JP2012057939A - Ejector-type refrigeration cycle device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To stably operate an ejector-type refrigeration cycle device, even if a change in a flow rate of driving flow passing through a nozzle part of an ejector occurs.SOLUTION: The ejector-type refrigeration cycle device includes: first and second radiators 121 and 122 for radiating one and the other refrigerants branched by a branch part 38 provided at a discharge part side of a first compression mechanism 11a, a suction side decompressor 39 and a suction-side evaporator 16 for decompressing, expanding and evaporating the refrigerant that is made to flow from the second radiator 122; and a second compression mechanism 21a sucking and compressing the outlet side refrigerant of the suction-side evaporator 16 and discharge the refrigerant to a refrigerant suction port 13b of the ejector 13.

Description

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

  2. Description of the Related Art Conventionally, an ejector refrigeration cycle having an ejector that functions as a refrigerant decompression unit and a refrigerant circulation unit is known. For example, Patent Documents 1 to 3 disclose an ejector-type refrigeration cycle in which a compressor discharge refrigerant dissipates heat by exchanging heat with outdoor air using a radiator and decompresses the dissipated high-pressure refrigerant at a nozzle portion of the ejector. ing.

  For example, in the ejector-type refrigeration cycle of Patent Document 1, a gas-liquid separator that separates the gas-liquid of the low-pressure refrigerant is disposed downstream of the diffuser portion of the ejector, and the gas-phase refrigerant outlet of the gas-liquid separator is connected to the compressor inlet side. And the liquid-phase refrigerant outlet is connected to the inlet of the suction-side evaporator, and the outlet of the suction-side evaporator is connected to the refrigerant suction port of the ejector.

  Further, in the ejector refrigeration cycle of Patent Document 2, a branching portion for branching the flow of the refrigerant flowing out from the radiator is provided on the upstream side of the nozzle portion of the ejector, and one of the refrigerants branched at the branching portion is ejected from the nozzle of the ejector. The other refrigerant is caused to flow into the refrigerant suction port side of the ejector.

  An outlet-side evaporator that evaporates the refrigerant that has flowed out of the diffuser portion is disposed downstream of the diffuser portion of the ejector, and a fixed throttle that decompresses and expands the refrigerant between the branch portion and the refrigerant suction port of the ejector. The suction side evaporator is arranged so that the refrigerating capacity can be exhibited in both evaporators.

  Further, in the ejector refrigeration cycle of Patent Document 3, a branching part for branching the flow of the refrigerant flowing out from the diffuser part is provided on the downstream side of the diffuser part of the ejector, and one of the refrigerants branched at the branching part is supplied to the outflow evaporator. The other refrigerant is caused to flow into the refrigerant suction port side of the ejector through the suction side evaporator. As a result, the refrigerating capacity can be exhibited in both evaporators.

  In an ejector applied to this type of ejector-type refrigeration cycle, high-pressure refrigerant is decompressed and expanded at the nozzle portion of the ejector, and the refrigerant on the downstream side of the evaporator is sucked from the refrigerant suction port by the pressure drop of the injected refrigerant. Thus, the loss of kinetic energy at the time of decompression expansion in the nozzle portion is recovered.

  The recovered kinetic energy (hereinafter referred to as “recovered energy”) is converted into pressure energy by the diffuser portion of the ejector to increase the pressure of the compressor suction refrigerant, thereby reducing the driving power of the compressor. The coefficient of performance (COP) of the ejector refrigeration cycle is improved.

Japanese Patent No. 3322263 Japanese Patent No. 3931899 JP 2008-107055 A

  However, in this type of ejector-type refrigeration cycle, the suction capacity of the ejector decreases as the flow rate of refrigerant (hereinafter referred to as drive flow) passing through the nozzle portion decreases, so the amount of recovered energy also decreases. End up. For this reason, the above-mentioned COP improvement effect will reduce with the flow volume fall of a drive flow.

  For example, in the ejector refrigeration cycle of Patent Document 1, when the pressure of the high-pressure refrigerant decreases as the outside air temperature decreases, the pressure difference between the high-pressure refrigerant and the low-pressure refrigerant decreases, and the flow rate of the ejector drive flow decreases. End up.

  When such a decrease in the flow rate of the drive flow occurs, not only the suction capacity of the ejector is reduced and the amount of recovered energy is reduced, but also the liquid-phase refrigerant is hardly supplied from the gas-liquid separator to the evaporator, and the cycle is The refrigeration capacity that can be exerted also decreases. As a result, the COP is significantly reduced as the driving flow rate decreases.

  Furthermore, if the suction capability of the ejector is reduced and refrigerant is no longer supplied to the evaporator, the low-pressure refrigerant cannot exhibit the endothermic effect in the evaporator, causing a problem that the cycle breaks down.

  This will be described in detail with reference to FIG. FIG. 40 is a Mollier diagram showing the state of the refrigerant in the ejector refrigeration cycle of Patent Document 1 (see FIG. 2 of Patent Document 1). 40 indicates the state of the refrigerant during normal operation, and the broken line indicates the state of the refrigerant when the above-described cycle failure occurs.

As is clear from FIG. 40, when the pressure difference between the high-pressure refrigerant and the low-pressure refrigerant decreases due to a decrease in the outside air temperature or the like (open arrow X _{40 in} FIG. ₄₀ ), the suction capacity of the ejector decreases. As a result, when the refrigerant is not supplied to the evaporator, the low-pressure refrigerant cannot exhibit the endothermic effect in the evaporator (open arrow Y _{40 in} FIG. ₄₀ ).

  For this reason, as shown by the broken line in FIG. 40, the amount of heat that the refrigerant can dissipate in the radiator becomes equivalent to the compression work of the compressor. As a result, the amount of heat cannot be substantially transferred from the low pressure side to the high pressure side via the refrigerant, and the cycle fails.

  On the other hand, in the ejector type refrigeration cycle of Patent Document 2, the refrigerant flow path from the branch portion to the refrigerant suction port via the fixed throttle and the suction side evaporator is connected in parallel to the nozzle portion of the ejector. Therefore, the refrigerant that has flowed into the suction-side evaporator can be led out to the refrigerant suction port by using the refrigerant suction and discharge capabilities of the compressor.

  Therefore, even if the flow rate of the drive flow is reduced due to the reduction in the pressure difference between the high-pressure refrigerant and the low-pressure refrigerant, and the amount of energy recovered by the ejector is reduced, the refrigerant is transferred to the suction side evaporator and the outflow side evaporator by the action of the compressor Can be supplied.

  Thereby, cycle failure like the ejector type refrigerating cycle of patent documents 1 can be avoided. However, it cannot be avoided that the amount of pressure increase in the diffuser portion and the COP decrease due to the decrease in the flow rate of the driving flow.

  Further, in the ejector refrigeration cycle of Patent Document 3, the refrigerant can be caused to flow in an annular manner in the order of compressor → radiator → ejector → outflow side evaporator → compressor. Therefore, even if the flow rate of the driving flow is reduced due to the reduction in the pressure difference between the high-pressure refrigerant and the low-pressure refrigerant, and the suction capacity of the ejector is reduced, the refrigerant can be supplied to the outflow evaporator by the action of the compressor.

  Thereby, cycle failure like the ejector type refrigerating cycle of patent documents 1 can be avoided. However, it is not possible to avoid a decrease in COP due to a decrease in the amount of pressure increase in the diffuser section and a decrease in COP due to the inability to supply refrigerant to the suction side evaporator as the flow rate of the driving flow decreases. .

  That is, in an ejector-type refrigeration cycle that uses an ejector as a refrigerant decompression means, if the flow rate fluctuation of the driving flow occurs, the cycle cannot be stably operated while exhibiting a high COP.

  The present invention has been made in view of the above points, and an object of the present invention is to stably operate an ejector refrigeration cycle even when the flow rate fluctuation of the drive flow of the ejector occurs.

  In order to achieve the above object, in the first aspect of the present invention, the first compression mechanism (11a) for compressing and discharging the refrigerant and the flow of the high-pressure refrigerant discharged from the first compression mechanism (11a) are branched. Branching portion (38) to be discharged, a first radiator (121) for radiating one refrigerant branched at the branching portion (38), and a nozzle portion for decompressing and expanding the refrigerant flowing out from the first radiator (121) The refrigerant is sucked from the refrigerant suction port (13b) by the flow of the high-speed jet refrigerant jetted from (13a), and the mixed refrigerant of the jetted refrigerant and the sucked refrigerant sucked from the refrigerant suction port (13b) 13d), the ejector (13) whose pressure is increased, the second radiator (122) for radiating the other refrigerant branched by the branching portion (38), and the refrigerant flowing out from the second radiator (122) is depressurized. Suction side decompression hand to inflate (39), a suction-side evaporator (16) for evaporating the refrigerant flowing out from the suction-side decompression means (39) and flowing it out to the refrigerant suction port (13b) side, and a suction-side evaporator (16) outlet-side refrigerant An ejector-type refrigeration cycle including a second compression mechanism (21a) that sucks, compresses and discharges.

  According to this, since the heat exchange capability (heat radiation performance) of the first radiator (121) and the second radiator (122) can be changed independently, for example, heat exchange of the second radiator (122) The capacity and the heat exchange capacity (endothermic performance) of the suction side evaporator (16) can be easily adapted. Therefore, it is easy to stabilize the operation of the cycle.

  Moreover, the enthalpy of the refrigerant | coolant which flows in into the nozzle part (13a) of an ejector (13) is employ | adopted as the 1st radiator (121) by the thing whose heat exchange capability is lower than a 2nd radiator (122). Unnecessary decrease can be avoided. Thereby, the amount of recovered energy in the nozzle part (13a) can be increased, and COP can be improved.

  Furthermore, since the second compression mechanism (21a) is provided, the suction capability of the ejector (13) can be assisted by the action of the second compression mechanism (21a). Therefore, the refrigerant can be reliably supplied to the suction-side evaporator (16) even if the operating condition is such that the suction capacity of the ejector (13) decreases as the drive flow rate of the ejector (13) decreases. it can.

  At this time, the refrigerant can be boosted by the boosting action of the two first and second compression mechanisms (11a, 21a) and the diffuser portion (13d) of the ejector (13). In contrast, the COP can be improved by reducing the driving power of the first and second compression mechanisms (11a, 21a).

  That is, by increasing the suction pressure of the first compression mechanism (11a) by the pressure increasing action of the diffuser part (13d), not only the compressor driving power of the first compression mechanism (11a) is reduced, but also the respective first Since the pressure difference between the suction pressure and the discharge pressure of the first and second compression mechanisms (11a, 21a) can be reduced, the compression efficiency of the first and second compression mechanisms (11a, 21a) can be improved.

  As a result, even if the flow rate fluctuation of the driving flow occurs and the boosting capability of the diffuser section (13d) decreases, the ejector refrigeration cycle can be stably operated in a state where a high COP is exhibited.

  This means that the high / low pressure difference of the cycle is kept large, such as a refrigeration cycle apparatus that reduces the refrigerant evaporation temperature of the suction side evaporator (16) to a very low temperature (for example, about −30 ° C. to −10). It is extremely effective in a refrigeration cycle apparatus that needs to be kept.

  According to a second aspect of the present invention, in the ejector-type refrigeration cycle according to the first aspect, the refrigerant that has flowed out of the diffuser section (13) is evaporated to flow out to the suction side of the first compression mechanism (11a). It is characterized by comprising a vessel (14).

  According to this, not only the suction side evaporator (16) but also the outflow side evaporator (14) can exhibit the refrigerating capacity. Further, in the suction side evaporator (16), the refrigerant evaporating pressure corresponds to the suction action of the injected refrigerant, and in the outflow side evaporator (14), the refrigerant evaporating pressure after being boosted by the diffuser portion (13d). Therefore, the refrigerant | coolant evaporation temperature of a suction side evaporator (16) and an outflow side evaporator (14) can be made into a different temperature.

  According to a third aspect of the present invention, in the ejector refrigeration cycle according to the first or second aspect, the ejector refrigeration cycle is arranged in a refrigerant passage extending from the first radiator (121) outlet side to the nozzle part (13a) inlet side, It is characterized by comprising a high-pressure side decompression means (17) for decompressing and expanding the refrigerant flowing out from one radiator (121).

  According to this, the refrigerant flowing into the nozzle part (13a) can be depressurized by the action of the high-pressure side depressurizing means (17) until it becomes a gas-liquid two-phase refrigerant. Therefore, compared with the case where the liquid refrigerant is introduced into the nozzle part (13a), the boiling of the refrigerant in the nozzle part (13a) can be promoted to improve the nozzle efficiency.

  As a result, the amount of pressure increase in the diffuser section (13d) can be increased to further improve the COP. In addition, nozzle efficiency is energy conversion efficiency at the time of converting the pressure energy of a refrigerant | coolant into a kinetic energy in a nozzle part (13a).

  Further, by configuring the high-pressure side pressure reducing means (17) with a variable throttle mechanism, it is possible to change the flow rate of the refrigerant flowing into the nozzle portion (13a) according to the cycle load fluctuation. As a result, the ejector-type refrigeration cycle can be operated while exhibiting a high COP even if load fluctuation occurs.

  According to a fourth aspect of the present invention, in the ejector refrigeration cycle according to any one of the first to third aspects, heat exchange is performed between the refrigerant flowing out of the second radiator (122) and the low-pressure side refrigerant of the cycle. An internal heat exchanger (34, 35) 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 can be increased and the COP can be improved.

  Further, in the internal heat exchanger (34, 35), the enthalpy of the refrigerant flowing out from the second radiator (122) is reduced, so that the enthalpy of the refrigerant flowing into the nozzle portion (13a) of the ejector (13) is unnecessary. Can be avoided. Thereby, the amount of recovered energy in the nozzle portion 13a can be increased, and COP can be improved.

  In the invention according to claim 5, in the ejector refrigeration cycle according to any one of claims 1 to 4, the second radiator (122) includes a condensing part (122b) and a condensing part ( 122b) has a gas-liquid separation part (122c) that separates the gas-liquid of the refrigerant that has flowed out from 122b) and a supercooling part (122d) that supercools the liquid-phase refrigerant that has flowed out from the gas-liquid separation part (122c) It is characterized by that.

  According to this, since the supercooled low enthalpy refrigerant can flow into the suction side evaporator (16), the enthalpy of the inlet side refrigerant and the enthalpy of the outlet side refrigerant of the suction side evaporator (16) can be reduced. COP can be improved by expanding the enthalpy difference (refrigeration capacity).

  According to a sixth aspect of the present invention, in the ejector-type refrigeration cycle according to any one of the first to fifth aspects, the suction-side decompression means (39) expands and decompresses the refrigerant, and the pressure of the refrigerant An expander that converts energy into mechanical energy and outputs the energy. According to this, the energy efficiency as the whole ejector-type refrigeration cycle can be improved by effectively utilizing the mechanical energy output from the expander.

Moreover, in invention of Claim 7, the 1st compression mechanism (11a) which compresses and discharges a refrigerant | coolant, The heat radiator (12) which radiates the high pressure refrigerant | coolant discharged from the 1st compression mechanism (11a), The first and second heat exchangers (51, 52) for exchanging heat between the refrigerant and the heat exchange target fluid and the high-speed jet refrigerant flow injected from the nozzle portion (13a) for decompressing and expanding the refrigerant, An ejector (13) that sucks from the suction port (13b) and boosts the mixed refrigerant of the injected refrigerant and the suction refrigerant sucked from the refrigerant suction port (13b) at the diffuser unit (13d), and the diffuser unit (13d) An outflow-side gas-liquid separator (55) that separates the gas-liquid of the refrigerant that has flowed out of the refrigerant, a second compression mechanism (21a) that compresses the refrigerant and discharges it to the refrigerant suction port (13b) side, and a refrigerant flow path Channel switching means for switching (53 54) and equipped with a,
The flow path switching means (53, 54) causes the refrigerant discharged from the first compression mechanism (11a) to flow through the radiator (12) → the first heat exchanger (51) → the nozzle portion (13a) in the first operation mode. At the same time, the refrigerant flow path is arranged so that the liquid refrigerant flowing out from the outflow side gas-liquid separator (55) flows in the order of the second heat exchanger (52) → the second compression mechanism (21a) → the refrigerant suction port (13d). In the switching and second operation mode, the refrigerant discharged from the first compression mechanism (11a) flows in the order of the radiator (12) → second heat exchanger (52) → nozzle portion (13a) and the outflow side gas-liquid separator ( 55) An ejector refrigeration cycle that switches the refrigerant flow path so that the liquid refrigerant flowing out of the first heat exchanger (51) → the second compression mechanism (21a) → the refrigerant suction port (13d) flows in this order. .

  According to this, in the first operation mode, the refrigerant on the downstream side of the radiator (12) is supplied to the first heat exchanger (51) to defrost the first heat exchanger (51), The heat exchange target fluid can be cooled in the heat exchanger (52). In the second operation mode, the second heat exchanger (52) is supplied with the refrigerant downstream of the radiator (12) to defrost the second heat exchanger (52). The heat exchange target fluid can be cooled at (16a).

  In other words, by alternately switching between the first operation mode and the second operation mode, while performing defrosting of one of the first and second heat exchangers (51, 52), The heat exchange target fluid can be cooled.

  Furthermore, since the second compression mechanism (21a) is provided, the suction capability of the ejector (13) can be assisted by the action of the second compression mechanism (21a) in any operation mode. Therefore, similarly to the first aspect of the present invention, even if the flow rate fluctuation of the driving flow occurs and the boosting capability of the diffuser portion (13d) is reduced, the ejector refrigeration cycle is stabilized in a state where a high COP is exhibited. Can be actuated.

  The invention according to claim 8 is the ejector refrigeration cycle according to claim 7, further comprising a high-pressure side pressure reducing means (57) for decompressing and expanding the refrigerant flowing into the nozzle portion (13a). According to this, similarly to the third aspect of the invention, the ejector-type refrigeration cycle can be operated while exhibiting a high COP even when a load change occurs.

  In the invention according to claim 9, in the ejector refrigeration cycle according to claim 7 or 8, the internal heat exchanger (36) for exchanging heat between the high-pressure refrigerant upstream of the nozzle portion (13a) and the low-pressure refrigerant of the cycle. 37). According to this, in each maneuver mode, the enthalpy difference (refrigeration capacity) between the enthalpy of the inlet-side refrigerant and the enthalpy of the outlet-side refrigerant is expanded, and the COP is increased. It can be improved.

  Here, in the ejector refrigeration cycle according to claim 4 or 9, specifically, the low-pressure side refrigerant of the cycle is sucked into the first compression mechanism (11a) as in the invention according to claim 10. It may be a refrigerant, or may be a refrigerant sucked into the second compression mechanism (21a) as in the invention described in claim 11.

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

  According to this, the refrigerant | coolant discharge capability of a 1st compression mechanism (11a) and the refrigerant | coolant discharge capability of a 2nd compression mechanism (21a) are adjusted independently, and either of a 1st, 2nd compression mechanism (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 thirteenth aspect of the present invention, in the ejector refrigeration cycle according to any one of the first to twelfth aspects, the first compression mechanism (11a) and the second compression mechanism (21a) are in the same housing. It is accommodated and it is comprised integrally. Accordingly, the first compression mechanism (11a) and the second compression mechanism (21a) can be reduced in size, and the entire ejector refrigeration cycle can be reduced in size.

  As in the invention described in claim 14, in the ejector refrigeration cycle according to any one of claims 1 to 13, the first compression mechanism (11a) is configured to increase the pressure of the refrigerant until the pressure becomes equal to or higher than the critical pressure. 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.

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 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 Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigeration cycle of 14th Embodiment. It is a whole block diagram of the ejector type refrigerating cycle of 15th Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigerating cycle of 15th Embodiment. It is a whole block diagram of the ejector type refrigerating cycle of 16th Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigerating cycle of 16th Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of 17th Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigerating cycle of 17th Embodiment. It is a whole block diagram of the ejector type refrigerating cycle of 18th Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigerating cycle of 18th Embodiment. It is a whole block diagram of the ejector type refrigerating cycle of 19th Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigerating cycle of 19th Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of 20th Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant of the ejector-type refrigerating cycle of a prior art.

  Among the embodiments described below, the first to eighth embodiments are embodiments as reference examples on which the present invention is based, whereas the ninth to twentieth embodiments are described in the claims. It is an embodiment of the invention.

(First embodiment)
An example in which the ejector refrigeration cycle 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. 1 is an overall configuration diagram of an ejector refrigeration cycle 10 of the present embodiment.

  In the ejector-type refrigeration cycle 10, the first compressor 11 sucks refrigerant, compresses and discharges it, and drives the first compression mechanism 11a having a fixed discharge capacity by the first electric motor 11b. Machine. Specifically, various compression mechanisms such as a scroll-type compression mechanism and a vane-type compression mechanism can be employed as the first compression mechanism 11a.

  The first electric motor 11b is controlled in its operation (number of rotations) by a control signal output from a control device described later, and may adopt either an AC motor or a DC motor. And the refrigerant | coolant discharge capability of the 1st compression mechanism 11a is changed by this rotation speed control. Therefore, the 1st electric motor 11b comprises the 1st discharge capability change means which changes the refrigerant | coolant discharge capability of the 1st compression mechanism 11a.

  A radiator 12 is connected to the discharge port side of the first compressor 11. The heat dissipator 12 heat-exchanges the high-pressure refrigerant discharged from the first compressor 11 and the outside air (outside air) blown by the cooling fan 12a to dissipate the high-pressure refrigerant and cool it. It is. 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 10 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. The refrigerant is mixed with refrigerating machine oil for lubricating the first and second compression means 11a and 21a, and the refrigerating machine oil circulates in the cycle together with the refrigerant.

  Further, a receiver (liquid receiver) as a high-pressure side gas-liquid separator that separates the gas-liquid refrigerant flowing out of the radiator 12 and stores excess liquid-phase refrigerant may be provided on the outlet side of the radiator 12. Good. Then, the saturated liquid phase refrigerant separated from the receiver may be led to the downstream side.

  A first branch portion 18 that branches the flow of the high-pressure refrigerant that has flowed out of the radiator 12 is connected to the outlet side of the radiator 12. The 1st branch part 18 is comprised by the three-way coupling which has three inflow / outflow ports, and makes one refrigerant | coolant inflow port and two into a refrigerant | coolant outflow port. 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.

  One refrigerant outlet of the first branch portion 18 is connected to a temperature type expansion valve 17 as a high pressure side pressure reducing means, and the other refrigerant outlet has a first electricity as a first suction side pressure reducing means to be described later. The type expansion valve 19 side is connected.

  The temperature type expansion valve 17 has a temperature sensing part (not shown) disposed in an outlet side refrigerant passage, which will be described later, on the outlet side evaporator 14 outlet, and the temperature and pressure of the outlet side evaporator 14 outlet side refrigerant are adjusted. Based on this, a variable throttle mechanism that detects the degree of superheat of the refrigerant on the outlet side of the outlet evaporator 14 and adjusts the valve opening (refrigerant flow rate) by a mechanical mechanism so that the degree of superheat becomes a predetermined value set in advance. is there.

  A nozzle portion 13 a of the ejector 13 is connected to the outlet side of the temperature type expansion valve 17. 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.

  More specifically, the ejector 13 includes a nozzle portion 13a and a nozzle portion that reduce the passage area of the intermediate-pressure refrigerant that has flowed out from one refrigerant outlet of the first branching portion 18 to expand the refrigerant in an isentropic manner. The refrigerant suction port 13b is disposed so as to communicate with the refrigerant ejection port 13a and sucks the refrigerant discharged from the second compressor 21 described later.

  Further, a mixing portion 13c for mixing the high-speed jet refrigerant jetted from the nozzle portion 13a and the sucked refrigerant sucked from the refrigerant suction port 13b is provided in the refrigerant flow downstream portion of the nozzle portion 13a and the refrigerant suction port 13b. In addition, a diffuser portion 13d forming a pressure increasing portion is provided on the refrigerant flow downstream side of the mixing portion 13c.

  The diffuser portion 13d is formed in a shape that gradually increases the refrigerant passage area, and functions to increase the refrigerant pressure by decelerating the refrigerant flow, that is, to convert the velocity energy of the refrigerant into pressure energy. A refrigerant inlet 28a of the second branch portion 28 is connected to the outlet side of the diffuser portion 13d.

  The basic configuration of the second branch unit 28 is the same as that of the first branch unit 18. The outflow side evaporator 14 is connected to one refrigerant outlet 28b of the second branch portion 28, and the second electric expansion valve 29 as a second suction side pressure reducing means is connected to the other refrigerant outlet 28c. ing.

  Furthermore, the second branch portion 28 of the present embodiment is configured so that the flow direction of the refrigerant flowing out from the one refrigerant outlet 28b to the outlet evaporator 14 side and the second electric expansion valve 29 from the other refrigerant outlet 28c. The flow direction of the refrigerant flowing out to the side 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 from the diffuser portion 13d outlet side to the refrigerant inflow port 28a. .

  Therefore, the refrigerant flowing into the second branch portion 28 flows out of the second branch portion 28 without unnecessarily reducing the flow velocity when the flow is branched. Thereby, the flow velocity (dynamic pressure) of the refrigerant that has flowed out of the ejector 13 at the second branch portion 28 is maintained. Of course, the 2nd branch part 28 is not limited to this, You may form in a substantially T shape.

  The outflow side evaporator 14 evaporates the low-pressure refrigerant and absorbs heat by exchanging heat between the refrigerant flowing out from one refrigerant outlet 28b of the second branch portion 28 and the air in the refrigerator circulated by the blower fan 14a. This is an endothermic heat exchanger that exerts its action. Therefore, the heat exchange target fluid in the outflow side evaporator 14 is the air in the refrigerator.

  The blower fan 14a is an electric blower whose rotation speed (amount of blown air) is controlled by a control voltage output from the control device. The suction port of the first compressor 11 is connected to the refrigerant outlet side of the outflow side evaporator 14.

  The second electric expansion valve 29 decompresses and expands the refrigerant flowing out from the other refrigerant outlet 28 c of the second branch portion 28. More specifically, the second electric expansion valve 29 includes a valve body configured to be able to change the throttle opening degree and an electric actuator including a stepping motor that changes the throttle opening degree of the valve body. This is a variable aperture mechanism.

  Furthermore, the second electric expansion valve 29 of the present embodiment can fully close the throttle passage. Therefore, when the second electric expansion valve 29 fully closes the throttle passage, the flow of the refrigerant flowing out from the diffuser portion 13d is not branched by the second branching portion 28, and the total flow rate is reduced to the outflow side evaporator. It flows out to the 14th side. The operation of the second electric expansion valve 29 is controlled by a control signal output from the control device.

  Further, as described above, the first electric expansion valve 19 is connected to the other refrigerant outlet of the first branch portion 18. The basic configuration of the first electric expansion valve 19 is the same as that of the second electric expansion valve 29. Therefore, when the first electric expansion valve 19 fully closes the throttle passage, the flow of the refrigerant flowing out of the radiator 12 is not branched by the first branching portion 18, and the total flow rate is the temperature expansion valve 17 side. Spill to

  The outlet side of the first and second electric expansion valves 19 and 29 is connected to a merging portion 20 for merging the refrigerant flows flowing out from the first and second electric expansion valves 19 and 29. The basic configuration of the merging unit 20 is the same as that of the second branch unit 28. 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.

  Furthermore, in the junction part 20 of this embodiment, the flow direction of the refrigerant | coolant which flows in into one refrigerant | coolant inlet 20b from the 1st electric expansion valve 19 and the 2nd electric expansion valve 29 to the other refrigerant | coolant inlet 20c. The flow direction of the refrigerant flowing in is substantially Y-shaped so as to be directed to 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 16.

  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 16 is connected to the refrigerant outlet 20 a of the junction 20. The suction-side evaporator 16 heat-exchanges the refrigerant flowing out from the merging portion 20 and the freezer air circulated and blown by the blower fan 16a, thereby evaporating the low-pressure refrigerant and exerting an endothermic effect. It is.

  Therefore, the heat exchange target fluid in the suction side evaporator 16 is freezer air. 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.

  The suction port of the second compressor 21 is connected to the outlet side of the suction side evaporator 16. The basic configuration of the second compressor 21 is the same as that of the first compressor 11. Accordingly, the second compressor 21 is an electric compressor that drives the fixed capacity type second compression mechanism 21a by the second electric motor 21b. Further, the second electric motor 21b constitutes a second discharge capacity changing means for changing the refrigerant discharge capacity of the second compression mechanism 21a.

  Further, as described above, the refrigerant suction port 13 b of the ejector 13 is connected to the discharge port of the second compressor 21.

  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, 12b, 14a, 16a, 19, 21a, 29, etc. To do.

  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 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 operation. Of course, you may comprise a 1st discharge capability control means and a 2nd discharge capability control means with a different control apparatus.

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

  Next, the operation of the present embodiment in the above configuration will be described with reference to 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 blower fans 14a and 16a, and the electric expansion valves 19 and 29. At this time, the control device controls the first and second electric expansion valves 19 and 29 to the throttle state or the fully closed state, whereby the following three types of cycle configurations can be realized.

  When the control device sets the first electric expansion valve 19 to the fully closed state and the second electric expansion valve 29 to the throttle state, it is possible to realize a cycle configuration in which the refrigerant flow is branched only by the second branch portion 28 ( Hereinafter, the operation mode in this cycle configuration is referred to as a low-pressure branch operation mode).

  When the control device sets the first electric expansion valve 19 to the throttle state and the second electric expansion valve 29 to the fully closed state, a cycle configuration in which the refrigerant flow is branched only by the first branch portion 18 can be realized ( Hereinafter, the operation mode in this cycle configuration is referred to as a high-pressure branch operation mode).

  When the control device places both the first and second electric expansion valves 19 and 29 in the throttle state, it is possible to realize a cycle configuration in which the refrigerant flow is branched at the first branching portion 18 and the second branching portion 28 at the same time (hereinafter referred to as the following) The operation mode in this cycle configuration is called the simultaneous branch operation mode).

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

  In addition, it does not require refrigeration capacity than during normal operation, the refrigerant flow circulating in the cycle is lower than during normal operation, or when the outside air temperature is lower than a predetermined reference temperature, When the high / low pressure difference of the cycle becomes smaller than a predetermined pressure difference, the simultaneous branching operation mode is switched.

The low-pressure branch operation mode, gas refrigerant (a ₂ points in FIG. 2) of the high-temperature high-pressure state discharged from the first compressor 11 radiator 12 flows into, was blown from the cooling fan 12a blown air (outside air) Heat is exchanged with the heat to condense and condense (point a ₂ → b _{2 in} FIG. 2). The refrigerant that has flowed out of the radiator 12 flows into the first branch portion 18.

At this time, since the first electric expansion valve 19 is in a fully closed state, the flow of the refrigerant flowing out of the radiator 12 is not branched in the first branch portion 18, and the total flow rate is the temperature type expansion valve. 17 flows into. The refrigerant that has flowed into the temperature type expansion valve 17 is decompressed and expanded in an enthalpy manner to a gas-liquid two-phase state (b ₂ point → c ₂ point in FIG. 2). At this time, the valve opening degree of the temperature type expansion valve 17 is adjusted so that the degree of superheat (g ₂ point in FIG. 2) of the outlet side refrigerant 14 outlet side refrigerant becomes a predetermined value.

The intermediate-pressure refrigerant flowing out of the thermal expansion valve 17, flows into the nozzle portion 13a of the ejector 13, isentropically decompressed to expand (c ₂ points in FIG 2 → d ₂ points). 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 injected refrigerant, the refrigerant discharged from the second compressor 21 is sucked from the refrigerant suction port 13b. (J ₂ point → e ₂ point in Fig. 2)
Further, the refrigerant injected from the nozzle portion 13a and the refrigerant sucked from the refrigerant suction port 13b are mixed in the mixing portion 13c of the ejector 13 and flow into the diffuser portion 13d (point d2 → point e2 in FIG. 2). . In the diffuser portion 13d, the passage energy is increased and the velocity energy of the refrigerant is converted into pressure energy, so that the pressure of the refrigerant increases (point e ₂ → point f _{2 in} FIG. 2).

  The refrigerant that has flowed out of the diffuser portion 13d is branched into a refrigerant flow that flows into the outflow-side evaporator 14 side and a refrigerant flow that flows into the second electric expansion valve 29 side at the second branch portion 28. At this time, the control device increases the refrigerant flow rate G1 flowing into the outflow side evaporator 14 side more than the refrigerant flow rate G2 flowing into the second electric expansion valve 29 side so that the refrigerant evaporation temperature in the suction side evaporator 16 is increased. The throttle opening of the second electric expansion valve 29 is adjusted so that the predetermined temperature is set in advance.

Refrigerant flowing into the discharge side evaporator 14 from the second branch portion 28, to be evaporated from the refrigerator air circulated blown by the blower fan 14a (f ₂ points in FIG 2 → g ₂ points). Thereby, the air in a refrigerator is cooled. Then, the refrigerant flowing out from the outflow side evaporator 14 is sucked into the first compressor 11 and compressed again (point g ₂ → point a _{2 in} FIG. 2).

On the other hand, the refrigerant that has flowed into the second electric expansion valve 29 from the second branch portion 28 is further decompressed and expanded in an isenthalpy manner to reduce its pressure (point f ₂ → point hα _{2 in} FIG. 2). Decompressed and expanded refrigerant in the second electric expansion valve 29, flows into the suction side evaporator 16, the blower fan 16a is evaporated by absorbing heat from the freezer in air circulated blown (h [alpha ₂ points 2 → i ₂ points). Thereby, the air in a freezer is cooled.

Then, the refrigerant flowing out from the suction side evaporator 16 is sucked into the second compressor 21 and compressed (i ₂ point → j ₂ point in FIG. 2). At this time, the control device controls the operations of the first and second electric motors 11b and 21b so that the COP of the ejector refrigeration cycle as a whole approaches a maximum. Specifically, in order to improve the compression efficiency of the first and second compression mechanisms 11a and 21a, the first and second compression mechanisms 11a and 21a are controlled so that the pressure increase amounts are substantially equal.

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

  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 compressors 11 and 21 increase, the temperature of the refrigerant rises due to the frictional heat, and the actual enthalpy increase ΔH2 Increases the compression efficiency.

Further, as described above, the refrigerant discharged from the second compressor 21 is sucked into the ejector 13 from the refrigerant suction port 13b (j ₂ point → e ₂ point in FIG. 2).

Next, in the high-pressure branch operation mode, similarly to the low-pressure branch operation mode, the refrigerant discharged from the first compressor 11 dissipates heat in the radiator 12 and condenses (point a ₂ → b _{2 in} FIG. 2). The refrigerant that has flowed out of the radiator 12 is branched at the first branch portion 18.

The high-pressure refrigerant that has flowed into the temperature-type expansion valve 17 from the first branch portion 18 is decompressed and expanded in an enthalpy manner at the temperature-type expansion valve 17 in the same manner as in the low-pressure branch operation mode, and enters a gas-liquid two-phase state (FIG. 2 b ₂ point → c ₂ point). Furthermore, the intermediate-pressure refrigerant flowing out of the thermal expansion valve 17 flows in the order of the nozzle portion 13a → the diffuser portion 13d → second branch portion 28 → the outflow-side evaporator 14 of the ejector 13 (c ₂ points in FIG 2 → d ₂ points → e ₂ points → f ₂ points).

In this operation mode, since the second electric expansion valve 29 is in a fully closed state, the flow of the refrigerant flowing out from the diffuser portion 13d is not branched in the second branching portion 28, and the total flow rate is reduced to the outflow side. It flows into the evaporator 14. The refrigerant that has flowed into the outflow side evaporator 14 from the second branch portion 28 absorbs heat from the air in the refrigerator and evaporates (f ₂ point → g ₂ point in FIG. 2), as in the low pressure branch operation mode. The air is sucked into the machine 11 and compressed again (point g ₂ → point a _{2 in} FIG. 2).

On the other hand, the high-pressure refrigerant that has flowed into the first electric expansion valve 19 from the first branch portion 18 is decompressed and expanded in an enthalpy manner by the first electric expansion valve 19 to reduce its pressure (dashed line in FIG. 2). (B ₂ point → hβ ₂ point).

  At this time, the valve opening degree (throttle opening degree) of the first electric expansion valve 19 is a flow rate ratio Ge / Gnoz between 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. However, the flow rate is adjusted so as to obtain an optimum flow rate ratio capable of exhibiting a high COP in the entire cycle. The refrigerant decompressed and expanded by the first electric expansion valve 19 flows into the junction 20.

The refrigerant that has flowed into the suction-side evaporator 16 from the merging portion 20 absorbs heat from the freezer air circulated by the blower fan 16a and evaporates (hβ ₂ point → i ₂ point indicated by a broken line in FIG. 2). Thereby, the air in a freezer is cooled. Further, the refrigerant that has flowed out of the suction-side evaporator 16 is compressed by the second compressor 21 (i ₂ point → j ₂ point in FIG. 2) as in the low-pressure branch operation mode, and the refrigerant suction port of the ejector 13 Suctioned to 13b (j ₂ point → e ₂ point in FIG. 2).

Next, in the simultaneous branch operation mode, similarly to the high pressure branch operation mode, the refrigerant that has flowed out of the radiator 12 is branched at the first branch portion 18. The refrigerant that has flowed out from the first branch portion 18 to the temperature expansion valve 17 side flows in the order of the temperature expansion valve 17 → the nozzle portion 13a of the ejector 13 → the diffuser portion 13d → the second branch portion 28 (point b _{2 in} FIG. 2). → c ₂ point → d ₂ point → e ₂ point → f ₂ point).

In the second branch portion 28, the flow of the refrigerant flowing out from the diffuser portion 13d is branched, as in the low pressure branch operation mode. The refrigerant that has flowed out from the second branch portion 28 toward the outflow side evaporator 14 flows in the order of the outflow side evaporator 14 → the first compressor 11 to cool the air in the refrigerator (point f ₂ → point g _{2 in} FIG. 2). ).

The refrigerant that has flowed out from the second branch portion 28 to the second electric expansion valve 29 side flows in the order of the second electric expansion valve 29 → the merging portion 20 (point c ₂ → point hβ ₂ indicated by a broken line in FIG. 2). Further, the refrigerant that has flowed out from the first branch portion 18 to the first electric expansion valve 19 side flows in the order of the first electric expansion valve 19 → the merging portion 20 (f ₂ point → hα ₂ point in FIG. 2).

Then, the flow of the refrigerant flowing out of the second electric expansion valve 29 and the flow of the refrigerant flowing out of the first electric expansion valve 19 are merged at the merging portion 20 (hβ ₂ point indicated by a thin line in FIG. 2 → hγ ₂ points, hα ₂ points → hγ ₂ points). The refrigerant that has flowed out from the merging portion 20 flows into the suction-side evaporator 16 and evaporates, similarly to the low-pressure branch operation mode and the high-pressure branch operation mode, to cool the freezer air.

Further, the refrigerant flowing out of the suction side evaporator 16 is compressed by the second compressor 21 and sucked into the refrigerant suction port 13b of the ejector 13 (hγ ₂ point → i ₂ point → j ₂ point in FIG. 2). → e ₂ points).

  Since the ejector refrigeration cycle 10 of the present embodiment operates as described above, the following effects can be exhibited.

  (A) In any of the operation modes, since the refrigerant flow is divided in at least one of the first branch portion 18 and the second branch portion 28, it is appropriate for both the outflow side evaporator 14 and the suction side evaporator 16. Refrigerant can be supplied. Therefore, both the outflow side evaporator 14 and the suction side evaporator 16 can exhibit a cooling action simultaneously.

  At this time, the refrigerant evaporation pressure of the outflow side evaporator 14 becomes the pressure after being increased by the second compressor 21 and the diffuser part 13d, while the refrigerant evaporation pressure of the suction side evaporator 16 is increased after being increased by the diffuser part 13d. Further, the pressure is reduced after the pressure is reduced by the second electric expansion valve 29.

  Therefore, the refrigerant evaporation pressure (refrigerant evaporation temperature) of the suction side evaporator 16 can be made sufficiently lower than the refrigerant evaporation pressure (refrigerant evaporation temperature) of the outflow side evaporator 14. As a result, the outflow evaporator 14 can be used for cooling in a low-temperature refrigerator, and the suction-side evaporator 16 can be used for cooling in a cryogenic freezer.

  (B) In any operation mode, since the second compressor 21 (second compression mechanism 21a) is provided, the suction capability of the ejector 13 can be assisted by the second compression mechanism 21a. Accordingly, the operating condition in which the pressure difference between the high-pressure refrigerant and the low-pressure refrigerant decreases and the drive flow of the ejector 13 decreases, that is, the suction capacity of the ejector 13 decreases, such as when the outside air temperature is low. Even under proper operating conditions, the refrigerant can be reliably supplied to the suction-side evaporator 16.

  At this time, since the pressure of the refrigerant can be increased by the pressure increasing action of the two first and second compression mechanisms 11a, 21a and the diffuser portion 13d of the ejector 13, the first, The COP can be improved by reducing the driving power of the second compression mechanisms 11a and 21a.

  That is, not only can the drive power of the first compression mechanism 11a be reduced by increasing the suction pressure of the first compression mechanism 11a by the pressure increasing action of the diffuser portion 13d, but also the first and second compression mechanisms 11a, 21a. Since the pressure difference between the suction pressure and the discharge pressure can be reduced, the compression efficiency of the first and second compression mechanisms 11a and 21a can be improved.

  Furthermore, in the present embodiment, the first and second electric motors 11b and 21b can independently change the refrigerant discharge capacities of the first and second compression mechanisms 11a and 21a. , 21a can be effectively improved.

  As a result, the ejector refrigeration cycle can be stably operated in a state in which a high COP is exhibited even if the flow rate fluctuation of the driving flow occurs and the boosting capability of the diffuser portion 13d is reduced.

  This is because, for example, the high / low pressure difference of the cycle is largely maintained as in the refrigeration cycle apparatus that reduces the refrigerant evaporation temperature of the suction side evaporator 16 to an extremely low temperature such as −30 to −10 ° C. as in the present embodiment. This is extremely effective in a refrigeration cycle apparatus that needs to be prepared.

  (C) In the low pressure branch operation mode, the refrigerant flow rate G1 flowing from the second branch portion 28 to the outflow side evaporator 14 side is greater than the refrigerant flow rate G2 flowing from the second branch portion 28 to the second temperature type expansion valve 29 side. Therefore, a larger amount of refrigerant can be radiated 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.

  (D) In the high pressure branch operation mode and the simultaneous branch operation mode, the first compressor 11 → the radiator 12 → the first branch portion 18 → the ejector 13 → the second branch portion 28 → the outflow side evaporator 14 → the first compressor 11 The refrigerant flows in this order, and further, the first compressor 11 → the radiator 12 → the first branching portion 18 → the first electric expansion valve 19 → the merging portion 20 → the suction side evaporator 16 → the second compressor 11 → the ejector 13 The refrigerant flows in the order of the second branch portion 28, the outflow side evaporator 14, and the first compressor 11.

  That is, since the flow of the refrigerant passing through the evaporator such as the outflow side evaporator 14 and the suction side evaporator 16 is annular, lubricating oil (refrigerator oil) for the first and second compressors 11 and 21 is used as the refrigerant. Even if mixed, it is possible to prevent the oil from staying in the outflow side evaporator 14 and the suction side evaporator 16.

  (E) In the simultaneous branching operation mode, it is possible to realize a cycle configuration in which the refrigerant flowing out from both the first electric expansion valve 19 and the second electric expansion valve 29 is supplied to the suction side evaporator 16. As a result, in contrast to the cycle configuration in which the refrigerant flowing out of one of the first electric expansion valve 19 and the second electric expansion valve 29 is supplied to the suction side evaporator 16, the refrigerant is supplied to the suction side evaporator 16. It becomes easy to increase the refrigerant flow rate.

  (F) Since the accumulator as the outflow side gas-liquid separator can be eliminated on the suction side of the first compressor 11 with respect to the ejector refrigeration cycle of Patent Document 1, the manufacturing cost of the ejector refrigeration cycle 10 as a whole is reduced. can do.

  (G) Since the temperature type expansion valve 17 which is a variable throttle mechanism is employed as the high pressure side pressure reducing means, the flow rate of the refrigerant flowing into the nozzle portion 13a of the ejector 13 can be changed according to the cycle load fluctuation. . As a result, the ejector-type refrigeration cycle can be operated while exhibiting a high COP even if load fluctuation occurs.

(H) Since the decompressed and expanded refrigerant at a temperature expansion valve 17 (c ₂ points in FIG. 2) is a gas-liquid two-phase state, and flows into the gas-liquid two-phase refrigerant to the nozzle section 13a of the ejector 13 be able to.

  Therefore, the boiling of the refrigerant in the nozzle portion 13a can be promoted and the nozzle efficiency can be improved as compared with the case where the liquid phase refrigerant is caused to flow into the nozzle portion 13a. As a result, the amount of recovered energy can be increased and the amount of pressure increase in the diffuser section 13d 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 with respect to the case where the liquid phase refrigerant is caused to flow into the nozzle part 13a, the processing of the nozzle part 13a is facilitated. As a result, the manufacturing cost of the ejector 13 can be reduced, and the manufacturing cost of the ejector refrigeration cycle 10 as a whole can be reduced.

(Second Embodiment)
In the present embodiment, as shown in the overall configuration diagram of FIG. 3, the arrangement of the temperature type expansion valve 17 is different from that of the first embodiment in that the refrigerant path extends from the radiator 12 outlet side to the first branch portion 18 inlet side. An example changed to is described. In FIG. 3, the same or equivalent parts as those in the first embodiment are denoted by the same reference numerals. The same applies to the following drawings. Other configurations are the same as those of the first embodiment.

  Next, the operation of the ejector refrigeration cycle 10 of the present embodiment will be described with reference to 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 type refrigeration cycle 10 of this embodiment is operated, the refrigerant flowing out of the radiator 12 is decompressed and expanded in an enthalpy manner by the temperature type expansion valve 17 in any operation mode, and is in a gas-liquid two-phase state. becomes (b ₄ points of Figure 4 → c ₄ points), and flows into the first branch portion 18.

In the high-pressure branch operation mode and the simultaneous branch operation mode, the high-pressure refrigerant that has flowed into the first electric expansion valve 19 from the first branch portion 18 is decompressed and expanded in an equal enthalpy manner by the first electric expansion valve 19. Then, the pressure is reduced (c ₄ point → hβ ₄ point indicated by a broken line in FIG. 4). Other operations are the same as those in the first embodiment.

  Therefore, also in the configuration of the present embodiment, the same effects as (A) to (H) of the first embodiment can be obtained.

(Third embodiment)
In this embodiment, as shown in the overall configuration diagram of FIG. 5, internal heat that causes the refrigerant flowing out of the radiator 12 and the low-pressure side refrigerant of the cycle to exchange heat with respect to the ejector refrigeration cycle 10 of the first embodiment. An example in which the exchanger 30 is added will be described.

  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 refrigerant that has flowed out of the radiator 12 in the present embodiment is a refrigerant that circulates in the refrigerant passage from the radiator 12 outlet side to the first branch portion 18 inlet side, and the low-pressure side refrigerant of the cycle is This refrigerant is sucked into the second compression mechanism 21a.

  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, when the ejector refrigeration cycle 10 of this embodiment is operated, as shown in the Mollier diagram of FIG. 6, the operation of the internal heat exchanger 30 causes the operation of the first embodiment in any operation mode. Thus, the enthalpy of the refrigerant on the suction side of the second compression mechanism 21a is increased (point i ₆ → i ′ _{6 in} FIG. 6), and the enthalpy of the refrigerant flowing into the first branch portion 18 is decreased (point b _{6 in} FIG. 6). → b ' ₆ points). Other operations are the same as those in the first embodiment in any operation mode.

  Therefore, also in the configuration of the present embodiment, the same effects as (A) to (H) of the first embodiment can be obtained. Furthermore, compared to the first embodiment, the enthalpy of the refrigerant flowing into the outflow side evaporator 14 and the suction side evaporator 16 can be reduced by the operation of the internal heat exchanger 30 in any operation mode.

  As a result, since the enthalpy difference between the enthalpy of the inlet side refrigerant and the enthalpy of the outlet side refrigerant of the outflow side evaporator 14 and the suction side evaporator 16 can be increased to increase the refrigeration capacity, the COP is further improved. it can.

(Fourth embodiment)
In the present embodiment, as shown in the overall configuration diagram of FIG. 7, internal heat that causes the refrigerant flowing out of the radiator 12 and the low-pressure side refrigerant of the cycle to exchange heat with respect to the ejector refrigeration cycle 10 of the first embodiment. An example in which the exchanger 31 is added will be described.

  The internal heat exchanger 31 performs heat exchange between the refrigerant flowing out of the radiator 12 passing through the high-pressure side refrigerant flow path 31a and the low-pressure side refrigerant of the cycle passing through the low-pressure side refrigerant flow path 31b. The basic configuration of the internal heat exchanger 31 is the same as that of the internal heat exchanger 30 of the third embodiment.

  More specifically, the refrigerant that has flowed out of the radiator 12 in the present embodiment is a refrigerant that circulates in the refrigerant passage from the first branching portion 18 outlet side to the first electric expansion valve 19 inlet side, and the low pressure of the cycle The side refrigerant is a refrigerant sucked into the second compression mechanism 21a. Other configurations are the same as those of the first embodiment.

Next, when the ejector refrigeration cycle 10 of the present embodiment is operated, as shown in the Mollier diagram of FIG. 8, the first heat exchanger 31 operates in the high-pressure branch operation mode and the simultaneous branch operation mode by the action of the internal heat exchanger 31. Compared to the embodiment, the enthalpies of the second compression mechanism 21a suction-side refrigerant and discharge-side refrigerant are increased (i ₈ point → i ′ ₈ point → j ′ ₈ point in FIG. 8), and the first electric expansion valve 19 is reached. The enthalpy of the flowing refrigerant decreases (b ₈ → b ′ ₈ points in FIG. 8). Other operations are the same as those in the first embodiment.

  Therefore, also in the configuration of the present embodiment, the same effects as (A) to (H) of the first embodiment can be obtained. Furthermore, since the enthalpy difference between the enthalpy of the inlet side refrigerant and the enthalpy of the outlet side refrigerant of the outflow side evaporator 14 and the suction side evaporator 16 can be increased and the refrigeration capacity can be increased as in the second embodiment. COP can be further improved.

  Furthermore, in the present embodiment, in the internal heat exchanger 31, the high-pressure refrigerant flowing through the refrigerant passage from the first branch portion 18 outlet side to the first electric expansion valve 19 inlet side and the second compression mechanism 21a are sucked. Since heat exchange with the low-pressure refrigerant is performed, the enthalpy of the refrigerant flowing from the first branch portion 18 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 described in more detail. As the enthalpy of the refrigerant flowing into the nozzle portion 13a increases, the slope of the isentropic line becomes gentle. 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 portion 13d can be increased, and a further COP improvement effect can be obtained.

(Fifth embodiment)
In the present embodiment, as shown in the overall configuration diagram of FIG. 9, internal heat that causes the refrigerant flowing out of the radiator 12 and the low-pressure side refrigerant of the cycle to exchange heat with respect to the ejector refrigeration cycle 10 of the first embodiment. An example in which the exchanger 32 is added will be described.

  The internal heat exchanger 32 performs heat exchange between the refrigerant flowing out of the radiator 12 passing through the high-pressure side refrigerant flow path 32a and the low-pressure side refrigerant of the cycle passing through the low-pressure side refrigerant flow path 32b. The basic configuration of the internal heat exchanger 32 is the same as that of the internal heat exchanger 30 of the third embodiment.

  More specifically, the refrigerant that has flowed out of the radiator 12 in the present embodiment is a refrigerant that circulates in the refrigerant passage from the radiator 12 outlet side to the first branch portion 18 inlet side, and the low-pressure side refrigerant of the cycle is The refrigerant is sucked into the first compression mechanism 11a. Other configurations are the same as those of the first embodiment.

Next, when the ejector refrigeration cycle 10 of this embodiment is operated, as shown in the Mollier diagram of FIG. 10, the operation of the internal heat exchanger 32 causes the operation of the first embodiment in any operation mode. Te, the enthalpy of the first compression mechanism 11a suction side refrigerant and the discharge-side refrigerant is increased (g ₁₀ points in FIG. 10 → g _'10 points), the enthalpy of the refrigerant flowing into the first branch portion 18 is reduced (FIG. 10 B ₁₀ points → b ′ ₁₀ points). Other operations are the same as those in the first embodiment.

  Therefore, also in the configuration of the present embodiment, the same effects as (A) to (H) of the first embodiment can be obtained. Furthermore, since the enthalpy difference between the enthalpy of the inlet side refrigerant and the enthalpy of the outlet side refrigerant of the outflow side evaporator 14 and the suction side evaporator 16 can be increased and the refrigeration capacity can be increased as in the second embodiment. COP can be further improved.

(Sixth embodiment)
In the present embodiment, as shown in the overall configuration diagram of FIG. 11, internal heat that causes the refrigerant flowing out of the radiator 12 and the low-pressure side refrigerant of the cycle to exchange heat with respect to the ejector refrigeration cycle 10 of the first embodiment. An example in which the exchanger 33 is added will be described.

  The internal heat exchanger 33 performs heat exchange between the refrigerant flowing out of the radiator 12 passing through the high-pressure side refrigerant flow path 33a and the low-pressure side refrigerant of the cycle passing through the low-pressure side refrigerant flow path 33b. The basic configuration of the internal heat exchanger 33 is the same as that of the internal heat exchanger 30 of the third embodiment.

  More specifically, the refrigerant that has flowed out of the radiator 12 in the present embodiment is a refrigerant that circulates in the refrigerant passage from the first branching portion 18 outlet side to the first electric expansion valve 19 inlet side, and the low pressure of the cycle The side refrigerant is a refrigerant that is sucked into the first compression mechanism 11a. Other configurations are the same as those of the first embodiment.

Next, when the ejector refrigeration cycle 10 of the present embodiment is operated, as shown in the Mollier diagram of FIG. 12, the first heat exchanger 33 operates in the high-pressure branch operation mode and the simultaneous branch operation mode by the action of the internal heat exchanger 33. to the embodiments, the enthalpy of the first compression mechanism 11a suction side refrigerant and the discharge-side refrigerant is increased (g ₁₂ points in FIG. 12 → g _'12 points), the enthalpy of the refrigerant flowing into the first electric expansion valve 19 Decreases (b ₁₂ points → b ′ ₁₂ points in FIG. 12). Other operations are the same as those in the first embodiment.

  Therefore, also in the configuration of the present embodiment, the same effects as (A) to (H) of the first embodiment can be obtained. Furthermore, the same COP improvement effect as in the fourth embodiment can be obtained.

(Seventh embodiment)
In the present embodiment, an example in which the configuration of the radiator 12 is changed with respect to the ejector refrigeration cycle 10 of the first embodiment as illustrated in the overall configuration diagram of FIG. 13 will be described.

  Specifically, the radiator 12 of the present embodiment includes a condensing unit 12b that condenses the refrigerant, a gas-liquid separation unit 12c (receiver unit) that separates the gas-liquid of the refrigerant that has flowed out of the condensing unit 12b, and a gas-liquid separation. This is configured as a so-called subcool condenser having a supercooling part 12d for supercooling the liquid-phase refrigerant flowing out from the part 12c. Other configurations are the same as those of the first embodiment.

When the ejector type refrigeration cycle 10 of the present embodiment is operated, as shown in the Mollier diagram of FIG. 14, in any operation mode, the refrigerant condensed in the condenser 12b of the radiator 12 is converted into the gas-liquid separator 12c. Gas-liquid separation. Further, the saturated liquid-phase refrigerant separated by the gas-liquid separator 12c is supercooled of at supercooling part 12d (b ₁₄ points in FIG. 14 → b _'14 points). Other operations are the same as those in the first embodiment.

  Therefore, also in the configuration of the present embodiment, the same effects as (A) to (H) of the first embodiment can be obtained. Furthermore, with respect to the first embodiment, the enthalpy of the refrigerant flowing into the outflow side evaporator 14 and the suction side evaporator 16 can be reduced in any operation mode.

  As a result, since the enthalpy difference between the enthalpy of the inlet side refrigerant and the enthalpy of the outlet side refrigerant of the outflow side evaporator 14 and the suction side evaporator 16 can be increased to increase the refrigeration capacity, the COP is further improved. it can.

Furthermore, for example, as in the case of using the internal heat exchanger 30 of the third embodiment, the enthalpy of the second compression mechanism 21a suction-side refrigerant (low-pressure side refrigerant of the cycle) is not unnecessarily increased ( ₁₄ points in FIG. 14). Therefore, it is possible to reduce the refrigerant evaporation pressure (refrigerant evaporation temperature) in the suction-side evaporator 16 with respect to the third embodiment by suppressing the density of the refrigerant sucked by the second compression mechanism 21a. .

(Eighth embodiment)
In each of the above-described embodiments, an example in which a normal chlorofluorocarbon refrigerant is employed as the refrigerant and a subcritical refrigeration cycle is configured has been described. However, in this embodiment, carbon dioxide is employed as the refrigerant, and the first compressor 11 discharges. An example in which a supercritical refrigeration cycle in which the refrigerant pressure is equal to or higher than the critical pressure of the refrigerant will be described. Furthermore, in the present embodiment, as shown in the overall configuration diagram of FIG. 15, the temperature type expansion valve 17 is eliminated from the first embodiment. Other configurations are the same as those of the first embodiment.

Next, the operation of the ejector refrigeration cycle 10 of this embodiment will be described with reference to the Mollier diagram of FIG. When the ejector refrigeration cycle 10 of this embodiment is operated, the refrigerant discharged from the first compressor 11 is radiated by the radiator 12 and cooled. At this time, the refrigerant passing through the radiator 12 dissipates heat in a supercritical state without condensing (a ₁₆ point → b ₁₆ point in FIG. 16).

The refrigerant that has flowed out of the radiator 12 flows into the first branch portion 18, and in the high-pressure branch operation mode and the simultaneous branch operation mode, the refrigerant flow that flows into the nozzle portion 13 a side of the ejector 13 and the first temperature type expansion valve 19. refrigerant flow divided into two parts flowing to the side (b ₁₆ points in FIG. 16). High-pressure refrigerant in the supercritical state flowing from the first branch portion 18 to the nozzle part 13a side is isentropically depressurized expanded in the nozzle portion 13a (b ₁₆ points in FIG. 16 → d ₁₆ points).

On the other hand, in the high-pressure branch operation mode and the simultaneous branch operation mode, the supercritical high-pressure refrigerant that has flowed out from the first branch portion 18 to the first temperature expansion valve 19 side is equivalently enthalpy in the first temperature expansion valve 19. reduced pressure is expanded to reduce its pressure (b ₁₆ points in FIG. ₁₆ → hβ 16 points). Other operations are the same as those in the first embodiment. Therefore, also in the configuration of the present embodiment, the same effects as (A) to (F) of the first embodiment can be obtained.

Further, in the supercritical refrigeration cycle, the high-pressure side refrigerant pressure is higher than that in the subcritical refrigeration cycle, so the high-low pressure difference in the cycle (the pressure difference between point b _{16 and} point d _{16 in} FIG. 16) increases, and the ejector 13 The amount of pressure reduction in the nozzle portion 13a increases. Thereby, since the difference (recovered energy amount) between the enthalpy of the refrigerant on the inlet side of the nozzle part 13a and the enthalpy of the refrigerant on the outlet side of the nozzle part 13a is increased, COP can be further improved.

(Ninth embodiment)
In the ejector refrigeration cycle 40 of the present embodiment, as shown in the overall configuration diagram of FIG. 17, a third configuration having the same configuration as that of the first branch portion 18 of the first embodiment is provided on the discharge port side of the first compressor 11. A branch portion 38 is arranged. The first radiator 121 is connected to one refrigerant outlet of the third branch portion 38, and the second radiator 122 is connected to the other refrigerant outlet.

  The first radiator 121 exchanges heat between the high-pressure refrigerant that has flowed out from one of the refrigerant outlets of the third branch portion 38 and the outside air (outside air) that is blown by the cooling fan 121a, thereby radiating the high-pressure refrigerant. It is a heat exchanger for heat dissipation that cools. The first heat radiator 122 exchanges heat between the high-pressure refrigerant that has flowed out from the other refrigerant outlet of the third branch portion 38 and the outside air (outside air) that is blown by the cooling fan 122a. It is a heat dissipation heat exchanger that dissipates heat and cools it.

  Furthermore, in the ejector type refrigeration cycle 40 of the present embodiment, the heat exchange capacity (heat radiation performance) of the first radiator 121 is reduced by reducing the heat exchange area of the first radiator 121 relative to the second radiator. The heat exchange capacity of the second radiator 122 is lowered. The cooling fans 121a and 122a are electric blowers in which the rotation speed (the amount of blown air) is controlled by a control voltage output from the control device.

  Moreover, the high pressure side gas-liquid separation which isolate | separates the gas-liquid of the refrigerant | coolant which flowed out from the 1st, 2nd heat radiator 121,122, respectively and stored the excess liquid phase refrigerant | coolant in the exit side of the 1st, 2nd heat radiator 121,122. You may provide the receiver (liquid receiver) as a container. Then, the saturated liquid phase refrigerant separated from the receiver may be led to the downstream side.

  On the outlet side of the first radiator 121, a temperature type expansion valve 17 is connected as high pressure side pressure reducing means similar to that of the first embodiment. Furthermore, the outlet side of the temperature type expansion valve 17 is connected to the inlet side of the nozzle portion 13a of the ejector 13 similar to that of the first embodiment. Further, an outlet-side evaporator 14 is connected to the outlet side of the diffuser portion 13d of the ejector 13.

  On the other hand, a fixed throttle 39 as a suction side pressure reducing means is connected to the outlet side of the second radiator 122. As the fixed throttle 39, specifically, an orifice or a capillary tube can be adopted. The suction side evaporator 16 similar to the first embodiment is connected to the outlet side of the fixed throttle 39. Other configurations are the same as those of the ejector refrigeration cycle 10 of the first embodiment.

Next, the operation of the present embodiment in the above configuration will be described with reference to 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, and the blower fans 14a and 16a. Thereby, the 1st compressor 11 suck | inhales a refrigerant | coolant, compresses and discharges. The state of the refrigerant at this time is point a ₁₈ in FIG.

The high-temperature and high-pressure gas-phase refrigerant discharged from the first compressor 11 flows into the third branching section 38, the refrigerant flow flowing into the first radiator 121 side, and the refrigerant flow flowing into the second radiator 122 side divided into two parts (a ₁₈ point in FIG. 18).

  Here, in this embodiment, the flow rate ratio Gr1 / Gr2 between the refrigerant flow rate Gr1 flowing into the first radiator 121 side and the refrigerant flow rate Gr2 flowing into the second radiator 122 side can exhibit a high COP as a whole cycle. The passage areas (pressure loss characteristics) of the respective refrigerant passages in the third branch portion 38 are determined so as to obtain the optimum flow rate ratio.

Refrigerant flowing into the first radiator 121 side, the cooling blower has been blown air from the fan 121a (outside air) and heat exchanger to be condensed heat radiation (a ₁₈ point of FIG. 18 → b1 ₁₈ points). The refrigerant having flowed into the second radiator 122 side, the cooling blower has been blown air from the fan 122a (outside air) and heat exchanger to be condensed heat radiation (a ₁₈ point of FIG. 18 → b2 ₁₈ points).

  At this time, since the heat exchange capability of the first radiator 121 is set lower than the heat exchange capability of the second radiator 122, the enthalpy of the refrigerant flowing out from the first radiator 121 is the second radiator 122. It becomes higher than the enthalpy of the refrigerant that has flowed out.

The refrigerant that has flowed out of the first radiator 121 flows into the temperature type expansion valve 17 and is decompressed and expanded in an enthalpy manner to a gas-liquid two-phase state (b1 ₁₈ point → c ₁₈ point in FIG. 18). At this time, the valve opening degree of the thermal expansion valve 17, the discharge side evaporator 14 outlet superheat degree of the refrigerant (g ₁₈ points in FIG. 18) is adjusted to be predetermined value.

The intermediate-pressure refrigerant flowing out of the thermal expansion valve 17 isentropically depressurized expanded in the nozzle portion 13a (c ₁₈ points in FIG. 18 → d ₁₈ points). Then, the refrigerant injected from the nozzle portion 13a and the suction refrigerant sucked from the refrigerant suction port 13b are mixed in the mixing portion 13c of the ejector 13 (d ₁₈ point → e ₁₈ point, j ₁₈ point in FIG. 18 → e ₁₈ points), is boosted by the diffuser part 13d (e ₁₈ points in FIG. 18 → f ₁₈ points).

Refrigerant flowing out of the diffuser unit 13d, and flows into the discharge side evaporator 14, and absorbs heat from the circulation blower has been refrigerator air by the blower fan 14a to evaporate (f ₁₈ points in FIG. 18 → g ₁₈ points). Thereby, the air in a refrigerator is cooled. Then, the refrigerant that has flowed out of the outflow side evaporator 14 is sucked into the first compressor 11 and compressed again (g ₁₈ point → a ₁₈ point in FIG. 18).

On the other hand, the refrigerant that has flowed out of the second radiator 122 is further decompressed and expanded in an isenthalpy manner at the fixed throttle 39, and the pressure is reduced (b2 ₁₈ point → h ₁₈ point in FIG. 18). Decompressed and expanded refrigerant at a fixed throttle 39, and flows into the suction side evaporator 16, absorbs heat from the freezer in air circulated blown by the blower fan 16a to evaporate (h ₁₈ points in FIG. 18 → i ₁₈ points ). Thereby, the air in a warehouse is cooled.

Then, the refrigerant that has flowed out of the suction side evaporator 16 is sucked into the second compressor 21 and compressed (i ₁₈ point → j ₁₈ point in FIG. 18), and enters the ejector 13 from the refrigerant suction port 13 b of the ejector 13. Suction is performed (j ₁₈ point → e ₁₈ point in FIG. 18). Other operations such as control of the pressure increase amounts of the first and second compression mechanisms 11a and 21a are the same as those in the first embodiment.

  Since the ejector refrigeration cycle 10 of this embodiment operates as described above, it is possible to obtain the same effects as (A), (B), (D), and (F) to (H) of the first embodiment. it can.

  Furthermore, since the heat dissipation performance of the first radiator 121 and the second radiator 122 can be changed independently, for example, the heat dissipation performance of the second radiator 122 and the heat absorption performance of the suction side evaporator 16 are easily adapted. And the heat dissipation performance of the first and second radiators 121 and 122 and the heat absorption performance of the outflow side evaporator 14 can be easily matched. Therefore, it is easy to stabilize the operation of the cycle.

  Moreover, since the heat exchange capability of the first radiator 121 is lower than the heat exchange capability of the second radiator (122), the enthalpy of the refrigerant flowing into the nozzle portion 13a of the ejector 13 is unnecessarily reduced. Can be avoided. Thereby, the recovery energy amount in the nozzle part 13a can be increased, and the COP improvement effect similar to 4th Embodiment can be acquired.

(10th Embodiment)
In the present embodiment, as shown in the overall configuration diagram of FIG. 19, heat is exchanged between the refrigerant flowing out of the second radiator 122 and the low-pressure side refrigerant of the cycle with respect to the ejector refrigeration cycle 10 of the ninth embodiment. An example in which the internal heat exchanger 34 is added will be described.

  The internal heat exchanger 34 exchanges heat between the refrigerant that has flowed out of the second radiator 122 that passes through the high-pressure side refrigerant flow path 34a and the low-pressure side refrigerant of the cycle that passes through the low-pressure side refrigerant flow path 34b. is there. The basic configuration of the internal heat exchanger 34 is the same as that of the internal heat exchanger 30 of the third embodiment. More specifically, the low-pressure side refrigerant of the cycle in the present embodiment is a refrigerant sucked into the second compression mechanism 21a. Other configurations are the same as those of the ninth embodiment.

Next, when the ejector refrigeration cycle 10 of this embodiment is operated, as shown in the Mollier diagram of FIG. 20, the second compression mechanism 21 a is compared with the ninth embodiment by the action of the internal heat exchanger 34. The enthalpy of the suction side refrigerant increases (i ₂₀ point in FIG. 20 → i ′ ₂₀ point), and the enthalpy of the refrigerant flowing into the fixed throttle 39 decreases (b2 ₂₀ → b2 ′ ₂₀ point in FIG. 20). Other operations are the same as those in the ninth embodiment.

  Therefore, also in the configuration of the present embodiment, the same effect as in the ninth embodiment can be obtained. Furthermore, the enthalpy of the refrigerant flowing into the suction side evaporator 16 can be reduced by the action of the internal heat exchanger 34. Thereby, since the enthalpy difference of the enthalpy of the inlet side refrigerant | coolant of the suction side evaporator 16 and the enthalpy of the outlet side refrigerant | coolant can be expanded and a refrigerating capacity can be increased, COP can be improved further.

(Eleventh embodiment)
In the present embodiment, as shown in the overall configuration diagram of FIG. 21, heat is exchanged between the refrigerant flowing out of the second radiator 122 and the low-pressure side refrigerant of the cycle with respect to the ejector refrigeration cycle 10 of the ninth embodiment. An example in which the internal heat exchanger 35 is added will be described.

  The internal heat exchanger 35 performs heat exchange between the refrigerant that has flowed out of the radiator 12 that passes through the high-pressure side refrigerant flow path 35a and the low-pressure side refrigerant of the cycle that passes through the low-pressure side refrigerant flow path 35b. The basic configuration of the internal heat exchanger 35 is the same as that of the internal heat exchanger 30 of the third embodiment. More specifically, the low-pressure side refrigerant of the cycle in the present embodiment is a refrigerant sucked into the first compression mechanism 11a. Other configurations are the same as those of the ninth embodiment.

Next, when the ejector refrigeration cycle 10 of the present embodiment is operated, as shown in the Mollier diagram of FIG. 22, the first compression mechanism 11a is compared with the ninth embodiment by the action of the internal heat exchanger 35. enthalpy of the suction-side refrigerant is increased (g ₂₂ points in FIG. 22 → g _'22 points), the enthalpy of the refrigerant flowing into the fixed throttle 39 is reduced (b2 ₂₂ points in FIG. 22 → b2' ₂₂ points). Other operations are the same as those in the ninth embodiment.

  Therefore, also in the configuration of the present embodiment, the same effect as in the ninth embodiment can be obtained. Furthermore, by the action of the internal heat exchanger 35, as in the tenth embodiment, the enthalpy difference between the enthalpy of the inlet side refrigerant and the enthalpy of the outlet side refrigerant of the suction side evaporator 16 is increased to increase the refrigerating capacity. Therefore, COP can be further improved.

(Twelfth embodiment)
In the present embodiment, an example in which the configuration of the second radiator 122 is changed with respect to the ejector refrigeration cycle 10 of the ninth embodiment will be described as shown in the overall configuration diagram of FIG.

  Specifically, the basic configuration of the second radiator 122 of the present embodiment is the same as that of the radiator 12 of the seventh embodiment. Therefore, the second radiator 122 of the present embodiment is a so-called subcooled condenser having a condensing unit 122b, a gas-liquid separating unit 122c (receiver unit), and a supercooling unit 122d. Other configurations are the same as those of the ninth embodiment.

When the ejector refrigeration cycle 10 of this embodiment is operated, as shown in the Mollier diagram of FIG. 24, the refrigerant condensed in the condensing part 122b of the second radiator 122 is separated into gas and liquid in the gas-liquid separating part 122c. Is done. Furthermore, the saturated liquid phase refrigerant separated by the gas-liquid separation unit 122c is supercooled by the supercooling unit 122d (b2 ₂₄ point → b2 ′ ₂₄ point in FIG. 24). Other operations are the same as those in the ninth embodiment.

  Therefore, also in the configuration of the present embodiment, the same effect as in the ninth embodiment can be obtained. Furthermore, since the enthalpy of the refrigerant flowing into the suction side evaporator 16 can be reduced, the enthalpy difference between the enthalpy of the inlet side refrigerant and the enthalpy of the outlet side refrigerant of the suction side evaporator 16 is increased to increase the refrigerating capacity. Therefore, COP can be further improved.

  Furthermore, similarly to the seventh embodiment, for example, the refrigerant evaporation pressure (refrigerant evaporation temperature) in the suction-side evaporator 16 can be reduced compared to the tenth embodiment.

(13th Embodiment)
In the present embodiment, as shown in the overall configuration diagram of FIG. 25, the temperature type expansion valve 17 is abolished with respect to the configuration of the ninth embodiment, and carbon dioxide is used as a refrigerant as in the eighth embodiment. An example in which a supercritical refrigeration cycle is configured will be described. Other configurations are the same as those of the ninth embodiment.

When the ejector type refrigeration cycle 10 of the present embodiment is operated, as shown in the Mollier diagram of FIG. 26, the refrigerant discharged from the first compressor 11 is branched at the third branch portion 38, and each of the branched refrigerants The first and second radiators 121 and 122 dissipate heat and cool. At this time, the refrigerant passing through the first and second radiators 121 and 122 radiates heat in a supercritical state without condensing (a ₂₆ point → b 1 ₂₆ point, a ₂₆ point → b 2 ₂₆ point in FIG. ₂₆ ).

The supercritical high-pressure refrigerant flowing out from the first radiator 121 is decompressed and expanded in an isentropic manner at the nozzle portion 13a (b1 ₂₆ point → d ₂₆ point in FIG. 26). On the other hand, the high-pressure supercritical refrigerant flowing out of the second radiator 122 is isenthalpic depressurize and expanded by the fixed throttle 39 reduces the pressure (b2 ₂₆ points in FIG. 26 → h ₂₆ points). Other operations are the same as those in the ninth embodiment.

  Therefore, also in the configuration of this embodiment, the same effects as (A), (B), (D), and (F) of the first embodiment can be obtained. Furthermore, as in the eighth embodiment, it is possible to obtain a COP improvement effect by increasing the amount of recovered energy in the nozzle portion 13a.

(14th Embodiment)
In the present embodiment, an example in which the ejector refrigeration cycle of the present invention is applied to a refrigeration apparatus will be described. FIG. 27 is an overall configuration diagram of the ejector refrigeration cycle 50 of the present embodiment.

  The ejector-type refrigeration cycle 50 includes two first and second heat exchangers 51 and 52, and among these, a first operation mode in which the first heat exchanger 51 cools the internal air that is a heat exchange target fluid, The second heat exchanger 52 is configured to be switchable between a second operation mode for cooling the internal air. In addition, the solid line arrow in FIG. 27 shows the flow of the refrigerant in the first operation mode, and the broken line arrow shows the flow of the refrigerant in the second operation mode.

  The operation mode is switched by switching the refrigerant flow path of the ejector refrigeration cycle 50 by the first and second electric four-way valves 53 and 54 that are refrigerant flow switching means. The operations of the first and second electric four-way valves 53 and 54 are controlled by control signals output from the control devices, respectively.

  The first electric four-way valve 53 is connected to the outlet side of the radiator 12 having the same configuration as that of the first embodiment. Then, the outlet side of the radiator 12 and the inlet side of the first heat exchanger 51 and the liquid phase refrigerant outlet side of the accumulator 55 described later and the inlet side of the second heat exchanger 52 are simultaneously connected. Between the refrigerant flow path (circuit shown by the solid line arrow in FIG. 27), the outlet side of the radiator 12 and the inlet side of the second heat exchanger 52, and the liquid-phase refrigerant outlet side of the accumulator 55 and the first heat exchanger. The refrigerant flow paths (circuits indicated by broken line arrows in FIG. 27) that are connected simultaneously to the inlet side of 51 are switched.

  The accumulator 55 is an outflow side gas-liquid separator that separates the gas-liquid refrigerant flowing out from the diffuser portion 13d of the ejector 13 and stores excess refrigerant. Further, in the present embodiment, a fixed throttle 59 that decompresses and expands the refrigerant flowing out to the first electric four-way valve 53 side is disposed between the liquid-phase refrigerant outlet of the accumulator 55 and the first electric four-way valve 53. ing. The basic configuration of the fixed aperture 59 is the same as that of the fixed aperture 39 of the ninth embodiment.

  On the other hand, the second electric four-way valve 54 is connected to the outlet side of the first and second heat exchangers 51 and 52. Then, the outlet side of the first heat exchanger 51 and the nozzle part 13a inlet side of the ejector 13 and the outlet side of the second heat exchanger 52 and the inlet side of the second compressor 21 are simultaneously connected. 27 between the refrigerant flow path (circuit shown by the solid line arrow in FIG. 27), the outlet side of the second heat exchanger 52 and the nozzle portion 13a inlet side of the ejector 13, and the outlet side of the first heat exchanger 51 and the second compression. The refrigerant flow path (circuit shown by the broken line arrow in FIG. 27) that connects the suction port side of the machine 21 at the same time is switched.

  Further, in the present embodiment, a temperature type expansion valve 57 as a high pressure side pressure reducing means for decompressing and expanding the refrigerant flowing into the nozzle portion 13a is disposed between the second electric four-way valve 54 and the nozzle portion 13a of the ejector 13. Has been.

  The temperature type expansion valve 57 has a temperature sensing part (not shown) arranged in the suction side refrigerant passage of the second compressor 21 of the ejector 13, and the temperature and pressure of the second compressor 21 suction side refrigerant. Based on the above, a variable throttle that detects the degree of superheat of the refrigerant on the suction side of the second compressor 21 and adjusts the valve opening (refrigerant flow rate) by a mechanical mechanism so that the degree of superheat becomes a predetermined value set in advance. Mechanism.

  The first and second heat exchangers 51 and 52 are utilization-side heat exchangers that exchange heat between the refrigerant flowing into the interior and the internal air circulated by the first and second blower fans 51a and 52a. . The first and second blower fans 51a and 52a are electric blowers in which the number of rotations (the amount of blown air) is controlled by a control voltage output from the control device.

  The accumulator 55 is an outflow-side gas-liquid separator that separates the gas-liquid refrigerant flowing out from the diffuser portion 13d of the ejector 13 and accumulates excess refrigerant in the cycle. Furthermore, the gas-phase refrigerant outlet of the accumulator 55 is connected to the refrigerant inlet side of the first compressor 11. 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 with reference to the Mollier diagram of FIG. In the ejector refrigeration cycle 50 of the present embodiment, the inside of the freezer is continuously cooled by switching between the first operation mode and the second operation mode at predetermined times.

a. First Operation Mode In the first operation mode, the control device operates the first and second electric motors 11b and 21b, the cooling fan 12a, and the first and second blower fans 51a and 52a.

  Further, the control device simultaneously performs between the outlet side of the radiator 12 and the inlet side of the first heat exchanger 51 and between the liquid phase refrigerant outlet side of the accumulator 55 and the inlet side of the second heat exchanger 52. The first electric four-way valve 53 is switched so as to be connected, between the outlet side of the first heat exchanger 51 and the inlet side of the nozzle portion 13a of the ejector 13, and the outlet side of the second heat exchanger 52 and the second compressor. The second electric four-way valve 54 is switched so as to be simultaneously connected to the inlet 21 side.

  Thereby, as shown by the solid line arrow in FIG. 27, the first compressor 11 → the radiator 12 (→ the first electric four-way valve 53) → the first heat exchanger 51 (→ the second electric four-way valve 54) → The refrigerant circulates in the order of the temperature type expansion valve 57 → the nozzle portion 13 a of the ejector 13 → the gas phase refrigerant outlet of the accumulator 55 → the first compressor 11, and the liquid phase refrigerant outlet of the accumulator 55 → the fixed throttle 59 (→ first electric A cycle in which the refrigerant circulates in the order of the four-way valve 53) → the second heat exchanger 52 (→ the second electric four-way valve 54) → the second compressor 21 → the refrigerant suction port 13b of the ejector 13 → the accumulator 55. .

Therefore, the refrigerant compressed by the first compression mechanism 11a is (a ₂₈ point in FIG. 28), the radiator 12 at is cooled by outside air heat exchanger which is blown by the blower fan 12a (a in Fig. 28 ₂₈ Point → b ₂₈ ), and flows into the first heat exchanger 51 via the first electric four-way valve 53.

Refrigerant flowing into the first heat exchanger 51 is cooled by exchanging the heat with air circulation blower has been refrigerator by the first blower fan 51a (b ₂₈ points in FIG. 28 → b _'28 points). At this time, the first heat exchanger 51 is defrosted.

The refrigerant that has flowed out of the first heat exchanger 51 flows into the temperature type expansion valve 57 through the second electric four-way valve 54. The refrigerant that has flowed into the temperature type expansion valve 57 is decompressed and expanded in an enthalpy manner to a gas-liquid two-phase state (b ′ ₂₈ point → c ₂₈ point in FIG. 2). At this time, the valve opening degree of the thermal expansion valve 17, the discharge side evaporator 14 outlet superheat degree of the refrigerant (g ₂₈ points in FIG. 28) is adjusted to be predetermined value.

The intermediate-pressure refrigerant flowing out of the thermal expansion valve 57, flows into the nozzle portion 13a of the ejector 13 is injected is decompressed and expanded isentropically to (c ₂₈ points in FIG. 2 → d ₂₈ points). Due to the suction action of the jet refrigerant, the refrigerant discharged from the second compression mechanism 21a is sucked from the refrigerant suction port 13b (j ₂₈ point → e ₂₈ point in FIG. 28).

Further, the refrigerant injected from the nozzle portion 13a and the refrigerant sucked from the refrigerant suction port 13b are mixed in the mixing portion 13c of the ejector 13, and the pressure is increased in the diffuser portion 13d (point e _{28 in} FIG. 28 → f ₂₈ points).

The refrigerant flowing from the diffuser portion 13d is separated into gas and liquid in the accumulator 55 (f ₂₈ points in FIG. 28 → g ₂₈ points, f ₂₈ points → g _'28 points), flows out from the gas phase refrigerant outlet of the accumulator 55 The vapor phase refrigerant thus sucked into the first compression mechanism 11a is compressed again. (G ₂₈ points → a ₂₈ points in Fig. 28)
On the other hand, the liquid-phase refrigerant that has flowed out from the liquid-phase refrigerant outlet of the accumulator 55 is further decompressed and expanded in an isenthalpy manner at the fixed restrictor 59 (g ′ ₂₈ point → h ₂₈ point in FIG. 28). It flows into the second heat exchanger 52 through the second heat exchanger 52 through 53. The refrigerant that has flowed into the second heat exchanger 52 absorbs heat from the indoor blown air circulated by the second blower fan 52a and evaporates (from point h _{28 to} point i _{28 in} FIG. 28). Thereby, the air in a warehouse is cooled.

  Then, the refrigerant flowing out of the second heat exchanger 52 is sucked into the second compression mechanism 21a via the second electric four-way valve 54 and compressed. At this time, similarly to the first embodiment, the control device controls the operations of the first and second electric motors 11b and 21b so that the COP as the entire ejector-type refrigeration cycle approaches the maximum.

  Therefore, in the first operation mode of the present embodiment, the refrigerant discharged from the first compression mechanism 11a is radiated by the radiator 12 and the first heat exchanger 51, and the refrigerant is evaporated by the second heat exchanger 52. The refrigerant flow path is switched. Thereby, the internal air can be cooled by the second heat exchanger 52 while defrosting the first heat exchanger 51.

b. Second Operation Mode In the second operation mode, the control device operates between the outlet side of the radiator 12 and the inlet side of the second heat exchanger 52 and the liquid phase refrigerant outlet side of the accumulator 55 and the first heat exchanger 52. The first electric four-way valve 53 is switched so as to be simultaneously connected to the inlet side of the first heat exchanger 51, and the first heat exchanger 51 is connected between the outlet side of the second heat exchanger 52 and the inlet side of the nozzle portion 13a of the ejector 13. The second electric four-way valve 54 is switched so that the outlet side of the second compressor 21 and the suction port side of the second compressor 21 are simultaneously connected.

  Thereby, as shown by the broken line arrow in FIG. 27, the first compressor 11 → the radiator 12 (→ the first electric four-way valve 53) → the second heat exchanger 52 (→ the second electric four-way valve 54) → The refrigerant circulates in the order of the temperature type expansion valve 57 → the nozzle portion 13 a of the ejector 13 → the gas phase refrigerant outlet of the accumulator 55 → the first compressor 11, and the liquid phase refrigerant outlet of the accumulator 55 → the fixed throttle 59 (→ first electric A cycle in which the refrigerant circulates in the order of the four-way valve 53) → the first heat exchanger 51 (→ the second electric four-way valve 54) → the second compressor 21 → the refrigerant suction port 13b of the ejector 13 → the accumulator 55. .

  Therefore, in the second operation mode of the present embodiment, contrary to the first operation mode, the refrigerant discharged from the first compression mechanism 11a is radiated by the radiator 12 and the second heat exchanger 52, and the first operation mode. The heat exchanger 51 switches to a refrigerant flow path for evaporating the refrigerant. For this reason, the state of the refrigerant in the second operation mode is the same as the Mollier diagram of FIG. Thereby, the inside air can be cooled by the first heat exchanger 51 while defrosting the second heat exchanger 52.

  Since the ejector refrigeration cycle 50 of this embodiment operates as described above, the same effects as (B), (G), and (H) of the first embodiment can be obtained. Furthermore, by alternately switching between the first operation mode and the second operation mode, the inside air is defrosted in either one of the first and second heat exchangers 51 and 52 while the other is in the chamber air Can be cooled. Therefore, the ejector refrigeration cycle 50 of the present embodiment can exhibit a continuously stable cooling capacity.

(Fifteenth embodiment)
In the present embodiment, as shown in the overall configuration diagram of FIG. 29, with respect to the ejector-type refrigeration cycle 50 of the fourteenth embodiment, the inside of the high-pressure refrigerant upstream of the nozzle portion 13a and the low-pressure refrigerant of the cycle exchanges heat. An example in which the heat exchanger 36 is added will be described. This internal heat exchanger 36 performs heat exchange between the high-pressure refrigerant upstream of the nozzle portion 13a passing through the high-pressure side refrigerant flow path 36a and the low-pressure side refrigerant of the cycle passing through the low-pressure side refrigerant flow path 30b. is there.

  The basic configuration of the internal heat exchanger 36 is the same as that of the internal heat exchanger 30 of the third embodiment. More specifically, the high-pressure refrigerant on the upstream side of the nozzle portion 13a in this embodiment is a refrigerant that circulates in the refrigerant passage from the second electric four-way valve 54 to the temperature-type expansion valve 57, and the low-pressure side refrigerant of the cycle is The refrigerant is sucked into the second compression mechanism 21a. Other configurations are the same as those in the fourteenth embodiment.

Next, when the ejector refrigeration cycle 50 of this embodiment is operated, the second compression mechanism 21a is compared with the fourteenth embodiment by the action of the internal heat exchanger 36 as shown in the Mollier diagram of FIG. The enthalpy of the suction side refrigerant increases (i ₃₀ point in FIG. 30 → i ′ ₃₀ point), and the enthalpy of the refrigerant flowing into the temperature type expansion valve 57 decreases (b ′ ₃₀ → b ″ ₃₀ point in FIG. 30). Other operations are the same as those in the fourteenth embodiment.

  Therefore, also in the configuration of the present embodiment, the same effect as that of the fourteenth embodiment can be obtained. Furthermore, in any of the operation modes, the enthalpy of the refrigerant flowing into the heat exchanger acting as an evaporator of the first and second heat exchangers 51 and 52 is reduced by the action of the internal heat exchanger 36. Can do.

  Thereby, since the enthalpy difference between the enthalpy of the inlet side refrigerant and the enthalpy of the outlet side refrigerant of the heat exchanger acting as an evaporator can be increased to increase the refrigeration capacity, the COP can be further improved.

(Sixteenth embodiment)
In the present embodiment, as shown in the overall configuration diagram of FIG. 31, with respect to the ejector refrigeration cycle 50 of the fourteenth embodiment, the internal heat exchange between the high-pressure refrigerant upstream of the nozzle portion 13 a and the low-pressure refrigerant of the cycle is performed. An example in which the heat exchanger 37 is added will be described. The internal heat exchanger 37 performs heat exchange between the high-pressure refrigerant on the upstream side of the nozzle portion 13a passing through the high-pressure side refrigerant flow path 37a and the low-pressure side refrigerant of the cycle passing through the low-pressure side refrigerant flow path 37b. is there.

  The basic configuration of the internal heat exchanger 36 is the same as that of the internal heat exchanger 30 of the third embodiment. More specifically, the high-pressure refrigerant on the upstream side of the nozzle portion 13a in this embodiment is a refrigerant that circulates in the refrigerant passage from the second electric four-way valve 54 to the temperature-type expansion valve 57, and the low-pressure side refrigerant of the cycle is The refrigerant is sucked into the second compression mechanism 21a. Other configurations are the same as those in the fourteenth embodiment.

Next, when the ejector-type refrigeration cycle 50 of this embodiment is operated, as shown in the Mollier diagram of FIG. 32, the first compression mechanism 11a is compared with the fourteenth embodiment by the action of the internal heat exchanger 37. The enthalpy of the suction side refrigerant increases (g ₃₂ point → g ′ ₃₂ point in FIG. 32), and the enthalpy of the refrigerant flowing into the temperature type expansion valve 57 decreases (b ′ ₃₂ → b ″ ₃₂ point in FIG. 32). Other operations are the same as those in the fourteenth embodiment.

  Therefore, also in the configuration of the present embodiment, the same effect as that of the fourteenth embodiment can be obtained. Further, in any operation mode, due to the action of the internal heat exchanger 37, the enthalpy difference between the enthalpy of the inlet side refrigerant and the enthalpy of the outlet side refrigerant of the heat exchanger acting as an evaporator is the same as in the fifteenth embodiment. COP can be further improved because the refrigerating capacity can be increased by expanding.

(17th Embodiment)
In the present embodiment, as shown in the overall configuration diagram of FIG. 33, the temperature type expansion valve 57 is abolished with respect to the configuration of the fourteenth embodiment, and carbon dioxide is used as a refrigerant as in the eighth embodiment. An example in which a supercritical refrigeration cycle is configured will be described. Other configurations are the same as those in the fourteenth embodiment.

When the ejector-type refrigeration cycle 50 of this embodiment is operated, the refrigerant discharged from the first compressor 11 is discharged from the radiator 12 and the first and second heat exchangers 51 and 52 as shown in the Mollier diagram of FIG. In the heat exchanger acting as a radiator, heat is released and cooled. At this time, the refrigerant passing through the radiator 12 and the heat exchanger acting as a radiator dissipates heat in a supercritical state without condensing (a _{34 points} → b _{34 points} → b ′ ₃₄ points in FIG. ₃₄ ).

The supercritical high-pressure refrigerant that has flowed out of the heat exchanger acting as a radiator among the first and second heat exchangers 51 and 52 flows into the nozzle portion 13a of the ejector through the second electric four-way valve 54. To do. Then, it is decompressed and expanded in an isentropic manner at the nozzle portion 13a (b ′ ₃₄ point → d ₃₄ point in FIG. 34). Other operations are the same as those in the fourteenth embodiment.

  Therefore, also in the configuration of the present embodiment, the same effect as that of the fourteenth embodiment can be obtained. Furthermore, as in the eighth embodiment, it is possible to obtain a COP improvement effect by increasing the amount of recovered energy in the nozzle portion 13a.

(Eighteenth embodiment)
In the present embodiment, as shown in the overall configuration diagram of FIG. 35 and the Mollier diagram of FIG. 36, with respect to the ejector refrigeration cycle 50 of the fourteenth embodiment, on the downstream side of the diffuser portion 13 d of the ejector 13, The example which has arrange | positioned the ventilation fan 14a which ventilates the fluid for heat exchange to the outflow side evaporator 14 and the outflow side evaporator 14 in the upstream of the accumulator 55 is demonstrated. Other configurations are the same as those in the fourteenth embodiment.

Operating the ejector-type refrigeration cycle 50 of this embodiment, not only operates similarly to the fourteenth embodiment, the outflow-side evaporator 14, as shown in the Mollier diagram of FIG. 36, f from f ₃₆ points 'The liquid phase refrigerant in the process up to ₃₆ points can be evaporated to exert an endothermic effect. Thereby, the blowing air from the blowing fan 14a can also be cooled.

  At this time, in the outflow side evaporator 14, the refrigerant evaporates at a temperature higher than the refrigerant evaporating temperature in the heat exchanger acting as an evaporator among the first and second heat exchangers 51 and 52. That is, of the first and second heat exchangers 51 and 52, the refrigerant evaporates in different temperature zones in the heat exchanger acting as an evaporator and the outflow side evaporator 14.

  Thereby, in this embodiment, while being able to acquire the same effect as 1st Embodiment, the store | warehouse | chamber in the refrigerator which preserve | saves food, a drink, etc. at the low temperature of 0 to 10 degreeC with the outflow side evaporator 14, for example. The internal air can also be cooled. Of course, the outflow evaporator 14 may be added to the ejector refrigeration cycle 50 of the fourteenth to seventeenth embodiments.

(Nineteenth embodiment)
In the present embodiment, as shown in the overall configuration diagram of FIG. 37, an internal heat exchanger 35 similar to that of the eleventh embodiment is added to the ejector refrigeration cycle 10 of the ninth embodiment, and the outflow side evaporation is performed. The example which abolished the apparatus 14 and the ventilation fan 14a is demonstrated.

  The internal heat exchanger 35 of the present embodiment exchanges heat between the high-pressure side refrigerant that flows through the refrigerant passage from the outlet side of the second radiator 122 to the inlet side of the fixed throttle 19 and the refrigerant sucked into the first compression mechanism 11a. is there. Other configurations are the same as those of the ninth embodiment.

When the ejector refrigeration cycle 10 of this embodiment is operated, the refrigerant flowing out of the diffuser portion 13d is caused to flow from the low pressure side of the internal heat exchanger 35 by the action of the internal heat exchanger 35 as shown in the Mollier diagram of FIG. It evaporates in the refrigerant flow path 35b, and the enthalpy of the first compression mechanism 11a suction-side refrigerant increases (f ₃₈ point → g ₃₈ point in FIG. 38). Further, the enthalpy of the refrigerant flowing out from the second radiator 122 decreases (b ′ ₃₈ points → b ″ ₃₈ points in FIG. 38).

  Other operations are the same as those in the ninth embodiment. Therefore, in this embodiment, not only can the cooling action be exhibited by the suction-side evaporator 16, but the same effects as (B), (D), (F) to (H) of the first embodiment can be obtained. it can.

  Furthermore, the COP improvement effect by not reducing the enthalpy of the refrigerant | coolant which flows in into the nozzle part 13a similarly to 9th Embodiment can be acquired. Furthermore, the COP improvement effect by the internal heat exchanger 32 similar to the eleventh embodiment can be obtained.

(20th embodiment)
In this embodiment, as shown in the overall configuration diagram of FIG. 39, an accumulator 55 and a suction-side gas-liquid separator 55a similar to those of the fourteenth embodiment are added to the ejector refrigeration cycle 10 of the nineteenth embodiment. Is.

  The suction-side gas-liquid separator 55a separates the gas-liquid refrigerant flowing out from the suction-side evaporator 16 and stores excess liquid-phase refrigerant in the cycle. Further, the suction port of the second compressor 21 is connected to the gas-phase refrigerant outlet of the suction side gas-liquid separator 55a. Other configurations are the same as those in the nineteenth embodiment.

  Therefore, when the ejector type refrigeration cycle 10 of this embodiment is operated, it operates similarly to the nineteenth embodiment and not only exhibits the cooling action in the suction side evaporator 16, but also (B) of the first embodiment. , (F) to (H), and the same COP improvement effect as the twenty-sixth embodiment can be obtained.

  Furthermore, the problems of the liquid compression of the first and second compressors 11 and 21 can be avoided by the action of the accumulator 55 and the suction side gas-liquid separator 55a. In this embodiment, an example in which both the accumulator 55 and the suction-side gas-liquid separator 55a are provided has been described. However, either one may be provided.

(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 and second compressors 11 and 21 have been described as adopting separate compressors, but the first and second compression mechanisms 11a and 21a and The first and second electric motors 11b and 21b may be configured integrally.

  For example, the first and second compression mechanisms 11a and 21a and the first and second electric motors 11b and 21b may be accommodated in the same housing and integrally configured. In this case, the rotation shafts of the first and second compression mechanisms 11a and 21a may be shared, and both compression mechanisms may be driven by a driving force supplied from a common drive source.

  Thereby, the 1st, 2nd compression mechanism 11a, 21a can be reduced in size, and size reduction as the whole ejector-type refrigerating cycle can be achieved.

  (2) In the above-described embodiment, an example in which an electric compressor is adopted as the first and second compressors 11 and 21 has been described. However, the formats of the first and second compressors 11 and 21 are not limited to this. .

  For example, you may employ | adopt the variable capacity type compressor which can adjust refrigerant | coolant discharge capability with the change of discharge capacity | capacitance by using an engine etc. as a drive source. In this case, the discharge capacity changing means becomes the discharge capacity changing means. Moreover, you may use the fixed capacity type compressor which adjusts a refrigerant | coolant discharge capability by changing the connection with a drive source intermittently by the interruption of an electromagnetic clutch. In this case, the electromagnetic clutch becomes the discharge capacity changing means.

  Further, the first and second compressors 11 and 21 may employ the same type of compression mechanism or different types of compression mechanisms.

  (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. Similarly, a variable throttle mechanism may be adopted as each suction side pressure reducing means.

  In the first to seventh and ninth to twelfth embodiments described above, the temperature is adjusted so that the degree of superheat of the refrigerant on the outlet side of the outflow evaporator 14 becomes a predetermined value as the high pressure side pressure reducing means. Although the type expansion valve 17 is employed, of course, a temperature type expansion valve 57 that adjusts the superheat degree of the refrigerant on the outlet side of the suction side evaporator 16 to a predetermined value set in advance may be employed.

  Furthermore, an electric expansion valve that can adjust the throttle opening (valve opening) by an external electric control signal may be employed as the high pressure side pressure reducing means. Furthermore, a fixed throttle mechanism having the same configuration as the fixed throttles 39 and 59 may be adopted as the high pressure side pressure reducing means without using the variable throttle mechanism. Furthermore, in the first to seventh, ninth to twelfth, fourteenth to sixteenth, and eighteenth to twentieth embodiments, the high pressure side pressure reducing means may be eliminated.

  Further, when the supercritical refrigeration cycle is configured as in the eighth, thirteenth, and seventeenth embodiments, the high-pressure side refrigerant pressure is changed to the radiator 12, the first radiator 121, the first, 1. Adopting a pressure control valve that adjusts to the target high pressure determined so that COP becomes substantially maximum based on the high-pressure side refrigerant temperature on the outlet side of the heat exchanger functioning as a radiator among the first and second heat exchangers. May be.

  As such a pressure control valve, specifically, it has a temperature sensing part provided on the outlet side of the heat exchanger that acts as a radiator, and a heat exchanger that acts as a radiator inside the temperature sensing part A configuration in which a pressure corresponding to the temperature of the high-pressure refrigerant on the outlet side is generated, and the valve opening is adjusted by a mechanical mechanism in accordance with the balance between the internal pressure of the temperature sensing portion and the refrigerant pressure on the outlet side of the heat exchanger acting as a radiator, etc. Can be adopted.

  (4) As the high-pressure side decompression unit and each low-pressure side decompression unit in each of the above-described embodiments, an expander that expands and decompresses the refrigerant and converts the refrigerant pressure energy into mechanical energy and outputs it is adopted. May be. As such an expander, specifically, a volume type compression mechanism such as a scroll type, a vane type, or a rolling piston type can be employed.

  Then, by flowing the refrigerant so as to flow backward with respect to the refrigerant flow when the positive displacement compression mechanism is used as the compression mechanism, mechanical energy can be output while the refrigerant is volume-expanded and depressurized. For example, if a rotary positive displacement compression mechanism is employed as an expander, rotational energy can be output as mechanical energy.

  Furthermore, if the mechanical energy output from the expander is used as an auxiliary power source for the first and second compression mechanisms, for example, the energy efficiency of the ejector refrigeration cycle 10 as a whole can be improved. The mechanical energy output from the expander may be used as a power source for external equipment.

  For example, if a generator is adopted as an external device, electric energy can be obtained. Moreover, if a flywheel is employ | adopted as an external apparatus, the mechanical energy output from the expander can be stored as a kinetic energy. Moreover, if a stroking device (spring spring) is employed as an external device, the mechanical energy output from the expander can be stored as elastic energy.

  (5) In contrast to the first to thirteenth embodiments described above, an accumulator 55 as an outflow side gas-liquid separator similar to the fourteenth to eighteenth embodiments may be provided on the suction side of the first compressor 11. Good. Thereby, only the gaseous-phase refrigerant | coolant isolate | separated with the accumulator can be supplied to the 1st compression mechanism 11a, and the problem of the liquid compression of the 1st compression mechanism 11a can be avoided.

  Similarly, a suction side gas-liquid separator 55a similar to that of the twentieth embodiment may be arranged with respect to the first to eighteenth embodiments. Thereby, only the gaseous-phase refrigerant | coolant isolate | separated with the suction side gas-liquid separator can be supplied to the 2nd compression mechanism 21a, and the problem of the liquid compression of the 2nd compression mechanism 21a can be avoided.

  (6) In the above-described first to thirteenth embodiments, the example in which different cooling target spaces (refrigerator space, freezer space) are cooled by the outflow side evaporator 14 and the suction side evaporator 16 has been described. The space to be cooled may be cooled. In this case, it is desirable that the outflow side evaporator 14 and the suction side evaporator 16 are assembled in an integrated structure, and the air blown from the blower fan is passed in the order of the outflow side evaporator 14 → the suction side evaporator 16.

  This is because the refrigerant evaporation pressure (refrigerant evaporation temperature) of the suction side evaporator 16 is lower than the refrigerant evaporation pressure (refrigerant evaporation temperature) of the outflow side evaporator 14 as described above. That is, by passing the blown air from the blower fan as described above, a temperature difference between the refrigerant evaporation temperature of the outflow side evaporator 14 and the suction side evaporator 16 and the blown air is secured, and the blown air is efficiently flowed. Can be cooled.

  Further, as a specific means for assembling the outflow side evaporator 14 and the suction side evaporator 16 into an integral structure, for example, the constituent parts of both the evaporators 14 and 16 are made of aluminum and integrated with a joining means such as brazing. You may join to. Furthermore, the structure couple | bonded integrally by mechanical engagement means, such as bolting, may be sufficient.

  Further, as the outflow side evaporator 14 and the suction side evaporator 16, a fin-and-tube type heat exchanger is adopted, and the fins of the outflow side evaporator 14 and the suction side evaporator 16 are made common so that the refrigerant can pass therethrough. It is good also as a structure which divides | segments into two evaporators by a path | pass structure (flow path structure).

  Further, when the outflow side evaporator 14 and the suction side evaporator 16 are configured to cool the same freezer, the refrigerant evaporation temperature of the suction side evaporator 16 arranged downstream of the blown air flow becomes frosted. The resulting temperature (below 0 ° C.) is reached. On the other hand, the absolute humidity of the blown air flowing into the suction side evaporator 16 can be reduced in advance by adjusting the refrigerant evaporation temperature in the outflow side evaporator 14.

  Thereby, generation | occurrence | production of the frost in the suction side evaporator 16 can be suppressed. Furthermore, since the flow of the blast air due to frost formation can be prevented, the fin pitch and the like of the suction side evaporator 16 can be reduced to reduce the size of the suction side evaporator 16. The same applies to the eighteenth embodiment.

  (7) In the first to eighth embodiments 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 operating the ejector-type refrigeration cycle 10, it is only necessary to switch to an operation mode that can exhibit the highest cycle efficiency among any of the operation modes.

  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 18 and 28 may be configured by three-way valves to switch the refrigerant flow path.

  Further, as the first and second suction side pressure reducing means, a fixed throttle mechanism similar to the fixed throttles 39 and 59 is adopted, and the first and second branch parts 18 and 28 and the first and second suction side pressure reducing means are An electromagnetic valve (open / close valve) for opening and closing the flow path may be provided between or downstream of the first and second suction side decompression means.

  (8) In the internal heat exchangers 30 to 37 of the above-described embodiments, 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 not mentioned. The refrigerant flow direction and the refrigerant flow direction in the low-pressure side refrigerant flow path may be parallel flows, and 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. It may be counterflow. In the twenty-sixth and twenty-seventh embodiments, the low-pressure side refrigerant of the cycle may be the second compressor 21 suction refrigerant.

  (9) In the above-described embodiment, the ejector refrigeration cycle 10, 40, 50 including only the first and second compression mechanisms 11a, 21a has been described. However, an additional compression mechanism may be provided. For example, with respect to the suction side evaporator 16 of the first embodiment, an additional evaporator is arranged in parallel, and an additional compression mechanism is provided so as to suck and compress only the refrigerant that has flowed out of this evaporator. May be.

  (10) In the above-described embodiment, the example in which the ejector refrigeration cycle 10 of the present invention is applied to a refrigeration / refrigeration apparatus and a refrigerator has been described, but the application of the present invention is not limited thereto. For example, the ejector refrigeration cycle 10 may be applied to other stationary refrigeration cycle apparatuses, vehicle air conditioners, and the like.

  (11) In the above-described embodiment, the suction side evaporator 16 is configured as a use side heat exchanger, and the radiator 12 is configured as an outdoor heat exchanger that radiates heat to the atmosphere side. 16 may be configured as an outdoor heat exchanger that absorbs heat from a heat source such as the atmosphere, and the heat radiator 12 may be configured as a heat pump cycle configured as an indoor heat exchanger that heats a refrigerant to be heated such as air or water.

11, 12 1st, 2nd compressor 11a, 21a 1st, 2nd compression mechanism 11b, 21b 1st, 2nd electric motor 12 Radiator 12b, 122b Condensing part 12c, 122c Gas-liquid separation part 12d, 122d Supercooling Part 13 Ejector 13a Nozzle part 13b Refrigerant suction port 13d Diffuser part 14 Outflow side evaporator 16 Suction side evaporator 17, 57 Thermal expansion valve 18, 28, 38 First, second and third branch parts 19, 29 First , Second electric expansion valve 20 merging section 27 pressure control valve 30-37 internal heat exchanger 39, 59 fixed throttle 51, 52 first, second heat exchanger 53, 54 first, second electric four-way valve 55 Accumulator 121, 122 First and second radiator

Claims (14)

A first compression mechanism (11a) for compressing and discharging the refrigerant;
A branch part (38) for branching the flow of the high-pressure refrigerant discharged from the first compression mechanism (11a);
A first radiator (121) for radiating heat from one of the refrigerants branched at the branch part (38);
The refrigerant is sucked from the refrigerant suction port (13b) by the flow of the high-speed jet refrigerant jetted from the nozzle portion (13a) for decompressing and expanding the refrigerant flowing out from the first radiator (121), and the jet refrigerant and the An ejector (13) for increasing the pressure of the mixed refrigerant with the suction refrigerant sucked from the refrigerant suction port (13b) at the diffuser section (13d);
A second radiator (122) for radiating heat from the other refrigerant branched at the branch portion (38);
Suction side decompression means (39) for decompressing and expanding the refrigerant flowing out of the second radiator (122);
A suction-side evaporator (16) for evaporating the refrigerant flowing out from the suction-side decompression means (39) and causing it to flow out toward the refrigerant suction port (13b);
An ejector-type refrigeration cycle comprising: a suction unit evaporator (16) and a second compression mechanism (21a) for sucking, compressing and discharging the outlet side refrigerant.   The ejector type according to claim 1, further comprising an outflow side evaporator (14) for evaporating the refrigerant flowing out from the diffuser section (13) to flow out to the suction side of the first compression mechanism (11a). Refrigeration cycle.   A high-pressure side pressure reducing means (17) disposed in a refrigerant passage from the outlet side of the first radiator (121) to the inlet side of the nozzle part (13a) and decompressing and expanding the refrigerant flowing out of the first radiator (121). The ejector-type refrigeration cycle according to claim 1 or 2, further comprising:   The internal heat exchanger (34, 35) for exchanging heat between the refrigerant flowing out of the second radiator (122) and the low-pressure side refrigerant of the cycle is provided. The ejector refrigeration cycle described.   The second radiator (122) includes a condensing unit (122b) that condenses the refrigerant, a gas-liquid separating unit (122c) that separates the gas and liquid of the refrigerant flowing out from the condensing unit (122b), and the gas-liquid separation. The ejector refrigeration cycle according to any one of claims 1 to 4, further comprising a supercooling section (122d) for supercooling the liquid-phase refrigerant flowing out of the section (122c).   6. The expansion device according to claim 1, wherein the suction side decompression means (39) is an expander that expands and decompresses the refrigerant and converts the pressure energy of the refrigerant into mechanical energy and outputs the mechanical energy. The ejector type refrigeration cycle according to claim 1. A first compression mechanism (11a) for compressing and discharging the refrigerant;
A radiator (12) for radiating heat from the high-pressure refrigerant discharged from the first compression mechanism (11a);
First and second heat exchangers (51, 52) for exchanging heat between the refrigerant and the heat exchange target fluid;
The suction sucked from the refrigerant suction port (13b) by sucking the refrigerant from the refrigerant suction port (13b) by the flow of the high-speed jet refrigerant jetted from the nozzle portion (13a) for decompressing and expanding the refrigerant. An ejector (13) for increasing the pressure of the refrigerant mixed with the refrigerant at the diffuser section (13d);
An outflow-side gas-liquid separator (55) for separating the gas-liquid of the refrigerant flowing out from the diffuser section (13d);
A second compression mechanism (21a) that compresses the refrigerant and discharges the refrigerant to the refrigerant suction port (13b) side;
Flow path switching means (53, 54) for switching the refrigerant flow path,
The flow path switching means (53, 54)
In the first operation mode, the refrigerant discharged from the first compression mechanism (11a) flows in the order of the radiator (12) → the first heat exchanger (51) → the nozzle part (13a) and the outflow side gas-liquid The refrigerant flow path is switched so that the liquid refrigerant flowing out of the separator (55) flows in the order of the second heat exchanger (52) → the second compression mechanism (21a) → the refrigerant suction port (13d),
In the second operation mode, the refrigerant discharged from the first compression mechanism (11a) flows in the order of the radiator (12) → the second heat exchanger (52) → the nozzle portion (13a) and the outflow side gas-liquid Switching the refrigerant flow path so that the liquid refrigerant flowing out of the separator (55) flows in the order of the first heat exchanger (51) → the second compression mechanism (21a) → the refrigerant suction port (13d). Characteristic ejector refrigeration cycle.   The ejector-type refrigeration cycle according to claim 7, further comprising high-pressure side decompression means (57) for decompressing and expanding the refrigerant flowing into the nozzle portion (13a).   The ejector refrigeration cycle according to claim 7 or 8, further comprising an internal heat exchanger (36, 37) for exchanging heat between the high-pressure refrigerant upstream of the nozzle portion (13a) and the low-pressure refrigerant of the cycle. .   The ejector refrigeration cycle according to claim 4 or 9, wherein the low-pressure side refrigerant of the cycle is a refrigerant sucked into the first compression mechanism (11a).   The ejector refrigeration cycle according to claim 4 or 9, wherein the low-pressure side refrigerant of the cycle is a refrigerant sucked into the second compression mechanism (21a). First discharge capacity changing means (11b) for changing the refrigerant discharge capacity of the first compression mechanism (11a);
Second discharge capacity changing means (21b) for changing the refrigerant discharge capacity of the second compression mechanism (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 mechanism (11a) and the second compression mechanism (21a). The ejector refrigeration cycle according to any one of claims 1 to 11, wherein the ejector refrigeration cycle is configured to be possible.   The said 1st compression mechanism (11a) and the said 2nd compression mechanism (21a) are accommodated in the same housing, and are comprised integrally, The one of Claim 1 thru | or 12 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 13, wherein the first compression mechanism (11a) increases the pressure of the refrigerant until the pressure becomes equal to or higher than a critical pressure.

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