Ejector, ejector refrigeration cycle, and method of operating nozzle device

Abstract

This ejector is provided with a nozzle section (30, 301) and a body section (40). The nozzle section (30, 301) has formed therein a nozzle passage (32) which leads from a nozzle inlet (32a), where decompression of a fluid is started, to an injection port (32e) from which the fluid is ejected. A first passage (321) and a second passage (322) are formed as the nozzle passage (32). The first passage (321) has a passage cross-sectional area that contracts going from the nozzle inlet (32a) toward the downstream side of a direction of flow of the fluid. The second passage (322) has a passage cross-sectional area that expands going from a first outlet (32b) of the first passage (321) toward the downstream side of the direction of flow of the fluid. In the first passage (321), bubbles are generated in the fluid in a liquid phase, and in the second passage (322), the fluid reaches a critical state.

Description

[Replenishment based on Rule 26 02.12.2024] Ejector, ejector type refrigeration cycle, and nozzle device operation method CROSS-REFERENCE TO RELATED APPLICATIONS

[0001] This application is based on Japanese Patent Application No. 2023-212957 filed on December 18, 2023, and Japanese Patent Application No. 2024-123251 filed on July 30, 2024, the contents of which are incorporated herein by reference.

[0002] The present disclosure relates to an ejector that exerts a fluid pressure increasing effect, an ejector-type refrigeration cycle including the ejector, and an operating method of a nozzle device applicable to the ejector.

[0003] Patent Document 1 discloses an ejector-type refrigeration cycle including an ejector. In the ejector-type refrigeration cycle, the pressure of the refrigerant drawn into the compressor can be increased above the refrigerant evaporation pressure in the evaporator by the fluid pressure increasing action of the ejector. As a result, the ejector-type refrigeration cycle can reduce the power consumption of the compressor and improve the coefficient of performance (COP) of the cycle.

[0004] In order to improve the pressure-boosting capability of the ejector, it is effective to improve the nozzle efficiency ηnoz of the nozzle portion of the ejector. The nozzle efficiency ηnoz is the energy conversion efficiency when converting the pressure energy of the fluid into velocity energy in a nozzle passage formed in the nozzle portion. The nozzle passage is a refrigerant passage that extends from the nozzle inlet, where decompression of the fluid begins in the nozzle portion, to the injection port from which the fluid is injected. Therefore, in order to improve the nozzle efficiency ηnoz, it is effective to increase the velocity of the injected refrigerant injected from the injection port.

[0005] Therefore, the ejector of Patent Document 1 employs a two-stage expansion nozzle that reduces the pressure of the refrigerant in two stages. More specifically, the nozzle of Patent Document 1 has a variable throttle as a first-stage nozzle and a fixed throttle as a second-stage nozzle.

[0006] In the ejector refrigeration cycle of Patent Document 1, supercooled liquid refrigerant is introduced into the inlet of the nozzle, generating fine bubbles in the liquid refrigerant near the wall surface of the throat of the first-stage nozzle. Furthermore, the liquid refrigerant containing the bubbles generated in the first-stage nozzle is introduced into the second-stage nozzle, generating more bubbles and growing the bubbles in the second-stage nozzle.

[0007] As a result, the ejector of Patent Document 1 attempts to increase the velocity of the refrigerant injected from the injection port of the nozzle portion, thereby improving the nozzle efficiency ηnoz.

[0008] Japanese Patent Application Laid-Open No. 2008-111662

[0009] However, according to the study by the present inventors, in the ejector of Patent Document 1, when a load fluctuation occurs in the ejector-type refrigeration cycle, the effect of improving the nozzle efficiency ηnoz by adopting a two-stage expansion nozzle unit cannot be fully obtained.The present inventors investigated the cause and found that the cause is that the ejector of Patent Document 1 adopts a variable throttle as the first-stage nozzle.

[0010] More specifically, when a load fluctuation occurs in the ejector-type refrigeration cycle apparatus, the flow rate of the refrigerant flowing into the nozzle of the ejector also fluctuates. Therefore, if a variable throttle is used as the first-stage nozzle, the flow rate of the refrigerant flowing through the nozzle passage can be appropriately adjusted by adjusting the passage cross-sectional area of ​​the throat of the first-stage nozzle in accordance with the load fluctuation of the ejector-type refrigeration cycle apparatus.

[0011] On the other hand, if the passage cross-sectional area of ​​the throat of the first-stage nozzle is changed in accordance with load fluctuations in the ejector-type refrigeration cycle, fluctuations will occur in the generation pattern and amount of bubbles generated in the liquid phase refrigerant near the wall surface of the throat of the first-stage nozzle.

[0012] Therefore, in the ejector of Patent Document 1, when a load fluctuation occurs in the ejector-type refrigeration cycle, the generation of bubbles in the liquid phase refrigerant in the nozzle passage may be insufficient, or the refrigerant in the nozzle passage may be in a thermodynamic non-equilibrium state, which may result in the nozzle efficiency ηnoz not being sufficiently improved and the fluid pressure increasing effect not being sufficient.

[0013] Here, the thermodynamic non-equilibrium state is a state in which the temperature of the liquid-phase refrigerant particles (hereinafter referred to as droplets) contained in the gas-liquid two-phase refrigerant is higher than the saturation temperature of the refrigerant at the same pressure, while the thermodynamic equilibrium state is a state in which the temperature of the droplets contained in the gas-liquid two-phase refrigerant is equal to the saturation temperature of the refrigerant at the same pressure.

[0014] If the fluid in the ejector is in a thermodynamic non-equilibrium state, the velocity of the fluid in the ejector cannot be sufficiently accelerated, and the ejector cannot exert a sufficient fluid pressure increasing effect.

[0015] In view of the above, a first object of the present disclosure is to provide an ejector that can exhibit high pressure-boosting capability regardless of fluctuations in the flow rate of fluid flowing into a nozzle portion.

[0016] A second object of the present disclosure is to provide an ejector refrigeration cycle device that can achieve a high coefficient of performance regardless of load fluctuations.

[0017] A third object of the present disclosure is to provide a method for operating a nozzle device that can improve nozzle efficiency regardless of fluctuations in the flow rate of the inflowing fluid.

[0018] The ejector according to a first aspect of the present disclosure includes a nozzle portion and a body portion. The nozzle portion reduces the pressure of a driving-side fluid and ejects it. The body portion is formed with a suction port, a mixing portion, and a refrigerant outlet. The suction port is a portion that sucks in a suction-side fluid. The mixing portion is a portion that mixes the ejected fluid ejected from the nozzle portion with the suction fluid sucked through the suction port. The refrigerant outlet is a portion that discharges the mixed fluid mixed in the mixing portion.

[0019] The nozzle section has a nozzle passage formed therein, extending from a nozzle inlet where pressure reduction of the fluid begins to an injection port from which the fluid is injected. The nozzle passage further includes a first passage and a second passage. The first passage has a cross-sectional area that decreases from the nozzle inlet toward the downstream side in the fluid flow direction. The second passage has a cross-sectional area that increases from a first outlet of the first passage toward the downstream side in the fluid flow direction.

[0020] In the first passage, gas bubbles are generated in the fluid in the liquid phase, and in the second passage, the fluid is brought to a critical state.

[0021] Here, the second passage is a fluid passage extending from the inlet of the second passage to the outlet of the second passage. That is, the second passage is a fluid passage extending from the first outlet corresponding to the so-called throat portion to the second outlet in the embodiment described below. Therefore, the second passage includes the first outlet and the second outlet. The passage cross-sectional area is the area of ​​a cross section perpendicular to the axial direction.

[0022] According to this, bubbles are generated in the liquid-phase fluid in the first passage, and the flow state of the fluid can be changed from a bubbly flow to a mist flow. Furthermore, in the second passage, the fluid that has become a mist flow reaches a critical state. Therefore, in the second passage, the velocity of the fluid that has become a mist flow can be made to approach the sonic velocity, thereby increasing the velocity of the injected fluid. This improves the nozzle efficiency ηnoz of the nozzle portion and improves the pressure-boosting capability of the ejector.

[0023] Here, bubbly flow refers to a fluid flow pattern in which bubbles exist in a supercooled liquid-phase fluid. During the decompression process of a bubbly flow, the bubble diameter expands and the bubbles grow. A mist flow refers to a fluid flow pattern in which particles of liquid-phase fluid (hereinafter referred to as droplets) exist in a saturated gas-phase fluid. During the decompression process of a mist flow, the droplets, attracted by the expanding saturated gas-phase refrigerant, gain velocity. With the same mass flow rate of refrigerant, a mist flow with a higher volume fraction of gas-phase fluid will have a faster velocity than a bubbly flow.

[0024] Furthermore, when the flow velocity of a compressible fluid reaches the speed of sound, the product of the density and velocity of the flow reaches a maximum. Therefore, the state in which the product of the density and velocity of the flow is maximum is referred to as the critical state. Furthermore, the flow rate of the fluid when the critical state is reached is referred to as the critical flow rate.

[0025] The critical flow rate changes depending on the cross-sectional area of ​​the second passage at which the fluid reaches the critical state. Therefore, when a flow rate fluctuation occurs in the driving-side fluid, the flow rate of the fluid flowing through the nozzle passage can be adjusted by changing the position in the second passage at which the fluid reaches the critical state. Therefore, it is not necessary to change the cross-sectional area of ​​the first outlet to change the flow rate of the fluid flowing through the nozzle passage.

[0026] Therefore, even if the flow rate of the fluid flowing into the nozzle passage fluctuates, the type and amount of bubbles generated in the liquid-phase fluid near the wall surface of the first outlet are unlikely to be affected. As a result, sufficient bubbles can be generated in the liquid-phase fluid, regardless of the flow rate fluctuation of the fluid flowing into the nozzle portion, and the velocity of the sprayed refrigerant can be increased.

[0027] That is, according to the ejector of the first aspect of the present disclosure, it is possible to provide an ejector that can exhibit high pressure-boosting capability regardless of fluctuations in the flow rate of the fluid flowing into the nozzle portion.

[0028] Also, an ejector according to a second aspect of the present disclosure includes a nozzle portion and a body portion. The nozzle portion reduces the pressure of a driving-side fluid and injects it. The body portion is formed with a suction port, a mixing portion, and a refrigerant outlet. The suction port is a portion that sucks in a suction-side fluid. The mixing portion is a portion that mixes the injection fluid injected from the nozzle portion with the suction fluid sucked through the suction port. The refrigerant outlet is a portion that discharges the mixed fluid mixed in the mixing portion.

[0029] The nozzle section has a nozzle passage formed therein, extending from a nozzle inlet where pressure reduction of the fluid begins to an injection port from which the fluid is injected. The nozzle passage further includes a first passage and a second passage. The first passage has a cross-sectional area that decreases from the nozzle inlet toward the downstream side in the fluid flow direction. The second passage has a cross-sectional area that increases from a first outlet of the first passage toward the downstream side in the fluid flow direction.

[0030] At the first outlet, the gas-liquid two-phase fluid is in a thermodynamic non-equilibrium state, and in the second passage, the fluid reaches a critical state, and further, in either the second passage or the mixing section, the fluid reaches a thermodynamic equilibrium state.

[0031] With this configuration, even if the refrigerant is in a thermodynamic non-equilibrium state in the first passage, it is in a thermodynamic equilibrium state in either the second passage or the mixing section, so that the pressure energy of the driving side fluid can be effectively converted into kinetic energy in either the second passage or the mixing section within the ejector.

[0032] Furthermore, when a flow rate fluctuation occurs in the driving-side fluid, the flow rate of the fluid flowing through the nozzle passage can be adjusted by changing the position in the second passage where the fluid reaches a critical state, so there is no need to change the passage cross-sectional area of ​​the first outlet in order to change the flow rate of the fluid flowing through the nozzle passage.

[0033] That is, according to the ejector of the second aspect of the present disclosure, it is possible to provide an ejector that can exhibit high pressure-boosting capability regardless of fluctuations in the flow rate of the fluid flowing into the nozzle portion.

[0034] In the ejector, the driving-side fluid does not necessarily have to be a liquid-phase fluid, but may be any fluid that can become a bubbly flow in the first passage. For example, the driving-side fluid may be a fluid in a supercritical state, in which the pressure of the fluid is equal to or higher than the critical point pressure and the temperature of the fluid is equal to or higher than the critical point temperature.

[0035] A third aspect of the present disclosure provides an ejector-type refrigeration cycle including a compressor, a heat radiating unit, an evaporating unit, and an ejector. The compressor compresses and discharges a refrigerant. The heat radiating unit dissipates heat from the refrigerant discharged from the compressor. The evaporating unit evaporates the refrigerant.

[0036] The ejector has a nozzle portion and a body portion. The nozzle portion reduces the pressure of the refrigerant flowing out from the heat dissipation portion and sprays it. The body portion is formed with a suction port, a mixing portion, and a refrigerant outlet. The suction port is a portion that sucks in the refrigerant flowing out from the evaporation portion. The mixing portion is a portion that mixes the refrigerant sprayed from the nozzle portion with the suction refrigerant sucked in from the suction port. The refrigerant outlet is a portion that discharges the mixed refrigerant mixed in the mixing portion to the suction port side of the compressor.

[0037] The nozzle section has a nozzle passage extending from a nozzle inlet where decompression of the refrigerant begins to an injection port from which the refrigerant is injected. The nozzle passage includes a first passage and a second passage. The first passage has a cross-sectional area that decreases from the nozzle inlet toward the downstream side in the refrigerant flow direction. The second passage has a cross-sectional area that increases from a first outlet of the first passage toward the downstream side in the refrigerant flow direction.

[0038] In the first passage, bubbles are generated in the liquid phase refrigerant, and in the second passage, the refrigerant reaches a critical state.

[0039] According to this, since the ejector of the first aspect of the present disclosure is provided, the ejector can exhibit high pressure-boosting capability regardless of fluctuations in the flow rate of the refrigerant flowing into the nozzle. That is, according to the ejector-type refrigeration cycle device of the third aspect of the present disclosure, it is possible to provide an ejector-type refrigeration cycle device that can exhibit a high coefficient of performance regardless of load fluctuations.

[0040] A fourth aspect of the present disclosure provides an ejector-type refrigeration cycle including a compressor, a heat radiating unit, an evaporating unit, and an ejector. The compressor compresses and discharges a refrigerant. The heat radiating unit dissipates heat from the refrigerant discharged from the compressor. The evaporating unit evaporates the refrigerant.

[0041] The ejector has a nozzle portion and a body portion. The nozzle portion reduces the pressure of the refrigerant flowing out from the heat dissipation portion and sprays it. The body portion is formed with a suction port, a mixing portion, and a refrigerant outlet. The suction port is a portion that sucks in the refrigerant flowing out from the evaporation portion. The mixing portion is a portion that mixes the refrigerant sprayed from the nozzle portion with the suction refrigerant sucked in from the suction port. The refrigerant outlet is a portion that discharges the mixed refrigerant mixed in the mixing portion to the suction port side of the compressor.

[0042] The nozzle section has a nozzle passage extending from a nozzle inlet where decompression of the refrigerant begins to an injection port from which the refrigerant is injected. The nozzle passage includes a first passage and a second passage. The first passage has a cross-sectional area that decreases from the nozzle inlet toward the downstream side in the refrigerant flow direction. The second passage has a cross-sectional area that increases from a first outlet of the first passage toward the downstream side in the refrigerant flow direction.

[0043] At the first outlet, the refrigerant in a gas-liquid two-phase state is in a thermodynamic non-equilibrium state, and in the second passage, the refrigerant reaches a critical state, and further, the refrigerant reaches a thermodynamic equilibrium state either in the second passage or in the mixing section.

[0044] According to this, since the ejector of the second aspect of the present disclosure is provided, the ejector can exhibit high pressure-boosting capability regardless of fluctuations in the flow rate of the fluid flowing into the nozzle portion. That is, according to the ejector-type refrigeration cycle device of the fourth aspect of the present disclosure, it is possible to provide an ejector-type refrigeration cycle device that can exhibit a high coefficient of performance regardless of load fluctuations.

[0045] In the ejector-type refrigeration cycle, the refrigerant flowing out from the heat dissipation portion does not necessarily have to be a liquid refrigerant, but may be a refrigerant that can become a bubbly flow in the first passage. For example, the refrigerant may be a supercritical refrigerant.

[0046] A fifth aspect of the present disclosure provides a method for operating a nozzle device including a nozzle portion that reduces the pressure of a fluid.

[0047] The nozzle section has a nozzle passage formed therein, extending from a nozzle inlet where pressure reduction of the fluid begins to an injection port from which the fluid is injected. The nozzle passage includes a first passage and a second passage. The first passage has a cross-sectional area that decreases from the nozzle inlet toward the downstream side in the fluid flow direction. The second passage has a cross-sectional area that increases from a first outlet of the first passage toward the downstream side in the fluid flow direction.

[0048] Bubbles are generated in the liquid phase fluid in the first passage, the fluid is brought to a critical state in the second passage, and the position at which the fluid reaches a critical state in the second passage is changed by adjusting the velocity of the fluid flowing into the nozzle inlet.

[0049] According to this, bubbles are generated in the liquid phase fluid in the first passage, and the flow state of the fluid can be changed from a bubbly flow to a mist flow. Furthermore, in the second passage, the mist flow fluid is brought to a critical state. Therefore, in the second passage, the velocity of the mist flow fluid can be brought close to the sonic velocity, and the velocity of the injected fluid can be increased. As a result, the nozzle efficiency ηnoz can be improved.

[0050] Furthermore, by adjusting the velocity of the fluid flowing into the nozzle inlet, the position in the second passage where the fluid reaches a critical state can be changed, so that the critical flow rate can be changed according to the cross-sectional area of ​​the passage at the position where the fluid reaches a critical state. In other words, there is no need to change the cross-sectional area of ​​the passage at the first outlet in order to change the flow rate of the fluid flowing through the nozzle passage.

[0051] Therefore, even if the flow rate of the fluid flowing into the nozzle portion fluctuates, the type and amount of bubbles generated in the liquid-phase fluid near the wall surface of the first outlet are not affected, and as a result, sufficient bubbles can be generated in the liquid-phase fluid, regardless of the flow rate fluctuation of the fluid flowing in, making it possible to increase the velocity of the jetted fluid.

[0052] That is, according to the operating method of the nozzle device of the fifth aspect of the present disclosure, it is possible to provide an operating method of the nozzle device that can improve the nozzle efficiency ηnoz regardless of fluctuations in the flow rate of the inflowing fluid.

[0053] A sixth aspect of the present disclosure provides a method for operating a nozzle device including a nozzle portion that reduces the pressure of a fluid.

[0054] The nozzle section has a nozzle passage formed therein, extending from a nozzle inlet where pressure reduction of the fluid begins to an injection port from which the fluid is injected. The nozzle passage includes a first passage and a second passage. The first passage has a cross-sectional area that decreases from the nozzle inlet toward the downstream side in the fluid flow direction. The second passage has a cross-sectional area that increases from a first outlet of the first passage toward the downstream side in the fluid flow direction.

[0055] The gas-liquid two-phase fluid that has reached a thermodynamic non-equilibrium state in the first passage is brought to a critical state and to a thermodynamic equilibrium state in the second passage, and further, by adjusting the velocity of the fluid flowing into the nozzle inlet, the position in the second passage where the fluid reaches a thermodynamic equilibrium state is changed.

[0056] According to this, even if the fluid flowing out of the first outlet is in a thermodynamic non-equilibrium state, the fluid is brought to a critical state and a thermodynamic equilibrium state in the second passage. Therefore, at the outlet of the second passage, the pressure energy of the fluid flowing into the nozzle inlet can be effectively converted into velocity energy, thereby increasing the velocity of the injected fluid. As a result, the nozzle efficiency ηnoz can be improved.

[0057] Furthermore, by adjusting the velocity of the fluid flowing into the nozzle inlet, the position where the fluid reaches a thermodynamic equilibrium state in the second passage is changed, thereby changing the velocity and density at which the fluid reaches a critical state, thereby changing the critical flow rate of the fluid. In other words, there is no need to change the cross-sectional area of ​​the first outlet in order to change the flow rate of the fluid flowing through the nozzle passage.

[0058] Therefore, even if the flow rate of the fluid flowing into the nozzle portion fluctuates, the type and amount of bubbles generated in the liquid-phase fluid near the wall surface of the first outlet are not affected, and as a result, sufficient bubbles can be generated in the liquid-phase fluid, regardless of the flow rate fluctuation of the fluid flowing in, making it possible to increase the velocity of the jetted fluid.

[0059] That is, according to the method for operating a nozzle device of the sixth aspect of the present disclosure, it is possible to provide a method for operating a nozzle device that can improve the nozzle efficiency ηnoz regardless of fluctuations in the flow rate of the inflowing fluid.

[0060] In the method for operating the nozzle device, the fluid flowing into the nozzle does not necessarily have to be a liquid-phase fluid, but may be a refrigerant that can become a bubbly flow in the first passage. For example, it may be a fluid in a supercritical state.

[0061] The above and other objects, features, and advantages of the present disclosure will become more apparent from the following detailed description taken in conjunction with the accompanying drawings. 1. A schematic overall configuration diagram of a vehicle refrigeration cycle device of a first embodiment. 2. An axial cross-sectional view of an ejector of the first embodiment. 3. An axial enlarged cross-sectional view of a nozzle portion of the first embodiment when the passage area of ​​a nozzle inlet is enlarged. 4. An axial cross-sectional view of a nozzle portion of the first embodiment when the passage area of ​​a nozzle inlet is reduced. 5. An axial enlarged cross-sectional view of a nozzle portion of the first embodiment when the nozzle passage is blocked. 6. A graph showing momentum-freezed phase change equilibrium sound velocity and momentum-equilibrium phase change equilibrium sound velocity versus dryness fraction. 7. A Mollier diagram showing changes in refrigerant state in an ejector-type refrigeration cycle of the first embodiment. 8. A schematic overall configuration diagram of a vehicle refrigeration cycle device of a second embodiment. 9. An axial cross-sectional view of an ejector of the second embodiment. 10. A Mollier diagram showing changes in refrigerant state during normal operation of the ejector-type refrigeration cycle of the second embodiment. 11. A Mollier diagram showing changes in refrigerant state during high-load operation of the ejector-type refrigeration cycle of the second embodiment. 12. A schematic overall configuration diagram of a vehicle refrigeration cycle device of a third embodiment. Fig. 10 is a schematic overall configuration diagram of a vehicle refrigeration cycle device of a fourth embodiment. Fig. 11 is a schematic overall configuration diagram of a vehicle refrigeration cycle device of a fifth embodiment. Fig. 12 is a Mollier diagram showing changes in the state of refrigerant during normal operation of the ejector refrigeration cycle of the fifth embodiment. Fig. 13 is a Mollier diagram showing changes in the state of refrigerant during high load operation of the ejector refrigeration cycle of the fifth embodiment. Fig. 14 is a schematic overall configuration diagram of a vehicle refrigeration cycle device of a sixth embodiment.

[0062] Hereinafter, multiple embodiments for carrying out the present disclosure will be described with reference to the drawings. In each embodiment, parts corresponding to matters described in the preceding embodiment will be assigned the same reference numerals, and duplicate descriptions may be omitted. In each embodiment, when only a part of the configuration is described, other previously described embodiments may be applied to the other parts of the configuration. In addition to combinations of parts that are specifically specified as being combinable in each embodiment, it is also possible to partially combine embodiments even if not specified, as long as there is no particular problem with the combination.

[0063] First Embodiment A first embodiment of the present disclosure will be described using Figures 1 to 7. In this embodiment, an ejector-type refrigeration cycle including an ejector according to the present disclosure is applied to a vehicle refrigeration cycle device 1 mounted on an electric vehicle. An electric vehicle is a vehicle that obtains driving power for traveling from an electric motor. The vehicle refrigeration cycle device 1 of this embodiment is mounted on the vehicle and cools on-board equipment that generates heat during operation.

[0064] The vehicle refrigeration cycle device 1 cools a battery 80 as an on-board device. The battery 80 is a secondary battery that stores power to be supplied to multiple on-board devices that operate electrically. The battery 80 is an assembled battery formed by electrically connecting multiple stacked battery cells in series or parallel. The battery cells in this embodiment are lithium-ion batteries.

[0065] The battery 80 generates heat during operation (i.e., during charging and discharging). The battery 80 is prone to a decrease in output at low temperatures and prone to deterioration at high temperatures. For this reason, the temperature of the battery 80 needs to be maintained within an appropriate temperature range (in this embodiment, 15°C or higher and 55°C or lower). Therefore, in the electric vehicle of this embodiment, the vehicle refrigeration cycle device 1 is used to cool the battery 80.

[0066] The vehicle refrigeration cycle device 1 includes an ejector-type refrigeration cycle 10, a coolant circuit 60, a control device 70, and the like.

[0067] First, an ejector refrigeration cycle 10 will be described using the overall configuration diagram of Figure 1. The ejector refrigeration cycle 10 is a vapor compression refrigeration cycle that cools the coolant circulating through a coolant circuit 60. The ejector refrigeration cycle 10 uses an HFO refrigerant (specifically, R1234yf) as the refrigerant. The ejector refrigeration cycle 10 constitutes a subcritical refrigeration cycle in which the pressure of the high-pressure side refrigerant does not exceed the critical pressure of the refrigerant.

[0068] The refrigerant is mixed with refrigerating machine oil to lubricate the compressor 11. The refrigerating machine oil is PAG oil (i.e., polyalkylene glycol oil) or POE (i.e., polyol ester) that is compatible with the liquid-phase refrigerant. A portion of the refrigerating machine oil circulates through the ejector-type refrigeration cycle 10 together with the refrigerant.

[0069] The compressor 11 in the ejector-type refrigeration cycle 10 draws in, compresses, and discharges refrigerant. The compressor 11 is an electric compressor that rotates a fixed-displacement compression mechanism with a fixed discharge capacity using an electric motor. The rotation speed (i.e., refrigerant discharge capacity) of the compressor 11 is controlled by a control signal output from a control device 70, which will be described later.

[0070] The discharge port of the compressor 11 is connected to the refrigerant inlet side of an outdoor heat exchanger 12. The outdoor heat exchanger 12 is a heat dissipation unit that exchanges heat between the refrigerant discharged from the compressor 11 and outside air blown by an outdoor air fan (not shown), thereby dissipating heat from the refrigerant and condensing it.

[0071] The inlet side of a branching section 13 is connected to the refrigerant outlet of the outdoor heat exchanger 12. The branching section 13 branches the flow of refrigerant flowing out from the outdoor heat exchanger 12. The branching section 13 is a three-way joint having three inlet and outlet ports that communicate with each other. The branching section 13 can be a joint formed by joining multiple pipes or a joint formed by providing multiple refrigerant passages in a metal block or a resin block.

[0072] One outlet of the branching portion 13 is connected to a refrigerant inlet 33 side of the ejector 14. The other outlet of the branching portion 13 is connected to an inlet side of an electric expansion valve 15.

[0073] The ejector 14 is a fluid transport unit that uses the action of the refrigerant sprayed from the nozzle unit 30 to suck the refrigerant through a suction port 41 formed in the body unit 40 and transport it toward a refrigerant outlet 44. Furthermore, the ejector 14 is a fluid pressurizing unit that pressurizes a mixed refrigerant obtained by mixing the sprayed refrigerant and the suction refrigerant sucked through the suction port 41. The detailed configuration of the ejector 14 will be described later.

[0074] The inlet side of the refrigerant passage of the first chiller 16a is connected to the refrigerant outlet 44 of the ejector 14. The first chiller 16a is a heat exchanger that exchanges heat between the refrigerant and the cooling water circulating through the cooling water circuit 60. The first chiller 16a is an evaporation heat exchanger that evaporates the refrigerant by causing the refrigerant flowing out of the ejector 14 to absorb heat contained in the cooling water pressure-fed from a cooling water pump 61 (described later). In this way, the first chiller 16a cools the cooling water. The suction side of the compressor 11 is connected to the outlet of the refrigerant passage of the first chiller 16a.

[0075] The electric expansion valve 15 is an electric variable throttle mechanism that reduces the pressure of the refrigerant by throttling the refrigerant passage. The electric expansion valve 15 has a valve body and a drive unit. The valve body changes the throttle opening of the electric expansion valve 15. The drive unit displaces the valve body. An electric actuator such as a stepping motor or a brushless DC motor can be used as the drive unit. The operation of the electric expansion valve 15 is controlled by a control signal output from the control device 70.

[0076] The outlet of the electric expansion valve 15 is connected to the inlet side of the refrigerant passage of the second chiller 16b. The second chiller 16b is a heat exchanger that exchanges heat between the refrigerant and the cooling water circulating through the cooling water circuit 60. The second chiller 16b is an evaporation unit that evaporates the refrigerant by having the refrigerant flowing out of the electric expansion valve 15 absorb heat contained in the cooling water flowing out from the cooling water passage of the first chiller 16a. In this way, the cooling water is cooled in the second chiller 16b. The suction port 41 side of the ejector 14 is connected to the outlet of the refrigerant passage of the second chiller 16b.

[0077] Next, a detailed configuration of the ejector 14 will be described with reference to Figures 2 to 6. The ejector 14 of this embodiment transports and pressurizes a refrigerant as a fluid. Therefore, in this embodiment, the term "fluid" will be replaced with the term "refrigerant." For example, a sprayed fluid will be referred to as a sprayed refrigerant. A suctioned fluid will be referred to as a suctioned refrigerant. A mixed fluid will be referred to as a mixed refrigerant. Others will be referred to in the same manner as appropriate.

[0078] The ejector 14 has a nozzle portion 30, a body portion 40, and an initial velocity control valve 50. The nozzle portion 30 reduces the pressure of the refrigerant flowing out of the exterior heat exchanger 12, accelerates it to supersonic speed, and sprays it. The nozzle portion 30 is formed by subjecting a tapered cylindrical metal member (stainless steel in this embodiment) to plastic processing or cutting. The nozzle portion 30, together with the initial velocity control valve 50, constitutes a nozzle device.

[0079] 2, an initial velocity adjustment space 31 and a nozzle passage 32 are formed inside the nozzle portion 30. A refrigerant inlet 33 is formed on the cylindrical outer wall surface of the nozzle portion 30, through which the subcooled liquid-phase refrigerant flowing out from the exterior heat exchanger 12 flows into the initial velocity adjustment space 31. The refrigerant inlet 33 flows in a direction perpendicular to the axial direction of the nozzle passage 32.

[0080] The initial velocity adjustment space 31 is a space for converting velocity energy of the supercooled liquid-phase refrigerant that has flowed into the space from the refrigerant inlet 33 into pressure energy. Specifically, in the initial velocity adjustment space 31, the passage area of ​​the refrigerant passage through which the supercooled liquid-phase refrigerant flows is made larger than the passage cross-sectional area of ​​the refrigerant inlet 33, thereby converting the velocity energy of the supercooled liquid-phase refrigerant into pressure energy.

[0081] The nozzle passage 32 is a refrigerant passage that extends from the nozzle inlet 32a to the injection port 32e. The nozzle inlet 32a is the location where decompression of the refrigerant begins. The injection port 32e is the location where the refrigerant that has passed through the nozzle passage 32 is injected into a mixing section 42 formed in the body section 40. The nozzle passage 32 is formed in the shape of a body of revolution.

[0082] A valve body 51 of an initial velocity control valve 50 is disposed in the initial velocity control space 31. The initial velocity control valve 50 is a velocity control unit that adjusts the initial velocity (more specifically, the axial velocity) of the refrigerant flowing from the nozzle inlet 32a into the nozzle passage 32. The initial velocity control valve 50 is fixed to one axial end of the nozzle portion 30 by means of screw fastening or the like. The initial velocity control valve 50 has a valve body 51 and a drive unit 52.

[0083] The valve body 51 is formed of a cylindrical member made of metal (stainless steel in this embodiment). The valve body 51 has a conical tip that tapers toward the nozzle passage 32. The valve body 51 is disposed coaxially with the nozzle passage 32. Therefore, when the vehicle refrigeration cycle device 1 is in operation, the nozzle inlet 32a of the nozzle passage 32 has an inlet shape of a truncated cone side surface, as shown in Figures 3 and 4, and the inlet area is Ain.

[0084] The drive unit 52 is an electric actuator that axially displaces the valve body 51. Specifically, a stepping motor, a brushless DC motor, or the like can be used as the drive unit 52. The operation of the initial speed adjustment valve 50 is controlled by a control signal output from the control device 70.

[0085] 3, when the drive unit 52 displaces the tip of the valve body 51 away from the nozzle passage 32, the inlet area Ain of the nozzle inlet 32a increases. Therefore, the amount of velocity energy of the refrigerant converted to pressure energy in the initial velocity adjustment space 31 increases, and the initial velocity of the refrigerant flowing into the nozzle inlet 32a can be reduced.

[0086] 4, when the drive unit 52 displaces the tip of the valve body 51 toward the nozzle passage 32, the inlet area Ain of the nozzle inlet 32a decreases. The amount of velocity energy of the refrigerant converted into pressure energy in the initial velocity adjustment space 31 decreases, and the initial velocity of the refrigerant flowing into the nozzle inlet 32a can be increased.

[0087] Therefore, the initial velocity control valve 50 changes the inlet area Ain of the nozzle inlet 32a without changing the passage cross-sectional area of ​​the nozzle passage 32 extending from the nozzle inlet 32a to the injection port 32e during operation of the vehicle refrigeration cycle device 1. This makes it possible to adjust the axial velocity of the refrigerant flowing into the nozzle inlet 32a.

[0088] In other words, when the vehicle refrigeration cycle device 1 is operating, the initial velocity control valve 50 can adjust the axial velocity of the refrigerant flowing into the nozzle inlet 32a by changing only the inlet area Ain of the nozzle inlet 32a.

[0089] In this embodiment, the dimensional specifications of the initial velocity adjustment space 31 and the displacement amount of the valve body portion 51 are set so that the initial velocity of the refrigerant flowing into the nozzle inlet 32a approaches approximately 0 m / s when the inlet area Ain of the nozzle inlet 32a reaches its maximum value. Even if the initial velocity of the refrigerant flowing into the nozzle inlet 32a is 0 m / s, the refrigerant will flow through the nozzle passage 32 as long as there is a pressure difference between the refrigerant pressure in the initial velocity adjustment space 31 and the refrigerant pressure in the mixing portion 42.

[0090] Furthermore, when the vehicle refrigeration cycle device 1 is stopped, the initial speed control valve 50 of this embodiment can close the nozzle passage 32 by having the drive unit 52 abut the tip of the valve body unit 51 against the passage wall surface of the nozzle passage 32, as shown in Figure 5.

[0091] As shown in FIGS. 3 to 5, the nozzle passage 32 of this embodiment has refrigerant passages such as a first passage 321, a second passage 322, a third passage 323, and a fourth passage 324 formed therein.

[0092] The first passage 321 is formed in a shape such that the passage cross-sectional area gradually decreases from the nozzle inlet 32a toward the downstream side of the refrigerant flow. The first passage 321 is a passage that generates fine bubbles in the supercooled liquid-phase refrigerant that has flowed into the nozzle passage 32, changing the state of the refrigerant flow from a bubble flow to a mist flow. Therefore, the first passage 321 is a bubble-generating section.

[0093] The first passage 321 is formed with an inlet-side reduced area portion 321a, a constant area portion 321b, and an outlet-side reduced area portion 321c.

[0094] The reduced inlet area section 321a reduces the cross-sectional area of ​​the passage from the nozzle inlet 32a toward the downstream side in the refrigerant flow direction. In the reduced inlet area section 321a, the cross-sectional area is reduced so that the temperature of the refrigerant flowing into the reduced inlet area section 321a approaches the saturation temperature in order to efficiently foam the supercooled liquid-phase refrigerant that has flowed into the nozzle passage 32 in the constant area section 321b.

[0095] The constant-area section 321b is a section where the cross-sectional area of ​​the passage is constant from the outlet of the reduced-area inlet section 321a toward the downstream side in the refrigerant flow direction. In the constant-area section 321b, bubbles are generated in the liquid refrigerant by nucleate boiling. Furthermore, the refrigerant flow phase is changed from a bubbly flow to a mist flow. In other words, the refrigerant flow phase is changed from a bubbly flow to a mist flow.

[0096] More specifically, in the constant area portion 321b, nucleate boiling occurs at the nano-level minute irregularities on the passage wall surface. The initial bubbles generated by nucleate boiling detach from the passage wall surface and become gas bubbles. The gas bubbles grow and expand in diameter as they move in the flow direction. When the bubble diameter exceeds the critical diameter that can be maintained by surface tension determined by the physical properties of the refrigerant, bubble collapse occurs. Then, the bubble collapse energy generated when the bubbles collapse shears the liquid refrigerant around the bubbles, generating droplets.

[0097] As a result, in the constant area portion 321b, the state of the refrigerant flow changes from a bubbly flow to a mist flow.

[0098] The outlet-side area reduction section 321c reduces the cross-sectional area of ​​the passage from the outlet of the constant-area section 321b toward the downstream side in the refrigerant flow direction. The outlet-side area reduction section 321c reduces the cross-sectional area of ​​the passage so that the temperature of the gas phase refrigerant and the temperature of the liquid droplets are in equilibrium (hereinafter referred to as temperature equilibrium state).

[0099] The second passage 322 is formed in a shape such that the cross-sectional area of ​​the passage increases from the first outlet 32b of the first passage 321 toward the downstream side in the refrigerant flow direction. The second passage 322 is a passage that brings the refrigerant that has become a mist flow in the first passage 321 to a critical state. In other words, the second passage 322 is a passage that brings the refrigerant in the mist flow generated in the first passage 321 to a critical state. The second passage 322 is a critical pressure generating section.

[0100] The refrigerant in the mist flow that flows from the first outlet 32b of the first passage 321 into the second passage 322 is in a temperature equilibrium state, but the axial velocity of the gas phase refrigerant and the axial velocity of the droplets are in a non-equilibrium state (hereinafter referred to as a velocity non-equilibrium state). Furthermore, the refrigerant that flows from the first outlet 32b into the second passage 322 has droplets with inertia that move toward the central axis, and the gas phase refrigerant that moves toward the passage wall surface while expanding its volume.

[0101] Therefore, the degree of velocity imbalance between the gas phase refrigerant and the liquid droplets in the second passage 322 is greater than that in the first outlet 32b. The velocity of the refrigerant in the mist flow in the velocity imbalance state is greater than the momentum frozen phase change equilibrium sound velocity a ｆｅ Therefore, in the second passage 322, the mass average velocity u of the refrigerant that has reached the critical state ｔｈ２ is the momentum frozen phase change equilibrium sound speed a ｆｅ The cross-sectional area of ​​the passage is enlarged so that it approaches

[0102] mass average velocity u ｔｈ２ is expressed by the following formula F1. Momentum frozen phase change equilibrium sound speed a ｆｅ are expressed by the following formulas F2 and F3.

[0103]

[0104] The subscript in in the formula F1 indicates the state of the refrigerant at the inlet of the second passage 322 (i.e., the first outlet 32b), and the subscript th indicates the state of the refrigerant that has reached the critical state. ｉｎ is the mass average velocity of the refrigerant at the first outlet 32b. A is the passage cross-sectional area. x is the axial distance. F is the friction loss of the two-phase flow on the passage wall surface.

[0105]

[0106]

[0107] ν Ｇ is the specific volume of the gas phase refrigerant. Ｌ is the specific volume of the liquid phase refrigerant. ν is the average specific volume of the gas-liquid two-phase refrigerant. S Ｇ is the specific entropy of the gas phase refrigerant.Ｌ is the specific entropy of the liquid phase refrigerant. T of the refrigerant is the temperature of the refrigerant. ｅｅ is the momentum equilibrium phase change equilibrium sound speed.

[0108] Here, in FIG. 6, momentum frozen phase change equilibrium sound speed a ｆｅ and momentum equilibrium phase change equilibrium sound speed a ｅｅ 6 shows the relationship between the quality fraction X and the speed of sound when the refrigerant R1234yf is at 50°C. As shown in FIG. 6, regardless of the quality fraction X, the momentum frozen phase change equilibrium sound speed a ｆｅ is the momentum equilibrium phase change equilibrium sound speed a ｅｅ It will be faster than

[0109] During the decompression and expansion process of the supercooled liquid refrigerant, it takes time for the bubbles generated by nucleate boiling to grow. Therefore, even if the pressure is reduced, bubbles do not immediately appear in the supercooled liquid refrigerant, which is called a pressure undershoot phenomenon. When the pressure undershoot phenomenon converts the pressure energy of the supercooled liquid refrigerant into velocity energy, the mass average velocity u ｔｈ２ is a relatively fast value.

[0110] Therefore, the refrigerant in the mist flow in the velocity non-equilibrium state has a velocity that is equal to the momentum frozen phase change equilibrium sound velocity a ｆｅ The refrigerant approaches the critical state by matching the flow rate of the refrigerant with the flow rate of the refrigerant. Furthermore, downstream of the position where the refrigerant reaches the critical state in the second passage 322, the gas phase refrigerant expands its volume toward the passage wall, so the refrigerant on the passage wall side has a higher void fraction, which is the volume ratio of the gas phase, than the refrigerant on the central axis side. As a result, the refrigerant with a high void fraction on the passage wall side becomes subsonic again.

[0111] Furthermore, the first outlet 32b is the first-stage throat portion where the cross-sectional area of ​​the refrigerant passage formed by the first passage 321 and the second passage 322 is the smallest. Therefore, the refrigerant passage formed by the first passage 321 and the second passage 322 corresponds to the first-stage nozzle in a two-stage expansion nozzle.

[0112] According to the study by the inventors, it has been confirmed that, in the first passage 321 and the second passage 322, by setting the absolute value of the area change rate of the second passage 322 to be larger than the absolute value of the area change rate of the first passage 321, it is easier to make the mist flow of refrigerant reach a critical state in the second passage 322. That is, by setting the absolute value of the area change rate of the second passage 322 to be larger than the absolute value of the area change rate of the first passage 321, the mass average velocity u of the refrigerant that has reached the critical state can be reduced. ｔｈ２ The momentum frozen phase change equilibrium sound speed a ｆｅ It has been confirmed that it is easy to approach

[0113] The absolute value of the area change rate of the first passage 321 is the absolute value of the difference between the passage cross-sectional area at the nozzle inlet 32a and the passage cross-sectional area at the first outlet 32b, divided by the axial length L1 of the first passage 321. The passage cross-sectional area is the area of ​​a cross section perpendicular to the axial direction. The starting point of the axial length L1 may be the center point of the opening circle that forms the nozzle inlet 32a in the nozzle portion 30, and the end point may be the center point of the cross section perpendicular to the axial direction of the refrigerant passage formed by the first outlet 32b.

[0114] The absolute value of the area change rate of the second passage 322 is the absolute value of the difference between the passage cross-sectional area at the inlet of the second passage 322 (i.e., the first outlet 32b) and the passage cross-sectional area at the second outlet 32c, divided by the axial length L2 of the second passage 322. The start point of the axial length L2 may be the center point of the axial cross section of the refrigerant passage formed by the first outlet 32b, and the end point may be the center point of the axial cross section of the refrigerant passage formed by the second outlet 32c.

[0115] Furthermore, it has been confirmed that for the first passage 321, if the passage cross-sectional area at the first outlet 32b is approximately 90% of the maximum passage cross-sectional area perpendicular to the axial direction at the nozzle inlet 32a, a mist flow can be generated effectively.

[0116] The third passage 323 is formed in a shape such that the cross-sectional area of ​​the passage decreases from the second outlet 32c of the second passage 322 toward the downstream side in the refrigerant flow direction. The third passage 323 atomizes droplets contained in the refrigerant in the mist flow flowing out from the second passage 322, and returns the refrigerant to a critical state at the third outlet 32d of the third passage 323. Therefore, the third passage 323 is a droplet atomization section.

[0117] As described above, the refrigerant mist flowing out from the second outlet 32c of the second passage 322 is in a velocity non-equilibrium state, and therefore, due to the influence of inertial force, liquid droplets tend to be unevenly distributed toward the central axis, and gas phase refrigerant tends to be unevenly distributed toward the wall surface. Therefore, when the refrigerant mist flowing out from the second outlet 32c flows into the third passage 323, the highly compressible gas phase refrigerant unevenly distributed toward the wall surface is accelerated due to the reduction in the cross-sectional area of ​​the passage.

[0118] As the difference in velocity between the accelerated gas refrigerant and the droplets increases, velocity relaxation occurs between the gas refrigerant and the droplets. When velocity relaxation occurs, shear force acts on the droplets from the accelerated gas refrigerant, causing the droplets to break down into smaller particles. Furthermore, the droplets, whose total surface area has increased as a result of being broken down into smaller particles, are subjected to drag from the gas refrigerant. As a result, the gas refrigerant and the droplets move at the same speed.

[0119] Therefore, in the third passage 323, a velocity relaxation phenomenon is caused to occur, and the axial velocity of the gas phase refrigerant and the axial velocity of the droplets are brought into equilibrium (hereinafter referred to as velocity equilibrium state). More specifically, in a state where the velocity relaxation phenomenon occurs, the local dynamic pressure fluctuation τ of the refrigerant in the mist flow can be expressed by the following formula F4. Furthermore, the Weber number We, which is the ratio of the dynamic pressure change τ to the surface tension of the droplets, can be expressed by the following formula F5.

[0120]

[0121]

[0122] ρ ｇ is the density of the gas phase refrigerant. ｇ is the local gas phase refrigerant velocity. ｌ is the droplet velocity, and u is the average velocity of the gas-liquid two-phase refrigerant.

[0123] When the Weber number We exceeds the critical Weber number Wec, the droplets in the mist flow break up and become finer. This allows the droplet diameter to approach the theoretical minimum diameter (specifically, about several microns). Therefore, in the third passage 323, the passage cross-sectional area is reduced so that the Weber number We exceeds the critical Weber number Wec (Wec = 12 in this embodiment).

[0124] Furthermore, in the third passage 323, a velocity equilibrium state is reached, and the mass average velocity u of the refrigerant that has reached the critical state ｔｈ３ is the momentum equilibrium phase change equilibrium sound velocity a ｅｅ The passage cross-sectional area of ​​the third outlet 32d is set so as to approach

[0125] The fourth passage 324 is formed in a shape such that the cross-sectional area of ​​the passage increases from the third outlet 32d of the third passage 323 toward the downstream side in the refrigerant flow direction. The fourth passage 324 is a passage that accelerates the mist flow of refrigerant that has reached sonic speed at the third outlet 32d of the third passage 323 to supersonic speed. Therefore, the fourth passage 324 is a supersonic acceleration section.

[0126] In the fourth passage 324, the cross-sectional area is enlarged so that the axial velocity of the gas phase refrigerant and the axial velocity of the droplets can be balanced and the refrigerant can be accelerated effectively. An injection port 32e is formed at the most downstream portion of the fourth passage 324 in the refrigerant flow direction.

[0127] Here, the third outlet 32d is the second-stage throat where the cross-sectional area of ​​the refrigerant passage formed by the third passage 323 and the fourth passage 324 is the smallest. Therefore, the refrigerant passage formed by the third passage 323 and the fourth passage 324 corresponds to the second-stage nozzle in a two-stage expansion nozzle.

[0128] According to the study by the inventors, it has been confirmed that in the second-stage nozzle, by setting the absolute value of the area change rate of the fourth passage 324 to be larger than the absolute value of the area change rate of the third passage 323, the velocity relaxation phenomenon is easily caused in the third passage 323. Furthermore, by setting it in the same way, the mass average velocity u of the refrigerant that has reached the critical state at the third outlet 32d can be reduced. ｔｈ３The momentum equilibrium phase change equilibrium sound speed a ｅｅ It has been confirmed that it is easy to approach

[0129] The absolute value of the area change rate of the third passage 323 is the absolute value of the difference between the passage cross-sectional area at the inlet of the third passage 323 (i.e., the second outlet 32c) and the passage cross-sectional area at the third outlet 32d, divided by the axial length L3 of the third passage 323. The start point of the axial length L3 may be the center point of the axial cross section of the refrigerant passage formed by the second outlet 32c, and the end point may be the center point of the axial cross section of the refrigerant passage formed by the third outlet 32d.

[0130] The absolute value of the area change rate of the fourth passage 324 is the absolute value of the difference between the passage cross-sectional area at the inlet of the fourth passage 324 (i.e., the third outlet 32d) and the passage cross-sectional area at the jet port 32e, divided by the axial length L4 of the fourth passage 324. The start point of the axial length L4 may be the center point of the axial cross section of the refrigerant passage formed by the third outlet 32d, and the end point may be the center point of the axial cross section of the refrigerant passage formed by the jet port 32e.

[0131] Furthermore, when comparing the first-stage nozzle and the second-stage nozzle, it has been confirmed that the velocity relaxation phenomenon is more likely to occur by setting the absolute value of the area change rate of the third passage 323 larger than the absolute value of the area change rate of the second passage 322. It has also been confirmed that the refrigerant mist flow can be accelerated effectively by setting the absolute value of the area change rate of the fourth passage 324 smaller than the absolute value of the area change rate of the second passage 322.

[0132] The body 40 is a cylindrical member that forms the outer shell of the ejector 14 and defines a refrigerant passage therein. The nozzle 30 is fixed to one longitudinal end of the body 40 by press-fitting, screw fastening, or other means. The central axis of the nozzle passage 32 of the nozzle 30 and the central axis of the refrigerant passage of the body 40 are coaxially arranged.

[0133] The body 40 is made of metal (aluminum alloy in this embodiment), or may be made of resin. The body 40 is provided with a suction port 41, a mixing section 42, an area expansion section 43, a refrigerant outlet 44, and the like.

[0134] The suction port 41 is formed on the cylindrical side surface of the body portion 40 at a position on the outer periphery of the nozzle portion 30. The suction port 41 is a through-hole that draws the refrigerant flowing out from the refrigerant passage of the second chiller 16b into the inside of the ejector 14. The ejector 14 draws the refrigerant through the suction port 41 by utilizing a pressure drop caused by an expansion wave generated by the injected refrigerant.

[0135] The mixing section 42 is a refrigerant passage that mixes the sprayed refrigerant sprayed from the nozzle section 30 with the suction refrigerant sucked through the suction port 41 to increase the pressure of the mixed refrigerant. The mixing section 42 is a refrigerant passage in the shape of a body of revolution that is arranged coaxially with the nozzle passage 32. Therefore, at the inlet of the mixing section 42, the sprayed refrigerant flows toward the central axis of the mixing section 42, and the suction refrigerant flows toward the inner wall surface of the mixing section 42. Therefore, the mixed refrigerant at the inlet of the mixing section 42 is in a non-equilibrium state.

[0136] The mixing section 42 is formed with a convergent mixing section 42a and a divergent mixing section 42b. The convergent mixing section 42a is located upstream of the divergent mixing section 42b in the refrigerant flow direction. The convergent mixing section 42a is formed in a truncated cone shape whose cross-sectional area decreases toward the downstream side in the refrigerant flow direction. The convergent mixing section 42a is a passage that accelerates the mixed refrigerant of the injected refrigerant and the suctioned refrigerant to two-phase sonic speed or faster.

[0137] The divergent mixing section 42b is connected downstream of the convergent mixing section 42a. The divergent mixing section 42b is formed in a truncated cone shape whose cross-sectional area increases downstream in the refrigerant flow direction. This section generates shock waves in the mixed refrigerant flowing out of the convergent mixing section 42a and then eliminates the generated shock waves.

[0138] For this reason, a neck portion 42c that minimizes the cross-sectional area of ​​the refrigerant passage formed in the mixing section 42 is formed at the connection between the convergent mixing section 42a and the divergent mixing section 42b of the mixing section 42. The neck portion 42c serves as the outlet of the convergent mixing section 42a and the inlet of the divergent mixing section 42b.

[0139] In order to effectively increase the pressure of the mixed refrigerant in the divergent mixing section 42b, the mixed refrigerant needs to be accelerated to a speed equal to or faster than the two-phase sonic speed in the convergent mixing section 42a. Therefore, in the convergent mixing section 42a of this embodiment, the cross-sectional area of ​​the passage is reduced so that the pressure of the mixed refrigerant in the neck section 42c in the convergent mixing section 42a is lower than the pressure Pnout of the injected refrigerant immediately after it is injected from the injection port 32e of the nozzle section 30.

[0140] In the tapered mixing section 42a, it is desirable to sufficiently mix the mixed refrigerant and the suction refrigerant to achieve an equilibrium state. Here, the equilibrium state in the tapered mixing section 42a means a state in which the mixed refrigerant has almost no temperature, pressure, or velocity distribution. Therefore, in this embodiment, the tapered mixing length LMIX1, which is the axial length of the tapered mixing section 42a, is set to satisfy the following formulas F6 and F7.

[0141]

[0142]

[0143] u ｓｎｉｎ is the mass average velocity of the refrigerant at the suction port 41. In other words, is the average velocity of the refrigerant in the axial direction at the inlet of the tapered mixing section 42a. ｌ１ is the density of droplets in the mixed refrigerant at the suction port 41. ｌ１ is the average diameter of the droplets at the suction port 41. ｇ１ is the viscosity of the gas phase refrigerant in the mixed refrigerant at the suction port 41.

[0144] Lv1 is the mass average velocity u of the mixed refrigerant at the inlet of the tapered mixing section 42a. ｓｎｉｎ The first relaxation distance Lv1 is calculated by multiplying the mass average velocity u ｓｎｉｎ This is a value obtained by multiplying the first velocity relaxation time τv1 required for the velocity of the gas phase refrigerant in the mixed refrigerant to become equal to the velocity of the droplets.

[0145] According to the study by the inventors, if the tapered mixing distance LMIX1 is set so as to satisfy the formulas F6 and F7, the dryness fraction X at the inlet of the tapered mixing section 42a can be ｓｎｉｎ It has been confirmed that even if the temperature changes over a wide range, the mixed refrigerant in the neck portion 42c reaches an equilibrium state and is accelerated to two-phase sonic speed or faster.

[0146] In the divergent mixing section 42b, the cross-sectional area of ​​the refrigerant mixture, which is in a sonic equilibrium state and flows in two phases, is expanded, generating an expansion wave in the refrigerant mixture immediately after the neck section 42c, i.e., in the vicinity of the inlet of the divergent mixing section 42b. The expansion wave generated in the divergent mixing section 42b generates a shock wave with its leading edge on the inner wall surface of the body 40. The energy of the shock wave then increases the refrigerant pressure.

[0147] Furthermore, the mixed refrigerant flowing through the divergent mixing section 42b generates pseudo-shock waves, in which multiple shock waves and expansion waves are repeatedly generated, depending on the Mach number of the mixed refrigerant. Specifically, when the Mach number of the mixed refrigerant at the center exceeds approximately 1.3, a pseudo-shock wave is generated in the center of the shock wave. The mixed refrigerant increases in pressure and enthalpy as it passes through the repeatedly generated shock waves.

[0148] When a pseudo-shock wave occurs, the central axis side static pressure, which is the refrigerant pressure on the central axis side of the diverging mixing section 42b, and the wall side static pressure, which is the refrigerant pressure on the wall side, change to different values, which causes energy loss due to friction of the mixed refrigerant.

[0149] Therefore, it is desirable to eliminate the pseudo-shock waves in the divergent mixing section 42b in order to improve the pressure-boosting capacity of the ejector 14. Furthermore, it has been found that the pseudo-shock waves can be eliminated by reducing the velocity of the mixed refrigerant to subsonic velocity.

[0150] Therefore, in the divergent mixing section 42b of this embodiment, the cross-sectional area of ​​the passage is enlarged so that the velocity of the mixed refrigerant at the outlet of the divergent mixing section 42b can rapidly become subsonic. Furthermore, in the divergent mixing section 42b, it is desirable to shorten the distance from when the pseudo-shock waves are generated until they disappear. Therefore, in this embodiment, the divergent mixing distance LMIX2, which is the axial length of the divergent mixing section 42b, is set to satisfy the following formulas F8 and F9.

[0151]

[0152]

[0153] u ｎｅｃｋ is the mass average velocity of the refrigerant in the neck 42c, or in other words, the average axial velocity of the refrigerant in the neck 42c. ｌ２ is the density of the droplets in the mixed refrigerant at the neck portion 42c. ｌ２ is the average diameter of the droplets at the neck 42c. ｇ２ is the viscosity of the gas phase refrigerant in the mixed refrigerant at the neck portion 42c.

[0154] Lv2 is the mass average velocity u of the mixed refrigerant in the neck portion 42c. ｎｅｃｋ The second relaxation distance Lv2 is calculated by multiplying the mass average velocity u ｎｅｃｋ represents the distance obtained by integrating the second velocity relaxation time τv2 required until the flow velocity of the gas phase refrigerant in the mixed refrigerant becomes equal to the flow velocity of the droplets.

[0155] According to the study by the inventors, if the divergent mixing distance LMIX2 is set so as to satisfy the formulas F8 and F9, the dryness fraction X ｎｅｃｋ It has been confirmed that the pseudo shock wave disappears in the divergent mixing section 42b even if the flow rate varies over a wide range.

[0156] Furthermore, when the Mach number of the refrigerant injected from the injection port 32e of the nozzle portion 30 exceeds approximately 1.3 to 1.5, the expansion wave generated at the injection port 32e is reflected by the velocity boundary layer with the suction flow, thereby generating a pseudo-shock wave inside the tapered mixing portion 42a.

[0157] In contrast, in the ejector 14 of this embodiment, the divergent mixing length LMIX2 is set to satisfy formulas F8 and F9. Therefore, in the divergent mixing section 42b, the pseudo-shock waves generated with the inner wall of the divergent mixing section 42b as the leading edge can be eliminated, and at the same time, the pseudo-shock waves generated inside the convergent mixing section 42a can also be eliminated. As a result, the ejector 14 can achieve a stable pressure increase effect.

[0158] The enlarged-area section 43 is connected to the downstream side of the mixing section 42. The enlarged-area section 43 is formed in a truncated cone shape that increases the cross-sectional area of ​​the passage toward the downstream side in the refrigerant flow direction. The enlarged-area section 43 forms a refrigerant passage that smoothly connects the outlet of the diverging mixing section 42b and the refrigerant outlet 44. The enlarged-area section 43 may be formed in a cylindrical shape with a constant cross-sectional area, as long as the cross-sectional area of ​​the passage does not decrease toward the flow direction of the mixed refrigerant.

[0159] Next, the coolant circuit 60 will be described. The coolant circuit 60 is a circuit that circulates coolant. In this embodiment, an ethylene glycol aqueous solution is used as the coolant. The coolant circuit 60 includes a coolant pump 61, a coolant passage for the first chiller 16a, a coolant passage for the second chiller 16b, a coolant passage for the battery 80, and the like.

[0160] The cooling water pump 61 is a cooling water pumping unit that sucks in and pumps out cooling water that has flowed out from the cooling water passage of the battery 80. The cooling water pump 61 is an electric water pump whose rotation speed (i.e., pumping capacity) is controlled by a control voltage output from the control device 70.

[0161] The inlet side of the cooling water passage of the first chiller 16a is connected to the discharge port of the cooling water pump 61. The inlet side of the cooling water passage of the second chiller 16b is connected to the outlet side of the cooling water passage of the first chiller 16a. The inlet side of the cooling water passage of the battery 80 is connected to the outlet side of the cooling water passage of the second chiller 16b.

[0162] The cooling water passage of the battery 80 is a passage through which the cooling water flowing out from the second chiller 16b flows to cool the battery 80. In other words, the cooling water passage of the battery 80 cools the battery 80 by exchanging heat between the cooling water and the battery cells.

[0163] The cooling water passages for the battery 80 are formed inside a dedicated battery case that houses multiple stacked battery cells. The cooling water passages for the battery 80 are configured so that multiple passages are connected in parallel inside the dedicated battery case so that all battery cells can be cooled evenly. The outlet of the cooling water passages for the battery 80 is connected to the intake side of the cooling water pump 61.

[0164] Therefore, when the control device 70 operates the cooling water pump 61, the cooling water pumped from the cooling water pump 61 circulates in the cooling water circuit 60 in the following order: the cooling water passage of the first chiller 16a, the cooling water passage of the second chiller 16b, the cooling water passage of the battery 80, and the intake port of the cooling water pump 61.

[0165] Next, the electrical control unit of the vehicle refrigeration cycle system 1 will be described. The control device 70 has a well-known microcomputer including a CPU, ROM, RAM, etc., and its peripheral circuits. The control device 70 performs various calculations and processes based on control programs stored in the ROM. The control device 70 then controls the operation of various control target devices connected to the output side based on the results of the calculations and processes.

[0166] A group of various control sensors is connected to the input side of the control device 70. The group of control sensors includes a discharge pressure sensor, a high-pressure temperature sensor, an intake pressure sensor, an intake temperature sensor, an evaporator temperature sensor, a coolant temperature sensor, and the like (not shown).

[0167] The discharge pressure sensor is a discharge pressure detection unit that detects a discharge pressure Pd, which is the pressure of the refrigerant discharged from the compressor 11. The high-pressure temperature sensor is a high-pressure temperature detection unit that detects a high-pressure temperature Td, which is the temperature of the refrigerant flowing out from the outdoor heat exchanger 12.

[0168] The suction pressure sensor is a suction pressure detection unit that detects the suction pressure Ps, which is the pressure of the suction refrigerant. The suction temperature sensor is a suction temperature detection unit that detects the suction temperature Ts, which is the temperature of the suction refrigerant suctioned into the compressor 11.

[0169] The evaporator temperature sensor is an evaporator temperature detector that detects the evaporator temperature Te, which is the temperature of the refrigerant flowing out from the second chiller 16 b. The coolant temperature sensor is a coolant temperature detector that detects the coolant temperature TwB, which is the temperature of the coolant flowing out from the coolant passage of the battery 80.

[0170] The control device 70 is an integrated unit that controls various control target devices connected to the output side. Therefore, the components (hardware and software) that control the operation of each control target device constitute the control unit that controls the operation of each control target device. For example, the component of the control device 70 that controls the refrigerant discharge capacity of the compressor 11 constitutes the discharge capacity control unit.

[0171] Next, the operation of the vehicle refrigeration cycle apparatus 1 of this embodiment configured as described above will be described. The control device 70 executes a control program stored in the ROM when a predetermined execution condition is met. In this embodiment, the execution condition is met when the start switch (also known as the ignition switch) of the vehicle system is turned on to start the vehicle system, or when the charging connector of the battery 80 is connected to the vehicle.

[0172] When the control program is executed, the control device 70 operates the cooling water pump 61 so as to achieve a predetermined reference pumping capacity. Furthermore, the control program reads detection signals from a group of control sensors at predetermined control cycles. Then, based on the read control signals, it determines whether cooling of the battery 80 is necessary.

[0173] The control program of this embodiment determines that cooling of the battery 80 is necessary when the coolant temperature TwB is equal to or higher than a predetermined reference cooling start temperature KTwBH while the battery 80 is not being cooled by the ejector refrigeration cycle 10. Also, when the coolant temperature TwB is equal to or lower than a predetermined reference cooling stop temperature KTwBL while the battery 80 is being cooled by the ejector refrigeration cycle 10, the control program determines that cooling of the battery 80 is not necessary.

[0174] When it is determined that the battery 80 needs to be cooled, the control device 70 cools the battery by controlling the operation of various components of the ejector-type refrigeration cycle 10. Specifically, the control device 70 controls the rotation speed of the compressor 11 so that the evaporator temperature Te approaches a predetermined reference evaporator temperature KTe.

[0175] The control device 70 also controls the throttle opening of the electric expansion valve 15 so that the degree of subcooling SC of the refrigerant flowing into the electric expansion valve 15 approaches the target degree of subcooling SCO. The control device 70 detects the degree of subcooling SC based on the discharge pressure Pd and the high-pressure temperature Td. The control device 70 determines the target degree of subcooling SCO based on the discharge pressure Pd by referring to a pre-stored control map. The control map determines the target degree of subcooling SCO so that the COP of the ejector refrigeration cycle 10 reaches a local maximum value.

[0176] The control device 70 also controls the operation of the initial speed control valve 50 of the ejector 14 so that the degree of superheat SH of the suction refrigerant approaches a predetermined reference degree of superheat KSH. The control device 70 detects the degree of superheat SH based on the suction pressure Ts and the suction temperature Ts.

[0177] Therefore, in the ejector refrigeration cycle 10, the state of the refrigerant changes as shown in the Mollier diagram of Fig. 7. That is, the discharge refrigerant discharged from the compressor 11 (point a7 in Fig. 7) flows into the outdoor heat exchanger 12. The discharge refrigerant that flows into the outdoor heat exchanger 12 releases heat to the outside air blown by the outside air fan, condenses, and becomes a subcooled liquid-phase refrigerant (from point a7 to point b7 in Fig. 7).

[0178] The flow of the subcooled liquid-phase refrigerant flowing out from the outdoor heat exchanger 12 is branched at the branching section 13. One of the refrigerant flows out from the branching section 13 and flows into the refrigerant inlet 33 of the ejector 14. Therefore, in the ejector 14, the subcooled liquid-phase refrigerant flowing out from the outdoor heat exchanger 12 serves as a driving fluid.

[0179] The refrigerant that flows into the initial velocity adjustment space 31 through the refrigerant inlet 33 of the ejector 14 has its initial velocity adjusted by the initial velocity adjustment valve 50 and flows into the nozzle passage 32. The refrigerant that flows into the nozzle passage 32 is decompressed in the first passage 321 and reaches a critical state in the second passage 322 (from point b7 to point c7 in FIG. 7). The velocity of the refrigerant that reaches a critical state in the second passage 322 is equal to the momentum frozen phase change equilibrium sound velocity a ｆｅ This becomes:

[0180] As described above, in the second passage 322, the velocity of the gas refrigerant and the velocity of the liquid droplets are not in equilibrium, so as the cross-sectional area of ​​the passage increases, the slip ratio increases and the volume fraction of liquid droplets in the refrigerant mist decreases. The slip ratio is the ratio of the velocity of the liquid droplets to the velocity of the gas refrigerant in the mist. Therefore, in the second passage 322, as the cross-sectional area of ​​the passage increases at the portion where the refrigerant reaches the critical state, the critical flow rate decreases.

[0181] Therefore, by reducing the initial velocity of the refrigerant flowing into the nozzle passage 32 with the initial velocity control valve 50, the position where the refrigerant reaches the critical state in the second passage 322 can be moved upstream in the refrigerant flow, thereby increasing the critical flow rate. Also, by increasing the initial velocity of the refrigerant flowing into the nozzle passage 32 with the initial velocity control valve 50, the position where the refrigerant reaches the critical state in the second passage 322 can be moved downstream in the refrigerant flow, thereby decreasing the critical flow rate.

[0182] The refrigerant that has reached the critical state in the second passage 322 flows into the third passage 323 while converting velocity energy into pressure energy. The mist flow refrigerant, in which the axial velocity of the gas phase refrigerant and the axial velocity of the droplets are in equilibrium in the third passage 323, reaches the critical state again at the third outlet 32d (from point c7 to point d7 in FIG. 7). The velocity of the refrigerant that has reached the critical state at the third outlet 32d is equal to the momentum equilibrium phase change equilibrium sound velocity a ｅｅ This becomes:

[0183] As mentioned above, momentum equilibrium phase change equilibrium sound speed a ｅｅ is the momentum frozen phase change equilibrium sound speed a ｆｅ That is, the axial velocity of the refrigerant that has reached the critical state at the third outlet 32d is lower than the axial velocity of the refrigerant that has reached the critical state in the second passage 322. Therefore, the pressure of the refrigerant that has reached the critical state at the third outlet 32d is higher than the pressure of the refrigerant that has reached the critical state in the second passage 322.

[0184] The refrigerant that has reached a critical state at the third outlet 32d is accelerated to a supersonic speed in the fourth passage 324 (from point d7 to point e7 in FIG. 7) and is injected into the mixer 42 from the injection port 32e.

[0185] In the nozzle portion 30 of this embodiment, the mist flow is accelerated by bringing the refrigerant to a critical state in the second passage 322 and then to a critical state at the third outlet 32d. This allows the decompression process of the refrigerant in the nozzle passage 32 (the process from point b7 to point e7 in Figure 7) to change along the isentropic curve shown by the thin dashed line in Figure 7, which is determined by the physical properties of the refrigerant.

[0186] As a result, the pressure energy of the subcooled liquid-phase refrigerant flowing out of the outdoor heat exchanger 12 can be effectively recovered and converted into velocity energy. In the Mollier diagram of Figure 7, the amount of energy recovered by the nozzle portion 30 is represented by Δh1.

[0187] The other refrigerant flowing out from the branching portion 13 flows into the electric expansion valve 15 and is isenthalpic-decompressed (from point b7 to point f7 in FIG. 7 ). The gas-liquid two-phase refrigerant decompressed by the electric expansion valve 15 flows into the second chiller 16b.

[0188] In the second chiller 16b, the refrigerant decompressed by the electric expansion valve 15 absorbs heat from the cooling water and evaporates (from point f7 to point g7 in FIG. 7). This cools the cooling water. The refrigerant flowing out of the second chiller 16b (point g7 in FIG. 7) is drawn through the suction port 41 of the ejector 14. Therefore, in the ejector 14 of this embodiment, the refrigerant flowing out of the second chiller 16b serves as a suction-side fluid.

[0189] Inside the ejector 14, the injected refrigerant with a relatively low enthalpy (point e7 in Figure 7) injected from the nozzle portion 30 and the suction refrigerant with a relatively high enthalpy (point g in Figure 7) sucked from the suction port 41 join together in the tapered mixing section 42a to become a mixed refrigerant (from point e7 to point h7, and from point g7 to point h7 in Figure 7).

[0190] The refrigerant mixed in the convergent mixing section 42a is in a non-equilibrium state, making it difficult to represent it as a single point on a Mollier diagram. Therefore, point h7 in Figure 7 shows a typical refrigerant state near the neck section 42c. This also applies to the following embodiments.

[0191] The mixed refrigerant that flows from the convergent mixing section 42a to the divergent mixing section 42b reaches an equilibrium state and is effectively pressurized by the action of the shock waves (from point h7 to point i7 in FIG. 7). The refrigerant pressurized in the divergent mixing section 42b flows through the expanded area section 43, flows out of the refrigerant outlet 44, and into the first chiller 16a.

[0192] In the first chiller 16a, the refrigerant flowing out from the ejector 14 absorbs heat from the cooling water and evaporates (from point i7 to point j7 in FIG. 7). This cools the cooling water. The refrigerant flowing out from the first chiller 16a (point j7 in FIG. 7) is drawn into the compressor 11 and compressed again (from point j7 to point a7 in FIG. 7).

[0193] In the cooling water circuit 60, cooling water pumped by a cooling water pump 61 flows into the cooling water passage of the first chiller 16a and is cooled. The cooling water cooled in the first chiller 16a flows into the cooling water passage of the second chiller 16b and is further cooled.

[0194] The cooling water cooled by the second chiller 16b flows into the cooling water passage of the battery 80. The cooling water that flows into the cooling water passage of the battery 80 exchanges heat with the battery 80, thereby cooling the battery 80. The cooling water that flows out from the cooling water passage of the battery 80 is sucked into the cooling water pump 61.

[0195] In the ejector refrigeration cycle 10, the refrigerant evaporation temperature of the second chiller 16b is lower than the refrigerant evaporation temperature of the first chiller 16a. Therefore, in the coolant circuit 60, the coolant pumped from the coolant pump 61 can be efficiently cooled in the order of the first chiller 16a to the second chiller 16b.

[0196] As described above, the vehicle refrigeration cycle device 1 can cool the battery 80. Furthermore, since the ejector-type refrigeration cycle device 10 includes the ejector 14, even if a load fluctuation occurs when the heat generation amount of the battery 80 changes, the ejector-type refrigeration cycle device 10 exhibits a sufficiently high COP regardless of the load fluctuation.

[0197] More specifically, the ejector 14 of this embodiment generates bubbles in the liquid refrigerant in the first passage 321, changing the refrigerant flow from a bubbly flow to a mist flow. Furthermore, in the second passage 322, the mist flow of refrigerant is brought to a critical state. Therefore, in the second passage 322, the velocity of the mist flow of refrigerant approaches the sonic velocity, thereby increasing the velocity of the refrigerant injected from the injection port 32e. As a result, the nozzle efficiency ηnoz is improved, and the pressure-boosting capability of the ejector 14 is improved.

[0198] Furthermore, the critical flow rate of the refrigerant that has reached the critical state varies depending on the cross-sectional area of ​​the portion of the second passage 322 where the refrigerant reaches the critical state. Therefore, when a flow rate fluctuation occurs in the driving side refrigerant, the flow rate of the refrigerant flowing through the nozzle passage 32 can be appropriately adjusted by changing the portion of the second passage 322 where the refrigerant reaches the critical state.

[0199] That is, it is not necessary to change the passage cross-sectional area of ​​the first outlet 32b corresponding to the first-stage throat of the first-stage nozzle in order to change the flow rate of the refrigerant flowing through the nozzle passage 32. Therefore, even if a flow rate fluctuation occurs in the driving side refrigerant, it is unlikely to affect the generation pattern and amount of bubbles generated in the liquid phase refrigerant near the wall surface of the first outlet 32b.

[0200] As a result, the ejector 14 of this embodiment can generate sufficient bubbles in the liquid refrigerant and increase the velocity of the sprayed refrigerant regardless of fluctuations in the flow rate of the refrigerant flowing into the nozzle portion 30. That is, a high pressure-boosting capability can be achieved regardless of fluctuations in the flow rate of the refrigerant flowing into the nozzle portion 30. Furthermore, since the ejector 14 is provided in the ejector-type refrigeration cycle 10, a sufficiently high COP can be achieved regardless of load fluctuations.

[0201] In the ejector 14 of this embodiment, the axial velocity of the refrigerant in the second passage 322 of the nozzle passage 32 is set to the momentum frozen phase change equilibrium sound velocity a ｆｅ This allows the refrigerant to reach a critical state in the second passage 322 even if a pressure undershoot phenomenon occurs when the supercooled liquid phase refrigerant is decompressed and boiled, causing the velocity of the refrigerant in a velocity non-equilibrium state flowing through the second passage 322 to increase.

[0202] Furthermore, in the ejector 14 of this embodiment, a constant area portion 321b is formed in the first passage 321 of the nozzle passage 32. The constant area portion 321b does not unnecessarily increase the refrigerant velocity. Therefore, bubbles that have detached from the passage wall surface can be quickly moved in the direction of the central axis and collapsed. In other words, the formation of the constant area portion 321b in the first passage 321 makes it easier to change the refrigerant flow from a bubbly flow to a mist flow.

[0203] Furthermore, in the ejector 14 of this embodiment, the nozzle passage 32 is formed with a third passage 323 and a fourth passage 324. Accordingly, in the third passage 323, the velocity energy of the refrigerant that has reached a critical state in the second passage 322 can be effectively utilized to atomize the droplets. Furthermore, because the refrigerant is brought to a critical state in the third outlet 32d, the atomized droplets can be accelerated to supersonic speed together with the gas-phase refrigerant in the fourth passage 324. As a result, the velocity of the injected refrigerant can be further increased, thereby improving the nozzle efficiency ηnoz.

[0204] In the ejector 14 of this embodiment, the axial velocity of the refrigerant at the third outlet 32d of the nozzle passage 32 is set to the momentum equilibrium phase change equilibrium sound velocity a ｅｅThis makes it possible to increase the pressure of the refrigerant at the third outlet 32d above the pressure of the refrigerant that has reached the critical state in the second passage 322. Therefore, in the fourth passage 324, the pressure energy of the refrigerant can be used to effectively accelerate the atomized droplets together with the gas-phase refrigerant.

[0205] In addition, in the tapered mixing section 42a of the ejector 14 of this embodiment, the tapered mixing distance LMIX1 is set to satisfy the above-mentioned formulas F6 and F7.

[0206] This allows the mixed refrigerant near the neck portion 42c downstream of the tapered mixing section 42a to approach an equilibrium state. That is, the axial velocity of the droplets contained in the mixed refrigerant near the neck portion 42c can be made equal to the axial velocity of the gas-phase refrigerant. Therefore, the sonic velocity of the mixed refrigerant near the neck portion 42c can be reduced from the gas-phase sonic velocity to the two-phase sonic velocity.

[0207] As a result, the Mach number M of the mixed refrigerant at the neck portion 42c ｎｅｃｈ , thereby improving the pressure-boosting capability of the ejector 14. Furthermore, in the ejector 14 of this embodiment, the difference between the velocity of the injected refrigerant and the velocity of the suctioned refrigerant is likely to increase due to the improvement in the nozzle efficiency ηnoz of the nozzle portion 30. Therefore, bringing the mixed refrigerant closer to an equilibrium state in the neck portion 42c is extremely effective in improving the pressure-boosting capability.

[0208] In addition, the divergent mixing section 42b of the ejector 14 of this embodiment is set so that the divergent mixing distance LMIX2 satisfies the above-described formulas F8 and F9. This allows the mixed refrigerant at the outlet of the divergent mixing section 42b to approach an equilibrium state. Therefore, the pseudo shock waves can be reliably eliminated in the divergent mixing section 42b.

[0209] Furthermore, in the ejector 14 of this embodiment, the refrigerant that has approached the equilibrium state can be caused to flow out from the refrigerant outlet 44 of the ejector 14 .

[0210] This allows the suction pressure sensor and the suction temperature sensor to detect the suction pressure Ps and suction temperature Ts of the suction refrigerant approaching equilibrium. Therefore, the control device 70 can accurately detect the superheat degree SH of the suction refrigerant and accurately bring the superheat degree SH of the suction refrigerant closer to the reference superheat degree KSH. In other words, liquid compression by the compressor 11 can be accurately suppressed.

[0211] The ejector refrigeration cycle device 10 of the present embodiment also includes an initial velocity control valve 50. This makes it possible to easily adjust the velocity of the refrigerant flowing into the nozzle inlet 32a of the ejector 14 regardless of load fluctuations, and to easily adjust the critical flow rate of the refrigerant in the second passage 322.

[0212] Furthermore, this embodiment discloses a method of operating a nozzle device that can improve the nozzle efficiency ηnoz regardless of fluctuations in the flow rate of the inflowing fluid.

[0213] That is, as a method of operating a nozzle device constituted by the nozzle portion 30 and the initial velocity adjustment valve 50, an operating method is disclosed in which bubbles are generated in the liquid phase fluid in the first passage 321, the fluid is brought to a critical state in the second passage 322, and further, the position at which the fluid reaches a critical state in the second passage is changed by adjusting the velocity of the fluid flowing into the nozzle inlet 32a.

[0214] This allows bubbles to be generated in the liquid fluid in the first passage 321, changing the flow state of the fluid from a bubbly flow to a mist flow. Furthermore, in the second passage 322, the mist flow fluid is brought to a critical state. Therefore, in the second passage 322, the velocity of the mist flow fluid can be brought closer to the sonic velocity, thereby increasing the velocity of the jet fluid. As a result, the nozzle efficiency ηnoz can be improved.

[0215] Furthermore, by adjusting the velocity of the driving-side fluid flowing into the nozzle inlet 32a, the position where the fluid reaches a critical state in the second passage 322 is changed, so that the critical flow rate can be changed according to the passage cross-sectional area of ​​the position where the fluid reaches a critical state. In other words, there is no need to change the passage cross-sectional area of ​​the first outlet 32b in order to change the flow rate of the fluid flowing through the nozzle passage 32.

[0216] Therefore, even if the flow rate of the driving side fluid fluctuates, the generation pattern and amount of bubbles generated in the liquid-phase fluid near the wall surface of the first outlet 32b are unlikely to be affected. As a result, the operating method of the nozzle device of this embodiment can improve the nozzle efficiency ηnoz regardless of the flow rate fluctuation of the driving side fluid. In addition, the nozzle device of this embodiment is equipped with an initial velocity adjustment valve 50. This makes it possible to easily adjust the critical flow rate of the refrigerant in the second passage 322.

[0217] Second Embodiment In this embodiment, as shown in the overall configuration diagram of FIG. 8, an example will be described in which an ejector-type refrigeration cycle 10 including an ejector 141 is applied to a vehicle refrigeration cycle device 1.

[0218] 9, the ejector 141 employs a nozzle portion 301 in contrast to the ejector 14 described in the first embodiment. The nozzle portion 301 of the ejector 141 does not include the third passage 323 and the fourth passage 324. Therefore, in the nozzle portion 301, the second outlet of the second passage 322 serves as the injection port 32e.

[0219] The ejector-type refrigeration cycle 10 of this embodiment employs carbon dioxide (i.e., R744) as the refrigerant. The ejector-type refrigeration cycle 10 of this embodiment constitutes a supercritical refrigeration cycle in which the pressure of the high-pressure side refrigerant is equal to or higher than the critical pressure of the refrigerant.

[0220] In the ejector refrigeration cycle 10 of this embodiment, a refrigerant in a supercritical state with a relatively high pressure flows into the nozzle 301 of the ejector 141. When the refrigerant in the supercritical state flows into the nozzle 301, the velocity of the droplets contained in the refrigerant that has become a two-phase gas-liquid refrigerant during the decompression process becomes faster than the sonic velocity of the two-phase gas-liquid refrigerant. As a result, there is a shortage of bubble nuclei necessary to evaporate the refrigerant that has become a liquid phase during the decompression process.

[0221] As a result, the droplets generated by decompression bubbling do not have enough time to dissipate heat into the surrounding gas-phase refrigerant, causing the temperature of the droplets to exceed the saturation temperature of the refrigerant. In other words, part of the pressure energy of the refrigerant that flows into the nozzle 301 cannot be converted into kinetic energy, and is instead stored as thermal energy in the droplets.

[0222] Therefore, in the ejector-type refrigeration cycle 10 of this embodiment, the refrigerant may be in a thermodynamic non-equilibrium state in the first passage 321 or the first outlet 32b, where the velocity difference between the liquid droplets and the gas phase refrigerant is likely to increase. If the refrigerant is in a thermodynamic non-equilibrium state, the velocity of the refrigerant in the ejector 141 cannot be sufficiently accelerated, and the ejector 141 cannot sufficiently exert its fluid pressure increasing effect.

[0223] Therefore, in the ejector 141, the cross-sectional area ratio CAin / CAth of the first passage 321 is set to 10 or less. The cross-sectional area ratio of the first passage 321 is the ratio of the passage cross-sectional area CAin at the nozzle inlet 32a to the passage cross-sectional area CAth at the first outlet 32b. This adjusts the slip ratio of the gas-liquid two-phase refrigerant in the first passage 321 or the first outlet 32b so that a thermodynamic nonequilibrium state is eliminated in the second passage 322.

[0224] More specifically, by setting the cross-sectional area ratio CAin / CAth of the first passage 321 to 10 or less, the acceleration rate of the liquid-phase refrigerant in the first passage 321 can be suppressed. This allows the position where the refrigerant starts to bubble closer to the first outlet 32b. Furthermore, in the mist flow generated after bubbling, a speed difference is created between the gas-phase refrigerant and the liquid droplets, promoting heat dissipation from the liquid droplets, whose temperature is higher than the saturation temperature, to the surrounding gas-phase refrigerant. This eliminates a thermodynamic non-equilibrium state in the second passage 322.

[0225] Furthermore, in the ejector 141, even if the refrigerant injected from the nozzle portion 301 enters a thermodynamic non-equilibrium state, the passage cross-sectional area of ​​the mixing portion 42 is changed so that the thermodynamic non-equilibrium state of the mixed refrigerant in the mixing portion 42 can be resolved.

[0226] Specifically, with a refrigerant such as R744, in which the ratio of the density of droplets to the density of the gas phase refrigerant is relatively small, the droplets contained in the sprayed refrigerant are easily dispersed within the mixing section 42. Therefore, the degree of reduction in the passage cross-sectional area of ​​the convergent mixing section 42a is set so as to sufficiently accelerate the speed of the suctioned refrigerant and promote heat dissipation from the suctioned refrigerant from the droplets.

[0227] As a result, even if the sprayed refrigerant reaches a thermodynamic non-equilibrium state, the droplets contained in the sprayed refrigerant are diffused within the mixing section 42, and the heat contained in the droplets is dissipated into the suctioned refrigerant that has been depressurized and accelerated, thereby making it possible to eliminate the thermodynamic non-equilibrium state in the mixing section 42.

[0228] In the tapered mixing section 42a of the present embodiment, the velocity of the suctioned refrigerant is sufficiently accelerated by making the pressure of at least a portion of the mixed refrigerant in the tapered mixing section 42a lower than the pressure of the sprayed refrigerant at the injection port 32e of the nozzle 301. In addition, in the tapered mixing section 42a of the present embodiment, the degree of reduction of the passage cross-sectional area is determined within a range in which the tapered mixing distance LMIX1 satisfies the formulas F6 and F7 described in the first embodiment.

[0229] Furthermore, in the ejector-type refrigeration cycle device 10 of this embodiment, unlike the first embodiment, the inlet side of the accumulator 17 is connected to the refrigerant outlet of the first chiller 16a. The accumulator 17 is a low-pressure gas-liquid separator that separates the refrigerant flowing out from the first chiller 16a into gas and liquid phases and stores the separated liquid-phase refrigerant as surplus refrigerant for the cycle. The gas-phase refrigerant outlet of the accumulator 17 is connected to the suction port side of the compressor 11. The remaining configuration of the vehicle refrigeration cycle device 1 is the same as that of the first embodiment.

[0230] Next, a description will be given of the operation of the vehicle refrigeration cycle device 1 of this embodiment having the above-described configuration. In the control program of this embodiment, similarly to the first embodiment, when it is determined that cooling of the battery 80 is necessary, the control device 70 controls the operation of various control target devices.

[0231] In this embodiment, the control device 70 controls the throttle opening of the electric expansion valve 15 so that the pressure P1 and temperature T1 of the refrigerant flowing into the electric expansion valve 15 approach the target pressure PO1 and target temperature TO1. The target pressure PO1 and target temperature TO1 are determined so that the COP of the cycle approaches its maximum value. The control device 70 also controls the operation of the initial speed control valve 50 so that the thermodynamic non-equilibrium state of the refrigerant is resolved.

[0232] Other devices to be controlled are controlled in the same manner as in Embodiment 1. Furthermore, in the ejector-type refrigeration cycle 10 of this embodiment, the change in the state of the refrigerant during normal operation, in which the flow rate of the refrigerant circulating through the cycle is relatively small, differs from the change in the state of the refrigerant during high-load operation, in which the flow rate of the refrigerant circulating through the cycle is relatively large.

[0233] First, during normal operation, the state of the refrigerant changes as shown in the Mollier diagram of Fig. 10. In Fig. 10, the state of the refrigerant at the same points in the cycle configuration as in Fig. 7 described in the first embodiment is indicated by the same reference characters (alphabet), and only the subscripts (numbers) have been changed to match the diagram numbers.

[0234] The discharged refrigerant (point a10 in FIG. 10 ) that has been pressurized above the critical pressure by the compressor 11 flows into the outdoor heat exchanger 12. The discharged refrigerant that has flowed into the outdoor heat exchanger 12 dissipates heat into the outside air blown by the outdoor air fan, thereby reducing the enthalpy (from point a10 to point b10 in FIG. 10 ). The flow of the subcooled liquid phase refrigerant that has flowed out of the outdoor heat exchanger 12 is branched at the branching section 13. One of the refrigerants flowing out of the branching section 13 flows into the refrigerant inlet 33 of the ejector 141.

[0235] The refrigerant that flows into the initial velocity adjustment space 31 through the refrigerant inlet 33 of the ejector 141 has its initial velocity adjusted by the initial velocity adjustment valve 50, and then flows into the nozzle passage 32. The refrigerant that flows into the nozzle passage 32 is decompressed in the first passage 321 and becomes a gas-liquid two-phase refrigerant in a thermodynamic non-equilibrium state (from point b10 to point c10 in FIG. 10 ). The refrigerant then reaches a critical state in the second passage 322. The second passage 322 also includes a first outlet 32b that serves as the inlet of the second passage 322.

[0236] Here, the thin dashed line in Fig. 10 is an isentropic line on the inlet side that indicates the same entropy as the refrigerant that has flowed into the nozzle passage 32. As shown in Fig. 10, the refrigerant that has reached a thermodynamic non-equilibrium state (point c10 in Fig. 10) deviates from the isentropic line on the inlet side because the droplets store thermal energy.

[0237] In contrast to this, in the ejector 141 of this embodiment, during normal operation, the passage cross-sectional area of ​​the refrigerant passage in the nozzle portion 301 is changed so that the thermodynamic non-equilibrium state of the refrigerant can be eliminated in the second passage 322. Furthermore, the initial velocity control valve 50 adjusts the initial velocity of the refrigerant flowing into the nozzle passage 32 so that the thermodynamic non-equilibrium state of the refrigerant can be eliminated in the second passage 322.

[0238] As a result, the refrigerant injected from the injection port 32e approaches the isentropic line on the inlet side (from point c10 to point e10 in FIG. 10 ). Furthermore, during normal operation, the initial velocity control valve 50 adjusts the initial velocity of the refrigerant, thereby changing the position in the second passage 322 where the refrigerant reaches a thermodynamic equilibrium state.

[0239] The mixed refrigerant of the refrigerant injected from the nozzle 301 and the refrigerant sucked from the suction port 41 also reaches a thermodynamic equilibrium state. Therefore, the state of most mixed refrigerants approaches point h10 in Fig. 10. Here, the thin dashed dotted line in Fig. 10 is a pressure-boosting isentropic line that indicates the same entropy as the refrigerant immediately after it flows out of the refrigerant outlet 44 of the ejector 141. Point h10 is located on the pressure-boosting isentropic line.

[0240] Then, the mixed refrigerant is isentropically pressurized by the action of the shock waves generated in the divergent mixing section 42b (from point h10 to point i10 in FIG. 10). The other operations are the same as those in the first embodiment.

[0241] Next, high-load operation will be described. During high-load operation, the flow rate of refrigerant circulating through the cycle increases, causing the speed of the refrigerant flowing through the refrigerant passage of the nozzle portion 301 to increase compared to normal operation. As a result, the second passage 322 is unable to resolve the thermodynamic non-equilibrium state of the gas-liquid two-phase refrigerant that occurs in the first passage 321 or the first outlet 32b of the nozzle portion 301. As a result, the refrigerant in the thermodynamic non-equilibrium state is sprayed from the spray port 32e.

[0242] Therefore, during high-load operation, as shown in the Mollier diagram of FIG. 11, the refrigerant that is in a thermodynamic non-equilibrium state in the first passage 321 or the first outlet 32b of the nozzle portion 301 (point c11 in FIG. 11) is injected in a state deviating from the isentropic line on the inlet side (from point c11 to point e11 in FIG. 11).

[0243] In the mixed refrigerant of the injected refrigerant and the suctioned refrigerant, the thermodynamic non-equilibrium state is not resolved. Therefore, the state of many mixed refrigerants approaches point h11 in Fig. 11. Point h11 is located on the side with higher enthalpy than the isentropic line on the pressure increasing side.

[0244] In contrast to this, in the ejector 141 of the present embodiment, during high load operation, the passage cross-sectional area of ​​the mixing section 42 is changed so that the thermodynamic nonequilibrium state of the mixed refrigerant can be eliminated in the mixing section 42. Therefore, the mixed refrigerant of the suction refrigerant and the sprayed refrigerant accelerated in the tapered mixing section 42a approaches the isentropic line on the pressure increasing side.

[0245] Then, the mixed refrigerant is isentropically pressurized by the action of the shock waves generated in the divergent mixing section 42b (from point h11 to point i11 in FIG. 11). The other operations are the same as those during normal operation.

[0246] As described above, the vehicle refrigeration cycle device 1 of this embodiment can cool the battery 80, similarly to the first embodiment. Furthermore, since the ejector-type refrigeration cycle device 10 of this embodiment is provided with the ejector 141, even if a load fluctuation occurs when the heat generation amount of the battery 80 changes, a sufficiently high COP is exhibited regardless of the load fluctuation.

[0247] More specifically, in the ejector 141 of this embodiment, even if the refrigerant is in a thermodynamic non-equilibrium state in the first passage 321, the refrigerant can be brought into a thermodynamic equilibrium state in either the second passage 322 or the mixing section 42. Therefore, the pressure energy of the refrigerant that has flowed into the nozzle portion 301 can be effectively converted into kinetic energy within the ejector 141 in either the second passage 322 or the mixing section 42.

[0248] Furthermore, even if there is a change in the pressure, temperature, or flow rate of the refrigerant flowing into the nozzle portion 301, the flow rate of the refrigerant flowing through the nozzle passage 32 can be adjusted by changing the position at which the refrigerant reaches a critical state in the second passage 322. Therefore, there is no need to change the passage cross-sectional area of ​​the first outlet 32b in order to change the flow rate of the fluid flowing through the nozzle passage 32.

[0249] As a result, in the ejector 141 of this embodiment, the pressure-increasing capability of the ejector 141 can be improved regardless of fluctuations in the flow rate of the refrigerant flowing into the nozzle portion 301 .

[0250] Furthermore, this embodiment discloses a method of operating a nozzle device that can improve the nozzle efficiency ηnoz during normal operation regardless of fluctuations in the flow rate of the inflowing fluid.

[0251] That is, as an operating method of the nozzle device constituted by the nozzle portion 301 and the initial velocity adjustment valve 50, during normal operation, a gas-liquid two-phase fluid that is in a thermodynamic non-equilibrium state in the first passage 321 is brought to a critical state and thermodynamic equilibrium state in the second passage 322. Furthermore, an operating method is disclosed in which the position where the fluid is in a thermodynamic equilibrium state in the second passage 322 is changed by adjusting the velocity of the fluid flowing into the nozzle inlet 32a.

[0252] With this, even if the fluid flowing out from the first outlet 32b is in a thermodynamic non-equilibrium state, the fluid is brought to a critical state and a thermodynamic equilibrium state in the second passage 322. Therefore, at the outlet of the second passage 322, the pressure energy of the fluid flowing into the nozzle inlet 32a is effectively converted into velocity energy, thereby increasing the velocity of the injected fluid. As a result, the nozzle efficiency ηnoz can be improved.

[0253] Furthermore, by adjusting the velocity of the fluid flowing into the nozzle inlet 32a, the position where the fluid reaches a thermodynamic equilibrium state in the second passage 322 is changed, thereby changing the velocity and density at which the fluid reaches a critical state, thereby changing the critical flow rate of the fluid. In other words, there is no need to change the passage cross-sectional area of ​​the first outlet 32b in order to change the flow rate of the fluid flowing through the nozzle passage 32.

[0254] Therefore, even if a flow rate fluctuation of the fluid flowing into the nozzle portion 301 occurs, it is unlikely to affect the generation pattern or amount of bubbles generated in the liquid-phase fluid near the wall surface of the first outlet 32b. As a result, it is possible to generate sufficient bubbles in the liquid-phase fluid and increase the velocity of the ejected fluid regardless of a flow rate fluctuation of the inflowing fluid. In other words, it is possible to improve the nozzle efficiency ηnoz regardless of a flow rate fluctuation of the inflowing fluid.

[0255] Furthermore, the nozzle device of this embodiment is provided with the initial velocity adjustment valve 50. This makes it possible to easily change the critical flow rate of the fluid in the second passage 322.

[0256] (Third Embodiment) In this embodiment, as shown in the overall configuration diagram of Fig. 12, an example will be described in which a pressure regulating valve 18 is added to the ejector-type refrigeration cycle 10 described in the second embodiment. The pressure regulating valve 18 is a pressure adjusting unit that adjusts the pressure of the refrigerant flowing into the nozzle inlet 32a. The basic configuration of the pressure regulating valve 18 is the same as that of the electric expansion valve 15. The rest of the configuration of the vehicle refrigeration cycle device 1 is the same as that of the second embodiment.

[0257] In the vehicle refrigeration cycle device 1 of this embodiment, when the control program determines that cooling of the battery 80 is necessary, the control device 70 controls the operation of the pressure regulating valve 18 so that the inlet pressure Pni of the refrigerant flowing into the nozzle inlet 32a approaches a predetermined reference inlet pressure KPni. Other controlled devices are controlled in the same manner as in the second embodiment.

[0258] Therefore, the ejector-type refrigeration cycle 10 of this embodiment operates in the same manner as the second embodiment, and can obtain the same effects as the second embodiment. That is, since the ejector 141 is provided, even if a load fluctuation occurs, a sufficiently high COP can be achieved regardless of the load fluctuation.

[0259] Furthermore, since the ejector refrigeration cycle device 10 of this embodiment is provided with the pressure regulating valve 18, the inlet pressure Pni can be made to approach the reference inlet pressure KPni. This stabilizes the inlet pressure Pni, allowing the initial velocity regulating valve 50 to accurately regulate the critical flow rate of the refrigerant in the second passage 322. Similarly, the operating method of the nozzle device of this embodiment allows the critical flow rate of the refrigerant in the second passage 322 to be accurately regulated.

[0260] Fourth Embodiment In this embodiment, an example will be described in which the ejector 14 is applied to an ejector-type refrigeration cycle 10a shown in the overall configuration diagram of Fig. 13. In this embodiment, the ejector-type refrigeration cycle 10a is applied to a vehicle air conditioner 1a.

[0261] In the ejector refrigeration cycle 10a, the discharge port of the compressor 11 is connected to the refrigerant inlet side of the outdoor heat exchanger 12. The refrigerant outlet of the outdoor heat exchanger 12 is connected to the refrigerant inlet 33 side of the ejector 14. The refrigerant outlet 44 of the ejector 14 is connected to the inlet side of the accumulator 17. The gas-phase refrigerant outlet of the accumulator 17 is connected to the suction port side of the compressor 11.

[0262] The liquid-phase refrigerant outlet of the accumulator 17 is connected to the inlet side of the electric expansion valve 15. The outlet of the electric expansion valve 15 is connected to the refrigerant inlet side of the interior evaporator 16c. The interior evaporator 16c exchanges heat between the refrigerant and the ventilation air sent into the vehicle cabin, which is the space to be air-conditioned. The interior evaporator 16c is an evaporation unit that causes the refrigerant flowing out from the ejector 14 to absorb heat from the ventilation air, thereby evaporating the refrigerant.

[0263] The refrigerant outlet of the interior evaporator 16c is connected to the suction port 41 side of the ejector 14. In this embodiment, R1234yf is used as the refrigerant for the ejector-type refrigeration cycle 10a, as in the first embodiment. Other basic configurations of the vehicle air conditioning system 1a are the same as those of the vehicle refrigeration cycle system 1 described in the first embodiment.

[0264] Next, the operation of the vehicle air conditioner 1a of this embodiment will be described. When the air conditioning switch is turned on while the start switch of the vehicle system is turned on, the control device 70 executes the air conditioning control program stored in the ROM.

[0265] When the air conditioning control program is executed, the control device 70 controls the operation of various components of the ejector-type refrigeration cycle 10a to cool the blown air. Specifically, the control device 70 controls the rotation speed of the compressor 11 so that the refrigerant evaporation temperature in the indoor evaporator 16c approaches the target evaporation temperature TEO.

[0266] The target evaporation temperature TEO is determined based on the target outlet temperature TAO of the air blown into the vehicle cabin, which is calculated using the outside air temperature, the inside air temperature, the amount of solar radiation, the interior temperature setting set by the occupant, etc.

[0267] The control device 70 also controls the throttle opening of the electric expansion valve 15 so that the degree of superheat SH1 of the refrigerant flowing out of the indoor evaporator 16c approaches a predetermined reference degree of superheat KSH1. The control device 70 also controls the operation of the initial speed control valve 50 so that the refrigerant flowing into the nozzle passage 32 reaches a critical state in the second passage 322.

[0268] Therefore, in the ejector-type refrigeration cycle 10a, the refrigerant discharged from the compressor 11 flows into the outdoor heat exchanger 12. In the outdoor heat exchanger 12, the refrigerant dissipates heat to the outside air and condenses. The refrigerant flowing out of the outdoor heat exchanger 12 becomes a driving-side fluid and flows into the refrigerant inlet 33 of the ejector 14.

[0269] In the ejector 14, similarly to the first embodiment, the mixed refrigerant of the refrigerant sprayed from the nozzle portion 30 and the suction refrigerant sucked through the suction port 41 is pressurized by the action of the shock wave and flows out from the refrigerant outlet 44. The refrigerant flowing out from the refrigerant outlet of the ejector 14 flows into the accumulator 17 and is separated into gas and liquid.

[0270] The liquid-phase refrigerant separated in the accumulator 17 flows into the electric expansion valve 15 and is decompressed. The refrigerant decompressed by the electric expansion valve 15 flows into the interior evaporator 16c. In the interior evaporator 16c, the refrigerant decompressed by the electric expansion valve 15 absorbs heat from the blown air and evaporates. This cools the blown air that is blown into the vehicle cabin.

[0271] The indoor evaporator 16c is sucked through a suction port 41 of the ejector 14. Therefore, in the ejector 14 of this embodiment, the refrigerant flowing out from the indoor evaporator 16c serves as a suction-side fluid. The gas-phase refrigerant separated in the accumulator 17 is sucked into the compressor 11 and compressed again.

[0272] As described above, the vehicle air conditioner 1a of this embodiment can cool the air to be blown into the vehicle compartment. Furthermore, the ejector-type refrigeration cycle 10a of this embodiment can obtain the same effects as those of the first embodiment.

[0273] That is, the ejector 14 of the present embodiment can exhibit a high pressure-boosting capability regardless of fluctuations in the flow rate of the refrigerant flowing into the nozzle portion 30. Furthermore, since the ejector-type refrigeration cycle 10a is provided with the ejector 14, it can exhibit a sufficiently high COP regardless of load fluctuations.

[0274] Fifth Embodiment In this embodiment, as shown in the overall configuration diagram of FIG. 14, an example will be described in which an ejector-type refrigeration cycle 10a including the ejector 141 described in the second embodiment is applied to a vehicle air conditioner 1a.

[0275] In this embodiment, the refrigerant of the ejector-type refrigeration cycle 10a is R744, as in the second embodiment. Therefore, the ejector-type refrigeration cycle 10a of this embodiment constitutes a supercritical refrigeration cycle. The rest of the configuration of the vehicle air conditioner 1a is the same as that of the third embodiment.

[0276] Next, the operation of the vehicle air conditioner 1a of this embodiment having the above-described configuration will be described. The basic operation of the vehicle air conditioner 1a of this embodiment is the same as that of the fourth embodiment. Furthermore, in the ejector-type refrigeration cycle 10a of this embodiment, the change in the state of the refrigerant during normal operation is different from the state of the refrigerant during high-load operation, as in the second embodiment.

[0277] First, during normal operation, as shown in the Mollier diagram of Fig. 15, the refrigerant in a supercritical state discharged from the compressor 11 (point a15 in Fig. 15) flows into the outdoor heat exchanger 12. The refrigerant that flows into the outdoor heat exchanger 12 releases heat to the outside air, reducing its enthalpy (from point a15 to point b15 in Fig. 15). The refrigerant that flows out of the outdoor heat exchanger 12 flows into the refrigerant inlet 33 of the ejector 141.

[0278] The refrigerant that flows into the initial velocity adjustment space 31 through the refrigerant inlet 33 of the ejector 141 has its initial velocity adjusted by the initial velocity adjustment valve 50, and then flows into the nozzle passage 32. The refrigerant that flows into the nozzle passage 32 is decompressed in the first passage 321 and becomes a gas-liquid two-phase refrigerant that is in a thermodynamic non-equilibrium state (from point b15 to point c15 in FIG. 15 ). Then, the refrigerant reaches a critical state at the first outlet 32b or the second passage 322.

[0279] As in the second embodiment, the thermodynamically non-equilibrium state of the refrigerant is resolved in the second passage 322. Therefore, the refrigerant injected from the injection port 32e approaches the inlet isentropic line (from point c15 to point e15 in FIG. 15 ). Most of the mixed refrigerant, consisting of the injected refrigerant from the nozzle 301 and the suction refrigerant sucked through the suction port 41, approaches point h15 in FIG. 15 , as in the second embodiment.

[0280] Then, the mixed refrigerant is isentropically pressurized by the action of the shock waves generated in the diverging mixing section 42b (from point h15 to point i15 in FIG. 15 ), and flows out from the refrigerant outlet 44 of the ejector 141. The refrigerant that has flowed out from the refrigerant outlet 44 of the ejector 141 flows into the accumulator 17 and is separated into gas and liquid (from point i15 to point j15 and from point i15 to point k15 in FIG. 15 ).

[0281] The liquid-phase refrigerant flowing out from the liquid-phase refrigerant outlet of the accumulator 17 flows into the electric expansion valve 15 and is decompressed (from point k15 to point f15 in FIG. 15). The refrigerant decompressed by the electric expansion valve 15 flows into the indoor evaporator 16c. In the indoor evaporator 16c, the refrigerant flowing out from the electric expansion valve 15 absorbs heat from the blown air and evaporates (from point f15 to point g15 in FIG. 15). This cools the blown air. The refrigerant flowing out from the indoor evaporator 16c is drawn into the suction port 41 of the ejector 141.

[0282] The gas phase refrigerant flowing out from the gas phase refrigerant outlet of the accumulator 17 is drawn into the compressor 11 and compressed again (from point j15 to point a15 in FIG. 15). The other operations are the same as those of the fourth embodiment.

[0283] Next, high-load operation will be described. During high-load operation, as shown in the Mollier diagram of Fig. 16, the refrigerant in a thermodynamically non-equilibrium state in the first passage 321 or the first outlet 32b of the nozzle portion 301 (point c16 in Fig. 16) is injected in a state deviating from the isentropic line on the inlet side (from point c16 to point e16 in Fig. 16). Most of the mixed refrigerant of the injected refrigerant and the suction refrigerant in the thermodynamically non-equilibrium state approaches point h16 in Fig. 16, as in the second embodiment.

[0284] The mixed refrigerant of the suction refrigerant and the sprayed refrigerant accelerated in the convergent mixing section 42a approaches the isentropic line of the pressure increase side, as in the second embodiment. Then, the shock waves generated in the divergent mixing section 42b cause the mixed refrigerant to be isentropically pressurized (from point h16 to point i16 in FIG. 16 ). The remaining operation is the same as in normal operation.

[0285] Therefore, the vehicle air conditioner 1a of this embodiment can also cool the air to be blown into the vehicle compartment, as in the fourth embodiment. Furthermore, since the ejector-type refrigeration cycle 10a of this embodiment is provided with the ejector 141, it can achieve a sufficiently high COP regardless of load fluctuations, as in the second embodiment.

[0286] Sixth Embodiment In this embodiment, as shown in the overall configuration diagram of Fig. 17, a pressure regulating valve 18 similar to that of the third embodiment is added to the ejector refrigeration cycle 10a described in the fifth embodiment. The rest of the configuration of the vehicle air conditioner 1a is the same as that of the fifth embodiment. Furthermore, in the vehicle air conditioner 1a of this embodiment, the control device 70 controls the operation of the pressure regulating valve 18, similar to that of the third embodiment. The other controlled devices are controlled in the same way as in the fifth embodiment.

[0287] Therefore, the ejector-type refrigeration cycle 10a of this embodiment operates in the same manner as the fifth embodiment, and can obtain the same effects as the fifth embodiment. That is, since the ejector 141 is provided, a sufficiently high COP can be achieved regardless of load fluctuations.

[0288] Furthermore, the ejector refrigeration cycle 10a of this embodiment is provided with a pressure regulating valve 18. Therefore, similar to the third embodiment, the critical flow rate of the refrigerant in the second passage 322 can be accurately adjusted by the initial velocity regulating valve 50. Similarly, according to the operating method of the nozzle device of this embodiment, the critical flow rate of the refrigerant in the second passage 322 can be accurately adjusted.

[0289] Other Embodiments The present disclosure is not limited to the above-described embodiments, and various modifications can be made as follows within the scope of the present disclosure.

[0290] In the above embodiment, an example in which the ejector according to the present disclosure is applied to an ejector-type refrigeration cycle has been described, but the application of the ejector according to the present disclosure is not limited to this. For example, the ejector according to the present disclosure may be applied to a vacuum pump that utilizes negative pressure generated at the suction port 41 to bring a predetermined space close to a vacuum state.

[0291] In the above-described embodiment, the ejector-type refrigeration cycle 10, 10a according to the present disclosure is applied to the vehicle refrigeration cycle device 1 and the vehicle air conditioner 1a, but the present disclosure is not limited to this. For example, the ejector-type refrigeration cycle 10, 10a may be applied to a stationary air conditioner, a refrigeration device, or a refrigerator.

[0292] In the above-described embodiment, the heat dissipation unit of the ejector-type refrigeration cycle 10, 10a according to the present disclosure may be used as a heating unit that heats an object to be heated using the refrigerant discharged from the compressor 11 as a heat source. In this case, the ejector-type refrigeration cycle 10, 10a may be applied to a heating device or a hot water supply device that heats water for daily use, etc. Furthermore, the ejector-type refrigeration cycle 10, 10a may be applied to a heating device for warming up the above-described in-vehicle equipment.

[0293] Furthermore, in the above embodiment, an example has been described in which the operating method of the nozzle device according to the present disclosure is applied to the nozzle portion 30 of the ejector 14, but the operating method of the nozzle device according to the present disclosure can be applied to a wide range of nozzle devices in which improved nozzle efficiency is required.

[0294] The configuration of the ejector according to the present disclosure is not limited to the configuration disclosed in the above embodiment.

[0295] For example, in the above embodiment, an example has been described in which the nozzle portion 30 of the ejector 14 and the initial speed adjustment valve 50, which is a speed adjustment portion, are integrally formed, but the nozzle portion 30 and the initial speed adjustment valve 50 may be formed separately. Also, an ejector that does not have the initial speed adjustment valve 50 may be adopted as the ejector 14, 141.

[0296] In the above embodiment, the ejector 141 is described in which the cross-sectional area ratio CAin / CAth is set to 10 or less to eliminate the thermodynamic nonequilibrium state in the second passage 322. However, the present invention is not limited to this. The second passage ratio L2 / φD1 may be set to 3 or more to eliminate the thermodynamic nonequilibrium state in the second passage 322. The second passage ratio L2 / φD1 is the ratio of the axial length L2 of the second passage 322 to the diameter φD1 of the first outlet 32b corresponding to the throat portion.

[0297] The configuration of the ejector refrigeration cycle according to the present disclosure is not limited to the configuration disclosed in the above embodiment, and may be configured to switch the refrigerant circuit depending on the operation mode, as long as the operation described in the above embodiment can be performed.

[0298] Furthermore, the circuit configuration of the ejector refrigeration cycle according to the present disclosure is not limited to the circuit configuration disclosed in the above-described embodiment. In the ejector refrigeration cycle 10 described in the first embodiment, a dryness adjusting unit may be employed instead of the branch unit 13. The dryness adjusting unit is a branch unit that can allow refrigerants with different drynesses to flow out.

[0299] More specifically, the quality adjustment unit may be a centrifugal separator that separates the refrigerant into gas and liquid by centrifugal force. The refrigerant with a relatively high quality at the center may flow out to the refrigerant inlet 33 of the ejector 14, and the refrigerant with a relatively low quality at the periphery may flow out to the electric expansion valve 15.

[0300] In the ejector refrigeration cycle 10 described in the first embodiment, an internal heat exchanger may be provided instead of the first chiller 16a. The internal heat exchanger is an internal heat exchange unit that exchanges heat between the high-pressure refrigerant from the heat dissipation unit and the refrigerant flowing out from the refrigerant outlet 44 of the ejector 14. This reduces the enthalpy of the refrigerant flowing into the second chiller 16b, thereby increasing the cooling capacity exerted by the second chiller 16b.

[0301] In the ejector-type refrigeration cycle 10 described in the first to third embodiments, different objects to be cooled may be cooled in the evaporation heat exchanger (first chiller 16a in the first embodiment) and the evaporation section (second chiller 16b in the first embodiment).

[0302] In the ejector-type refrigeration cycle 10a described in the fourth to sixth embodiments, an evaporation heat exchanger may be connected in parallel to the evaporation section (the indoor evaporator 16c in the fourth embodiment), and the evaporation heat exchanger may be configured to cool an object to be cooled that is different from the object to be cooled in the evaporation section.

[0303] In the above embodiment, the outdoor heat exchanger 12 is used as the heat radiating unit, but the present invention is not limited to this. For example, the heat radiating unit may be configured by arranging a heat medium pump, a water-refrigerant heat exchanger, a heat radiating heat exchanger, etc. in a heat medium circuit that circulates the heat medium.

[0304] The heat medium pump is a pump that pumps a heat medium into the water passage of the water-refrigerant heat exchanger. The basic configuration of the heat medium pump may be the same as that of the cooling water pump 61. The same type of fluid as the cooling water can be used as the heat medium.

[0305] The water-refrigerant heat exchanger is a heat exchanger that exchanges heat between the high-pressure refrigerant discharged from the compressor 11 and the heat medium pumped from the heat medium pump. The heat dissipation heat exchanger is a heating heat exchanger that exchanges heat between the high-temperature heat medium heated in the water-refrigerant heat exchanger and a fluid to be dissipated. The fluid to be dissipated is not limited to outside air, but may be a heated object such as blown air to be sent to a space to be air-conditioned.

[0306] In the above embodiment, the battery 80 is used as the vehicle-mounted device to be cooled by the ejector-type refrigeration cycle, but the device is not limited to this. The vehicle-mounted device may also be configured to adjust the temperature of a motor generator, an inverter, a sensor processing unit, a transaxle, an ADAS control device, etc.

[0307] A motor generator is an electric motor that functions as both a motor that outputs driving force for driving and a generator. An inverter is an electric circuit device that supplies power to the motor generator, etc. A sensor processing unit is a control device that integrates environmental sensor interfaces and communication functions for autonomous driving and energy-saving driving. A transaxle is a power transmission mechanism that integrates a transmission, differential gear, etc. An ADAS control device is a control device for an advanced driver assistance system.

[0308] In the third and sixth embodiments, the ejector refrigeration cycles 10, 10a are described as including the pressure regulating valve 18 and the ejector 14, 141 having the initial speed regulating valve 50. However, the present invention is not limited to this. For example, the ejector 14, 141 may be used without the initial speed regulating valve 50 and may be provided with the pressure regulating valve 18. In this case, the speed of the refrigerant flowing into the nozzle portion 30, 301 may be adjusted by the pressure regulating valve 18.

[0309] Furthermore, the group of control sensors connected to the input side of the control device 70 is not limited to the detection units disclosed in the above embodiment, and various detection units may be added as necessary.

[0310] In the above-described embodiment, the refrigerants used in the ejector-type refrigeration cycles 10 and 10a are R1234yf and R744, but the present invention is not limited to these. For example, R134a, R600a, R410A, R404A, R32, R407C, R290, etc. may also be used. A mixed refrigerant containing a mixture of two or more of these refrigerants may also be used.

[0311] In the above embodiment, an example in which an ethylene glycol aqueous solution is used as the coolant has been described, but the present invention is not limited to this. For example, the heating heat medium and the low-temperature heat medium may be a solution containing dimethylpolysiloxane or nanofluid, an antifreeze solution, an aqueous liquid refrigerant containing alcohol, or a liquid medium containing oil.

[0312] The control aspects of the ejector refrigeration cycle according to the present disclosure are not limited to the control aspects disclosed in the above-described embodiments.

[0313] For example, in the vehicle refrigeration cycle system 1 of the first embodiment, the control device 70 may control the operation of the initial speed control valve 50 of the ejector 14 so that the degree of supercooling SC of the refrigerant flowing into the ejector 14 approaches the target degree of supercooling SCO. Furthermore, the control device 70 may control the throttle opening of the electric expansion valve 15 so that the degree of superheat SH of the suction refrigerant approaches a predetermined reference degree of superheat KSH. Furthermore, as shown in Fig. 12, a throttle device may be provided upstream of the ejector 14, and the opening of the throttle device may also be controlled to adjust the pressure of the refrigerant flowing into the first passage 321.

[0314] The refrigeration cycle device disclosed in this specification has the following features. (Item 1) An ejector comprising: a nozzle portion (30, 301) for injecting a driving-side fluid to reduce its pressure; a suction port (41) for sucking a suction-side fluid; a mixing portion (42) for mixing the jet fluid jetted from the nozzle portion with the suction fluid sucked from the suction port; and a body portion (40) having formed therein a refrigerant outlet (44) for discharging the mixed fluid mixed in the mixing portion, wherein the nozzle portion has a nozzle passage (32) extending from a nozzle inlet (32a) where the decompression of the fluid starts to an injection port (32e) from which the fluid is ejected, and the nozzle passage includes a first passage (321) for reducing the passage cross-sectional area from the nozzle inlet toward a downstream side in a flow direction of the fluid, and a second passage (322) for increasing the passage cross-sectional area from a first outlet (32b) of the first passage toward a downstream side in the flow direction of the fluid, wherein bubbles are generated in the fluid in the first passage, and the fluid reaches a critical state in the second passage. (Item 2) The velocity of the fluid in the second passage is determined by the momentum frozen phase change equilibrium sound velocity a ｆｅ 2. The ejector according to item 1,

[0315] However, the momentum frozen phase change equilibrium sound speed a ｆｅ is defined by the following equation (1).

[0316]

[0317] ν Ｇ is the specific volume of the gas phase fluid. Ｌ is the specific volume of the liquid phase fluid. ν is the average specific volume of the gas-liquid two-phase fluid. S Ｇ is the specific entropy of the gas phase refrigerant. Ｌ is the specific entropy of the liquid-phase refrigerant. T is the temperature of the fluid. x is the axial distance. (Item 3) The ejector according to Item 1 or 2, wherein the first passage is formed with a constant area portion (321b) where the passage cross-sectional area is constant. (Item 4) The nozzle passage includes a third passage (323) where the passage cross-sectional area decreases from a second outlet (32c) of the second passage toward the downstream side in the flow direction, and a fourth passage (324) where the passage cross-sectional area increases from a third outlet (32d) of the third passage toward the downstream side in the flow direction of the fluid, and the fluid at the third outlet reaches a critical state. (Item 5) The ejector according to any one of Items 1 to 3, wherein the velocity of the fluid at the third outlet is equal to or greater than the momentum equilibrium phase change equilibrium sonic velocity a ｅｅ 5. The ejector according to item 4,

[0318] However, the momentum equilibrium phase change equilibrium sound speed a ｅｅ is defined by the following equation (2).

[0319]

[0320] ν Ｇ is the specific volume of the gas phase fluid. Ｌ is the specific volume of the liquid phase fluid. ν is the average specific volume of the gas-liquid two-phase fluid. S Ｇ is the specific entropy of the gas phase refrigerant. Ｌwhere ρ is the specific entropy of the liquid-phase refrigerant, T is the temperature of the fluid, and x is the axial distance. (Item 6) The ejector according to Item 4 or 5, wherein the velocity of the fluid that has reached a critical state at the third outlet is slower than the velocity of the fluid that has reached a critical state in the second passage. (Item 7) A refrigerant compressor includes a nozzle portion (30, 301) for decompressing and injecting a driving-side fluid, a suction port (41) for sucking a suction-side fluid, a mixing portion (42) for mixing the jet fluid jetted from the nozzle portion with the suction fluid sucked from the suction port, and a body portion (40) formed with a refrigerant outlet (44) for discharging the mixed fluid mixed in the mixing portion, wherein the nozzle portion is formed with a nozzle passage (32) extending from a nozzle inlet (32a) where decompression of the fluid starts to an injection port (32e) from which the fluid is jetted, and the nozzle passage is formed with a first passage (321) for reducing the passage cross-sectional area from the nozzle inlet toward the downstream side in the flow direction of the fluid, and a second passage (322) for increasing the passage cross-sectional area from a first outlet (32b) of the first passage toward the downstream side in the flow direction of the fluid, wherein the gas-liquid two-phase fluid is in a thermodynamic non-equilibrium state at the first outlet, and the fluid reaches a critical state at the second passage, The ejector according to any one of items 1 to 7, further comprising a velocity adjusting section (50) that adjusts the velocity of the fluid flowing into the nozzle inlet without changing the cross-sectional area of ​​the nozzle passage, wherein the fluid is in a thermodynamic equilibrium state in either the second passage or the mixing section.(Item 9) An ejector (14) comprising: a compressor (11) that compresses and discharges a refrigerant; a heat dissipation section (12) that dissipates heat from the refrigerant discharged from the compressor; an evaporation section (16b, 16c) that evaporates the refrigerant; a nozzle section (30, 301) that depressurizes the refrigerant flowing out from the heat dissipation section (12); a suction port (41) that sucks the refrigerant flowing out from the evaporation section (16b, 16c); a mixing section (42) that mixes the injected refrigerant sprayed from the nozzle section with the suction refrigerant sucked through the suction port; and a body section (40) that has a refrigerant outlet (44) that discharges the mixed refrigerant mixed in the mixing section to a suction port side of the compressor, wherein the nozzle section has a nozzle passage (32) formed therein, the nozzle passage extending from a nozzle inlet (32a) where decompression of the refrigerant begins to a spray port (32e) from which the refrigerant is sprayed. The nozzle passage includes a first passage (321) that reduces the passage cross-sectional area from the nozzle inlet toward the downstream side in the flow direction of the refrigerant, and a second passage (322) that increases the passage cross-sectional area from a first outlet (32b) of the first passage toward the downstream side in the flow direction of the refrigerant, wherein the first passage generates bubbles in the liquid phase refrigerant, and the second passage causes the refrigerant to reach a critical state. (Item 10) The nozzle passage includes a third passage (323) that reduces the passage cross-sectional area from a second outlet (32c) of the second passage toward the downstream side in the flow direction of the refrigerant, and a fourth passage (324) that increases the passage cross-sectional area from a third outlet (32d) of the third passage toward the downstream side in the flow direction of the refrigerant, wherein the refrigerant reaches a critical state at the third outlet.(Item 11) An ejector (14) comprising: a compressor (11) that compresses and discharges a refrigerant; a heat dissipation section (12) that dissipates heat from the refrigerant discharged from the compressor; an evaporation section (16b, 16c) that evaporates the refrigerant; a nozzle section (30, 301) that decompresses and sprays the refrigerant flowing out of the heat dissipation section (12), and a suction port (41) that sucks the refrigerant flowing out of the evaporation section (16b, 16c), a mixing section (42) that mixes the sprayed refrigerant sprayed from the nozzle section with the suction refrigerant sucked through the suction port, and a refrigerant outlet (44) that discharges the mixed refrigerant mixed in the mixing section to a suction port side of the compressor; and a nozzle passage (32) that extends from a nozzle inlet (32a) where decompression of the refrigerant begins to a spray port (32e) from which the refrigerant is sprayed in the nozzle section. The nozzle passage includes a first passage (321) that reduces the passage cross-sectional area from the nozzle inlet toward the downstream side in the flow direction of the refrigerant, and a second passage (322) that increases the passage cross-sectional area from a first outlet (32b) of the first passage toward the downstream side in the flow direction of the refrigerant, wherein the refrigerant in a gas-liquid two-phase state is in a thermodynamic non-equilibrium state at the first outlet, the refrigerant reaches a critical state in the second passage, and the refrigerant reaches a thermodynamic equilibrium state in either the second passage or the mixing section. (Item 12) The ejector refrigeration cycle according to any one of Items 9 to 11, further comprising: a velocity adjusting unit (50) that adjusts the velocity of the refrigerant flowing into the nozzle inlet without changing the passage cross-sectional area within the nozzle passage. (Item 13) The ejector refrigeration cycle according to any one of Items 9 to 12, further comprising: a pressure adjusting unit (18) that adjusts the pressure of the refrigerant flowing into the nozzle inlet.(Item 14) A method for operating a nozzle device comprising: a nozzle section (30, 301) for decompressing a fluid; a nozzle passage (32) formed in the nozzle section from a nozzle inlet (32a) where decompression of the fluid begins to an injection port (32e) from which the fluid is injected; a first passage (321) for reducing a passage cross-sectional area from the nozzle inlet toward a downstream side in a flow direction of the fluid; and a second passage (322) for increasing a passage cross-sectional area from a first outlet (32b) of the first passage toward a downstream side in the flow direction of the fluid; the method comprises generating bubbles in the fluid in a liquid phase in the first passage; bringing the fluid into a critical state in the second passage; and changing a position in the second passage where the fluid reaches a critical state by adjusting a speed of the fluid flowing into the nozzle inlet. (Item 15) The method of operating a nozzle device according to Item 14, wherein the nozzle passage includes a third passage (323) that reduces a passage cross-sectional area from a second outlet (32c) of the second passage toward a downstream side in a flow direction of the fluid, and a fourth passage (324) that increases a passage cross-sectional area from a third outlet (32d) of the third passage toward a downstream side in a flow direction of the fluid, and the fluid is brought into a critical state at the third outlet. (Item 16) A method for operating a nozzle device comprising: a nozzle section (30, 301) for decompressing a fluid; a nozzle passage (32) formed in the nozzle section from a nozzle inlet (32a) where decompression of the fluid begins to an injection port (32e) from which the fluid is injected; a first passage (321) for reducing a passage cross-sectional area from the nozzle inlet toward a downstream side in a flow direction of the fluid; and a second passage (322) for increasing a passage cross-sectional area from a first outlet (32b) of the first passage toward a downstream side in the flow direction of the fluid; the method comprises bringing the gas-liquid two-phase fluid, which has reached a thermodynamic non-equilibrium state in the first passage, into a critical state and a thermodynamic equilibrium state in the second passage; and changing the position in the second passage where the fluid is in a thermodynamic equilibrium state by adjusting the velocity of the fluid flowing into the nozzle inlet.(Item 17) A method for operating a nozzle device according to any one of items 14 to 16, further comprising a speed adjusting unit (50) that adjusts the speed of the fluid flowing into the nozzle inlet without changing the cross-sectional area of ​​the nozzle passage. (Item 18) A method for operating a nozzle device according to any one of items 14 to 17, further comprising a pressure adjusting unit (18) that adjusts the pressure of the fluid flowing into the nozzle inlet.

[0321] Although the present disclosure has been described with reference to the embodiments, it is understood that the present disclosure is not limited to the embodiments or structures. The present disclosure also encompasses various modifications and equivalent modifications. In addition, various combinations and forms, including only one element, more than one element, or less than one element, are also within the scope and spirit of the present disclosure.

Claims

1. An ejector comprising: a nozzle portion (30, 301) for reducing the pressure of a driving side fluid and injecting it; a suction port (41) for sucking in a suction side fluid; a mixing portion (42) for mixing a jet fluid jetted from the nozzle portion with a suction fluid sucked from the suction port; and a body portion (40) in which a refrigerant outlet (44) for discharging the mixed fluid mixed in the mixing portion is formed; the nozzle portion is formed with a nozzle passage (32) extending from a nozzle inlet (32a) where the pressure reduction of the fluid begins to a jet port (32e) from which the fluid is jetted; the nozzle passage includes a first passage (321) for reducing the passage cross-sectional area from the nozzle inlet toward the downstream side in the flow direction of the fluid, and a second passage (322) for increasing the passage cross-sectional area from a first outlet (32b) of the first passage toward the downstream side in the flow direction of the fluid; and the first passage generates bubbles in the fluid in the liquid phase, and the second passage causes the fluid to reach a critical state.

2. The velocity of the fluid in the second passage is defined as the momentum frozen phase change equilibrium sound velocity a ｆｅ The ejector according to claim 1, wherein the momentum frozen phase change equilibrium sound velocity a ｆｅ is defined by the following formula (1). ν Ｇ is the specific volume of the gas phase fluid. ν Ｌ is the specific volume of the liquid phase fluid. ν is the average specific volume of the gas-liquid two-phase fluid. S Ｇ is the specific entropy of the gas phase refrigerant. Ｌ is the specific entropy of the liquid phase refrigerant; T is the temperature of the fluid; x is the axial distance.

3. The ejector according to claim 1, wherein the first passage is formed with a constant area portion (321b) in which the cross-sectional area of ​​the passage is constant.

4. An ejector as described in claim 1, comprising, as the nozzle passages, a third passage (323) which reduces the passage cross-sectional area from the second outlet (32c) of the second passage toward the downstream side in the flow direction of the fluid, and a fourth passage (324) which increases the passage cross-sectional area from the third outlet (32d) of the third passage toward the downstream side in the flow direction of the fluid, and in which the fluid at the third outlet reaches a critical state.

5. The velocity of the fluid at the third outlet is equal to the momentum equilibrium phase change equilibrium sound speed a ｅｅ The ejector according to claim 4, wherein the momentum equilibrium phase change equilibrium sound velocity a ｅｅ is defined by the following formula (2). ν Ｇ is the specific volume of the gas phase fluid. ν Ｌ is the specific volume of the liquid phase fluid. ν is the average specific volume of the gas-liquid two-phase fluid. S Ｇ is the specific entropy of the gas phase refrigerant. Ｌ is the specific entropy of the liquid phase refrigerant; T is the temperature of the fluid; x is the axial distance.

6. The ejector according to claim 4, wherein the velocity of said fluid that has reached a critical state at said third outlet is slower than the velocity of said fluid that has reached a critical state at said second passage.

7. A refrigerant compressor comprising: a nozzle portion (30, 301) for decompressing and spraying a driving side fluid; a suction port (41) for sucking a suction side fluid; a mixing portion (42) for mixing a spray fluid sprayed from the nozzle portion with a suction fluid sucked from the suction port; and a body portion (40) in which a refrigerant outlet (44) for discharging the mixed fluid mixed in the mixing portion is formed; the nozzle portion is formed with a nozzle passage (32) extending from a nozzle inlet (32a) where the decompression of the fluid begins to a spray port (32e) from which the fluid is sprayed; the nozzle passage is formed with a first passage (321) for reducing the passage cross-sectional area from the nozzle inlet toward the downstream side in the flow direction of the fluid, and a second passage (322) for increasing the passage cross-sectional area from a first outlet (32b) of the first passage toward the downstream side in the flow direction of the fluid; the first outlet is a state in which the gas-liquid two-phase fluid is in a thermodynamic non-equilibrium state; the second passage is a state in which the fluid reaches a critical state; An ejector in which the fluid is in a thermodynamic equilibrium state in either the second passage or the mixing section.

8. An ejector according to any one of claims 1 to 7, further comprising a velocity adjusting section (50) for adjusting the velocity of the fluid flowing into the nozzle inlet without changing the cross-sectional area of ​​the nozzle passage.

9. An ejector (14) comprising: a compressor (11) for compressing and discharging a refrigerant; a heat dissipation section (12) for dissipating heat from the refrigerant discharged from the compressor; an evaporation section (16b, 16c) for evaporating the refrigerant; a nozzle section (30, 301) for decompressing and spraying the refrigerant flowing out from the heat dissipation section (12); and a suction port (41) for sucking in the refrigerant flowing out from the evaporation section (16b, 16c), a mixing section (42) for mixing the sprayed refrigerant sprayed from the nozzle section with the suction refrigerant sucked from the suction port, and a refrigerant outlet (44) for discharging the mixed refrigerant mixed in the mixing section to the suction port side of the compressor, wherein a nozzle passage (32) is formed in the nozzle section from a nozzle inlet (32a) where the decompression of the refrigerant begins to a spray port (32e) from which the refrigerant is sprayed, The nozzle passage includes a first passage (321) that reduces a passage cross-sectional area from the nozzle inlet toward a downstream side in the flow direction of the refrigerant, and a second passage (322) that increases a passage cross-sectional area from a first outlet (32b) of the first passage toward a downstream side in the flow direction of the refrigerant, and an ejector-type refrigeration cycle in which bubbles are generated in the refrigerant in a liquid phase in the first passage, and the refrigerant reaches a critical state in the second passage.

10. An ejector-type refrigeration cycle as described in claim 9, comprising, as the nozzle passage, a third passage (323) which reduces the passage cross-sectional area from the second outlet (32c) of the second passage toward the downstream side in the flow direction of the refrigerant, and a fourth passage (324) which increases the passage cross-sectional area from the third outlet (32d) of the third passage toward the downstream side in the flow direction of the refrigerant, and in which the refrigerant reaches a critical state at the third outlet.

11. An ejector (14) comprising: a compressor (11) for compressing and discharging a refrigerant; a heat dissipation section (12) for dissipating heat from the refrigerant discharged from the compressor; an evaporation section (16b, 16c) for evaporating the refrigerant; a nozzle section (30, 301) for reducing the pressure of the refrigerant flowing out from the heat dissipation section (12) and spraying it; and a suction port (41) for sucking in the refrigerant flowing out from the evaporation section (16b, 16c), a mixing section (42) for mixing the sprayed refrigerant sprayed from the nozzle section with the suction refrigerant sucked from the suction port, and a refrigerant outlet (44) for discharging the mixed refrigerant mixed in the mixing section to the suction port side of the compressor, wherein a nozzle passage (32) is formed in the nozzle section from a nozzle inlet (32a) where the pressure reduction of the refrigerant begins to a spray port (32e) from which the refrigerant is sprayed, The nozzle passage includes a first passage (321) that reduces a passage cross-sectional area from the nozzle inlet toward a downstream side in the flow direction of the refrigerant, and a second passage (322) that increases a passage cross-sectional area from a first outlet (32b) of the first passage toward a downstream side in the flow direction of the refrigerant, the refrigerant in a gas-liquid two-phase state is in a thermodynamic non-equilibrium state at the first outlet, the refrigerant reaches a critical state in the second passage, and the refrigerant is in a thermodynamic equilibrium state in either the second passage or the mixing section.

12. An ejector-type refrigeration cycle device as claimed in any one of claims 9 to 11, further comprising a speed adjusting section (50) for adjusting the speed of the refrigerant flowing into the nozzle inlet without changing the cross-sectional area of ​​the nozzle passage.

13. The ejector-type refrigeration cycle according to any one of claims 9 to 11, further comprising a pressure adjusting section (18) for adjusting the pressure of the refrigerant flowing into the nozzle inlet.

14. A method for operating a nozzle device comprising a nozzle section (30, 301) for reducing the pressure of a fluid, the nozzle section having a nozzle passage (32) formed therein extending from a nozzle inlet (32a) where reduction in pressure of the fluid begins to an injection port (32e) from which the fluid is injected, the nozzle passage being formed with a first passage (321) for reducing the passage cross-sectional area from the nozzle inlet toward the downstream side in the flow direction of the fluid, and a second passage (322) for increasing the passage cross-sectional area from a first outlet (32b) of the first passage toward the downstream side in the flow direction of the fluid, the method comprising the steps of: generating bubbles in the liquid fluid in the first passage; bringing the fluid to a critical state in the second passage; and changing the position in the second passage where the fluid reaches the critical state by adjusting the speed of the fluid flowing into the nozzle inlet.

15. A method for operating a nozzle device as described in claim 14, comprising as the nozzle passages a third passage (323) which reduces the passage cross-sectional area from the second outlet (32c) of the second passage toward the downstream side in the flow direction, and a fourth passage (324) which increases the passage cross-sectional area from the third outlet (32d) of the third passage toward the downstream side in the flow direction of the fluid, and bringing the fluid into a critical state at the third outlet.

16. A method for operating a nozzle device comprising a nozzle section (30, 301) for reducing the pressure of a fluid, the nozzle section having a nozzle passage (32) formed from a nozzle inlet (32a) where the pressure reduction of the fluid begins to an injection port (32e) from which the fluid is injected, the nozzle passage being a first passage (321) for reducing the passage cross-sectional area from the nozzle inlet toward the downstream side in the flow direction of the fluid, and a second passage (322) for increasing the passage cross-sectional area from a first outlet (32b) of the first passage toward the downstream side in the flow direction of the fluid, the method comprising: bringing the gas-liquid two-phase fluid that has become in a thermodynamic non-equilibrium state in the first passage into a critical state and a thermodynamic equilibrium state in the second passage; and changing the position at which the fluid is in a thermodynamic equilibrium state in the second passage by adjusting the speed of the fluid flowing into the nozzle inlet.

17. A method for operating a nozzle device as described in any one of claims 14 to 16, further comprising a velocity adjusting section (50) for adjusting the velocity of the fluid flowing into the nozzle inlet without changing the cross-sectional area of ​​the passage in the nozzle passage.

18. A method for operating a nozzle apparatus according to any one of claims 14 to 16, further comprising a pressure adjusting section (18) for adjusting the pressure of the fluid flowing into the nozzle inlet.

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