refrigerant pressure in the suction-side evaporator 19 may exceed the pressure resistance of the suction-side evaporator 19.

  Therefore, in the present embodiment, the refrigerant passage area of the on-off valve 26 is formed small, and the pressure of the refrigerant flowing into the suction side evaporator 19 is lowered until it becomes lower than the pressure resistance capability of the suction side evaporator 19. .

  Therefore, when the on-off valve 26 having no pressure reducing function is arranged in the bypass passage 25, it is desirable to arrange the bypass passage side pressure reducing means in the bypass passage 25. As the bypass passage side pressure reducing means, a fixed throttle mechanism composed of a capillary tube, an orifice or the like can be adopted.

  Next, the operation of this embodiment in the above configuration will be described. In the ejector refrigeration cycle 100 of the present embodiment, it is possible to switch between a normal operation mode in which the interior is cooled and a defrost operation mode in which the suction side evaporator 19 is defrosted. In the normal operation mode, the control device closes the on-off valve 26. As a result, the operation is exactly the same as in the first embodiment.

  In the defrosting operation mode, the control device stops the operation of the cooling fan 12a, makes the variable throttle mechanism 18 fully closed, and opens the on-off valve 26. Thereby, the high-temperature refrigerant discharged from the first compression means 11 a of the compressor 11 flows into the bypass passage 25.

  At this time, in this embodiment, the compressor 10 → the bypass passage 25 → the suction side evaporation with respect to the pressure loss of the refrigerant circuit circulating in the order of the compressor 10 → the radiator 12 → the ejector 13 → the accumulator 17 → the compressor 10. Since the pressure loss of the refrigerant circuit circulating in the order of the container 19 → the ejector 13 → the accumulator 17 → the compressor 10 is set small, most of the refrigerant discharged from the compressor 10 flows into the bypass passage 25.

  Of course, a three-way valve is arranged at the inlet side connection part or the outlet side connection part of the bypass passage 25, and in the normal operation mode, the refrigerant discharged from the compressor 10 is allowed to flow out only to the radiator 12 side, and the defrosting operation is performed. In the mode, the refrigerant flow path may be switched so that the refrigerant discharged from the compressor 10 flows out only to the bypass passage 25 side.

  In addition, a normal auxiliary opening / closing valve not having a pressure reducing function is arranged in the refrigerant passage extending from the inlet side connection portion of the bypass passage 25 to the inlet side of the radiator 12, and the auxiliary opening / closing valve is opened in the normal operation mode. In the defrosting operation mode, the refrigerant flow path may be switched by closing the auxiliary on-off valve.

When the high-temperature and high-pressure refrigerant flowing into the bypass passage 25 passes through the on-off valve 26, it is decompressed and expanded in an enthalpy manner (o ₂₆ point → p ₂₆ point). The high-temperature and low-pressure gas-phase refrigerant that has passed through the on-off valve 26 flows into the suction-side evaporator 19 without flowing into the accumulator 17 side because the throttle opening of the variable throttle mechanism 18 is fully closed.

Refrigerant flowing into the suction side evaporator 19, to radiate the heat to the suction side evaporator 19 (p ₂₆ points → Q ₂₆ points). As a result, the suction side evaporator 19 is defrosted. The refrigerant radiated by the suction side evaporator 19 is depressurized by the pressure loss when passing through the refrigerant suction port 13b → the diffuser portion 13c → the accumulator 17 of the ejector (q ₂₆ points → r ₂₆ points).

The gas-phase refrigerant flowing out from the accumulator 17 is sucked into the compressor 10 and compressed again (r ₂₆ point → o ₂₆ point).

  As described above, in the normal operation mode of this embodiment, the same effect as that of the first embodiment can be obtained. In the defrosting operation mode, the suction side evaporator 19 can be defrosted.

  In this embodiment, the variable throttle mechanism 18 is employed as the suction side pressure reducing means, and the throttle opening of the variable throttle mechanism 18 is fully closed during the defrosting operation. Of course, a fixed throttle is used as the suction side pressure reducing means. And a check valve that is disposed between the outlet side of the suction side pressure reducing means and the connection portion of the bypass passage 25 and allows only the refrigerant to flow from the suction side pressure reducing means side to the suction side evaporator 19 side. May be provided.

(14th Embodiment)
Next, a thirteenth embodiment of the present invention will be described with reference to FIG. In the present embodiment, the ejector refrigeration cycle of the present invention is applied to a cold storage. FIG. 27 is an overall configuration diagram of an ejector refrigeration cycle 600 of the present embodiment.

  The ejector refrigeration cycle 600 is configured to be switchable between a cooling operation mode for cooling the internal air that is a heat exchange target fluid and a heating operation mode for heating the internal air. In addition, the solid line arrow in FIG. 5 shows the flow of the refrigerant in the cooling operation mode, and the broken line arrow shows the flow of the refrigerant in the heating operation mode.

  In the ejector refrigeration cycle 600 of the present embodiment, the first electric four-way valve 51 is connected to the discharge port (discharge port 10d) of the first compression means 11a of the compressor 10. The first electric four-way valve 51 is a refrigerant flow switching means whose operation is controlled by a control signal output from the control device.

  Specifically, the first electric four-way valve 51 is provided between the discharge port 10d side of the compressor 10 and the outdoor heat exchanger 53 inlet side, and between the variable throttle mechanism 18 outlet side and the use side heat exchanger 54. Between the refrigerant flow path (circuit indicated by a solid line arrow in FIG. 27), the compressor 10 discharge port 10d side and the use side heat exchanger 54 inlet side, and the variable throttle mechanism 18 outlet side and outdoor heat exchanger. The refrigerant flow path (circuit indicated by the broken line arrow in FIG. 5) that connects the 53 inlet side at the same time is switched.

  As in the refrigerant flow path indicated by the solid arrow in FIG. 27, the outdoor heat exchanger 53 is connected to the discharge port 10d of the compressor 10 in the cooling operation mode via the first electric four-way valve 51. The outdoor heat exchanger 53 is a heat exchanger that exchanges heat between the refrigerant passing through the inside and the outdoor air blown by the blower fan 53a. The blower fan 53a is an electric blower in which the rotation speed (the amount of blown air) is controlled by a control voltage output from the control device.

  Further, a second electric four-way valve 52 is connected to the outlet side of the outdoor heat exchanger 53 in the cooling operation mode. The second electric four-way valve 52 is refrigerant flow switching means whose operation is controlled by a control signal output from the control device, and the basic configuration is the same as that of the first electric four-way valve 51. is there.

  Specifically, the second electric four-way valve 52 is provided between the outdoor heat exchanger 53 outlet side and the fixed throttle 14 inlet side which is the high pressure side pressure reducing means, and between the use side heat exchanger 54 outlet side and the suction side branch portion. The refrigerant flow path (the circuit indicated by the solid line arrow in FIG. 27) that connects the 22 inlet side at the same time, the outdoor heat exchanger 53 outlet side and the suction side branch portion 22 inlet side, and the use side heat exchanger 54 The refrigerant flow path (circuit indicated by the broken line arrow in FIG. 27) that connects the outlet side and the fixed throttle 14 inlet side at the same time is switched.

  The use-side heat exchanger 54 is a heat exchanger that exchanges heat between the refrigerant passing through the inside and the indoor blown air that is the heat exchange target fluid blown by the blower fan 54a. The blower fan 54a is an electric blower in which the rotation speed (the amount of blown air) is controlled by a control voltage output from the control device.

  The variable throttle mechanism 18 of the present embodiment is disposed between the liquid refrigerant outlet side of the accumulator 17 and the first electric four-way valve 51. Other configurations are the same as those of the first embodiment.

  Next, the operation of this embodiment in the above configuration will be described. In the ejector refrigeration cycle 600 of the present embodiment, a cooling operation mode for cooling the internal air and a heating operation mode for heating the internal air can be switched.

  The cooling operation mode is executed when the cooling operation mode is selected by the operation switch of the operation panel. In the cooling operation mode, the control device operates the first and second electric motors 11b and 21b, the flow rate adjustment valve 38, and the blower fans 53a and 54a.

  Further, the control device simultaneously connects between the discharge port 10d side of the compressor 10 and the outdoor heat exchanger 53 inlet side, and between the variable throttle mechanism 18 outlet side and the use side heat exchanger 54 inlet side. The first electric four-way valve 51 is switched to simultaneously connect the outlet side of the outdoor heat exchanger 53 and the inlet side of the fixed throttle 14 and the outlet side of the use side heat exchanger 54 and the inlet side of the suction side branch portion 22 at the same time. Thus, the second electric four-way valve 52 is switched.

  Thereby, as shown by the solid line arrow in FIG. 27, the discharge port 10d of the compressor 10 (→ first electric four-way valve 51) → outdoor heat exchanger 53 (→ second electric four-way valve 52) → fixed throttle 14 The refrigerant circulates in the order of the nozzle portion 13a, the gas-phase refrigerant outlet of the accumulator 17, the intermediate pressure port 10c of the compressor 10, and the liquid-phase refrigerant outlet of the accumulator 17. The variable throttle mechanism 18 (→ first electric four-way valve 51 ) → use side heat exchanger 54 (→ second electric four-way valve 52) → suction side branch 22 arrow refrigerant suction port 13b → accumulator 17 in order of refrigerant circulation, suction side branch 22 → flow rate adjustment valve 38 → The suction port 10b of the compressor 10 is connected.

  Therefore, the first compression means 11a of the compressor 10 compresses the intermediate pressure refrigerant obtained by mixing the refrigerant discharged from the second compression means 21a and the gas-phase refrigerant flowing out of the ejector 13 and separated by the accumulator 17. It will be discharged.

  Further, the refrigerant discharged from the first compression means 11a radiates heat in the outdoor heat exchanger 53, and the refrigerant radiated in the outdoor heat exchanger 53 flows into the fixed throttle 14 which is a high pressure side pressure reducing means. Further, the liquid-phase refrigerant separated by the accumulator 17 flows into the use side heat exchanger 54, and the refrigerant evaporated in the use side heat exchanger 54 is sucked from the refrigerant suction port 13b.

  As a result, in the cooling operation mode of the present embodiment, the outdoor heat exchanger 53 corresponds to the radiator 12 in the first embodiment, and the use side heat exchanger 54 corresponds to the suction side evaporator 19 in the first embodiment. It becomes a structure, can operate | move similarly to 1st Embodiment, can cool the air in a store | warehouse | chamber, and can acquire the effect similar to 1st Embodiment.

  On the other hand, the heating operation mode is executed when the heating operation mode is selected by the operation switch of the operation panel. In the heating operation mode, the control device operates the first and second electric motors 11b and 21b, the flow rate adjustment valve 38, and the blower fans 53a and 54a.

  Further, the control device simultaneously connects between the discharge port 10d side of the compressor 10 and the use side heat exchanger 54 inlet side, and between the variable throttle mechanism 18 outlet side and the outdoor heat exchanger 53 inlet side. The first electric four-way valve 51 is switched to simultaneously connect between the outdoor heat exchanger 53 outlet side and the suction side branching section 22 inlet side and between the use side heat exchanger 54 outlet side and the fixed throttle 14 inlet side. Thus, the second electric four-way valve 52 is switched.

  Thereby, as shown by the broken line arrow in FIG. 27, the discharge port 10d (→ first electric four-way valve 51) of the compressor 10 → the use side heat exchanger 54 (→ second electric four-way valve 52) → fixed throttle The refrigerant circulates in the order of 14 → nozzle portion 13a → gas phase refrigerant outlet of accumulator 17 → intermediate pressure port 10c of compressor 10, and liquid phase refrigerant outlet of accumulator 17 → variable throttle mechanism 18 (→ first electric four-way valve) 51) → outdoor heat exchanger 53 (→ second electric four-way valve 52) → suction side branch 22 → refrigerant suction port 13b → accumulator 17 in order of refrigerant circulation, suction side branch 22 → flow rate adjusting valve 38 → The suction port 10b of the compressor 10 is connected.

  Therefore, the first compression means 11a of the compressor 10 compresses the intermediate pressure refrigerant obtained by mixing the refrigerant discharged from the second compression means 21a and the gas-phase refrigerant flowing out of the ejector 13 and separated by the accumulator 17. It will be discharged.

  Further, the refrigerant discharged from the first compression means 11a radiates heat in the use side heat exchanger 54, and the refrigerant radiated in the use side heat exchanger 54 flows into the fixed throttle 14 which is a high pressure side pressure reducing means. Further, the liquid-phase refrigerant separated by the accumulator 17 flows into the outdoor heat exchanger 53, and the refrigerant evaporated in the outdoor heat exchanger 53 is sucked from the refrigerant suction port 13b.

  Therefore, in the cooling operation mode of the present embodiment, the outdoor heat exchanger 53 corresponds to the suction side evaporator 19 in the first embodiment, and the use side heat exchanger 54 corresponds to the radiator 12 in the first embodiment. Thus, the internal air can be heated by operating in the same manner as in the first embodiment, and the same effect as in the first embodiment can be obtained.

  As a result, when the ejector refrigeration cycle 600 of the present embodiment is operated, the compressor 10 by the boosting action of the ejector 13 is the same as in the first embodiment in any of the cooling operation mode and the heating operation mode. Both the power reduction effect and the compression efficiency improvement effect similar to the economizer refrigeration cycle can be sufficiently obtained, and the COP of the ejector refrigeration cycle can be effectively improved.

(Other embodiments)
The present invention is not limited to the above-described embodiment, and can be variously modified as follows.

  (1) In each of the above-described embodiments, the first and second compression units 11a and 21a are driven by different first and second electric motors 11b and 21b, respectively. The compression shafts of the means 11a and 21a may be shared, and both compression means may be driven by a driving force supplied from a common drive source.

  Further, in each of the above-described embodiments, the first and second compression means 11a and 21a are configured by different compression mechanisms, but an example has been described. However, the first and second compression means 11a and 21a are configured by the same compression mechanism. You can also

  For example, a scroll-type compression mechanism can be employed in which a communication port communicating with a compression chamber in a predetermined compression process is provided, and an intermediate pressure refrigerant is introduced through the communication port. In this case, the part where the compression chamber is located upstream (low pressure side) from the communication port is the second compression means 21a, and the part located downstream (high pressure side) from the communication port is the first compression means 11a. do it.

  (2) In the above-described embodiments, the first and second compression units 11a and 21a are driven by the first and second electric motors 11b and 21b. However, for example, an engine or the like may be used as a drive source. .

  Further, as the compression means, a variable capacity compression mechanism that can adjust the refrigerant discharge capacity by changing the discharge capacity may be adopted. In this case, the discharge capacity changing means becomes the discharge capacity changing means. Moreover, you may use the fixed capacity type compression mechanism which adjusts a refrigerant | coolant discharge capability by changing connection with a drive source intermittently by intermittent connection of an electromagnetic clutch. In this case, the electromagnetic clutch becomes the discharge capacity changing means.

  Furthermore, as the first and second compression units 11a and 21a, the same type of compression mechanism may be employed, or different types of compression mechanisms may be employed.

  (3) In the above-described embodiment, the fixed ejector 13 in which the throttle passage area of the nozzle portion 13a is fixed is adopted as the ejector 13. However, the variable ejector configured to be able to change the throttle passage area of the nozzle portion. May be adopted.

  In the above-described embodiment, as the flow rate adjustment valve 38, a flow rate adjustment valve with a fully closed function that can fully close the refrigerant passage may be adopted. When the flow rate adjustment valve is fully closed, the operation similar to that of the ejector refrigeration cycle of the prior art can be realized by stopping the operation of the second compression means 21a.

  (4) In the above-described embodiment, an example in which a normal chlorofluorocarbon refrigerant is employed as the refrigerant has been described. However, the type of the refrigerant is not limited to this. For example, hydrocarbon refrigerant, carbon dioxide, etc. may be used. Furthermore, the ejector refrigeration cycle of the present invention may be configured as a supercritical refrigeration cycle in which the high-pressure side refrigerant pressure exceeds the critical pressure of the refrigerant.

  Further, when the ejector-type refrigeration cycle 100 to 600 is a supercritical refrigeration cycle, the high pressure side refrigerant pressure is set as the high pressure side decompression means, and the COP is substantially maximum based on the high pressure side refrigerant temperature on the outlet side of the radiator 12. You may employ | adopt the pressure control valve adjusted to the target high pressure determined so that it may become.

  As such a pressure control valve, specifically, there is a temperature sensing part provided on the outlet side of the radiator 12, and a pressure corresponding to the temperature of the high-pressure refrigerant on the outlet side of the radiator 12 inside the temperature sensing part. The valve opening degree can be adjusted by a mechanical mechanism based on the balance between the internal pressure of the temperature sensing portion and the refrigerant pressure on the outlet side of the radiator 12.

  (5) In the above-described embodiment, the throttle mechanism is employed as the high pressure side pressure reducing means and the low pressure side pressure reducing means, but as the high pressure side pressure reducing means and the low pressure side pressure reducing means, the refrigerant is volume expanded to reduce the pressure, You may employ | adopt the expander which converts the pressure energy of a refrigerant | coolant into mechanical energy, and outputs it.

  As such an expander, specifically, a volume type compression means such as a scroll type, a vane type, or a rolling piston type can be employed. Then, by flowing the refrigerant so that it flows backward with respect to the refrigerant flow when the positive displacement compression means is used as the compression means, mechanical energy can be output while the refrigerant is volume-expanded and depressurized.

  (6) The means described in each embodiment described above can be applied to other embodiments. For example, an accumulator 17 similar to that of the first embodiment may be provided for the sixth and eighth to thirteenth embodiments. For example, the same high pressure side decompression means as the second embodiment may be provided for the sixth to thirteenth embodiments. Furthermore, for example, the internal heat exchanger 30 similar to the second and third embodiments may be provided for the sixth, eighth to tenth, twelfth and thirteenth embodiments.

  As the internal heat exchanger 30, a counter flow type heat exchanger in which the refrigerant flow direction in the high-pressure side refrigerant flow path and the refrigerant flow direction in the low-pressure side refrigerant flow path are different from each other may be adopted. A parallel flow heat exchanger in which the refrigerant flow direction in the refrigerant flow path and the refrigerant flow direction in the low-pressure side refrigerant flow path are the same may be employed.

  Further, in the tenth and twelfth embodiments, the outflow side evaporator 16 and the blower fan 16a may be eliminated and the internal heat exchanger 30 may be employed.

  For example, in the embodiment shown in FIG. 28, the high-pressure side refrigerant that has flowed out from the auxiliary radiator 12e and the low-pressure that has flowed out from the diffuser portion 13c are eliminated from the tenth embodiment by disposing the outflow-side evaporator 16 and the blower fan 16a. This is an example in which an internal heat exchanger 30 for exchanging heat with the side refrigerant, that is, the low-pressure side refrigerant sucked into the second compression means 21a is employed.

  On the other hand, as shown in FIG. 29, from the high-pressure side refrigerant that has flowed out of the auxiliary radiator 12e and the low-pressure side refrigerant that flows through the refrigerant flow paths shown in the regions α, β, and γ, that is, from the suction-side evaporator 19. An internal heat exchanger 30 that exchanges heat with the low-pressure refrigerant that has flowed out may be employed. In this case, it is desirable to provide the accumulator 17 on the outlet side of the diffuser portion 13c.

  Further, for example, in the embodiment shown in FIG. 30, the outflow side evaporator 16 and the blower fan 16a are abolished with respect to the twelfth embodiment, and the high pressure side refrigerant and the diffuser portion 13c that have flowed out of the branch refrigerant radiator 12d. This is an example in which the internal heat exchanger 30 is used to exchange heat with the low-pressure side refrigerant that has flowed out, that is, the low-pressure side refrigerant sucked into the second compression means 21a.

  On the other hand, as shown in FIG. 31, the high-pressure side refrigerant flowing out from the branching refrigerant radiator 12d and the low-pressure side refrigerant flowing through the refrigerant flow paths shown in the regions α, β, γ, that is, the suction-side evaporator 19 You may employ | adopt the internal heat exchanger 30 which heat-exchanges with the low voltage | pressure side refrigerant | coolant which flowed out from. In this case, it is desirable to provide the accumulator 17 on the outlet side of the diffuser portion 13c.

  (7) In the sixth and eighth to twelfth embodiments described above, the example in which different cooling target spaces (refrigerator space, freezer space) are cooled by the outflow side evaporator 16 and the suction side evaporator 19 has been described. However, the same cooling target space may be cooled. In this case, it is desirable that the outflow side evaporator 16 and the suction side evaporator 19 are assembled in an integrated structure, and the air blown from the blower fan is passed in the order of the outflow side evaporator 16 → the suction side evaporator 19.

  This is because the refrigerant evaporation pressure (refrigerant evaporation temperature) of the suction side evaporator 19 is lower than the refrigerant evaporation pressure (refrigerant evaporation temperature) of the outflow side evaporator 16. That is, by passing the blown air from the blower fan as described above, the temperature difference between the refrigerant evaporation temperature of the outflow side evaporator 16 and the suction side evaporator 19 and the blown air is secured, and the blown air is efficiently flowed. Can be cooled.

  (8) In the ninth embodiment described above, the high-pressure branch operation mode is switched during high-load operation, the low-pressure branch operation mode is switched during normal operation, and the simultaneous branch operation mode is switched during low-load operation. Of course, switching of each operation mode is not limited to this.

  For example, the high-pressure branch operation mode may be switched during high-load operation, the simultaneous branch operation mode during normal operation, and the low-pressure branch operation mode during low-load operation. That is, when the ejector refrigeration cycle 400 is operated, the operation mode may be switched to an operation mode in which the highest cycle efficiency can be exhibited.

  Moreover, it is good also as a cycle structure which switches a high voltage | pressure branch operation mode and a low voltage | pressure branch operation mode, without implement | achieving simultaneous branch operation mode. In this case, the first and second branch portions 22 and 28 may be configured by three-way valves to switch the refrigerant flow path.

  (9) In the thirteenth embodiment, switching between the normal operation mode and the defrosting operation mode is performed based on the operation signal of the changeover switch on the operation panel, but switching between the normal operation mode and the defrosting operation mode is performed. It is not limited to this.

  For example, the control device may alternately switch between the normal operation mode and the defrosting operation mode every predetermined time. That is, when the normal operation mode is continued for a predetermined first reference time or longer, the mode is switched to the defrosting operation mode, and when the defrost operation mode is continued for a predetermined second reference time or longer, the normal operation is performed. You may make it switch to mode.

  Moreover, if the bypass passage 25 and the on-off valve 26 described above are applied to the sixth to twelfth embodiments in the same manner as the twelfth embodiment, the defrosting operation mode can be realized in the ejector refrigeration cycles 200 to 500. .

  (10) In the fourteenth embodiment, the example in which the first and second electric four-way valves 51 and 52 are employed as the refrigerant flow path switching means has been described. However, the refrigerant flow path switching means is not limited to this. For example, as shown in FIG. 32, instead of the electric four-way valve 51, two electric three-way valves 51a may be combined, and as shown in FIG. 33, four on-off valves (solenoid valves). You may comprise combining 51b.

  (11) In each of the above-described embodiments, the example in which the ejector refrigeration cycle 100 to 500 of the present invention is applied to a refrigerator, a freezing / refrigeration apparatus, and a cold storage container has been described. However, the application of the present invention is not limited thereto. . For example, the ejector refrigeration cycle may be applied to an air conditioner, another stationary refrigeration cycle apparatus, a vehicle air conditioner, and the like.

  (12) In each of the first to thirteenth embodiments described above, the outflow side evaporator 16 and the suction side evaporator 19 are configured as use side heat exchangers, and the radiator 12 is configured as an outdoor heat exchanger that radiates heat to the atmosphere side. However, as in the thirteenth embodiment, the outflow side evaporator 16 and the suction side heat exchanger 16 are configured as outdoor heat exchangers that absorb heat from a heat source such as the atmosphere, and the radiator 12 is made of air or water. It is good also as a heat pump cycle comprised as an indoor side heat exchanger which heats a to-be-heated refrigerant.

It is a whole block diagram of the ejector-type refrigerating cycle of 1st Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigerating cycle of 1st Embodiment. It is a whole block diagram of the ejector type refrigerating cycle of 2nd Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigerating cycle of 2nd Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of 3rd Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigerating cycle of 3rd Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of 4th Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigerating cycle of 4th Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of 5th Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigerating cycle of 5th Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of 6th Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigerating cycle of 6th Embodiment. It is a whole block diagram of the ejector type refrigerating cycle of 7th Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigerating cycle of 7th Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of 8th Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigerating cycle of 8th Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of 9th Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigerating cycle of 9th Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of 10th Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigerating cycle of 10th Embodiment. It is a whole block diagram of the ejector type refrigerating cycle of 11th Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigerating cycle of 11th Embodiment. It is a whole block diagram of the ejector type refrigerating cycle of 12th Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigerating cycle of 12th Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of 13th Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant in the defrost operation mode of the ejector-type refrigerating cycle of 13th Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of 14th Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle in other embodiment. It is a whole block diagram which shows the modification of the ejector-type refrigerating cycle in other embodiment. It is a whole block diagram of another ejector type refrigerating cycle in other embodiments. It is a whole block diagram which shows the modification of another ejector-type refrigerating cycle in other embodiment. It is a block diagram of the flow-path switching means in other embodiment. It is a block diagram of another flow-path switching means in other embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of a prior art.

Explanation of symbols

11a, 21a 1st, 2nd compression means 11b, 21b 1st, 2nd electric motor 12, 12d Radiator, branch refrigerant radiator 12e Auxiliary radiator 13 Ejector 13a Nozzle part 13b Refrigerant suction port 14 Fixed throttle 16 Outflow side evaporation Unit 17 Accumulator 18, 28 First and second variable throttle mechanisms 19 Suction side evaporator 22 Suction side branch portion 25 Bypass passage 26 On-off valve 30, 31 Internal heat exchangers 32, 42 First and second branch portions 51, 52 First, second electric four-way valve 53 Outdoor heat exchanger 54 User side heat exchanger

Claims (30)

First compression means (11a) for compressing and discharging the refrigerant;
A radiator (12) for radiating heat from the high-pressure refrigerant discharged from the first compression means (11a);
The high-pressure refrigerant flowing out from the radiator (12) is sucked from the refrigerant suction port (13b) by the flow of high-speed jet refrigerant jetted from the nozzle portion (13a) that decompresses and expands the jet refrigerant and the refrigerant. An ejector (13) that mixes and raises the suction refrigerant sucked from the suction port (13b);
A suction side evaporator (19) for evaporating the refrigerant;
A suction-side branching part (22) for branching the flow of the refrigerant flowing out from the suction-side evaporator (19) and allowing one of the branched refrigerants to flow out toward the refrigerant suction port (13b);
A flow rate adjusting means (38) for adjusting the flow rate of the other refrigerant branched at the suction side branch portion (22);
Second compression means (21a) for compressing and discharging the refrigerant whose flow rate has been adjusted by the flow rate adjustment means (38),
The first compression means (11a) compresses and discharges an intermediate pressure refrigerant obtained by mixing the refrigerant discharged from the second compression means (21a) and the refrigerant flowing out of the ejector (13). Ejector refrigeration cycle. An outflow-side gas-liquid separator (17) that separates the gas-liquid refrigerant flowing out of the ejector (13) and causes the separated gas-phase refrigerant to flow out to the suction side of the first compression means 11a;
And a suction-side decompression means (18) that decompresses and expands the liquid refrigerant separated by the outflow-side gas-liquid separator (17) and flows it out to the inlet side of the suction-side evaporator (19). The ejector type refrigeration cycle according to claim 1.   An outlet side evaporator (16) is provided between the outlet side of the ejector (13) and the suction side of the first compression means (11a) and evaporates the refrigerant flowing out of the ejector (13). The ejector-type refrigeration cycle according to claim 2. A branch part (42) for branching the flow of the refrigerant flowing out of the ejector (13);
Outflow side evaporation that is arranged between one outlet side of the branch part (42) and the suction side of the second compression means (21a) and evaporates one refrigerant branched by the branch part (42). A vessel (16),
A suction-side pressure reducing means (18) for decompressing and expanding the other refrigerant branched at the branching portion (42) and flowing it out to the inlet side of the suction-side evaporator (19). 2. The ejector refrigeration cycle according to 1.   The ejector type according to any one of claims 1 to 4, further comprising an internal heat exchanger (30) for exchanging heat between the refrigerant flowing out of the radiator (12) and the low-pressure side refrigerant of the cycle. Refrigeration cycle.   The ejector refrigeration cycle according to claim 5, wherein the low-pressure side refrigerant of the cycle is a refrigerant that has flowed out of the suction-side evaporator (19).   The ejector refrigeration cycle according to claim 5, wherein the low-pressure side refrigerant of the cycle is a refrigerant sucked into the second compression means (21a). A branch part (32) for branching the flow of the refrigerant flowing out of the radiator (12);
Suction side decompression means (18) for decompressing and expanding the refrigerant flowing into the suction side evaporator (19),
The nozzle part (13a) decompresses and expands one of the refrigerants branched at the branch part (32),
2. The ejector refrigeration cycle according to claim 1, wherein the suction side decompression unit (18) decompresses and expands the other refrigerant branched at the branch portion (32).   An outlet side evaporator (16) is provided between the outlet side of the ejector (13) and the suction side of the first compression means (11a) and evaporates the refrigerant flowing out of the ejector (13). The ejector-type refrigeration cycle according to claim 8.   The auxiliary radiator (12e) is provided, wherein the other refrigerant branched by the branch part (32) is radiated to flow out to the suction side pressure reducing means (18) inlet side. The ejector-type refrigeration cycle described in 1.   11. The internal heat exchanger (30) for exchanging heat between the other refrigerant branched at the branch portion (32) and the low-pressure side refrigerant of the cycle is provided. The ejector refrigeration cycle described.   The internal heat exchanger (31) for exchanging heat between the refrigerant in the decompression / expansion process in the suction side decompression means (18) and the low-pressure side refrigerant of the cycle is provided. The ejector refrigeration cycle described.   The ejector refrigeration cycle according to claim 11 or 12, wherein the low-pressure side refrigerant of the cycle is a refrigerant that has flowed out of the suction-side evaporator (19).   The ejector refrigeration cycle according to claim 11 or 12, wherein the low-pressure side refrigerant of the cycle is a refrigerant sucked into the second compression means (21a). A first branch portion (32) configured to be able to branch the flow of the refrigerant flowing out of the radiator (12);
A second branch portion (42) configured to be able to branch the flow of the refrigerant flowing out of the ejector (13);
One refrigerant disposed between the one outlet side of the second branch part (42) and the suction side of the second compression means (21a) evaporates one refrigerant branched by the second branch part (42). An outflow evaporator (16) for
First and second suction side decompression means (18, 28) for decompressing and expanding the refrigerant flowing into the suction side evaporator (19),
The nozzle part (13a) decompresses and expands one of the refrigerants branched at the first branch part (32),
The first suction side decompression means (18) decompresses and expands the other refrigerant branched at the first branch part (32),
The second suction side decompression means (28) decompresses and expands the other refrigerant branched at the second branch portion (42),
The said suction side evaporator (19) evaporates at least one among the refrigerant | coolants expanded and decompressed by the said 1st, 2nd suction side decompression means (18, 28), The feature of Claim 1 characterized by the above-mentioned. Ejector type refrigeration cycle.   The ejector according to claim 15, further comprising an internal heat exchanger (30) for exchanging heat between the other liquid-phase refrigerant branched at the first branch portion (32) and the low-pressure side refrigerant of the cycle. Refrigeration cycle.   The ejector-type refrigeration cycle according to claim 16, wherein the low-pressure side refrigerant in the cycle is a refrigerant that has flowed out of the suction-side evaporator (19).   The ejector refrigeration cycle according to claim 16, wherein the low-pressure side refrigerant of the cycle is a refrigerant sucked into the second compression means (21a). A branch portion (32) for branching the flow of the high-pressure refrigerant discharged from the first compression means (11a) and allowing one of the branched refrigerant to flow out toward the radiator (12);
A branch refrigerant radiator (12d) for radiating heat from the other refrigerant branched at the branch section (32);
2. A suction-side decompression means (18) that decompresses and expands the high-pressure refrigerant that has flowed out of the branching refrigerant radiator (12d) and flows it out to an inlet side of the suction-side evaporator (19). The ejector-type refrigeration cycle described in 1.   An outlet side evaporator (16) is provided between the outlet side of the ejector (13) and the suction side of the first compression means (11a) and evaporates the refrigerant flowing out of the ejector (13). The ejector-type refrigeration cycle according to claim 19.   21. The ejector refrigeration cycle according to claim 19 or 20, further comprising an internal heat exchanger (30) for exchanging heat between the refrigerant flowing out of the branched refrigerant radiator (12d) and the low-pressure side refrigerant of the cycle.   The ejector-type refrigeration cycle according to claim 21, wherein the low-pressure side refrigerant of the cycle is a refrigerant that has flowed out of the suction-side evaporator (19).   The ejector refrigeration cycle according to claim 21, wherein the low-pressure side refrigerant of the cycle is a refrigerant sucked into the second compression means (21a).   The ejector-type refrigeration cycle according to any one of claims 1 to 23, further comprising high-pressure side decompression means (14) for decompressing and expanding the refrigerant flowing out to the inlet side of the nozzle portion (13a). A bypass passage (25) for guiding the high-pressure side refrigerant of the cycle discharged from the first compression means (11a) to the suction-side evaporator (19);
The ejector-type refrigeration cycle according to any one of claims 1 to 24, further comprising opening / closing means (26) for opening and closing the bypass passage (25). First compression means (11a) for compressing and discharging the refrigerant;
A use side heat exchanger (54) for exchanging heat between the refrigerant and the fluid to be heat exchanged;
An outdoor heat exchanger (53) for exchanging heat between the refrigerant and the outside air;
A refrigerant channel switching means (51, 52) for switching between a refrigerant channel in a cooling operation mode for cooling the fluid for heat exchange and a refrigerant channel in a heating operation mode for heating the fluid for heat exchange;
By the flow of the high-speed injection refrigerant injected from the nozzle part (13a) that decompresses and expands the refrigerant radiated in one of the outdoor heat exchanger (53) and the use side heat exchanger (54) An ejector (13) that sucks the refrigerant from the refrigerant suction port (13b), mixes the injected refrigerant and the suction refrigerant sucked from the refrigerant suction port (13b), and boosts the pressure;
Of the outdoor heat exchanger (53) and the use side heat exchanger (54), the flow of the refrigerant evaporated in the other heat exchanger is branched, and one of the branched refrigerants is supplied to the refrigerant suction port (13b). ) Suction side branch (22) to flow out to the side,
A flow rate adjusting means (38) for adjusting the flow rate of the other refrigerant branched at the suction side branch portion (22);
Second compression means (21a) for compressing and discharging the refrigerant whose flow rate has been adjusted by the flow rate adjustment means (38),
The first compression means (11a) compresses and discharges an intermediate pressure refrigerant obtained by mixing the refrigerant discharged from the second compression means (21a) and the refrigerant flowing out of the ejector (13),
The refrigerant flow switching means (51, 52)
In the cooling operation mode, the refrigerant discharged from the first compression means (11a) is radiated by the outdoor heat exchanger (53), and the refrigerant evaporated by the use side heat exchanger (54) is Switch to the refrigerant flow path leading to the suction side branch (22) side,
In the heating operation mode, the refrigerant discharged from the first compression means (11a) is radiated by the use side heat exchanger (54), and the refrigerant evaporated by the outdoor heat exchanger (53) is 2. An ejector refrigeration cycle, wherein the refrigerant flow path is switched to a refrigerant flow path that leads to the suction side branch section (22) side. Furthermore, an outflow side gas-liquid separator (17) for separating the gas-liquid of the refrigerant flowing out from the ejector (13) is provided,
The gas-phase refrigerant outlet of the outflow side gas-liquid separator (17) is connected to the suction port side of the second compression means (21a),
The refrigerant flow switching means (51, 52)
In the cooling operation mode, the refrigerant radiated by the outdoor heat exchanger (53) is caused to flow into the nozzle part (13a), and the liquid phase refrigerant separated by the outflow side gas-liquid separator (17) is changed to the Switch to the refrigerant flow path to flow into the use side heat exchanger (54),
In the heating operation mode, the refrigerant radiated by the use side heat exchanger (54) is caused to flow into the nozzle part (13a), and the liquid phase refrigerant separated by the outflow side gas-liquid separator (17) is supplied. 27. The ejector refrigeration cycle according to claim 26, wherein the refrigerant flow path is switched to a refrigerant flow path flowing into the outdoor heat exchanger (53). First discharge capacity changing means (11b) for changing the refrigerant discharge capacity of the first compression means (11a);
Second discharge capacity changing means (21b) for changing the refrigerant discharge capacity of the second compression means (21a),
The first discharge capacity changing means (11b) and the second discharge capacity changing means (21b) independently change the refrigerant discharge capacity of the first compression means (11a) and the second compression means (21a). The ejector refrigeration cycle according to any one of claims 1 to 27, wherein the ejector refrigeration cycle is configured to be possible.   The said 1st compression means (11a) and the said 2nd compression means (21a) are accommodated in the same housing, and are comprised integrally, The one of Claims 1 thru | or 28 characterized by the above-mentioned. The ejector-type refrigeration cycle described in 1.   The ejector refrigeration cycle according to any one of claims 1 to 29, wherein the first compression means (11a) increases the pressure of the refrigerant until it reaches a critical pressure or more.

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