JP2008190769A - Ejector type refrigerating cycle - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide an ejector type refrigerating cycle which simultaneously achieves both of improvement effect of nozzle efficiency and refrigerating capacity expansion effect of a suction-side evaporator, even when an operating condition is changed. <P>SOLUTION: The distributed-air amount of a first cooling fan 12a is controlled to adjust a degree of supercooling of an outlet refrigerant of a first radiator 12 for cooling a discharged refrigerant of a compressor 11 to be 3K or less. Further the flow of the outlet refrigerant of the first radiator 12 is branched, one of the branched refrigerants is allowed to flow into a nozzle portion 14a of the ejector 14, and the other is allowed to flow into a second radiator 16. Further the distributed-air amount of a second cooling fan 16a is controlled to adjust a degree of supercooling of an outlet refrigerant of the second radiator 16 to be 10K or more, and the refrigerant is allowed to flow into a suction-side evaporator 19. Thus both of the improvement effect of nozzle efficiency of the nozzle portion 14a and the refrigerating capacity expansion effect of the suction-side evaporator 19 are simultaneously achieved even when the operating condition is changed. <P>COPYRIGHT: (C)2008,JPO&INPIT

Description

  The present invention relates to an ejector-type refrigeration cycle having an ejector.

  Conventionally, in Patent Document 1, the flow of the refrigerant is branched at a branch portion arranged on the downstream side of the radiator that dissipates the refrigerant discharged from the compressor and on the upstream side of the nozzle portion of the ejector, and one of the branched refrigerants is arranged on the nozzle portion side. An ejector refrigeration cycle is disclosed in which the other refrigerant flows into the refrigerant suction port side of the ejector.

  In the ejector refrigeration cycle of Patent Document 1, a downstream evaporator is disposed downstream of the diffuser portion of the ejector, and a throttle mechanism for decompressing and expanding the refrigerant between the branch portion and the refrigerant suction port of the ejector and suction side evaporation. An evaporator is arranged so that the refrigerant can exert an endothermic effect in both evaporators.

Further, the outlet side of the downstream evaporator is connected to the suction side of the compressor, and the refrigerant boosted by the boosting action of the diffuser section is sucked into the compressor, thereby reducing the driving power of the compressor and reducing the cycle efficiency ( COP) is being improved.
JP 2005-308380 A

  By the way, in this type of ejector, by allowing a saturated liquid phase refrigerant or a gas-liquid two-phase refrigerant to flow into the nozzle portion, boiling of the refrigerant in the nozzle portion can be promoted, and the nozzle efficiency can be improved. The nozzle efficiency is energy conversion efficiency when converting the pressure energy of the refrigerant into kinetic energy.

  And by improving this nozzle efficiency, since the flow velocity of the refrigerant | coolant flow injected from a nozzle part can be increased and the recovery energy amount of an ejector can be increased, cycle efficiency can be improved further.

  Therefore, the present inventors have previously disclosed an ejector type in which a supercooler integrated condenser is employed as a radiator in the cycle of Patent Document 1 in Japanese Patent Application No. 2006-219770 (hereinafter referred to as the prior application example). A refrigeration cycle is proposed.

  Note that the subcooler-integrated condenser is a condenser that radiates and condenses the refrigerant discharged from the compressor, a receiver that separates the gas-liquid of the refrigerant on the downstream side of the condenser and stores the liquid-phase refrigerant, and a receiver. It is what is called a subcool type condenser comprised including the supercooling part which further thermally radiates the liquid phase refrigerant | coolant isolate | separated by (4).

  Furthermore, in the cycle of the prior application example, a branch portion is formed at the liquid-phase refrigerant outlet portion of the receiver portion, one branched saturated liquid-phase refrigerant is allowed to flow into the nozzle portion of the ejector, and the other saturated liquid-phase refrigerant is passed through. It flows into the cooling part.

  As a result, the nozzle efficiency is improved, and the supercooled refrigerant whose enthalpy has been reduced in the supercooling section is caused to flow into the suction side evaporator via the decompression means, thereby expanding the refrigeration capacity of the suction side evaporator. ing. That is, in the cycle of the prior application example, the cycle efficiency is improved with respect to the cycle of Patent Document 1 by the effect of improving the nozzle efficiency and the effect of expanding the refrigerating capacity of the suction side evaporator.

  However, when the ejector refrigeration cycle of the prior application example is actually operated, the above cycle efficiency improvement effect may not be sufficiently obtained depending on the operation conditions of the cycle.

  Then, when the present inventors investigated the cause, it turned out that it is a cause that the supercooling degree of a supercooling part exit refrigerant | coolant changes with the driving | running conditions of a cycle. The reason is that when the degree of supercooling of the refrigerant at the outlet of the supercooling portion changes, the flow rate of the refrigerant passing through the decompression means changes.

  To explain this in more detail, for example, when the ejector-type refrigeration cycle of the prior application example is applied to an air conditioner, if the outdoor air temperature for heat exchange with the refrigerant in the supercooler integrated condenser increases, The amount of heat dissipated by the refrigerant in the condenser integrated condenser is reduced, and the degree of supercooling of the refrigerant at the outlet of the supercooling section is lowered.

  And when the supercooling degree of a supercooling part exit refrigerant | coolant falls, since the density of the refrigerant | coolant which flows in into a pressure reduction means falls, the refrigerant | coolant flow rate which passes a pressure reduction means falls. Furthermore, the present inventors investigated the details of the relationship between the degree of supercooling of the refrigerant at the inlet of the decompression means and the flow rate of the refrigerant passing through the decompression means, and obtained the investigation results shown in FIG.

  5 represents the degree of supercooling of the refrigerant at the inlet of the decompression means. Note that the degree of supercooling indicates the absolute value of the amount of temperature decrease from the saturation temperature corresponding to the condensation pressure of the refrigerant. Therefore, the absolute temperature of the refrigerant decreases as the degree of supercooling increases. Further, the degree of supercooling 0 (K) means a saturated liquid phase state. The vertical axis in FIG. 5 indicates the flow rate of the refrigerant passing through the pressure reducing means when the pressure reducing means inlet pressure and the outlet pressure are set to predetermined values and the pressure difference between the pressure reducing means inlet and outlet is constant.

  Further, in FIG. 5, assuming that the mass flow rate at the degree of supercooling 10 (K) is 100, the change in the refrigerant flow rate with respect to the change in the degree of supercooling is rendered dimensionless and plotted with ●. For reference, the refrigerant flow rate in the gas-liquid two-phase state is plotted with ◯. In this investigation, a capillary tube, which is a fixed throttle, is used as the pressure reducing means.

  As shown in FIG. 5, it can be seen that the flow rate of the refrigerant passing through the decompression means decreases as the degree of supercooling decreases, and in particular, the flow rate decreases significantly when approaching a saturated liquid phase state. In other words, in the cycle of the prior application example, when the degree of supercooling of the refrigerant decreases and approaches the saturated liquid phase state, the flow rate of the refrigerant passing through the decompression means is greatly reduced, and the flow rate of the refrigerant flowing into the suction side evaporator is also reduced. It will decline.

  For this reason, if the degree of supercooling of the refrigerant at the outlet of the supercooling section decreases due to the operating conditions of the cycle, the refrigerating capacity of the suction side evaporator also decreases, and the effect of improving cycle efficiency due to the expansion of the refrigerating capacity of the suction side evaporator described above is obtained. It becomes impossible.

  Further, for example, when the outdoor air temperature becomes high under high load operation conditions, the amount of surplus refrigerant stored in the receiver unit decreases, and the saturated liquid phase refrigerant cannot be sufficiently supplied to the supercooling unit. For this reason, the state of the refrigerant flowing into the supercooling section also changes. Such a change in the state of the refrigerant flowing into the supercooling section also causes a change in the degree of supercooling of the refrigerant at the outlet of the supercooling section.

  In view of the above points, an object of the present invention is to provide an ejector-type refrigeration cycle that can simultaneously obtain the effect of improving nozzle efficiency and the effect of expanding the refrigeration capacity of the suction-side evaporator even when operating conditions change. .

  In order to achieve the above object, in the present invention, the compressor (11) that compresses and discharges the refrigerant, and the first and second radiators (12) that release heat from the high-pressure refrigerant compressed by the compressor (11). 16), and an ejector (14) for sucking the refrigerant from the refrigerant suction port (14b) by a high-speed refrigerant flow injected from the nozzle part (14a) for decompressing and expanding the refrigerant on the downstream side of the first radiator (12), The second radiator (16) decompression means (18) for decompressing and expanding the downstream-side refrigerant, and the suction-side evaporator (19) for evaporating the downstream-side refrigerant and outflowing to the upstream side of the refrigerant suction port (14b) ), And the heat dissipation capability of the second radiator (16) can be adjusted independently of the heat dissipation capability of the first radiator (12), and the heat dissipation capability of the first radiator (12) The capacity of the refrigerant flowing into the nozzle part (14a) is saturated liquid phase Or the ejector type which is adjusted so that it may become a gas-liquid two phase state, and the heat dissipation capability of the 2nd radiator (16) is adjusted so that the state of the refrigerant which flows into decompression means (18) may be in a supercooled state. Features a refrigeration cycle.

  According to this, the heat dissipation capability of the second radiator (16) can be independently adjusted with respect to the heat dissipation capability of the first radiator (12), and the refrigerant flows into the nozzle portion (14a). Since the state is adjusted so as to be a saturated liquid phase state or a gas-liquid two phase state, the above-described effect of improving the nozzle efficiency can be obtained even if the operating conditions change.

  Furthermore, since the heat radiation capability of the second radiator (16) is adjusted so that the state of the refrigerant flowing into the decompression means (18) becomes a supercooled state, the flow rate of the refrigerant passing through the decompression means (18) decreases. Can be suppressed. As a result, it is possible to obtain the effect of expanding the refrigerating capacity of the suction side evaporator (19) even if the operating conditions change.

  Therefore, even if the operating conditions change, it is possible to simultaneously obtain the effect of improving the nozzle efficiency of the nozzle section (14a) and the effect of expanding the refrigeration capacity of the suction side evaporator (19).

  The ejector-type refrigeration cycle having the above-described characteristics includes a branch portion (A) that branches the flow of the high-pressure refrigerant, and the branch portion (A) branches the flow of the high-pressure refrigerant that has flowed out of the first radiator (12). Then, one of the branched refrigerants may flow into the nozzle part (14a), and the other refrigerant may flow into the second radiator (16).

  According to this, the refrigerant flow path through which the refrigerant flows and the compressor (11) in the order of the compressor (11) → the first radiator (12) → the branching portion (A) → the nozzle portion (14a) → the compressor (11). ) → first radiator (12) → branch (A) → second radiator (16) → pressure reducing means (18) → suction side evaporator (19) → refrigerant suction port (14b) in this order A refrigerant flow path can be configured.

  Therefore, in the second radiator (16), the refrigerant in the saturated liquid phase state or the gas-liquid two-phase state on the downstream side of the first radiator (12) is radiated, so that the refrigerant discharged from the compressor (11) is directly introduced. Compared with the case where heat is radiated, the heat exchange capacity of the second radiator (16) can be reduced. Thereby, the size reduction effect of a 2nd heat radiator (16) can also be acquired.

  The ejector-type refrigeration cycle having the above-described characteristics includes a branching section (B) that branches the flow of the high-pressure refrigerant, and the branching section (B) branches the flow of the high-pressure refrigerant discharged from the compressor (11). One of the branched refrigerants may flow into the first radiator (12), and the other refrigerant may flow into the second radiator (16).

  According to this, the refrigerant flow path through which the refrigerant flows and the compressor (11) in the order of the compressor (11) → the branching portion (B) → the first radiator (12) → the nozzle portion (14a) → the compressor (11). ) → Branch (B) → Second radiator (16) → Depressurization means (18) → Suction side evaporator (19) → Refrigerant suction port (14b) → Compressor (11) A road can be constructed.

  Therefore, the heat dissipation capability of the second radiator (16) can be easily configured with a cycle that can be adjusted completely independently of the heat dissipation capability of the first radiator (12).

  Further, the ejector refrigeration cycle having the above-described characteristics may include an outflow side evaporator (15a) for evaporating the refrigerant on the downstream side of the ejector (14). According to this, refrigerating capacity can be exhibited in both the outflow side evaporator (15a) and the suction side evaporator (19).

  Moreover, in the ejector-type refrigeration cycle having the above-described features, the first heat dissipation capacity adjusting means (12a) for adjusting the heat dissipation capacity of the first radiator (12) and the first subcooling of the refrigerant discharged from the first radiator (12). First subcooling degree detecting means (21, 22) for detecting the degree of temperature, and first control means (20a) for controlling the operation of the first heat radiation capacity adjusting means (12a), the first control means (20a) May be configured to control the operation of the first heat radiation capacity adjusting means (12a) so that the first degree of subcooling is equal to or lower than a predetermined first set value.

  According to this, the refrigerant flowing into the nozzle part (14a) is in a saturated liquid phase state due to the pressure loss of the refrigerant passage from the outlet side of the first radiator (12) to the nozzle part (14a) of the ejector (14). In addition, since the first set value can be set, the state of the refrigerant flowing into the nozzle portion (14a) can be surely set to the saturated liquid phase state or the gas-liquid two phase state.

  Further, in the ejector-type refrigeration cycle having the above-described characteristics, the second heat dissipating capacity adjusting means (16a) for adjusting the heat dissipating capacity of the second heat dissipator (16) and the second subcooling of the refrigerant discharged from the second heat dissipator (16). A second supercooling degree detecting means (23, 24) for detecting the degree of heat and a second control means (20b) for controlling the operation of the second heat radiation capacity adjusting means (16a), and the second control means (20b). May be configured to control the operation of the second heat radiation capacity adjusting means (16a) so that the second subcooling degree is equal to or higher than a predetermined second set value.

  According to this, since the second set value can be set so that the degree of change in the refrigerant flow rate passing through the pressure reducing means (18) is reduced due to the degree of supercooling, the refrigerant flow rate passing through the pressure reducing means (18). Can be effectively suppressed.

  Further, in the ejector-type refrigeration cycle having the above-described features, specifically, the first heat dissipation capacity adjusting means is a first blower (12a) that blows air that exchanges heat with the refrigerant in the first radiator (12). May be. Further, the second heat radiation capacity adjusting means may be a second blower (16a) that blows air that exchanges heat with the refrigerant in the second heat radiator (16).

  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)
FIG. 1 shows an example in which an ejector refrigeration cycle 10 according to a first embodiment of the present invention is applied to a vehicle air conditioner. In the ejector refrigeration cycle 10 of the present embodiment, a compressor 11 that sucks in refrigerant, compresses and discharges the refrigerant is rotated by a vehicle travel engine (not shown) via a pulley and a belt.

  The compressor 11 may be a variable capacity compressor that can adjust the refrigerant discharge capacity by changing the discharge capacity, or a fixed capacity type compressor that adjusts the refrigerant discharge capacity by changing the operating rate of the compressor operation by switching the electromagnetic clutch. Any of the machines may be used. Further, if an electric compressor is used as the compressor 11, the refrigerant discharge capacity can be adjusted by adjusting the rotation speed of the electric motor.

  A first radiator 12 is connected to the discharge side of the compressor 11. The first radiator 12 exchanges heat between the high-temperature and high-pressure refrigerant discharged from the compressor 11 and the outside air (air outside the passenger compartment) blown by the first cooling fan 12a to dissipate the high-temperature and high-pressure refrigerant. It is a vessel.

  The first cooling fan 12a is an electric blower in which the rotation speed (the amount of blown air) is controlled by a control voltage output from the air conditioning control device 20 described later. And the heat dissipation capability of the 1st heat radiator 12 is adjusted with the amount of blowing air of this 1st cooling fan 12a. Therefore, the 1st cooling fan 12a comprises the 1st thermal radiation capability adjustment means of this embodiment.

  In the ejector refrigeration cycle of the present embodiment, a chlorofluorocarbon refrigerant is employed as the refrigerant, and a subcritical cycle in which the high-pressure refrigerant does not rise above the supercritical pressure is configured. And the amount of blowing air of the 1st cooling fan 12a is controlled so that the 1st radiator 12 may function as a condenser which condenses a refrigerant.

  More specifically, the heat dissipation capability of the first radiator 12 is adjusted so that the degree of supercooling of the refrigerant at the outlet of the first radiator 12 is equal to or less than a predetermined first set value (specifically, 3K). Details of the first set value will be described later.

  A branch portion A that branches the flow of the refrigerant is connected to the outlet side of the first radiator 12. The branch portion A can be easily configured by a three-way joint or the like. Furthermore, one of the refrigerants branched at the branching part A flows into the nozzle part side pipe 13a connecting the branching part A and the nozzle part 14a of the ejector 14, and the other refrigerant is a refrigerant suction port of the branching part A and the ejector 14. It flows into the suction port side piping 13b connecting 14b.

  The ejector 14 is a depressurizing unit that depressurizes the refrigerant, and is a refrigerant circulating unit that circulates the refrigerant by a suction action of a refrigerant flow injected at a high speed.

  The ejector 14 is configured to reduce the passage area of the refrigerant flowing in through the nozzle portion side pipe 13a and communicate with the nozzle portion 14a that decompresses and expands the refrigerant in an isentropic manner and the refrigerant injection port of the nozzle portion 14a. It has the refrigerant | coolant suction port 14b which attracts | sucks the gaseous-phase refrigerant | coolant which has been arrange | positioned and flowed out from the suction side evaporator 19 mentioned later.

  Furthermore, a mixing unit 14c that is disposed downstream of the nozzle unit 14a and the refrigerant suction port 14b and mixes the high-speed refrigerant flow from the nozzle unit 14a and the suction refrigerant from the refrigerant suction port 14b, and the mixing unit 14c It has a diffuser portion 14d that is disposed on the downstream side and forms a pressure increasing portion that decelerates the refrigerant flow and increases the refrigerant pressure.

  The diffuser portion 14d is formed in a shape that gradually increases the passage area of the refrigerant, and has the function of decelerating the refrigerant flow to increase the refrigerant pressure, that is, the function of converting the velocity energy of the refrigerant into pressure energy.

  An outlet-side evaporator 15 is connected to the outlet side of the diffuser portion 14 d of the ejector 14. The outflow side evaporator 15 performs heat exchange between the low-pressure refrigerant decompressed by the nozzle portion 14a of the ejector 14 and the blown air blown from the blower fan 15a, and cools the blown air by causing the low-pressure refrigerant to absorb heat. It is a heat exchanger.

  The blower fan 15 a is an electric blower whose rotational speed is controlled by a control voltage output from the air conditioning controller 20. Furthermore, an accumulator 15 b that separates the gas-liquid of the refrigerant flowing out from the outflow side evaporator 15 and stores excess refrigerant is connected to the outlet side of the outflow side evaporator 15. Furthermore, the suction side of the compressor 11 is connected to the gas-phase refrigerant outlet of the accumulator 15b via a low-pressure side refrigerant flow path 17b of the internal heat exchanger 17 described later.

  On the other hand, in the suction port side pipe 13b into which the other refrigerant branched at the branching section A flows, the second radiator 16, the high-pressure side refrigerant flow path 17a of the internal heat exchanger 17, the fixed throttle 18, the suction side evaporator 19 is provided. First, the second radiator 16 causes the refrigerant to radiate heat by exchanging heat between the refrigerant flowing into the suction port side pipe 13b from the branch portion A and the outside air (air outside the vehicle compartment) blown by the second cooling fan 16a. It is a heat exchanger for heat dissipation.

  The second cooling fan 16 a is an electric blower whose rotation speed (amount of blown air) is controlled by a control voltage output from the air conditioning control device 20. And the heat dissipation capability of the 2nd heat radiator 16 is adjusted with the amount of blowing air of this 2nd cooling fan 16a. Therefore, the 2nd cooling fan 16a comprises the 2nd heat dissipation capability adjustment means of this embodiment. Thereby, the heat dissipation capability of the second radiator 16 is configured to be independently adjustable with respect to the heat dissipation capability of the first radiator 12.

  As described above, in the present embodiment, since the degree of supercooling of the refrigerant at the outlet of the first radiator 12 is adjusted to be equal to or lower than the first set value, the second radiator 16 performs supercooling to supercool the refrigerant. It functions as a vessel. More specifically, the heat dissipation capability of the second radiator 16 is adjusted so that the degree of supercooling of the refrigerant at the outlet of the second radiator 16 is equal to or higher than a predetermined second set value (specifically, 10K). Details of the second set value will be described later.

  A high-pressure side refrigerant flow path 17 a of the internal heat exchanger 17 is disposed on the downstream side of the second radiator 16. The internal heat exchanger 17 performs heat exchange between the high-temperature and high-pressure refrigerant passing through the high-pressure side refrigerant flow path 17a and the low-temperature and low-pressure refrigerant passing through the low-pressure side refrigerant flow path 17b. The refrigerant flowing into the suction-side evaporator 19 is cooled by heat exchange between the refrigerants, so that the enthalpy difference (refrigeration capacity) of the refrigerant between the refrigerant inlet and outlet in the suction-side evaporator 19 can be increased. .

  Various configurations can be adopted as the specific configuration of the internal heat exchanger 17. Specifically, a configuration in which the refrigerant pipes that form the high-pressure side refrigerant flow path 17a and the low-pressure side refrigerant flow path 17b are brazed and joined together to exchange heat, or the inside of the outer pipe that forms the high-pressure side refrigerant flow path 17a. It is possible to adopt a double-pipe heat exchanger configuration in which the low-pressure side refrigerant flow path 17b is disposed.

  A fixed throttle 18 is disposed on the downstream side of the high-pressure refrigerant passage 17a of the internal heat exchanger 17. The fixed throttle 18 is a pressure reducing unit that decompresses and expands the refrigerant, and also a flow rate adjusting unit that adjusts the flow rate of the refrigerant flowing into the suction side evaporator 19. Furthermore, in this embodiment, a capillary tube is employed as the fixed throttle 18. Of course, other types of fixed throttles (eg, orifices) may be employed.

  A suction side evaporator 19 is disposed downstream of the fixed throttle 18. The suction-side evaporator 19 is a heat-absorbing heat exchanger that cools the blown air by heat-exchanging the low-pressure refrigerant decompressed by the fixed throttle 18 and the blown air blown from the blower fan 15a and absorbing heat to the low-pressure refrigerant. is there. Further, the outlet side of the suction side evaporator 19 is connected to the refrigerant suction port 14b.

  In addition, the outflow side evaporator 15 and the suction side evaporator 19 of this embodiment are assembled | attached to the integral structure. Accordingly, the blown air blown by the blower fan 15a flows in the direction of the arrow Z, and is first cooled by the outflow side evaporator 15 and then cooled by the suction side evaporator 19 and flows into the space to be cooled (vehicle interior). . Therefore, in the present embodiment, the same cooling target space can be cooled by the outflow side evaporator 15 and the suction side evaporator 19.

  Next, an outline of the electric control unit of the present embodiment will be described. The air conditioning control device 20 includes a known microcomputer including a CPU, a ROM, a RAM, and the like and peripheral circuits thereof. The air conditioning control device 20 performs various calculations and processes based on the control program stored in the ROM, and controls the operations of the various electric actuators 12a, 15a, 16a and the like described above.

  In addition, although the air-conditioning control apparatus 20 of this embodiment comprises the control means which controls both the 1st cooling fan 12a and the 2nd cooling fan 16a, among the air-conditioning control apparatuses 20, the 1st cooling fan 12a is comprised. The hardware and software to be controlled are referred to as first control means 20a, and the hardware and software for controlling the second cooling fan 16a are referred to as second control means 20b. Of course, you may comprise the 1st control means 20a and the 2nd control means 20b by a separate control apparatus.

  Further, the air conditioning control device 20 includes a first temperature sensor 21 that detects the outlet refrigerant temperature Tr1 of the first radiator 12, a first pressure sensor 22 that detects the outlet refrigerant pressure Pr1 of the first radiator 12, and a second radiator 16. Detection signals such as a second temperature sensor 23 for detecting the outlet refrigerant temperature Tr2 and a second pressure sensor 24 for detecting the first radiator 12 outlet medium pressure Pr2 are input.

  In addition to the detection signals of the sensors, various operation signals from an operation panel (not shown) are input to the air conditioning control device 20. The operation panel is provided with an operation switch for operating the vehicle refrigeration apparatus, a temperature setting switch for setting the cooling temperature of the space to be cooled, and the like.

  Next, the operation of the present embodiment in the above-described configuration will be described with reference to FIG. FIG. 2 is a Mollier diagram schematically showing the state of the refrigerant in the ejector refrigeration cycle 10 of the present embodiment. First, when the operation switch of the operation panel is turned on and the rotational driving force of the vehicle running engine is transmitted to the compressor 11, the compressor 11 is activated.

  The high-temperature and high-pressure refrigerant (point a in FIG. 2) discharged from the compressor 11 flows into the first radiator 12 and is cooled and condensed by the outside air blown by the first cooling fan 12a (a in FIG. 2). Point → b). At this time, the first control means 20a calculates the degree of supercooling of the first radiator 12 outlet refrigerant based on the first radiator 12 outlet refrigerant temperature Tr1 and the pressure Pr1, and the calculated value is the first set value (3K). ) The rotational speed (the amount of blown air) of the first cooling fan 12a is controlled so as to be as follows.

  Therefore, in the present embodiment, the first temperature sensor 21 and the first pressure sensor 22 constitute a first supercooling degree detection means for detecting the first supercooling degree of the refrigerant at the outlet of the first radiator 12.

  Here, details of the first set value will be described. The first set value is a value set so that the refrigerant flowing into the nozzle portion 14a is in a saturated liquid phase state due to the pressure loss in the refrigerant passage from the outlet side of the first radiator 12 to the nozzle portion 14a of the ejector 14. is there. Therefore, if the degree of supercooling of the refrigerant at the outlet of the first radiator 12 is equal to or lower than the first set value, the refrigerant flowing into the nozzle portion 14a is surely in a saturated liquid phase state or a gas-liquid two phase state.

  In addition, in order to make the refrigerant flowing into the nozzle part 14a into the saturated liquid phase state or the gas-liquid two phase state more reliably, the first control means 20a can detect the minimum supercooling degree based on the temperature Tr1 and the pressure Pr1. May be the first set value. In addition, a throttle mechanism is added to the refrigerant path from the outlet side of the first radiator 12 to the nozzle portion 14a of the ejector 14 so that the refrigerant flowing into the nozzle portion 14a becomes a saturated liquid phase refrigerant at this minimum supercooling degree. That's fine.

  Next, the flow of the supercooled refrigerant that has flowed out of the first radiator 12 includes a refrigerant flow that flows into the nozzle portion side pipe 13a side at the branch portion A, and a refrigerant flow that flows into the suction port side pipe 13b side. Fork.

  The refrigerant that has flowed from the branch portion A into the nozzle portion side pipe 13 a flows into the nozzle portion 14 a of the ejector 14. At this time, as described above, the refrigerant flowing into the nozzle portion 14a is in a saturated liquid phase state or a gas-liquid two phase state (point b → point c in FIG. 2).

  The refrigerant flowing into the nozzle portion 14a is decompressed and expanded in an isentropic manner (point c → point d in FIG. 2). 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 a refrigerant | coolant injection port. Due to the refrigerant pressure drop at this time, the refrigerant (gas phase refrigerant) after passing through the suction side evaporator 19 is sucked from the refrigerant suction port 14b.

  The refrigerant injected from the nozzle portion 14a and the refrigerant sucked into the refrigerant suction port 14b are mixed in the mixing portion 14c on the downstream side of the nozzle portion 14a (point d → point e in FIG. 2) and flow into the diffuser portion 14d. In the diffuser part 14d, the refrigerant velocity (expansion) energy is converted into pressure energy due to expansion of the passage area, so that the pressure of the refrigerant rises (point e → point f in FIG. 2).

  The refrigerant that has flowed out of the diffuser portion 14 d of the ejector 14 flows into the outflow side evaporator 15. In the outflow side evaporator 15, the low-pressure refrigerant absorbs heat from the blown air (arrow Z) of the blower fan 15a and evaporates. The refrigerant that has flowed out of the outflow side evaporator 15 is separated into gas and liquid by the accumulator 15b (point f → point g in FIG. 2).

  The gas-phase refrigerant flowing out from the gas-phase refrigerant outlet of the accumulator 15b flows into the low-pressure side refrigerant flow path 17b of the internal heat exchanger 17, and the high-temperature high-pressure refrigerant flowing into the high-pressure side refrigerant flow path 17a of the internal heat exchanger 17 and Heat exchange is performed (point g → point h in FIG. 2). And the gaseous-phase refrigerant | coolant which flowed out out of the low voltage | pressure side refrigerant | coolant flow path 17b is suck | inhaled by the compressor 11, and is compressed again (the h point-> a point of FIG. 2).

  On the other hand, the refrigerant flow flowing into the suction port side pipe 13b from the branch part A flows into the second radiator 16 and is cooled by the outside air blown by the second cooling fan 16a (point b in FIG. 2 → i). point). At this time, the second control means 20b of the air conditioning control device 20 determines that the degree of subcooling of the second radiator 16 outlet refrigerant is the second set value (10K) based on the temperature Tr2 and the pressure Pr2 of the second radiator 16 outlet refrigerant. ) The rotational speed (amount of blown air) of the second cooling fan 16a is controlled so as to be above.

  Therefore, in the present embodiment, the second temperature sensor 23 and the second pressure sensor 24 constitute second supercooling degree detection means for detecting the second supercooling degree of the refrigerant at the outlet of the second radiator 16.

  Here, details of the second set value will be described. As shown in FIG. 5 described above, as the degree of supercooling of the refrigerant flowing into the fixed throttle 18 increases, the degree of change of the flow rate passing through the fixed throttle 18 with respect to the degree of supercooling decreases.

  Therefore, the second set value is set so that the flow rate of the refrigerant passing through the fixed throttle 18 connected to the downstream side of the second radiator 16 does not change greatly even if the degree of supercooling of the refrigerant at the outlet of the second radiator 16 changes. Is the value set to. Therefore, if the degree of supercooling of the refrigerant at the outlet of the second radiator 16 is equal to or higher than the second set value, the cycle efficiency is not significantly reduced.

  Then, the flow of the supercooled refrigerant flowing out of the second radiator 16 is further cooled in the high-pressure side refrigerant flow path 17a of the internal heat exchanger 17 to reduce enthalpy (point i in FIG. 2 → j point). The supercooled refrigerant that has flowed out of the high-pressure side refrigerant passage 17a is decompressed and expanded by the fixed throttle 18 to become a low-pressure refrigerant (point j to point k in FIG. 2), and this low-pressure refrigerant flows into the suction-side evaporator 19. To do.

  In the suction side evaporator 19, the low-pressure refrigerant absorbs heat from the blown air (arrow Z) of the blower fan 15 a after passing through the outflow side evaporator 15 and evaporates (point k → l in FIG. 2). The gas-phase refrigerant evaporated in the suction side evaporator 19 is sucked into the ejector 14 from the refrigerant suction port 14b (point l → point e in FIG. 2).

  Since the ejector refrigeration cycle 10 of the present embodiment operates as described above, the refrigerant downstream of the diffuser portion 14d of the ejector 14 is supplied to the outflow evaporator 15, and the downstream refrigerant of the fixed throttle 18 is supplied to the suction side evaporator 19. Therefore, the cooling function can be exhibited at the outflow side evaporator 15 and the suction side evaporator 19 at the same time.

  At that time, the refrigerant evaporation pressure of the outflow side evaporator 15 becomes the pressure after being increased by the diffuser portion 14d, while the refrigerant evaporation pressure of the suction side evaporator 19 is the lowest pressure immediately after the pressure reduction in the nozzle portion 14a. be able to. Thereby, the refrigerant evaporation pressure (refrigerant evaporation temperature) of the suction side evaporator 19 can be made lower than the refrigerant evaporation pressure (refrigerant evaporation temperature) of the outflow side evaporator 15.

  Therefore, with respect to the flow direction Z of the blown air of the blower fan 15a, the outflow side evaporator 15 having a high refrigerant evaporation temperature is arranged on the upstream side, and the suction side evaporator 19 having a low refrigerant evaporation temperature is arranged on the downstream side. Thus, a temperature difference between the refrigerant evaporation temperature and the blown air in the outflow side evaporator 15 and a temperature difference between the refrigerant evaporation temperature and the blown air in the suction side evaporator 19 can be secured.

  As a result, since the cooling performance of both the outflow side evaporator 15 and the suction side evaporator 19 can be effectively exhibited, the cooling performance for the cooling target space (vehicle interior) is combined with the outflow side evaporator 15 and the suction side evaporator 19. Can effectively improve. Further, since the suction pressure of the compressor 11 can be increased by the boosting action in the diffuser portion 14d, the amount of compression work of the compressor 11 can be reduced by an amount corresponding to the boosting action, and a power saving effect can be exhibited.

  Furthermore, in this embodiment, since the heat dissipation capability of the first radiator 12 is adjusted so that the state of the refrigerant flowing into the nozzle portion 14a of the ejector 14 becomes a saturated liquid phase state or a gas-liquid two-phase state, Even if a change in operating conditions such as an increase in temperature occurs, the nozzle efficiency of the nozzle portion 14a is not reduced.

  Moreover, since the heat dissipation capability of the second radiator 16 is adjusted so that the state of the refrigerant flowing into the fixed throttle 18 becomes a supercooled state, a decrease in the flow rate of the refrigerant passing through the fixed throttle 18 can be avoided. As a result, the refrigerating capacity of the suction side evaporator 19 does not decrease even if the operating conditions change. That is, according to the present embodiment, the effect of improving the nozzle efficiency of the nozzle portion 14a and the effect of expanding the refrigerating capacity of the suction side evaporator 19 can be obtained at the same time even if the operating conditions change.

  Furthermore, in this embodiment, since the branch part A is arrange | positioned in the downstream of the 1st heat radiator 12, in order of the compressor 11-> 1st heat radiator 12-> branch part A-> nozzle part 14a-> compressor 11. Circulating refrigerant flow path and compressor 11 → first radiator 12 → branch portion A → second radiator 16 → pressure reducing means 18 → suction side evaporator 19 → refrigerant suction port 14b → compressor 11 are circulated in this order. A refrigerant flow path can be configured.

  Therefore, the refrigerant flowing into the suction side evaporator 19 can be cooled by both the first and second radiators 12 and 16. That is, since the second radiator 16 radiates the refrigerant in the saturated liquid phase downstream of the first radiator 12, the second radiator 16, compared with the case where the refrigerant discharged from the compressor 11 is directly introduced to radiate heat. The heat exchange capacity and heat dissipation capacity can be reduced. As a result, the second radiator 16 and the second cooling fan 12a can be reduced in size.

(Second Embodiment)
In the first embodiment, the branch section A that branches the flow of the refrigerant on the downstream side of the first radiator 12 is provided, but in this embodiment, the flow of the refrigerant discharged from the compressor 11 is changed as shown in the overall configuration diagram of FIG. A branching section B that branches off is provided. That is, the first radiator 12 and the second radiator 16 are connected in parallel to the flow of the refrigerant discharged from the compressor 11. Other configurations are the same as those of the first embodiment.

  Therefore, in this embodiment, the refrigerant flow path in which the refrigerant flows in the order of the compressor 11 → the branch portion B → the first radiator 12 → the nozzle portion 14a → the compressor 11 and the compressor 11 → the branch portion B → the second radiator. The refrigerant flow path through which the refrigerant flows is configured in the order of 16 → pressure reducing means 18 → suction side evaporator 19 → refrigerant suction port 14b.

  And if the ejector type refrigerating cycle 10 of this embodiment is operated, the state of a refrigerant | coolant will change like 1st Embodiment. That is, the refrigerant flowing from the branching section B to the nozzle 14a side of the ejector 14 is point a → b point → c point → d point → e point → f point → g point → h point → a in the Mollier diagram of FIG. Cycles in the order of the points. On the other hand, the refrigerant flowing into the second radiator 16 side from the branch portion B flows in the order of point a → i point → j point → k point → l point → e point in FIG.

  Accordingly, similarly to the first embodiment, the cooling function can be exhibited simultaneously in the outflow side evaporator 15 and the suction side evaporator 19, and the compressor 11 can be saved by the pressure increasing action of the diffuser portion 14d. Even if the operating conditions change, the effect of improving the nozzle efficiency of the nozzle portion 14a and the effect of expanding the refrigerating capacity of the suction side evaporator 19 can be obtained at the same time.

  Furthermore, in this embodiment, since the flow of the refrigerant discharged from the compressor 11 is branched at the branching section B, the degree of supercooling of the refrigerant on the downstream side of the first radiator 12 is determined by the rotational speed of the first cooling fan 12a (the amount of blown air). ) Can be adjusted only by control, and the degree of supercooling of the refrigerant on the downstream side of the second radiator 16 can be adjusted only by controlling the number of rotations (the amount of blown air) of the second cooling fan 16a.

  Therefore, the first and second cooling fans 12a and 16a can independently adjust the heat dissipation capabilities of the first and second radiators 12 and 16, and the first and second cooling fans 12a and 16a can be easily controlled. It becomes.

(Third embodiment)
In the present embodiment, an example in which the ejector refrigeration cycle 10 of the present invention is applied to a commercial refrigeration system will be described. In the present embodiment, as shown in the overall configuration diagram of FIG. 4, the first cooling fan 12a, the first control means 20a, the first temperature sensor 21 and the first pressure sensor 22 are eliminated from the first embodiment. Yes.

  Since this commercial refrigeration apparatus is used in a stationary state, the change in the heat radiation capacity of the first radiator 12 due to the temperature change in the environment in which the apparatus is installed is less with respect to the vehicle air conditioner and the like. Therefore, in the present embodiment, the heat radiation area of the first radiator 12 is set so that the refrigerant at the outlet of the first radiator 12 is in a supercooled state in the temperature range assumed in the environment where the business refrigeration refrigerator is installed. Is set.

  Therefore, even if the ejector refrigeration cycle 10 of the present embodiment is operated, the refrigerant at the outlet of the first radiator 12 is in a supercooled state, but from the branch portion A to the second radiator 16 due to a change in the operation state of the cycle. The degree of supercooling of the incoming refrigerant can vary.

  On the other hand, in this embodiment, since the heat dissipation capability of the second radiator 16 can be adjusted independently of the heat dissipation capability of the first radiator by the action of the second cooling fan 16a, the second heat dissipation. The supercooling degree of the outlet refrigerant can be adjusted to the second supercooling degree (10K) or more. As a result, the same effect as that of the first embodiment can be obtained.

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

  (1) In the above-described embodiment, the first temperature sensor 21 and the first pressure sensor 22 constitute the first supercooling degree detection means, and the second temperature sensor 23 and the second pressure sensor 24 detect the second supercooling degree. Although the means is configured, the superheat degree detecting means can be configured as follows.

  For example, the first supercooling degree detection means may be constituted by a temperature sensor that detects the refrigerant discharge refrigerant temperature and the first temperature sensor 21. Thereby, since a 1st supercooling degree detection means can be comprised only with a temperature sensor, the manufacturing cost of the ejector-type refrigeration cycle 10 can be reduced with respect to the case where a pressure sensor is used.

  Similarly, the second supercooling degree detection means may be constituted by a temperature sensor that detects the refrigerant temperature at the inlet of the second radiator 16 and a temperature sensor that detects the refrigerant temperature at the outlet of the second radiator 16. Of course, the first temperature sensor 21 may be used as a temperature sensor for detecting the refrigerant temperature at the inlet of the second radiator 16. The first temperature sensor 21 may be a temperature sensor that detects outdoor air temperature (first cooling fan 12a blown air temperature).

  Furthermore, the difference between the first radiator 12 outlet refrigerant pressure Pr1 and the second radiator 16 outlet refrigerant pressure Pr2 is due to the pressure loss in the branch portions A and B and the first and second radiators 12 and 16. In a cycle in which Pr1 and Pr2 are substantially equal, any one of the pressure sensors may be eliminated.

  Further, in a cycle in which Pr1 and Pr2 are substantially equal, the compressor 11 discharge refrigerant pressure may be set to Pr1 and PR2. In this case, the first and second supercooling degree detection means can be configured using a pressure sensor that detects the compressor 11 discharge refrigerant pressure. Furthermore, the refrigerant discharge pressure of the compressor 11 may be calculated from the detection value of the temperature sensor that detects the refrigerant suction temperature of the compressor 11 and the detection value of the rotation speed sensor that detects the rotation speed of the compressor 11.

  (2) In the above-described embodiment, the first and second heat radiation capacity adjusting means are constituted by the first and second cooling fans 12a and 16a, but the heat radiation capacity adjusting means is not limited to this.

  For example, as the first and second heat radiation capacity adjusting means, a bypass passage is provided to bypass the refrigerant passing through the first and second heat radiators 12 and 16 downstream of the first and second heat radiators, thereby adjusting the heat radiation capacity. May be. Furthermore, you may adjust the thermal radiation capability by providing the interruption | blocking mechanism which interrupts | blocks the flow of the ventilation air of the 1st, 2nd cooling fans 12a and 16a.

  (3) In the above-described embodiment, the ejector 14 and the fixed throttle 18 in which the refrigerant passage area is fixed are employed. However, as the ejector 14, a variable ejector capable of adjusting the refrigerant passage area of the nozzle portion 14a is employed. Also good. As a specific example of this variable nozzle portion, a mechanism may be used in which a needle is inserted into the passage of the variable nozzle portion and the position of the needle is controlled by an electric actuator to adjust the refrigerant passage area. Furthermore, instead of the fixed throttle 18, a variable throttle mechanism capable of adjusting the refrigerant passage area may be adopted as the pressure reducing means.

  (4) In the above-described embodiment, the example in which the internal heat exchanger 17 and the fixed throttle 18 are configured separately has been described, but the internal heat exchanger 17 and the fixed throttle 18 may be configured integrally. Specifically, the high-pressure side refrigerant flow path 17a of the internal heat exchanger 17 may be configured with a fixed throttle 18 (capillary tube).

  (5) In the first and second embodiments described above, the accumulator 15b is disposed on the downstream side of the outflow side evaporator 15. However, the accumulator 15b is abolished and the subcooling is performed on the downstream side of the first radiator 12. A receiver that stores the liquid refrigerant in the state may be provided. Furthermore, you may form the branch part A in the liquid phase refrigerant | coolant storage part of a receiver.

  In this case, for example, a known temperature type expansion valve may be arranged in the nozzle part side passage 13a, and the superheat degree of the refrigerant sucked in the compressor 11 may be set to a predetermined value by the temperature type expansion valve. Further, an electric expansion valve is arranged in the nozzle section side passage 13a, the degree of superheat is calculated from the detected values of the pressure sensor and the temperature sensor arranged on the suction side of the compressor 11, and the refrigerant sucked into the compressor 11 by the electric expansion valve. The degree of superheat may be a predetermined value.

  (6) In the above-described embodiment, the outflow side evaporator 15 and the suction side evaporator 19 are assembled in an integrated structure. As specific means, for example, the configuration of the outflow side evaporator 15 and the suction side evaporator 19 The parts may be made of aluminum and joined to the unitary structure by brazing.

  Furthermore, it may be configured to be integrally coupled with an interval of 10 mm or less by mechanical engagement means such as bolt tightening. Further, a fin-and-tube type heat exchanger is adopted as the outflow side evaporator 15 and the suction side evaporator 19, and the fins of the outflow side evaporator 15 and the suction side evaporator 19 are made common and are in contact with the fins. You may integrate as a structure divided | segmented by.

  (7) In the above-described embodiment, the same cooling target space is cooled by the outflow side evaporator 15 and the suction side evaporator 19, but different cooling target spaces may be cooled by both the evaporators 15 and 19. Good. Alternatively, the outflow side evaporator 15 may be eliminated and the suction side evaporator 19 alone may be used for cooling.

  Furthermore, a branch pipe that connects the suction side of the compressor 11 from the downstream side of the first radiator 12 or the second radiator 16 may be provided, and a throttle mechanism and a branch passage evaporator may be arranged in the branch pipe. According to this, in addition to the outflow side evaporator 15 and the suction side evaporator 19, the refrigerating capacity can be exhibited also by the branch passage evaporator.

  (8) In the above-described embodiment, an example in which the ejector refrigeration cycle 10 of the present invention is applied to a vehicle air conditioner and a commercial refrigeration refrigerator is described, but the application of the present invention is not limited to this. For example, the present invention may be applied to a household refrigerator, a vending machine cooling device, a showcase with a refrigeration function, and the like. Further, an HC refrigerant may be adopted as the refrigerant.

  (9) In the above-described embodiment, the first and second radiators 12 and 16 are outdoor heat exchangers that exchange heat between the refrigerant and the outside air, and the outflow side evaporator 15 and the suction side evaporator 19 are indoor side heat. Although applied as an exchanger for cooling the vehicle interior, the outflow side evaporator 15 and the suction side evaporator 19 are configured as outdoor heat exchangers that absorb heat from a heat source such as outside air, and the first and second radiators You may apply this invention to the heat pump cycle which comprises 12 and 16 as an indoor side heat exchanger which heats to-be-heated fluids, such as air or water.

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 1st Embodiment. It is a whole block diagram 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 graph which shows the relationship between the supercooling degree of a decompression means inlet_port | entrance refrigerant | coolant, and a decompression means passage refrigerant | coolant flow rate.

Explanation of symbols

DESCRIPTION OF SYMBOLS 11 ... Compressor, 12 ... 1st heat radiator, 12a ... 1st cooling fan,
14 ... Ejector, 14a ... Nozzle part, 14b ... Refrigerant suction port,
15 ... Outflow side evaporator, 16 ... Second radiator, 16a ... Second cooling fan,
18 ... fixed throttle, 19 ... suction side evaporator, 20 ... air conditioning controller,
20a ... 1st control means, 20b ... 2nd control means, 21 ... 1st temperature sensor,
22 ... 1st pressure sensor, 23 ... 2nd temperature sensor, 24 ... 2nd pressure sensor,
A, B ... Branch

Claims (8)

A compressor (11) for compressing and discharging the refrigerant;
First and second radiators (12, 16) for radiating high-pressure refrigerant compressed by the compressor (11);
An ejector (14) for sucking the refrigerant from the refrigerant suction port (14b) by a high-speed refrigerant flow injected from the nozzle part (14a) for decompressing and expanding the refrigerant on the downstream side of the first radiator (12);
Decompression means (18) for decompressing and expanding the second radiator (16) downstream refrigerant;
A suction side evaporator (19) for evaporating the refrigerant on the downstream side of the pressure reducing means (18) and flowing out to the upstream side of the refrigerant suction port (14b);
The heat dissipation capability of the second radiator (16) can be independently adjusted with respect to the heat dissipation capability of the first radiator (12),
The heat dissipation capability of the first radiator (12) is adjusted so that the state of the refrigerant flowing into the nozzle portion (14a) becomes a saturated liquid phase state or a gas-liquid two phase state,
The ejector refrigeration cycle is characterized in that the heat radiation capacity of the second radiator (16) is adjusted so that the state of the refrigerant flowing into the decompression means (18) becomes a supercooled state. A branch part (A) for branching the flow of the high-pressure refrigerant,
The branch part (A) branches the flow of the high-pressure refrigerant that has flowed out of the first radiator (12), allows one of the branched refrigerants to flow into the nozzle part (14a), and allows the other refrigerant to flow into the nozzle part (14a). The ejector refrigeration cycle according to claim 1, wherein the ejector refrigeration cycle is adapted to flow into the second radiator (16). A branch part (B) for branching the flow of the high-pressure refrigerant,
The branch section (B) branches the flow of the high-pressure refrigerant discharged from the compressor (11), causes one of the branched refrigerant to flow into the first radiator (12), and the other refrigerant to flow. The ejector refrigeration cycle according to claim 1, wherein the ejector refrigeration cycle is adapted to flow into the second radiator (16). The ejector refrigeration cycle according to any one of claims 1 to 3, further comprising an outflow side evaporator (15a) for evaporating the downstream side refrigerant of the ejector (14). First heat radiation capacity adjusting means (12a) for adjusting the heat radiation capacity of the first heat radiator (12);
First supercooling degree detection means (21, 22) for detecting the first supercooling degree of the refrigerant at the outlet of the first radiator (12);
First control means (20a) for controlling the operation of the first heat radiation capacity adjustment means (12a),
The first control means (20a) controls the operation of the first heat radiation capacity adjusting means (12a) so that the first subcooling degree is equal to or less than a predetermined first set value. The ejector refrigeration cycle according to any one of claims 1 to 4, wherein Second heat radiation capacity adjusting means (16a) for adjusting the heat radiation capacity of the second heat radiator (16);
Second supercooling degree detection means (23, 24) for detecting the second supercooling degree of the refrigerant at the outlet of the second radiator (16);
Second control means (20b) for controlling the operation of the second heat radiation capacity adjusting means (16a),
The second control means (20b) controls the operation of the second heat radiation capacity adjusting means (16a) so that the second subcooling degree is equal to or higher than a predetermined second set value. The ejector refrigeration cycle according to any one of claims 1 to 5, wherein The said 1st heat radiation capability adjustment means is a 1st air blower (12a) which blows the air which heat-exchanges with a refrigerant | coolant in the said 1st heat radiator (12), The any one of Claim 1 thru | or 6 characterized by the above-mentioned. Ejector type refrigeration cycle described in 1. The said 2nd heat dissipation capability adjustment means is a 2nd air blower (16a) which ventilates the air which heat-exchanges with a refrigerant | coolant in the said 2nd heat radiator (16), The any one of Claim 1 thru | or 7 characterized by the above-mentioned. Ejector type refrigeration cycle described in 1.

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