WO2014203460A1 - Cycle de réfrigération d'éjecteur - Google Patents

Cycle de réfrigération d'éjecteur Download PDF

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Publication number
WO2014203460A1
WO2014203460A1 PCT/JP2014/002784 JP2014002784W WO2014203460A1 WO 2014203460 A1 WO2014203460 A1 WO 2014203460A1 JP 2014002784 W JP2014002784 W JP 2014002784W WO 2014203460 A1 WO2014203460 A1 WO 2014203460A1
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WO
WIPO (PCT)
Prior art keywords
refrigerant
ejector
refrigeration cycle
gas
nozzle
Prior art date
Application number
PCT/JP2014/002784
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English (en)
Japanese (ja)
Inventor
西嶋　春幸
健太 茅野
高野　義昭
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株式会社デンソー
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Publication of WO2014203460A1 publication Critical patent/WO2014203460A1/fr

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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04F—PUMPING OF FLUID BY DIRECT CONTACT OF ANOTHER FLUID OR BY USING INERTIA OF FLUID TO BE PUMPED; SIPHONS
    • F04F5/00—Jet pumps, i.e. devices in which flow is induced by pressure drop caused by velocity of another fluid flow
    • F04F5/14—Jet pumps, i.e. devices in which flow is induced by pressure drop caused by velocity of another fluid flow the inducing fluid being elastic fluid
    • F04F5/16—Jet pumps, i.e. devices in which flow is induced by pressure drop caused by velocity of another fluid flow the inducing fluid being elastic fluid displacing elastic fluids
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04F—PUMPING OF FLUID BY DIRECT CONTACT OF ANOTHER FLUID OR BY USING INERTIA OF FLUID TO BE PUMPED; SIPHONS
    • F04F5/00—Jet pumps, i.e. devices in which flow is induced by pressure drop caused by velocity of another fluid flow
    • F04F5/14—Jet pumps, i.e. devices in which flow is induced by pressure drop caused by velocity of another fluid flow the inducing fluid being elastic fluid
    • F04F5/16—Jet pumps, i.e. devices in which flow is induced by pressure drop caused by velocity of another fluid flow the inducing fluid being elastic fluid displacing elastic fluids
    • F04F5/20—Jet pumps, i.e. devices in which flow is induced by pressure drop caused by velocity of another fluid flow the inducing fluid being elastic fluid displacing elastic fluids for evacuating
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B41/00—Fluid-circulation arrangements
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B41/00—Fluid-circulation arrangements
    • F25B41/30—Expansion means; Dispositions thereof
    • F25B41/39—Dispositions with two or more expansion means arranged in series, i.e. multi-stage expansion, on a refrigerant line leading to the same evaporator
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B5/00—Compression machines, plants or systems, with several evaporator circuits, e.g. for varying refrigerating capacity
    • F25B5/02—Compression machines, plants or systems, with several evaporator circuits, e.g. for varying refrigerating capacity arranged in parallel
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2341/00—Details of ejectors not being used as compression device; Details of flow restrictors or expansion valves
    • F25B2341/001—Ejectors not being used as compression device
    • F25B2341/0011—Ejectors with the cooled primary flow at reduced or low pressure
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2341/00—Details of ejectors not being used as compression device; Details of flow restrictors or expansion valves
    • F25B2341/001—Ejectors not being used as compression device
    • F25B2341/0014—Ejectors with a high pressure hot primary flow from a compressor discharge
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2400/00—General features or devices for refrigeration machines, plants or systems, combined heating and refrigeration systems or heat-pump systems, i.e. not limited to a particular subgroup of F25B
    • F25B2400/04—Refrigeration circuit bypassing means
    • F25B2400/0403—Refrigeration circuit bypassing means for the condenser
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2400/00—General features or devices for refrigeration machines, plants or systems, combined heating and refrigeration systems or heat-pump systems, i.e. not limited to a particular subgroup of F25B
    • F25B2400/04—Refrigeration circuit bypassing means
    • F25B2400/0409—Refrigeration circuit bypassing means for the evaporator
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B2400/00—General features or devices for refrigeration machines, plants or systems, combined heating and refrigeration systems or heat-pump systems, i.e. not limited to a particular subgroup of F25B
    • F25B2400/04—Refrigeration circuit bypassing means
    • F25B2400/0411—Refrigeration circuit bypassing means for the expansion valve or capillary tube

Definitions

  • the present disclosure relates to an ejector-type refrigeration cycle including an ejector. It relates to ejectors.
  • an ejector-type refrigeration cycle which is a vapor compression refrigeration cycle apparatus including an ejector, is known.
  • the refrigerant flowing out of the evaporator is sucked by the suction action of the high-speed jet refrigerant jetted from the nozzle part of the ejector, and the jet refrigerant and suction are sucked by the diffuser part (pressure booster) of the ejector.
  • the diffuser part pressure booster
  • the power consumption of the compressor is reduced and the coefficient of performance of the cycle (in comparison with the normal refrigeration cycle apparatus in which the refrigerant evaporation pressure in the evaporator and the suction refrigerant pressure in the compressor are substantially equal) COP) is improved.
  • Patent Document 1 includes two evaporators, and the refrigerant that has flowed out of the evaporator having the higher refrigerant evaporation pressure flows into the nozzle portion of the ejector.
  • a cycle configuration is disclosed in which the refrigerant flowing out of the evaporator having the lower refrigerant evaporation pressure is sucked by the suction action of the injected refrigerant.
  • an object of the present disclosure is to sufficiently improve the coefficient of performance of an ejector-type refrigeration cycle in which an evaporator downstream refrigerant flows into a nozzle portion of an ejector.
  • the ejector refrigeration cycle of the present disclosure includes a compressor, a radiator, a first decompression unit, a second decompression unit, a first evaporator, and a second evaporator.
  • An evaporator and an ejector are provided.
  • the compressor compresses and discharges the refrigerant.
  • the radiator dissipates heat from the refrigerant discharged from the compressor.
  • the first decompression unit and the second decompression unit decompress the refrigerant on the downstream side of the radiator.
  • the first evaporator evaporates the refrigerant decompressed by the first decompression unit.
  • the second evaporator evaporates the refrigerant decompressed by the second decompression unit.
  • the ejector has a booster.
  • the booster sucks the second evaporator downstream refrigerant from the refrigerant suction port by the suction action of the high-speed jet refrigerant injected from the nozzle that depressurizes the first evaporator downstream refrigerant, and sucks the jet refrigerant and the suction refrigerant.
  • the pressure is increased by mixing with a refrigerant.
  • the ejector refrigeration cycle of the present disclosure further includes a gas-liquid supply unit that causes the refrigerant flowing out from the first evaporator to flow into the nozzle unit in a gas-liquid two-phase state.
  • the gas-liquid supply part since the gas-liquid supply part is provided, the gas-liquid two-phase refrigerant can surely flow into the nozzle part of the ejector. Therefore, it is possible to reliably suppress the occurrence of the condensation delay and the deterioration of the refrigerant pressure increase performance in the diffuser portion of the ejector.
  • the jet refrigerant injected from the nozzle portion can also be a gas-liquid two-phase refrigerant.
  • the jet refrigerant injected from the nozzle portion can also be a gas-liquid two-phase refrigerant.
  • FIG. 7 is a sectional view taken along line VII-VII in FIG. 6.
  • the inventors of the present application investigated the cause, and in the configuration in which the gas-phase refrigerant flowing out of the evaporator flows into the nozzle portion of the ejector as in the ejector-type refrigeration cycle of Patent Document 1, (a) the injected refrigerant and The reason is that the mixed refrigerant with the suction refrigerant becomes a gas-liquid two-phase refrigerant with high dryness, and (b) the vapor phase refrigerant is condensed while being decompressed in the refrigerant passage formed in the nozzle part. It was found that.
  • the mixed refrigerant is a gas-liquid two-phase refrigerant having a relatively high dryness x (for example, a gas-liquid two-phase refrigerant having a dryness x of 0.8 or more)
  • the gas-liquid two-phase refrigerant is placed near the diffuser section or This is because a shock wave is generated in the diffuser portion, and the refrigerant pressurization performance in the diffuser portion of the ejector becomes unstable.
  • This shock wave is generated when the flow velocity of the two-phase fluid in the gas-liquid two-phase state shifts from the two-phase sound velocity ⁇ h or higher (supersonic state) to a value lower than the two-phase sound velocity ⁇ h (subsonic state). Is.
  • the two-phase sound velocity ⁇ h is a sound velocity of a fluid in a gas-liquid mixed state in which a gas phase fluid and a liquid phase fluid are mixed, and is defined by the following formula F1.
  • ⁇ h [P / ⁇ ⁇ (1 ⁇ ) ⁇ ⁇ l ⁇ ] 0.5 (F1)
  • ⁇ in the formula F1 is a void ratio, and indicates a volume ratio of voids (bubbles) contained per unit volume. More specifically, the void ratio ⁇ is defined by the following formula F2.
  • ⁇ x / ⁇ x + ( ⁇ g / ⁇ l) ⁇ (1 ⁇ x) ⁇ (F2)
  • ⁇ g in the formulas F1 and F2 is a gas phase fluid density
  • ⁇ l is a liquid phase fluid density
  • P is a pressure of a two-phase fluid.
  • FIGS. 20 and 21 schematically show an axial cross section of a general ejector.
  • the same reference numerals as those of the ejector 18 are attached to portions that perform the same or equivalent functions as those of the ejector 18 described in the embodiments described later.
  • a gas-liquid two-phase refrigerant having a relatively low dryness x (for example, a gas-liquid two-phase refrigerant having a dryness x of 0.5 or less) is caused to flow into the nozzle portion 18a of the ejector 18.
  • the dryness x of the refrigerant immediately before being injected from the refrigerant injection port 18c of the nozzle portion 18a is determined by the refrigerant flowing into the nozzle portion 18a. It becomes a value lower than the dryness x.
  • the injected refrigerant injected from the refrigerant injection port 18c of the nozzle portion 18a is mixed with the suction refrigerant in a gas phase state, thereby rapidly increasing the dryness x while decreasing the flow velocity.
  • the two-phase sound speed ⁇ h of the mixed refrigerant of the injection refrigerant and the suction refrigerant also rises abruptly.
  • the mixed refrigerant has a flow velocity that is higher than the two-phase sound velocity ⁇ h immediately after being injected from the refrigerant injection port 18c.
  • the shock wave generated when the flow velocity of the two-phase refrigerant is changed from the supersonic state to the subsonic state is generated in the vicinity of the refrigerant injection port 18c of the nozzle portion 18a. For this reason, the influence which a shock wave has on the refrigerant
  • a gas-liquid two-phase refrigerant having a relatively high dryness x for example, a gas-liquid two-phase refrigerant having a dryness x of 0.8 or more
  • a gas-liquid two-phase refrigerant having a dryness x of 0.8 or more flows into the nozzle portion 18a
  • the nozzle portion The dryness x of the refrigerant immediately before being injected from the refrigerant injection port 18c of 18a also increases. For this reason, the degree of increase in the dryness x when the injected refrigerant is mixed with the suction refrigerant and becomes the mixed refrigerant, compared with the case where the gas-liquid two-phase refrigerant having a relatively low dryness x flows into the nozzle portion 18a. Becomes smaller.
  • the degree of increase in the two-phase sonic velocity ⁇ h of the mixed refrigerant is also reduced, and the location where the mixed refrigerant has a value lower than the two-phase sonic velocity ⁇ h (location where the shock wave is generated) It is easier to separate from the refrigerant injection port 18c than when a gas-liquid two-phase refrigerant having a relatively low dryness x flows into 18a.
  • the diffuser portion 18g of the ejector 18 cannot exhibit the desired refrigerant pressurization performance, and the ejector-type refrigeration cycle of Patent Document 1 cannot sufficiently obtain the COP improvement effect due to the provision of the ejector.
  • the ejector-type refrigeration cycle of Patent Document 1 if the dryness x of the mixed refrigerant is 0.8 or more, the refrigerant pressurization performance tends to be unstable. It has been confirmed.
  • the diffuser portion 18g of the ejector 18 when the dryness x of the mixed refrigerant increases and becomes a gas-liquid two-phase refrigerant having a high dryness of 0.995 or more, the diffuser portion 18g of the ejector 18 The desired refrigerant pressurization performance cannot be exhibited. Furthermore, the flow rate of the suction refrigerant sucked from the refrigerant suction port 18d of the ejector 18 may decrease.
  • FIGS. 22 and FIG. 23 schematically show an axial section of a general ejector as in FIGS. 20 and 21 described above.
  • the gas-phase refrigerant in the injection refrigerant is decelerated while being mixed with the suction refrigerant.
  • the liquid phase refrigerant (that is, droplets) in the jet refrigerant is accelerated by the inertial force when jetted from the refrigerant jet port 18c of the nozzle portion 18a.
  • the inertial force of the droplet is represented by an integrated value of the weight of the droplet and the velocity of the droplet at the refrigerant ejection port 18c.
  • the pressure energy of the mixed refrigerant is converted into velocity energy, and the pressure of the mixed refrigerant is connected to the refrigerant suction port 18d as shown by the solid line in the lower graph of FIG. It is possible to lower the pressure of the refrigerant flowing out from the evaporator. Further, due to the pressure drop of the mixed refrigerant, the gas-phase refrigerant flowing out of the evaporator can be sucked.
  • the speed of the liquid droplets changes substantially the same as that of the gas-phase refrigerant. For this reason, the droplets in the mixed refrigerant cannot be sufficiently accelerated, and the pressure of the mixed refrigerant is unlikely to decrease as shown by the solid line in the lower graph of FIG. As a result, the suction refrigerant flow rate of the ejector 18 decreases.
  • the expansion wave, the injected refrigerant, and the suction refrigerant generated when the injected refrigerant is injected from the refrigerant injection port 18c are
  • a plurality of periodic shock waves called barrel shock waves as shown in FIG. 24 may be generated in the mixed refrigerant.
  • Such barrel shock waves periodically change the flow rate of the mixed refrigerant from the supersonic state to the subsonic state, and from the subsonic state to the supersonic state, so that the velocity energy of the mixed refrigerant is greatly lost. End up. Therefore, the barrel shock wave causes the suction refrigerant flow rate of the ejector 18 to be greatly reduced and causes the operating sound to be generated in the ejector 18.
  • FIG. 24 is an explanatory diagram for explaining the barrel shock wave, and is a schematic enlarged cross-sectional view around the refrigerant injection port 18c of the nozzle portion 18a of the ejector 18 of the prior art.
  • the vapor phase refrigerant is condensed while reducing the pressure in the refrigerant passage formed in the nozzle portion.
  • the vapor phase refrigerant is condensed while being reduced in the refrigerant passage formed in the nozzle portion.
  • the diffuser of the ejector is reduced by depressurizing the refrigerant so as to straddle the saturated gas line at the nozzle portion. The reason why the unit cannot exhibit the desired refrigerant pressurization performance will be described.
  • FIG. 25 is a Mollier diagram showing changes in the state of the refrigerant when a condensation delay occurs.
  • the refrigerant in the same state as in FIG. 3 is assigned the same reference numeral (alphabet) as in FIG. Only the subscripts (numbers) are changed. The same applies to other Mollier diagrams.
  • the refrigerant in the region where the enthalpy has slightly decreased from the saturated gas line is in a metastable state in which the refrigerant cannot be condensed unless the temperature is lowered than the refrigerant on the saturated gas line at the same pressure. Therefore, when the gas-phase refrigerant is caused to flow into the nozzle portion 18a, a condensation delay is generated in which condensation does not start until the temperature of the metastable refrigerant is lowered to some extent.
  • the enthalpy of the injected refrigerant increases as compared with the case where the refrigerant is isentropically expanded at the nozzle portion 18a (corresponding to ⁇ hx in FIG. 25).
  • the amount of increase in enthalpy corresponds to the amount of latent heat released as latent heat when the refrigerant flows through the refrigerant passage formed in the nozzle portion 18a. Therefore, if the amount of latent heat released increases, the nozzle portion A shock wave is generated in the flow of the refrigerant flowing through the refrigerant passage 18a.
  • shock wave generated by releasing the latent heat of the refrigerant lowers the flow rate of the injected refrigerant, so that the pressure increase performance of the refrigerant in the diffuser portion 18g is lowered.
  • an object of the present disclosure is to sufficiently improve the coefficient of performance of an ejector-type refrigeration cycle in which an evaporator downstream refrigerant flows into a nozzle portion of an ejector.
  • the refrigerant that has flowed out of the evaporator is configured to flow into the nozzle portion of the ejector, and the diffuser portion (pressure boosting portion) of the ejector is prevented from becoming unable to exhibit the desired refrigerant pressure increasing performance.
  • An ejector refrigeration cycle that can be used will be described. Further, the embodiments described below include a precondition of the present disclosure and a reference form. (First embodiment) The first embodiment will be described with reference to FIGS.
  • an ejector refrigeration cycle 10 including an ejector 18 is applied to a vehicle refrigeration cycle apparatus.
  • the ejector-type refrigeration cycle 10 has a function of cooling indoor air blown into the vehicle interior, and the internal air blown into the in-vehicle refrigerator (cool box) disposed in the vehicle interior. It performs the function of cooling.
  • the compressor 11 sucks refrigerant and compresses and discharges it until it becomes high-pressure refrigerant.
  • the compressor 11 of the present embodiment is an electric compressor configured by housing a fixed capacity type compression mechanism and an electric motor that drives the compression mechanism in one housing.
  • various compression mechanisms such as a scroll-type compression mechanism and a vane-type compression mechanism can be adopted. Further, the operation (rotation speed) of the electric motor is controlled by a control signal output from a control device to be described later, and either an AC motor or a DC motor may be adopted.
  • the compressor 11 may be an engine-driven compressor that is driven by a rotational driving force transmitted from a vehicle travel engine via a pulley, a belt, or the like.
  • This type of engine-driven compressor includes a variable displacement compressor that can adjust the refrigerant discharge capacity by changing the discharge capacity, and a fixed type that adjusts the refrigerant discharge capacity by changing the operating rate of the compressor by intermittently connecting the electromagnetic clutch.
  • a capacity type compressor or the like can be employed.
  • the ejector refrigeration cycle 10 employs an HFC refrigerant (specifically, R134a) as a refrigerant, and a vapor compression subcritical refrigeration cycle in which the high-pressure side refrigerant pressure does not exceed the refrigerant critical pressure. It is composed. Furthermore, refrigeration oil for lubricating the compressor 11 is mixed in the refrigerant, and a part of the refrigeration oil circulates in the cycle together with the refrigerant.
  • HFC refrigerant specifically, R134a
  • a vapor compression subcritical refrigeration cycle in which the high-pressure side refrigerant pressure does not exceed the refrigerant critical pressure. It is composed. Furthermore, refrigeration oil for lubricating the compressor 11 is mixed in the refrigerant, and a part of the refrigeration oil circulates in the cycle together with the refrigerant.
  • the refrigerant inlet side of the radiator 12 is connected to the discharge port side of the compressor 11.
  • the radiator 12 is a heat exchanger for radiating heat by exchanging heat between the high-pressure refrigerant discharged from the compressor 11 and outside air (outside air) blown by the cooling fan 12a to dissipate the high-pressure refrigerant and cool it.
  • the cooling fan 12a is an electric blower whose rotational speed (air amount) is controlled by a control voltage output from the control device.
  • the inlet side of the high-stage expansion device 13 as the first decompression unit is connected to the refrigerant outlet side of the radiator 12.
  • the high-stage expansion device 13 has a temperature sensing unit that detects the degree of superheat of the first evaporator 15 outlet-side refrigerant based on the temperature and pressure of the first evaporator 15 outlet-side refrigerant, and the first evaporator 15 outlet This is a temperature type expansion valve that adjusts the cross-sectional area of the throttle passage by a mechanical mechanism so that the degree of superheat of the side refrigerant falls within a predetermined reference range.
  • the outlet side of the high stage side throttle device 13 is connected to the refrigerant inlet of the branching section 14 that branches the flow of the refrigerant flowing out from the high stage side throttle device 13.
  • the branch portion 14 is configured by a three-way joint having three inflow / outflow ports, and one of the three inflow / outflow ports is a refrigerant inflow port, and the remaining two are refrigerant outflow ports.
  • Such a three-way joint may be formed by joining pipes having different pipe diameters, or may be formed by providing a plurality of refrigerant passages in a metal block or a resin block.
  • the refrigerant inlet side of the first evaporator 15 is connected to one refrigerant outlet of the branch part 14.
  • the first evaporator 15 evaporates the low-pressure refrigerant by exchanging heat between the low-pressure refrigerant decompressed by the high-stage expansion device 13 and the indoor air blown into the vehicle interior from the first blower fan 15a.
  • An endothermic heat exchanger that exhibits an endothermic effect.
  • the 1st ventilation fan 15a is an electric blower by which rotation speed (air amount) is controlled by the control voltage output from a control apparatus.
  • the low-stage side throttle device 16 is a fixed throttle with a fixed throttle opening, and specifically, a nozzle, an orifice, a capillary tube, or the like can be adopted.
  • the refrigerant inlet side of the second evaporator 17 is connected to the outlet side of the low stage side expansion device 16.
  • the second evaporator 17 exchanges heat between the low-pressure refrigerant decompressed by the low-stage expansion device 16 and the internal air that is circulated into the cool box from the second blower fan 17a. This is an endothermic heat exchanger that evaporates to exert an endothermic effect.
  • the basic configuration of the second evaporator 17 is the same as that of the first evaporator 15.
  • the refrigerant flowing into the second evaporator 17 is depressurized by the high stage side expansion device 13 and then further depressurized by the low stage side expansion device 16, the refrigerant evaporation pressure in the second evaporator 17.
  • the (refrigerant evaporation temperature) is lower than the refrigerant evaporation pressure (refrigerant evaporation temperature) in the first evaporator 15.
  • the 2nd ventilation fan 17a is an electric blower by which rotation speed (air amount) is controlled by the control voltage output from a control apparatus.
  • the inlet side of the nozzle portion 18 a of the ejector 18 is connected to the refrigerant outlet side of the first evaporator 15.
  • the ejector 18 functions as a decompression unit that decompresses the refrigerant on the downstream side of the first evaporator 15. Further, the ejector 18 functions as a refrigerant circulation section (refrigerant transport section) that sucks (transports) the refrigerant by the suction action of the jetted refrigerant jetted at a high speed and circulates in the cycle.
  • the ejector 18 has a nozzle portion 18a and a body portion 18b.
  • the nozzle portion 18a is formed of a substantially cylindrical metal (for example, a stainless alloy) that gradually tapers in the flow direction of the refrigerant, and the like, and the refrigerant is passed through a refrigerant passage (throttle passage) formed inside. It expands under reduced pressure entropy.
  • the refrigerant passage formed in the nozzle portion 18a is provided with a throat portion (minimum passage cross-sectional area) having the smallest refrigerant passage cross-sectional area.
  • the refrigerant passage is further provided with a divergent portion in which the sectional area of the refrigerant passage gradually increases toward the refrigerant injection port 18c for injecting the refrigerant from the throat. That is, the nozzle portion 18a of the present embodiment is configured as a so-called Laval nozzle.
  • the injection refrigerant injected from the refrigerant injection port 18c is in a gas-liquid two-phase state, and the refrigerant immediately before being injected from the refrigerant injection port 18c.
  • a flow rate of at least the two-phase sound speed ⁇ h (supersonic state) described in Formula F1 is employed.
  • the body portion 18b is formed of a substantially cylindrical metal (for example, aluminum) or resin, and functions as a fixing member for supporting and fixing the nozzle portion 18a therein, and forms an outer shell of the ejector 18.
  • the nozzle portion 18a is fixed by press-fitting or the like so as to be housed inside the one end side in the longitudinal direction of the body portion 18b.
  • a refrigerant suction port 18d provided in a portion corresponding to the outer peripheral side of the nozzle portion 18a in the outer peripheral side surface of the body portion 18b so as to penetrate the inside and the outside and communicate with the refrigerant injection port 18c of the nozzle portion 18a. Is formed.
  • the refrigerant suction port 18d is a through hole that sucks the refrigerant that has flowed out of the second evaporator 17 due to the suction action of the refrigerant injected from the refrigerant injection port 18c of the nozzle portion 18a into the ejector 18.
  • a mixing portion 18e, a suction passage 18f, and a diffuser portion 18g are formed inside the body portion 18b.
  • the injection refrigerant injected from the refrigerant injection port 18c and the suction refrigerant sucked from the refrigerant suction port 18d are mixed.
  • the suction passage 18f guides the suction refrigerant sucked from the refrigerant suction port 18d to the mixing unit 18e.
  • the diffuser part 18g is a pressure increasing part that pressurizes the mixed refrigerant mixed in the mixing part 18e.
  • the suction passage 18f is formed by a space between the outer peripheral side around the tapered tip of the nozzle portion 18a and the inner peripheral side of the body portion 18b.
  • the refrigerant passage cross-sectional area of the suction passage 18f is gradually reduced in the refrigerant flow direction. Thereby, the flow rate of the suction refrigerant flowing through the suction passage 18f is gradually increased, and the energy loss (mixing loss) when the suction refrigerant and the injection refrigerant are mixed in the mixing unit 18e is reduced.
  • the mixing portion 18e is formed by a space in the inner space of the body portion 18b ranging from the refrigerant injection port 18c of the nozzle portion 18a to the inlet portion 18h of the diffuser portion 18g in the axial section of the nozzle portion 18a. Furthermore, the distance La in the axial direction of the nozzle portion 18a from the refrigerant injection port 18c to the inlet portion 18h in the mixing portion 18e is determined so that the flow velocity of the refrigerant flowing into the inlet portion 18h is equal to or lower than the two-phase sound velocity ⁇ h. .
  • the area has a total value of the circular opening cross-sectional area of the refrigerant injection port 18c and the annular refrigerant passage cross-sectional area of the suction passage 18f in the axially vertical cross section of the nozzle portion 18a including the refrigerant injection port 18c.
  • the distance La is determined so as to satisfy the following formula F3.
  • the equivalent diameter ⁇ Da may be 9 mm and the distance La may be 7 mm.
  • the mixing portion 18e of the present embodiment is formed in a shape that gradually reduces the cross-sectional area of the refrigerant passage toward the downstream side of the refrigerant flow. More specifically, it is formed in a shape combining a truncated cone shape that gradually reduces the refrigerant passage cross-sectional area toward the downstream side of the refrigerant flow and a cylindrical shape that makes the refrigerant passage cross-sectional area constant. Further, the refrigerant passage cross-sectional area of the inlet portion 18h of the diffuser portion 18g is formed to be smaller than the refrigerant passage cross-sectional area of the refrigerant injection port 18c.
  • the axial length of the nozzle portion 18a of the cylindrical portion of the mixing portion 18e is Lb
  • the diameter of the cylindrical portion (corresponding to the diameter of the inlet portion 18h of the diffuser portion 18g) is ⁇ Db.
  • the distance Lb is determined so as to satisfy the following formula F4.
  • the diameter ⁇ Db may be 7 mm and the distance Lb may be 6 mm.
  • the diffuser portion 18g is disposed so as to be continuous with the outlet of the mixing portion 18e, and is formed so that the refrigerant passage cross-sectional area gradually increases.
  • the diffuser part 18g fulfills the function of converting the velocity energy of the mixed refrigerant flowing out from the mixing unit 18e into pressure energy, that is, the function of increasing the pressure of the mixed refrigerant by decelerating the flow rate of the mixed refrigerant.
  • the wall surface shape of the inner peripheral wall surface of the body portion 18b forming the diffuser portion 18g of the present embodiment is formed by combining a plurality of curves as shown in FIG. And since the degree of spread of the refrigerant passage cross-sectional area of the diffuser portion 18g gradually increases in the refrigerant flow direction and then decreases again, the refrigerant can be increased in an isentropic manner.
  • the suction port of the compressor 11 is connected to the refrigerant outlet side of the diffuser portion 18g of the ejector 18.
  • a control device (not shown) includes a known microcomputer including a CPU, a ROM, a RAM, and the like and peripheral circuits thereof.
  • the control device performs various calculations and processes based on the control program stored in the ROM, and controls the operations of the various control target devices 11, 12a, 15a, 17a, etc. connected to the output side.
  • the control device includes sensors such as an inside air temperature sensor, an outside air temperature sensor, a solar radiation sensor, a first evaporator temperature sensor, a second evaporator temperature sensor, an outlet side temperature sensor, an outlet side pressure sensor, and an inside temperature sensor.
  • the groups are connected, and the detection values of these sensor groups are input.
  • the inside air temperature sensor detects the cabin temperature.
  • the outside air temperature sensor detects the outside air temperature.
  • the solar radiation sensor detects the amount of solar radiation in the passenger compartment.
  • the first evaporator temperature sensor detects the blown air temperature (evaporator temperature) of the first evaporator 15.
  • the second evaporator temperature sensor detects the blown air temperature (evaporator temperature) of the second evaporator 17.
  • the outlet side temperature sensor detects the temperature of the radiator 12 outlet side refrigerant.
  • the outlet side pressure sensor detects the pressure of the radiator 12 outlet side refrigerant.
  • the internal temperature sensor detects the temperature in the cool box (internal temperature).
  • an operation panel (not shown) disposed near the instrument panel in the front part of the vehicle interior is connected to the input side of the control device, and operation signals from various operation switches provided on the operation panel are input to the control device.
  • various operation switches provided on the operation panel there are provided an air conditioning operation switch for requesting air conditioning in the vehicle interior, a vehicle interior temperature setting switch for setting the vehicle interior temperature, and the like.
  • the control device of the present embodiment is configured such that a control unit that controls the operation of various control target devices connected to the output side is integrally configured.
  • a configuration (hardware and software) that controls the operation of each control target device constitutes a control unit of each control target device.
  • operation of the compressor 11 comprises the discharge capability control part.
  • the control device operates the electric motor of the compressor 11, the cooling fan 12a, the first blower fan 15a, the second blower fan 17a, and the like. Thereby, the compressor 11 sucks the refrigerant, compresses it, and discharges it.
  • the refrigerant that has flowed out of the radiator 12 flows into the high-stage expansion device 13 and is decompressed in an enthalpy manner (b3 point ⁇ c3 point in FIG. 3).
  • the throttle opening degree of the high stage side expansion device 13 is adjusted so that the degree of superheat of the refrigerant on the outlet side of the first evaporator 15 (point d3 in FIG. 3) is within a predetermined range.
  • the flow of the refrigerant depressurized by the high stage side expansion device 13 is branched by the branching section 14.
  • One refrigerant branched in the branching section 14 flows into the first evaporator 15 and evaporates by absorbing heat from the indoor air blown by the first blower fan 15a (point c3 ⁇ d3 in FIG. 3). ). Thereby, indoor air is cooled.
  • the other refrigerant branched at the branching portion 14 flows into the low-stage expansion device 16 and is further depressurized isoenthalpiously (point c3 ⁇ point e3 in FIG. 3).
  • the refrigerant depressurized by the low-stage expansion device 16 flows into the second evaporator 17 and evaporates by absorbing heat from the internal air circulated by the second blower fan 17a (point e3 in FIG. 3). ⁇ f3 point). As a result, the internal air is cooled.
  • the superheated gas phase refrigerant flowing out of the first evaporator 15 flows into the nozzle portion 18a of the ejector 18 and is isentropically decompressed and injected (point d3 ⁇ point g3 in FIG. 3).
  • the refrigerant flowing out of the second evaporator 17 is sucked from the refrigerant suction port 18d of the ejector 18 by the suction action of the injection refrigerant.
  • the refrigerant injected from the nozzle portion 18a and the suction refrigerant sucked from the refrigerant suction port 18d are mixed by the mixing portion 18e of the ejector 18 and flow into the diffuser portion 18g (g3 ⁇ h3 point in FIG. 3, f3). Point ⁇ h3 point).
  • the velocity energy of the refrigerant is converted into pressure energy by expanding the refrigerant passage cross-sectional area.
  • the pressure of the mixed refrigerant of the injection refrigerant and the suction refrigerant increases (point h3 ⁇ point i3 in FIG. 3).
  • the refrigerant flowing out from the diffuser portion 18g is sucked into the compressor 11 and compressed again (point i3 ⁇ point a3 in FIG. 3).
  • the ejector refrigeration cycle 10 operates as described above, and can cool indoor air blown into the vehicle interior and internal air circulated into the cool box.
  • the refrigerant evaporation pressure (refrigerant evaporation temperature) of the second evaporator 17 is lower than the refrigerant evaporation pressure (refrigerant evaporation temperature) of the first evaporator 15, so that different temperatures are provided in the vehicle interior and the interior of the cool box.
  • the ejector refrigeration cycle 10 since the refrigerant whose pressure has been increased by the diffuser portion 18g of the ejector 18 is sucked into the compressor 11, the power consumption of the compressor 11 is reduced and the coefficient of performance (COP) of the cycle is improved. Can be made.
  • the dryness of the mixed refrigerant in the mixing portion 18e. x tends to be a relatively high value (for example, dryness x is 0.8 or more).
  • the mixed refrigerant becomes a gas-liquid two-phase refrigerant having a relatively high dryness x, as described with reference to FIGS. 20 and 21, the refrigerant pressure increase performance in the diffuser portion 18g becomes unstable.
  • the axial distance La of the nozzle portion 18a from the refrigerant injection port 18c of the nozzle portion 18a to the inlet portion 18h of the diffuser portion 18g in the mixing portion 18e is the refrigerant flowing into the inlet portion 18h. Is determined to be equal to or lower than the two-phase sound velocity ⁇ h.
  • the shape of the mixing portion 18e is formed so as to gradually reduce the refrigerant passage cross-sectional area toward the downstream side of the refrigerant flow. Furthermore, the refrigerant passage cross-sectional area of the inlet portion 18h of the diffuser portion 18g is set smaller than the refrigerant passage cross-sectional area of the refrigerant injection port 18c of the nozzle portion 18ac.
  • the flow rate of the mixed refrigerant is reduced to the two-phase sonic speed ⁇ h or less until it reaches the inlet 18h of the diffuser unit 18g so as to effectively decelerate the flow rate of the mixed refrigerant. It is trying to become.
  • the shape of the mixing portion 18e is a truncated cone shape that gradually reduces the refrigerant passage sectional area toward the refrigerant flow downstream side, and a cylindrical shape that makes the refrigerant passage sectional area constant. And the flow rate of the mixed refrigerant can be effectively decelerated by determining the distance Lb so as to satisfy the above formula F4.
  • the energy conversion efficiency (ejector efficiency ⁇ ej) in the ejector 18 can be greatly improved over the prior art.
  • the COP improvement effect due to the provision of the ejector 18 can be sufficiently obtained.
  • the ejector efficiency ⁇ ej is defined by the following formula F5.
  • ⁇ ej ⁇ hd ⁇ (Gn + Ge) ⁇ / ( ⁇ iej ⁇ Gn) (F5)
  • Gn is a flow rate of the injected refrigerant injected from the nozzle portion 18 a of the ejector 18, and is a flow rate of the refrigerant flowing through the first evaporator 15.
  • Ge is a flow rate of the suction refrigerant sucked from the refrigerant suction port 18d of the ejector 18 and is a flow rate of the refrigerant flowing through the second evaporator 17.
  • ⁇ hd is the amount of increase in enthalpy when the refrigerant is isentropically boosted in the diffuser portion 18g of the ejector 18 as shown in FIG. 3, and ⁇ iej is the ejector 18 as shown in FIG.
  • the amount of enthalpy is reduced when the pressure is reduced isentropically by the nozzle portion 18a.
  • a tapered portion 18i is formed as a refrigerant passage formed in the nozzle portion 18a so as to gradually reduce the refrigerant passage cross-sectional area toward the refrigerant injection port 18c. That is, the nozzle portion 18a of the present embodiment is configured as a so-called tapered nozzle. Furthermore, the injection part 18j is formed in the most downstream side of the refrigerant path formed in the nozzle part 18a of this embodiment.
  • the injection unit 18j is a space that guides the refrigerant from the most downstream portion of the tapered portion 18i toward the refrigerant injection port 18c. Therefore, the spray shape or the spreading direction of the jetted refrigerant injected from the refrigerant jet port 18c can be changed by the angle (spreading angle) ⁇ n in the axial section of the nozzle part 18a of the jetting part 18j. That is, the injection unit 18j can also be expressed as a space that defines the injection direction of the refrigerant injected from the refrigerant injection port 18c.
  • the injection section 18j is formed so that its inner diameter is constant or gradually expands toward the downstream side of the refrigerant flow.
  • the angle ⁇ n of the injection portion 18j in the axial section of the nozzle portion 18a is set to 0 °. That is, the injection part 18j of the present embodiment is formed by a cylindrical space extending in the axial direction of the nozzle part 18a and having a constant refrigerant passage cross-sectional area. In FIG. 5, the angle ⁇ n is illustrated as a slight value (about 1 °) in order to clarify the angle ⁇ n.
  • the axial length in which the injection portion 18j is formed in the refrigerant passage formed in the nozzle portion 18a is Lc, and the equivalent diameter of the opening area of the refrigerant injection port 18c is ⁇ Dc.
  • the distance Lc is determined so as to satisfy the following formula F6. Lc / ⁇ Dc ⁇ 1 (F6)
  • the refrigerant passage formed inside is formed as described above, so that the refrigerant injected from the refrigerant injection port 18c to the mixing portion 18e is freely expanded.
  • the refrigerant is injected from the refrigerant injection port.
  • the flow rate of the refrigerant immediately before the discharge tends to increase, and the diffuser portion 18g of the ejector 18 may not be able to exhibit the desired refrigerant pressure increase performance.
  • a general ejector collects a loss of kinetic energy when the refrigerant is decompressed at the nozzle portion by sucking the refrigerant from the refrigerant suction port by the suction action of the injected refrigerant.
  • the amount of energy recovered that is, the amount of decrease in the enthalpy indicated by ⁇ iej in FIG. 3 increases when the pressure of the refrigerant flowing into the nozzle unit is constant. It increases with.
  • V V0 + (2 ⁇ ⁇ iej) 0.5 (F7)
  • V0 is the initial speed of the refrigerant flowing into the nozzle portion.
  • the gas-phase refrigerant flowing at high speed in the refrigerant passage formed in the nozzle portion is condensed, and the gas-liquid two-phase refrigerant having a high gas-liquid density ratio (for example, the gas-liquid two-phase refrigerant having a gas-liquid density ratio of 200 or more). If it becomes, the wall surface friction of a refrigerant
  • the injection unit 18j is provided in the nozzle unit 18a configured as a tapered nozzle, and the mixed refrigerant injected from the refrigerant injection port 18c to the mixing unit 18e is freely expanded. Therefore, it is possible to accelerate the injected refrigerant in the mixing unit 18e without providing a divergent part like a Laval nozzle. That is, the refrigerant can be accelerated without causing wall friction between the refrigerant and the refrigerant passage that occurs when the refrigerant is supersonically accelerated at the divergent portion of the Laval nozzle.
  • the wall friction between the refrigerant and the refrigerant passage can be reduced, the loss of kinetic energy of the refrigerant flowing through the refrigerant passage can be suppressed, and the flow rate of the injected refrigerant can be suppressed from decreasing.
  • the energy loss of the refrigerant in the nozzle portion 18a can be reduced, and a decrease in the refrigerant pressure increase performance of the ejector 18 can be suppressed.
  • the refrigerant pressurization performance in the diffuser portion 18g can be stabilized, and the ejector efficiency ⁇ ej in the ejector 18 can be improved. Therefore, in the ejector type refrigeration cycle 10 of the present embodiment, the COP improvement effect due to the provision of the ejector 18 can be sufficiently obtained.
  • the example in which the angle ⁇ n of the injection portion 18j in the axial section of the nozzle portion 18a is 0 ° has been described.
  • the angle ⁇ n may be set larger than 0 °. That is, the injection unit 18j may be formed by a truncated cone space in which the refrigerant passage cross-sectional area gradually increases in the refrigerant flow direction.
  • This embodiment demonstrates the example which changed the structure of the ejector 18 with respect to 1st Embodiment, as shown in FIG. 6, FIG.
  • the refrigerant that has flowed from the refrigerant inflow port 18l to the upstream side of the refrigerant flow from the throat (minimum passage cross-sectional area) in the refrigerant passage formed in the nozzle portion 18a.
  • the throat minimum passage cross-sectional area
  • the refrigerant passage formed in the nozzle portion 18a Is provided with a swirl space 18k for swirling around the axis of the nozzle portion 18a.
  • this swirling space 18k is formed inside a cylindrical portion 18m provided on the upstream side of the refrigerant flow of the nozzle portion 18a. Accordingly, the cylindrical portion 18m constitutes a swirling space forming member, and in this embodiment, the swirling space forming member and the nozzle portion are integrally formed.
  • the swirling space 18k is formed in a rotating body shape, and its central axis extends coaxially with the nozzle portion 18a.
  • the rotating body shape is a three-dimensional shape formed when a plane figure is rotated around one straight line (central axis) on the same plane. More specifically, the swirl space 18k of the present embodiment is formed in a substantially cylindrical shape.
  • the refrigerant inflow passage 18n that connects the refrigerant inlet 18l and the swirl space 18k is, as seen from the central axis direction of the swirl space 18k, in the tangential direction of the inner wall surface of the swirl space 18k, as shown in FIG. It extends.
  • the refrigerant flowing into the swirl space 18k from the refrigerant inlet 18l flows along the inner wall surface of the swirl space 18k and swirls in the swirl space 18k.
  • the refrigerant pressure on the central axis side is lower than the refrigerant pressure on the outer peripheral side in the swirling space 18k. Therefore, in the present embodiment, during normal operation, the refrigerant on the central axis side in the swirl space 18k is closer to the gas-liquid two-phase side than the saturated gas line, that is, the refrigerant on the central axis side in the swirl space 18k The pressure of the refrigerant on the central axis side in the swirling space 18k is reduced so as to start the condensation.
  • Such adjustment of the refrigerant pressure on the central axis side in the swirling space 18k can be realized by adjusting the swirling flow velocity of the refrigerant swirling in the swirling space 18k. Further, the swirling flow velocity is adjusted by adjusting the ratio of the flow passage cross-sectional area between the passage cross-sectional area of the refrigerant inflow passage 18n and the axial cross-sectional area of the swirling space 18k, or on the upstream side of the nozzle portion 18a. This can be done by adjusting the aperture of the high-stage expansion device 13 arranged.
  • the ejector-type refrigeration cycle 10 of the present embodiment is configured such that a superheated gas phase refrigerant that has flowed out of the first evaporator 15 flows into the nozzle portion 18a of the ejector 18.
  • the refrigerant is condensed and accelerated while reducing the pressure in the refrigerant passage formed in the nozzle portion 18a of the ejector 18.
  • the refrigerant is swirled in the swirling space 18k, whereby the refrigerant on the swiveling center axis side in the swirling space 18k is decompressed to start condensation, and condensation nuclei are formed.
  • the gas-liquid two-phase refrigerant can flow into the nozzle portion 18a. Therefore, it is possible to suppress the occurrence of a condensation delay in the refrigerant at the nozzle portion 18a.
  • the nozzle efficiency ⁇ noz in the nozzle portion 18a can be greatly improved over the prior art. Furthermore, even in the ejector 18 that condenses and accelerates the refrigerant while reducing the pressure in the refrigerant passage formed in the nozzle portion 18a, it is possible to suppress a decrease in the refrigerant pressure increase performance in the diffuser portion 18g.
  • the nozzle efficiency ⁇ noz is energy conversion efficiency when the pressure energy of the refrigerant is converted into kinetic energy in the nozzle portion 18a.
  • the refrigerant pressurization performance in the diffuser portion 18g can be stabilized, and the ejector efficiency ⁇ ej in the ejector 18 can be improved. Therefore, in the ejector type refrigeration cycle 10 of the present embodiment, the COP improvement effect due to the provision of the ejector 18 can be sufficiently obtained.
  • the ejector 18 of the present embodiment even when the refrigerant flowing into the swirling space 18k is a gas-liquid two-phase refrigerant, the refrigerant pressure on the center side in the swirling space 18k is reduced to reduce the nozzle. Since the boiling of the refrigerant flowing into the throat (minimum passage cross-sectional area) of the portion 18a can be promoted, the nozzle efficiency ⁇ noz can be improved. (Fourth embodiment) This embodiment demonstrates the example which changed the structure of the ejector-type refrigerating cycle with respect to 1st Embodiment.
  • the branch portion 14 is disposed on the outlet side of the radiator 12, and one of the refrigerants branched at the branch portion 14 is made high.
  • the pressure is reduced until it becomes a low-pressure refrigerant by the stage side expansion device 13, and flows into the refrigerant inlet side of the first evaporator 15.
  • the other refrigerant branched by the branching section 14 is decompressed by the low stage side expansion device 16 until it becomes a low-pressure refrigerant, and flows into the refrigerant inlet side of the second evaporator 17.
  • the throttle opening of the low stage side throttle device 16 is set to be smaller than the throttle opening of the high stage side throttle device 13, which is lower than the pressure reduction amount in the high stage side throttle device 13.
  • the amount of pressure reduction in the expansion device 16 is large. Therefore, the refrigerant evaporation pressure (refrigerant evaporation temperature) in the second evaporator 17 is lower than the refrigerant evaporation pressure (refrigerant evaporation temperature) in the first evaporator 15.
  • Other configurations are the same as those of the first embodiment.
  • the flow of the refrigerant flowing out of the radiator 12 is branched at the branching section 14.
  • One refrigerant branched by the branching section 14 is decompressed by the high stage side expansion device 13 (b10 point ⁇ c10 point in FIG. 10) and flows into the first evaporator 15.
  • the other refrigerant branched by the branching section 14 is decompressed by the low-stage expansion device 16 (b10 point ⁇ e10 point in FIG. 10) and flows into the second evaporator 17. Subsequent operations are the same as those in the first embodiment.
  • the ejector 18 exhibits the same effect as that of the first embodiment, so that the COP improvement effect due to the provision of the ejector 18 can be sufficiently obtained.
  • a fixed throttle with a fixed throttle opening is adopted as the high stage side throttle device 13, and the low stage side throttle device 16 is used.
  • a temperature expansion valve is used.
  • a liquid storage tank (liquid storage part) 19 for storing excess refrigerant in the cycle is disposed between the refrigerant outlet side of the first evaporator 15 and the inlet side of the nozzle part 18 a of the ejector 18.
  • the up and down arrows in FIG. 12 indicate the up and down directions in a state where the liquid storage tank 19 is mounted on the vehicle.
  • the liquid storage tank 19 has a main body 19a, a refrigerant inflow port 19b, a refrigerant outflow port 19c, and the like.
  • the main body 19a is formed of a cylindrical member extending in the vertical direction and closed at both ends, and functions as a liquid storage part.
  • the refrigerant inflow port 19b allows the refrigerant that has flowed out of the first evaporator 15 to flow into the main body 19a.
  • the refrigerant outflow port 19c allows the gas-liquid two-phase refrigerant to flow out from the main body portion 19a to the nozzle portion 18a side of the ejector 18.
  • the refrigerant inflow port 19b is connected to the cylindrical side surface of the main body portion 19a and is constituted by a refrigerant pipe extending in the tangential direction of the cylindrical side surface of the main body portion 19a.
  • the refrigerant outflow port 19c is connected to the lower end surface (bottom surface) in the axial direction of the main body 19a, and is constituted by a refrigerant pipe extending coaxially with the main body 19a over the inside and outside of the main body 19a.
  • the upper end portion of the refrigerant outflow port 19c extends to the upper side from the connection portion of the refrigerant inflow port 19b. Further, a liquid phase refrigerant introduction hole 19d through which the liquid phase refrigerant stored in the main body portion 19a flows into the refrigerant outflow port 19c is formed below the refrigerant outflow port 19c.
  • the refrigerant flowing into the main body 19a from the refrigerant inflow port 19b is the cylinder of the main body 19a. Swirls along the inner wall surface. The gas-liquid refrigerant is separated by the action of centrifugal force generated by the swirling flow.
  • the separated liquid phase refrigerant falls downward due to the action of gravity and is stored in the main body 19a as an excess refrigerant.
  • the separated gas-phase refrigerant is mixed with the liquid-phase refrigerant flowing into the refrigerant outflow port 19c from the liquid-phase refrigerant introduction hole 19d when flowing out to the inlet side of the nozzle portion 18a via the refrigerant outflow port 19c. And flows out as a gas-liquid two-phase refrigerant.
  • the gas phase refrigerant flowing in from the refrigerant inflow port 19b is not separated into gas and liquid, and the refrigerant flows out. It flows out to the inlet side of the nozzle part 18a through the port 19c.
  • the gas-phase refrigerant that has flowed into the refrigerant outflow port 19c is mixed with the liquid-phase refrigerant that has flowed into the refrigerant outflow port 19c from the liquid-phase refrigerant introduction hole 19d and flows out as a gas-liquid two-phase refrigerant.
  • the liquid storage tank 19 of the present embodiment constitutes a gas-liquid supply unit that causes the refrigerant that has flowed out of the first evaporator 15 to flow out into the gas-liquid two-phase state to the inlet side of the nozzle unit 18a. More specifically, the liquid storage tank 19 mixes the liquid-phase refrigerant stored in the main body 19a and the refrigerant that has flowed out of the first evaporator 15, and flows it out to the inlet side of the nozzle portion 18a.
  • the dryness x of the mixed refrigerant obtained by mixing the injection refrigerant and the suction refrigerant in the mixing section 18e is also a relatively high value.
  • the dryness x is 0.8 or more).
  • the diffuser portion 18g of the ejector 18 cannot exhibit the desired refrigerant pressure-increasing performance. Furthermore, the flow rate of the suction refrigerant sucked from the refrigerant suction port 18d of the ejector 18 may decrease.
  • the ejector refrigeration cycle 10b of the present embodiment includes the liquid storage tank 19 as a gas-liquid supply unit, so that the gas-liquid two-phase refrigerant can surely flow into the nozzle unit 18a of the ejector 18. Can do. Therefore, it is possible to reliably suppress the occurrence of the condensation delay.
  • the injected refrigerant that is injected from the refrigerant injection port 18c is also surely converted into the gas-liquid two-phase refrigerant. It can suppress that dryness x raises. Therefore, it is possible to suppress the refrigerant pressure increase performance in the diffuser portion 18g from becoming unstable and the suction refrigerant flow rate sucked from the refrigerant suction port 18d from being lowered.
  • the dryness x of the injected refrigerant can be reduced to reduce the two-phase sonic speed ⁇ h of the mixed refrigerant, it occurs when the flow speed of the two-phase refrigerant changes from the supersonic state to the subsonic state.
  • the shock wave can be a gas-weak shock wave. Therefore, it is possible to effectively suppress the refrigerant pressure increase performance in the diffuser portion 18g from becoming unstable.
  • the COP can be sufficiently improved even in the cycle configuration in which the refrigerant on the downstream side of the first evaporator 15 flows into the nozzle portion 18a of the ejector 18. .
  • the gas-liquid supply unit is constituted by the liquid storage tank 19, the gas-liquid supply to the nozzle unit 18a of the ejector 18 can be reliably performed with a very simple configuration without incurring a complicated cycle configuration. Two-phase refrigerant can be introduced.
  • a temperature type expansion valve that is a variable throttle mechanism is adopted as the low stage side throttle device 16, and the degree of superheat of the refrigerant flowing out from the second evaporator 17 is a predetermined standard. It is within the range. In other words, the throttle opening degree of the low stage side expansion device 16 of the present embodiment is adjusted so that the refrigerant flowing out from the second evaporator 17 is equal to or less than a predetermined reference superheat degree.
  • the mixed refrigerant obtained by mixing the jet refrigerant that is in the gas-liquid two-phase state and the suction refrigerant that is in the gas phase state below the reference superheat degree is dried. It is possible to reliably suppress the degree x from increasing. Furthermore, the throttle opening degree of the low stage side expansion device 16 may be adjusted so that the refrigerant flowing out from the second evaporator 17 becomes a saturated gas phase refrigerant or a gas-liquid two-phase refrigerant.
  • the refrigerant pressurization performance in the diffuser portion 18g can be stabilized, and the ejector efficiency ⁇ ej in the ejector 18 can be improved.
  • the COP improvement effect by providing the ejector 18 can be sufficiently obtained.
  • a discharge refrigerant passage 20a that guides the gas-phase refrigerant discharged from the compressor 11 into the liquid storage tank 19 is added.
  • the discharge refrigerant passage 20a is desirably provided with a throttle portion for preventing the refrigerant pressure in the liquid storage tank 19 from increasing. Therefore, in the present embodiment, the discharge refrigerant passage 20a is constituted by a capillary tube.
  • the liquid storage tank 19 that is the gas-liquid supply unit of the present embodiment mixes the liquid-phase refrigerant stored in the liquid storage tank 19 and the gas-phase refrigerant discharged from the compressor 11, so that the nozzle 18 a Let it flow out to the inlet side.
  • Other configurations and operations are the same as those of the fifth embodiment. Even if the gas-liquid supply unit is configured as in the present embodiment, the same effect as in the fifth embodiment can be obtained.
  • a condensed refrigerant passage 20b that guides the liquid-phase refrigerant that has flowed out of the radiator 12 into the liquid storage tank 19 is added.
  • the condensing refrigerant passage 20b is desirably provided with a throttle portion for preventing the refrigerant pressure in the liquid storage tank 19 from increasing. Therefore, in the present embodiment, the condensed refrigerant passage 20b is constituted by a capillary tube.
  • the liquid storage tank 19 that is the gas-liquid supply unit of the present embodiment mixes the liquid-phase refrigerant that has flowed out of the radiator 12 and the gas-phase refrigerant that has flowed out of the first evaporator 15 to the inlet side of the nozzle portion 18a. It is configured to flow into Other configurations and operations are the same as those of the fifth embodiment. Even if the gas-liquid supply unit is configured as in the present embodiment, the same effect as in the fifth embodiment can be obtained.
  • the ejector 18 disclosed by 2nd, 3rd, 8th, 9th embodiment may apply to the ejector-type refrigerating cycle 10a of this embodiment.
  • the ejector 18 of the second embodiment has a cylindrical portion 18m provided on the upstream side of the refrigerant flow of the nozzle portion 18a as in the third embodiment.
  • a swirling space 18k for swirling the refrigerant flowing in from the refrigerant inlet 18l is provided.
  • Other configurations and operations of the ejector 18 and the ejector refrigeration cycle 10 are the same as those in the second embodiment.
  • the refrigerant is swirled in the swirling space 18k so that the gas-liquid two-phase refrigerant in which condensed nuclei are generated flows into the nozzle portion 18a.
  • the nozzle efficiency ⁇ noz can be improved. Therefore, it is possible to suppress a decrease in the refrigerant pressure increase performance in the diffuser portion 18g.
  • the refrigerant injected from the refrigerant injection port 18c of the nozzle portion 18a is freely expanded, it is possible to suppress an increase in wall friction. Therefore, the energy loss of the refrigerant in the nozzle portion 18a can be reduced, and the decrease in the refrigerant pressure increase performance of the ejector 18 can be suppressed.
  • the refrigerant pressure-increasing performance in the diffuser portion 18g can be stabilized, and the ejector efficiency ⁇ ej in the ejector 18 can be improved. Therefore, in the ejector type refrigeration cycle 10 of the present embodiment, the COP improvement effect due to the provision of the ejector 18 can be sufficiently obtained.
  • the nozzle portion 18a of the ejector 18 the fixed nozzle in which the refrigerant passage cross-sectional area of the minimum passage cross-sectional area portion formed at the inlet portion of the injection portion 18j is fixed has been described. In the embodiment, as shown in FIG. 16, an example will be described in which a variable nozzle configured to change the refrigerant passage cross-sectional area of the minimum passage cross-sectional area portion is adopted.
  • the ejector 18 of the present embodiment includes a needle valve 18y and a stepping motor 18x.
  • the needle valve 18y is a valve body that changes the refrigerant passage cross-sectional area of the nozzle portion 18a.
  • the stepping motor 18x is a drive unit that displaces the needle valve 18y.
  • the needle valve 18y is formed in the shape of a needle whose central axis is arranged coaxially with the central axis of the nozzle portion 18a. More specifically, the needle valve 18y is formed in a shape that tapers toward the downstream side of the refrigerant flow, and the taper tip portion on the most downstream side is further on the downstream side of the refrigerant flow than the refrigerant injection port 18c of the nozzle portion 18a. It arrange
  • the stepping motor 18x is disposed on the refrigerant inlet 18l side of the nozzle portion 18a, and displaces the needle valve 18y in the axial direction of the nozzle portion 18a. As a result, the cross-sectional area of the refrigerant passage having an annular cross section formed between the inner peripheral wall surface of the nozzle portion 18a and the outer peripheral wall surface of the needle valve 18y is changed.
  • the operation of the stepping motor 18x is controlled by a control signal output from the control device.
  • the nozzle part 18a is comprised as a variable nozzle, the refrigerant
  • the nozzle portion 18a of the present embodiment is configured as a plug nozzle, the injected refrigerant can be injected from the refrigerant injection port 18c to the mixing portion 18e along the outer surface of the needle valve 18y. Therefore, even if the flow rate of the refrigerant flowing into the nozzle portion 18a changes, the injected refrigerant can be easily expanded freely, and the kinetic energy of the refrigerant flowing through the refrigerant passage is reduced by reducing the wall friction between the refrigerant and the refrigerant passage. Loss can be suppressed.
  • the needle valve 18y of the present embodiment is disposed so as to penetrate through the swirling space 18k, so that friction between the refrigerant swirling in the swirling space 18k and the inner wall of the nozzle portion 18a occurs. It is easy to produce condensed nuclei.
  • the needle valve 18y employs a shape that tapers toward the downstream side of the refrigerant flow.
  • a taper shape that tapers toward the upstream side of the refrigerant flow may be employed.
  • This embodiment demonstrates the example which changed the structure of the ejector-type refrigerating cycle 10a with respect to 4th Embodiment.
  • a high-stage ejector 131 is employed as the first pressure reducing unit in place of the high-stage expansion device 13.
  • the basic configuration of the high-stage ejector 131 is the same as that of the ejector 18 described above. Therefore, similarly to the ejector 18, the high-stage ejector 131 has a high-stage nozzle portion 131a and a high-stage body portion 131b.
  • the high-stage nozzle portion 131a depressurizes the refrigerant.
  • the high-stage body section 131b is formed with a high-stage refrigerant suction port 131d that sucks the refrigerant that has flowed out of the first evaporator 15, and a high-stage diffuser section (high-stage booster section) 131g that boosts the mixed refrigerant.
  • the high-stage ejector 131 since the liquid-phase refrigerant condensed by the radiator 12 can be caused to flow into the high-stage nozzle portion 131a of the high-stage ejector 131 of the present embodiment, the high-stage ejector 131 has a high-stage side.
  • the gas-liquid two-phase refrigerant having a high degree of dryness does not flow into the nozzle portion 131a so that the high stage side diffuser portion 131g cannot exhibit the desired boosting performance.
  • the high-stage ejector 131 is not of the same configuration as the ejector 18 described above, and the ejector-type refrigeration cycle 10a is introduced when the liquid-phase refrigerant is introduced into the high-stage nozzle portion 131a.
  • the thing set so that high COP can be demonstrated as a whole is adopted.
  • a gas-liquid separator 21 that separates the gas-liquid refrigerant flowing out from the high stage side diffuser portion 131g of the high stage side ejector 131 is connected to the high stage side diffuser portion 131g outlet side of the high stage side ejector 131.
  • a refrigerant inlet of the first evaporator 15 is connected to the liquid-phase refrigerant outlet of the gas-liquid separator 21 via a fixed throttle 22, and a high-stage ejector is connected to the refrigerant outlet of the first evaporator 15.
  • a refrigerant suction port 131 is connected.
  • the gas-phase refrigerant outlet of the gas-liquid separator 21 is connected to the inlet side of the nozzle portion 18 a of the ejector 18.
  • Other configurations are the same as those of the fourth embodiment.
  • the flow of the liquid-phase refrigerant that has flowed out of the radiator 12 is branched at the branch portion 14.
  • One refrigerant branched by the branching portion 14 flows into the high-stage nozzle portion 131a of the high-stage ejector 131 and is isentropically decompressed and injected.
  • the refrigerant flowing out of the first evaporator 15 is sucked from the high stage side refrigerant suction port 131d of the high stage side ejector 131 by the suction action of the injection refrigerant.
  • the mixed refrigerant of the refrigerant injected from the high stage side nozzle part 131a and the suction refrigerant sucked from the high stage side refrigerant suction port 131d flows into the high stage side diffuser part 131g and is pressurized.
  • the refrigerant that has flowed out of the high stage side diffuser portion 131g flows into the gas-liquid separator 21 and is gas-liquid separated. Then, the liquid phase refrigerant separated by the gas-liquid separator 21 flows into the first evaporator 15 via the fixed throttle 22. On the other hand, the gas-phase refrigerant separated by the gas-liquid separator 21 flows into the nozzle portion 18 a of the ejector 18. Other operations are the same as those in the fourth embodiment.
  • the same effects as those of the fourth embodiment can be obtained. Furthermore, the power consumption of the compressor 11 can be reduced by the boosting action of the high-stage ejector 131, and the COP of the entire cycle can be further improved.
  • the ejector refrigeration cycle 10a that employs the high-stage ejector 131 as the first decompression unit is not limited to the cycle configuration shown in FIG. 18, but may be configured as shown in FIG. 19, for example.
  • the refrigerant inlet side of the first evaporator 15 is connected to the outlet side of the high stage side diffuser part 131g of the high stage side ejector 131, and further, the branch part ( A second branch portion 14 a that further branches the refrigerant flow is connected to the other refrigerant outlet of the first branch portion 14.
  • the refrigerant inlet of the third evaporator 23 is connected to one refrigerant outlet of the second branch portion 14a via the fixed throttle 132, and the high stage ejector 131 of the third evaporator 23 is connected to the refrigerant outlet of the third evaporator 23.
  • the high stage side refrigerant suction port 131d is connected.
  • the third evaporator 23 performs heat exchange between the low-pressure refrigerant decompressed by the fixed throttle 132 and the air blown from the third blower fan 23a, thereby evaporating the low-pressure refrigerant and exerting an endothermic effect. It is an exchanger.
  • the refrigerant inlet of the second evaporator 17 is connected to the other refrigerant outlet of the second branch portion 14 a via the low-stage expansion device 16.
  • Other configurations are the same as those of the fourth embodiment. Even with such a cycle configuration, the COP of the entire cycle can be further improved by the boosting action of the high-stage ejector 131. (Other embodiments)
  • the present disclosure is not limited to the above-described embodiment, and can be variously modified as follows without departing from the spirit of the present disclosure.
  • the ejector refrigeration cycle 10, 10 a, 10 b including the ejector 18 is applied to a vehicle refrigeration cycle apparatus, the indoor air is cooled by the first evaporator 15, and the second evaporator
  • application of the ejector-type refrigerating cycle 10, 10a, 10b is not limited to this.
  • the front seat air blown to the vehicle front seat side is cooled by the first evaporator 15, and the second evaporator 17 sends the air to the vehicle rear seat side.
  • the rear seat air may be cooled.
  • the first evaporator 15 blows air to food / drinking water at a low temperature (specifically, 0 ° C. to 10 ° C.). Air in the refrigerator compartment is cooled, and in the second evaporator 17, the air in the freezer compartment blown to the freezer compartment where the food is stored frozen at an extremely low temperature (specifically, -20 ° C to -10 ° C). You may make it cool.
  • a low temperature specifically, 0 ° C. to 10 ° C.
  • the gas-liquid refrigerant flowing out of the diffuser 18g is separated between the outlet side of the diffuser 18g of the ejector 18 and the inlet side of the compressor 11 and separated.
  • An accumulator that causes the gas-phase refrigerant to flow out to the suction port side of the compressor 11 may be disposed.
  • a beneficiary device may be disposed on the refrigerant outlet side of the radiator 12 so as to separate the gas-liquid refrigerant flowing out of the radiator 12 and flow out the liquid-phase refrigerant downstream.
  • an internal heat exchanger for exchanging heat between the high-temperature refrigerant flowing out from the radiator 12 and the low-temperature refrigerant sucked into the compressor 11 may be arranged.
  • an auxiliary pump for refrigerant pressure feeding may be provided between the refrigerant outlet side of the second evaporator 17 and the refrigerant suction port 18 d of the ejector 18.
  • the high stage side throttle device 13 and the low stage side throttle device 16 include a valve body configured to be able to change the throttle opening and an electric actuator including a stepping motor that changes the throttle opening of the valve body.
  • An electric variable aperture mechanism may be employed.
  • the heat radiator 12 including the heat exchanging unit that performs heat exchange between the refrigerant discharged from the compressor 11 and the outside air has been described.
  • a condensing unit that condenses the refrigerant discharged from the compressor 11 by exchanging heat between the refrigerant discharged from the compressor 11 and the outside air a modulator unit that separates the gas-liquid of the refrigerant flowing out from the condensing unit, and a modulator unit A so-called subcool type condenser having a supercooling section for supercooling the liquid phase refrigerant by exchanging heat between the liquid phase refrigerant flowing out from the outside air and the outside air may be employed.
  • the constituent members such as the body portion 18b of the ejector 18 are formed of metal
  • the material is not limited as long as the functions of the respective constituent members can be exhibited. That is, you may form these structural members with resin.
  • the opening diameter of the refrigerant injection port 18c may be set smaller than the opening diameter of the inlet portion 18h.
  • the inlet portion 18h when the opening diameter of the inlet portion 18h is set larger than the opening diameter of the refrigerant injection port 18c, the inlet portion 18h is provided with a protrusion that protrudes into the refrigerant passage so that the refrigerant passage of the inlet portion 18h is cut off.
  • the area may be smaller than the refrigerant passage cross-sectional area of the refrigerant injection port 18c.
  • R134a is employed as the refrigerant
  • the refrigerant is not limited to this.
  • R600a, R1234yf, R410A, R404A, R32, R1234yfxf, R407C, etc. can be employed.
  • the above embodiments may be appropriately combined within a feasible range.
  • the gas-liquid supply unit described in the fifth to seventh embodiments may be applied to the ejector refrigeration cycle 10a described in the fourth embodiment.
  • the ejector 18 disclosed in the second, third, eighth, and ninth embodiments may be applied as the ejector 18 of the ejector refrigeration cycle 10a described in the tenth embodiment.
  • the radiator 12 is an outdoor heat exchanger that exchanges heat between the refrigerant and the outside air
  • the first and second evaporators 15 and 17 are used as utilization side heat exchangers that cool the air.
  • the first and second evaporators 15 and 17 are configured as outdoor heat exchangers that absorb heat from a heat source such as outside air
  • the radiator 12 is an indoor heat exchanger that heats a heated fluid such as air or water.
  • the present disclosure may be applied to a heat pump cycle configured as follows.

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  • Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Thermal Sciences (AREA)
  • Fluid Mechanics (AREA)
  • Jet Pumps And Other Pumps (AREA)

Abstract

Selon la présente invention, un réservoir de stockage (19) qui stocke le réfrigérant excédentaire du cycle sous forme de réfrigérant en phase liquide est disposé entre le côté sortie d'un premier évaporateur (15) et le côté entrée de la buse (18a) d'un éjecteur (18), le réfrigérant en phase liquide stocké dans le réservoir de stockage (19) étant mélangé avec réfrigérant en phase gazeuse s'écoulant hors du premier évaporateur (15), et le réfrigérant dans un état à deux phases gaz-liquide étant amené à pénétrer dans la buse (18a). En conséquence, le retard de condensation dans la buse (18a) est supprimé, une augmentation de la sécheresse (x) du réfrigérant mélangé résultant du réfrigérant entrant prélevé dans un deuxième évaporateur (17) et de réfrigérant injecté à partir de la buse (18a) est supprimée, et le coefficient de performance (COP) d'un cycle de réfrigération d'éjecteur (10b) est suffisamment augmentée.
PCT/JP2014/002784 2013-06-18 2014-05-27 Cycle de réfrigération d'éjecteur WO2014203460A1 (fr)

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Cited By (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP3246637A4 (fr) * 2015-01-16 2018-12-26 Mitsubishi Electric Corporation Dispositif à cycle frigorifique
US10603985B2 (en) 2014-09-04 2020-03-31 Denso Corporation Liquid ejector and ejector refrigeration cycle

Families Citing this family (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP6277869B2 (ja) 2014-05-30 2018-02-14 株式会社デンソー エジェクタ式冷凍サイクル

Citations (7)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP2003269399A (ja) * 2002-03-12 2003-09-25 National Institute Of Advanced Industrial & Technology エジェクタ式二元冷凍装置及びその装置に用いるエジェクタ
JP2007046806A (ja) * 2005-08-08 2007-02-22 Denso Corp エジェクタ式サイクル
JP2009133624A (ja) * 2005-03-14 2009-06-18 Mitsubishi Electric Corp 冷凍空調装置
JP2009236330A (ja) * 2008-03-25 2009-10-15 Calsonic Kansei Corp 冷却システム
WO2012074650A1 (fr) * 2010-11-30 2012-06-07 Carrier Corporation Éjecteur
JP2012149790A (ja) * 2011-01-17 2012-08-09 Mitsubishi Electric Corp 冷凍サイクル装置及び流路切替装置及び流路切替方法
WO2014076905A1 (fr) * 2012-11-16 2014-05-22 株式会社デンソー Appareil à cycle de réfrigération

Patent Citations (7)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JP2003269399A (ja) * 2002-03-12 2003-09-25 National Institute Of Advanced Industrial & Technology エジェクタ式二元冷凍装置及びその装置に用いるエジェクタ
JP2009133624A (ja) * 2005-03-14 2009-06-18 Mitsubishi Electric Corp 冷凍空調装置
JP2007046806A (ja) * 2005-08-08 2007-02-22 Denso Corp エジェクタ式サイクル
JP2009236330A (ja) * 2008-03-25 2009-10-15 Calsonic Kansei Corp 冷却システム
WO2012074650A1 (fr) * 2010-11-30 2012-06-07 Carrier Corporation Éjecteur
JP2012149790A (ja) * 2011-01-17 2012-08-09 Mitsubishi Electric Corp 冷凍サイクル装置及び流路切替装置及び流路切替方法
WO2014076905A1 (fr) * 2012-11-16 2014-05-22 株式会社デンソー Appareil à cycle de réfrigération

Cited By (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US10603985B2 (en) 2014-09-04 2020-03-31 Denso Corporation Liquid ejector and ejector refrigeration cycle
EP3246637A4 (fr) * 2015-01-16 2018-12-26 Mitsubishi Electric Corporation Dispositif à cycle frigorifique

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