JP5962596B2 - Ejector refrigeration cycle - Google Patents

Description

The present invention relates to an ejector-type refrigeration cycle including an ejector.
It relates to ejectors.

  2. Description of the Related Art Conventionally, an ejector-type refrigeration cycle that is a vapor compression refrigeration cycle apparatus including an ejector is known.

  In this type of ejector-type refrigeration cycle, the refrigerant flowing out of the evaporator is sucked by the suction action of the high-speed jet refrigerant jetted from the nozzle part of the ejector, and the jet refrigerant and suction are sucked by the diffuser part (pressure booster) of the ejector. By converting the kinetic energy of the mixed refrigerant with the refrigerant into pressure energy, the mixed refrigerant is boosted and discharged to the suction side of the compressor.

  Thereby, in the ejector type refrigeration cycle, the power consumption of the compressor is reduced and the coefficient of performance of the cycle (in comparison with the normal refrigeration cycle apparatus in which the refrigerant evaporation pressure in the evaporator and the suction refrigerant pressure in the compressor are substantially equal) COP) is improved.

  Furthermore, as a specific configuration of such an ejector-type refrigeration cycle, for example, Patent Document 1 includes two evaporators, and the refrigerant that has flowed out of the evaporator having the higher refrigerant evaporation pressure flows into the nozzle portion of the ejector. A cycle configuration is disclosed in which the refrigerant flowing out of the evaporator having the lower refrigerant evaporation pressure is sucked by the suction action of the injected refrigerant.

JP 2012-149790 A

  However, according to the study by the present inventors, when the ejector-type refrigeration cycle of Patent Document 1 is actually operated, the diffuser portion of the ejector cannot exhibit the desired refrigerant pressurization performance, and is provided with an ejector. In some cases, the COP improvement effect could not be sufficiently obtained.

  Then, when the present inventors investigated the cause, in the structure which makes the gaseous-phase refrigerant | coolant which flowed out of the evaporator flow in to the nozzle part of an ejector like the ejector-type refrigeration cycle of patent document 1, (a) The reason is that the mixed refrigerant with the suction refrigerant becomes a gas-liquid two-phase refrigerant with high dryness, and (b) the vapor phase refrigerant is condensed while being decompressed in the refrigerant passage formed in the nozzle part. It was found that.

  The reason will be described below. First, (a) the reason why the diffuser part of the ejector cannot exhibit the desired refrigerant boosting performance because the mixed refrigerant of the injection refrigerant and the suction refrigerant becomes a gas-liquid two-phase refrigerant with high dryness. explain.

  The reason is that if the mixed refrigerant is a gas-liquid two-phase refrigerant having a relatively high dryness x (for example, a gas-liquid two-phase refrigerant having a dryness x of 0.8 or more), the gas-liquid two-phase refrigerant is This is because a shock wave is generated in the vicinity of the diffuser part or in the diffuser part, and the refrigerant pressure increase performance in the diffuser part of the ejector becomes unstable.

  This shock wave is generated when the flow velocity of the two-phase fluid in the gas-liquid two-phase state shifts from the two-phase sound velocity αh or higher (supersonic state) to a value lower than the two-phase sound velocity αh (subsonic state). Is.

Here, the two-phase sound velocity αh is a sound velocity of a fluid in a gas-liquid mixed state in which a gas phase fluid and a liquid phase fluid are mixed, and is defined by the following formula F1.
αh = [P / {α × (1−α) × ρl}] ^0.5 (F1)
Note that α in the formula F1 is a void ratio, and indicates a volume ratio of voids (bubbles) contained per unit volume. More specifically, the void ratio α is defined by the following formula F2.
α = x / {x + (ρg / ρl) × (1−x)} (F2)
Further, ρg in the formulas F1 and F2 is a gas phase fluid density, ρl is a liquid phase fluid density, and P is a pressure of a two-phase fluid.

  The reason why the refrigerant pressure increase performance in the diffuser portion of the ejector becomes unstable due to the shock wave will be described with reference to FIGS. 20 and 21. FIG. 20 and FIG. 21, the axial section of a typical ejector is schematically shown in the upper part, but for the sake of clarity, the same or equivalent ejector 18 described in the embodiments described later is used. The same reference numerals as those of the ejector 18 are given to the portions that perform the above functions.

  First, a gas-liquid two-phase refrigerant having a relatively low dryness x (for example, a gas-liquid two-phase refrigerant having a dryness x of 0.5 or less) is caused to flow into the nozzle portion 18a of the ejector 18. In this case, when the refrigerant expands isentropically in the nozzle portion 18a, the dryness x of the refrigerant immediately before being injected from the refrigerant injection port 18c of the nozzle portion 18a is determined by the refrigerant flowing into the nozzle portion 18a. It becomes a value lower than the dryness x.

  And the injection refrigerant | coolant injected from the refrigerant | coolant injection port 18c of the nozzle part 18a mixes with the suction | inhalation refrigerant | coolant which is in a gaseous-phase state, and raises dryness x rapidly, reducing the flow rate. As a result, as indicated by a thick broken line in FIG. 20, the two-phase sound speed αh of the mixed refrigerant of the injection refrigerant and the suction refrigerant also rises abruptly.

  As a result, when a gas-liquid two-phase refrigerant having a relatively low dryness x is flowed into the nozzle portion 18a, the mixed refrigerant has a flow velocity that is higher than the two-phase sound velocity αh immediately after being injected from the refrigerant injection port 18c. The shock wave generated when the flow velocity of the two-phase refrigerant is changed from the supersonic state to the subsonic state is generated in the vicinity of the refrigerant injection port 18c of the nozzle portion 18a. For this reason, the influence which a shock wave has on the refrigerant | coolant pressurization performance of the diffuser part 18g is small.

  Next, when a gas-liquid two-phase refrigerant having a relatively high dryness x (for example, a gas-liquid two-phase refrigerant having a dryness x of 0.8 or more) flows into the nozzle portion 18a, the nozzle portion The dryness x of the refrigerant immediately before being injected from the refrigerant injection port 18c of 18a also increases. For this reason, the degree of increase in the dryness x when the injected refrigerant is mixed with the suction refrigerant and becomes the mixed refrigerant, compared with the case where the gas-liquid two-phase refrigerant having a relatively low dryness x flows into the nozzle portion 18a. Becomes smaller.

  Accordingly, as shown by the thick broken line in FIG. 21, the degree of increase in the two-phase sonic velocity αh of the mixed refrigerant is also reduced, and the location where the mixed refrigerant has a value lower than the two-phase sonic velocity αh (location where the shock wave is generated) is the nozzle portion. It is easier to separate from the refrigerant injection port 18c than when a gas-liquid two-phase refrigerant having a relatively low dryness x flows into 18a.

  When the location where the shock wave is generated moves away from the refrigerant injection port 18c and near the inlet portion of the diffuser portion 18g or into the diffuser portion 18g, the flow velocity of the mixed refrigerant flowing in the diffuser portion 18g due to the action of the shock wave. Becomes unstable, and the refrigerant pressurization performance in the diffuser portion 18g becomes unstable.

  As a result, the diffuser portion 18g of the ejector 18 cannot exhibit the desired refrigerant pressure-increasing performance, and the COP improvement effect due to the provision of the ejector cannot be sufficiently obtained in the ejector refrigeration cycle of Patent Document 1. According to the study by the present inventors, in the ejector-type refrigeration cycle of Patent Document 1, when the dryness x of the mixed refrigerant is 0.8 or more, the refrigerant pressurization performance tends to become unstable. It has been confirmed.

  Further, according to the study by the present inventors, when the dryness x of the mixed refrigerant increases and becomes a gas-liquid two-phase refrigerant having a high dryness of 0.995 or more, the diffuser portion 18g of the ejector 18 is In addition to not being able to exhibit the desired refrigerant pressure increase performance, the flow rate of the suction refrigerant sucked from the refrigerant suction port 18d of the ejector 18 may decrease.

  The reason for this is that, in a gas-liquid two-phase refrigerant with high dryness, the shear force that the liquid-phase refrigerant in the mixed refrigerant receives from the gas-phase refrigerant increases, and the average particle size of droplets (liquid-phase refrigerant particles) in the mixed refrigerant This is because the diameter becomes small.

  The reason why the suction refrigerant flow rate of the ejector is reduced due to the average particle size of the droplets in the mixed refrigerant being reduced will be described with reference to FIGS. 22 and FIG. 23 schematically show an axial section of a general ejector as in FIGS. 20 and 21 described above.

  First, when a gas-liquid two-phase refrigerant that is not highly dry is flowed into the nozzle portion 18a of the ejector 18, the gas-phase refrigerant in the injection refrigerant is decelerated while being mixed with the suction refrigerant. On the other hand, the liquid phase refrigerant (that is, droplets) in the jet refrigerant is accelerated by the inertial force when jetted from the refrigerant jet port 18c of the nozzle portion 18a. The inertial force of the droplet is represented by an integrated value of the weight of the droplet and the velocity of the droplet at the refrigerant ejection port 18c.

  As the droplets are accelerated in this way, the pressure energy of the mixed refrigerant is converted into velocity energy, and the pressure of the mixed refrigerant is connected to the refrigerant suction port 18d as shown by the solid line in the lower graph of FIG. It is possible to lower the pressure of the refrigerant flowing out from the evaporator. Further, due to the pressure drop of the mixed refrigerant, the gas-phase refrigerant flowing out of the evaporator can be sucked.

  However, when a highly dry gas-liquid two-phase refrigerant flows into the nozzle portion 18a of the ejector 18, not only the drag received by the droplets in the mixed refrigerant from the gas-phase refrigerant increases, but also the average droplet size. Since the diameter is reduced and the weight of the droplet is reduced, the inertial force of the droplet is also reduced.

  Accordingly, when the gas-liquid two-phase refrigerant having a high degree of dryness is flowed into the nozzle portion 18a, the speed of the droplet changes substantially the same as the gas-phase refrigerant. For this reason, the droplets in the mixed refrigerant cannot be sufficiently accelerated, and the pressure of the mixed refrigerant is unlikely to decrease as shown by the solid line in the lower graph of FIG. As a result, the suction refrigerant flow rate of the ejector 18 decreases.

  Further, in the region where the mixed refrigerant becomes a gas-phase refrigerant and the refrigerant passage area of the mixing unit 18e does not change, the expansion wave generated when the injection refrigerant is injected from the refrigerant injection port 18c, the injection refrigerant, and the suction refrigerant merge. As a result of the collision of the compression waves generated in the above, a plurality of periodic shock waves called barrel shock waves as shown in FIG. 24 may be generated in the mixed refrigerant.

  Such barrel shock waves periodically change the flow rate of the mixed refrigerant from the supersonic state to the subsonic state, and from the subsonic state to the supersonic state, so that the velocity energy of the mixed refrigerant is greatly lost. End up. Therefore, the barrel shock wave causes the suction refrigerant flow rate of the ejector 18 to be greatly reduced and causes the operating sound to be generated in the ejector 18.

  FIG. 24 is an explanatory diagram for explaining the barrel shock wave, and is a schematic enlarged cross-sectional view around the refrigerant injection port 18c of the nozzle portion 18a of the ejector 18 of the related art.

  Next, (b) the gas-phase refrigerant is condensed while reducing the pressure in the refrigerant passage formed in the nozzle portion, that is, from the point d3 to the point g3 in the Mollier diagram of FIG. 3 described in an embodiment described later. As shown in the depressurization process leading to, the reason why the diffuser portion of the ejector cannot exhibit the desired refrigerant pressurization performance by depressurizing the refrigerant so as to straddle the saturated gas line at the nozzle portion will be described.

  The reason for this is that in such a decompression process, when the gas-phase refrigerant flowing through the refrigerant passage formed in the nozzle portion 18a is condensed, it becomes saturated as shown by point d25 → g25 in FIG. This is because condensation may not start immediately and the refrigerant may be supersaturated, resulting in a condensation delay.

  FIG. 25 is a Mollier diagram showing changes in the state of the refrigerant when a condensation delay occurs. The refrigerant in the same state as in FIG. 3 is assigned the same reference numeral (alphabet) as in FIG. Only the subscripts (numbers) are changed. The same applies to other Mollier diagrams.

  Here, the cause of such condensation delay will be described. In view of the intermolecular force of van der Waals force, as shown in the Mollier diagram of FIG. It can be drawn as a curve deviated from the line.

  For this reason, the refrigerant | coolant of the area | region where enthalpy decreased slightly rather than the saturated gas line will be in the metastable state which cannot condense a refrigerant | coolant, unless it lowers temperature than the refrigerant | coolant on the saturated gas line of the same pressure. Therefore, when the gas-phase refrigerant is caused to flow into the nozzle portion 18a, a condensation delay is generated in which condensation does not start until the temperature of the metastable refrigerant is lowered to some extent.

  Further, when the condensation delay occurs, the enthalpy of the injected refrigerant increases compared to the case where the refrigerant is isentropically expanded at the nozzle portion 18a (corresponding to Δhx in FIG. 25). The amount of increase in enthalpy corresponds to the amount of latent heat released as latent heat when the refrigerant flows through the refrigerant passage formed in the nozzle portion 18a. Therefore, if the amount of latent heat released increases, the nozzle portion A shock wave is generated in the flow of the refrigerant flowing through the refrigerant passage 18a.

  And the shock wave generated by releasing the latent heat of the refrigerant lowers the flow rate of the injected refrigerant, so that the pressure increase performance of the refrigerant in the diffuser portion 18g is lowered.

  In view of the above points, an object of the present invention is to sufficiently improve the coefficient of performance of an ejector-type refrigeration cycle in which an evaporator downstream refrigerant flows into a nozzle portion of an ejector.

The present invention has been devised in order to achieve the above object. In the invention according to claim 1, the compressor (11) compresses and discharges the refrigerant, and the compressor (11) discharges the refrigerant. The radiator (12) that radiates the refrigerant, the first decompression means (13) and the second decompression means (16) that decompress the refrigerant on the downstream side of the radiator (12), and the decompression by the first decompression means (13). A first evaporator (15) for evaporating the generated refrigerant, a second evaporator (17) for evaporating the refrigerant depressurized by the second depressurization means (16), and a refrigerant on the downstream side of the first evaporator (15) The refrigerant on the downstream side of the second evaporator (17) is sucked from the refrigerant suction port (18d) by the suction action of the high-speed jet refrigerant jetted from the nozzle part (18a) for reducing the pressure, and the jet refrigerant and the refrigerant suction port (18d) The pressure increasing unit (1) for increasing the pressure by mixing the suction refrigerant sucked from And a ejector (18) having a g),
Further, it is characterized by an ejector type refrigeration cycle provided with a two-phase changing means (19) for causing the refrigerant flowing out from the first evaporator (15) to flow into the nozzle part (18a) in a gas-liquid two-phase state.

  According to this, since the two-phase conversion means (19) is provided, the gas-liquid two-phase refrigerant can surely flow into the nozzle portion (18a) of the ejector (18). Therefore, it is possible to reliably suppress the occurrence of the condensation delay and the deterioration of the refrigerant pressure increase performance in the diffuser portion (18g) of the ejector (18).

Further, the gas-liquid two-phase refrigerant is caused to flow into the nozzle part (18a) of the ejector (18) and isentropically depressurized so that the jet refrigerant injected from the nozzle part (18a) is also gas-liquid two-phase refrigerant. Therefore, it is possible to suppress an increase in the dryness x of the mixed refrigerant of the injection refrigerant and the suction refrigerant. Therefore, an increase in the dryness x of the mixed refrigerant causes the refrigerant in the diffuser portion (18g) to increase. It can suppress that pressure | voltage rise performance becomes unstable and the suction | inhalation refrigerant | coolant flow rate attracted | sucked from a refrigerant | coolant suction port (18d) falls. As a result, the coefficient of performance (COP) of the ejector refrigeration cycle in which the refrigerant on the downstream side of the first evaporator (15) flows into the nozzle portion (18a) of the ejector (18) can be sufficiently improved.

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

It is a whole block diagram of the ejector-type refrigerating cycle of 1st Embodiment. It is an axial sectional view of the ejector of the first embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant at the time of operating the ejector-type refrigerating cycle of 1st Embodiment. It is a graph which shows the ejector efficiency of the ejector of 1st Embodiment. It is an axial sectional view of the ejector of the second embodiment. It is an axial sectional view of an ejector of a 3rd embodiment. It is VII-VII sectional drawing of FIG. It is a graph which shows the nozzle efficiency of the ejector of 3rd Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of 4th Embodiment. It is a Mollier diagram which shows the state of the refrigerant | coolant at the time of operating the ejector-type refrigerating cycle of 5th Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of 5th Embodiment. It is sectional drawing of the liquid storage tank of 5th Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of 6th Embodiment. It is a whole block diagram of the ejector type refrigerating cycle of 7th Embodiment. It is an axial sectional view of an ejector according to an eighth embodiment. It is axial direction sectional drawing of the ejector of 9th Embodiment. It is an axial sectional view of the ejector of the modification of 9th Embodiment. It is a whole block diagram of the ejector-type refrigerating cycle of 10th Embodiment. It is a whole block diagram of the ejector type refrigerating cycle of the modification of 10th Embodiment. It is explanatory drawing for demonstrating the location where a shock wave generate | occur | produces in an ejector at the time of driving | operating of a general ejector-type refrigerating cycle. It is explanatory drawing for demonstrating the location where a shock wave generate | occur | produces in an ejector at the time of the driving | operation from which the dryness of the refrigerant | coolant which flows in into a nozzle part becomes comparatively high. It is explanatory drawing for demonstrating the pressure change of the mixed refrigerant at the time of operation | movement of a general ejector-type refrigerating cycle. It is explanatory drawing for demonstrating the pressure change of the mixed refrigerant at the time of the driving | operation from which the dryness of the refrigerant | coolant which flows into a nozzle part becomes comparatively high. It is explanatory drawing for demonstrating a barrel shock wave. It is a Mollier diagram which shows the state of the refrigerant | coolant at the time of a condensation delay having arisen in the nozzle part of the ejector.

  In the embodiment described below, in an ejector-type refrigeration cycle configured to allow the refrigerant that has flowed out of the evaporator to flow into the nozzle portion of the ejector, the diffuser portion (pressure increase portion) of the ejector cannot exhibit the desired refrigerant pressure increase performance. Means for suppressing this will be described. Furthermore, the embodiment described below includes a form that serves as a premise and a form that serves as a reference of the invention described in the claims.

(First embodiment)
The first embodiment will be described with reference to FIGS. In the present embodiment, an ejector refrigeration cycle 10 including an ejector 18 is applied to a vehicle refrigeration cycle apparatus. Specifically, the ejector refrigeration cycle 10 has a function of cooling indoor air blown into the vehicle interior, and an internal air blown into an in-vehicle refrigerator (cool box) disposed in the vehicle interior. Serves to cool the air.

  In the ejector refrigeration cycle 10 shown in the overall configuration diagram of FIG. 1, a compressor 11 sucks refrigerant and compresses and discharges it until it becomes a high-pressure refrigerant. Specifically, the compressor 11 of the present embodiment is an electric compressor configured by housing a fixed capacity type compression mechanism and an electric motor that drives the compression mechanism in one housing.

  As this compression mechanism, various compression mechanisms such as a scroll-type compression mechanism and a vane-type compression mechanism can be employed. Further, the operation (rotation speed) of the electric motor is controlled by a control signal output from a control device to be described later, and either an AC motor or a DC motor may be adopted.

  Furthermore, the compressor 11 may be an engine-driven compressor that is driven by a rotational driving force transmitted from a vehicle traveling engine via a pulley, a belt, or the like. This type of engine-driven compressor includes a variable displacement compressor that can adjust the refrigerant discharge capacity by changing the discharge capacity, and a fixed type that adjusts the refrigerant discharge capacity by changing the operating rate of the compressor by intermittently connecting the electromagnetic clutch. A capacity type compressor or the like can be employed.

  The ejector refrigeration cycle 10 employs an HFC refrigerant (specifically, R134a) as a refrigerant, and a vapor compression subcritical refrigeration cycle in which the high-pressure side refrigerant pressure does not exceed the refrigerant critical pressure. It is composed. Furthermore, refrigeration oil for lubricating the compressor 11 is mixed in the refrigerant, and a part of the refrigeration oil circulates in the cycle together with the refrigerant.

  The refrigerant inlet side of the radiator 12 is connected to the discharge port side of the compressor 11. The radiator 12 is a heat exchanger for radiating heat by exchanging heat between the high-pressure refrigerant discharged from the compressor 11 and outside air (outside air) blown by the cooling fan 12a to dissipate the high-pressure refrigerant and cool it. The cooling fan 12a is an electric blower in which the number of rotations (the amount of blown air) is controlled by a control voltage output from the control device.

  The refrigerant outlet side of the radiator 12 is connected to the inlet side of the high-stage expansion device 13 as the first pressure reducing means. The high-stage expansion device 13 has a temperature sensing unit that detects the degree of superheat of the first evaporator 15 outlet-side refrigerant based on the temperature and pressure of the first evaporator 15 outlet-side refrigerant, and the first evaporator 15 outlet It is a temperature type expansion valve that adjusts the throttle passage area by a mechanical mechanism so that the degree of superheat of the side refrigerant falls within a predetermined reference range.

  The outlet side of the high stage side throttle device 13 is connected to the refrigerant inlet of the branching section 14 that branches the flow of the refrigerant flowing out from the high stage side throttle device 13. The branch portion 14 is configured by a three-way joint having three inflow / outflow ports, and one of the three inflow / outflow ports is a refrigerant inflow port, and the remaining two are refrigerant outflow ports. Such a three-way joint may be formed by joining pipes having different pipe diameters, or may be formed by providing a plurality of refrigerant passages in a metal block or a resin block.

  The refrigerant inlet side of the first evaporator 15 is connected to one refrigerant outlet of the branch portion 14. The first evaporator 15 evaporates the low-pressure refrigerant by exchanging heat between the low-pressure refrigerant decompressed by the high-stage expansion device 13 and the indoor blowing air blown into the vehicle interior from the first blower fan 15a. This is an endothermic heat exchanger that exerts an endothermic effect. The 1st ventilation fan 15a is an electric blower by which rotation speed (blowing air amount) is controlled by the control voltage output from a control apparatus.

  In addition, the other refrigerant outlet of the branch portion 14 is connected to the inlet side of the low stage side expansion device 16 as the second decompression means. The low-stage side throttle device 16 is a fixed throttle with a fixed throttle opening, and specifically, a nozzle, an orifice, a capillary tube, or the like can be adopted.

  The refrigerant inlet side of the second evaporator 17 is connected to the outlet side of the low stage expansion device 16. The second evaporator 17 exchanges heat between the low-pressure refrigerant decompressed by the low-stage expansion device 16 and the internal blown air circulated from the second blower fan 17a into the cool box, thereby reducing the low-pressure refrigerant. It is a heat exchanger for heat absorption which evaporates and exhibits endothermic action. The basic configuration of the second evaporator 17 is the same as that of the first evaporator 15.

  Here, since the refrigerant flowing into the second evaporator 17 is depressurized by the high stage side expansion device 13 and then further depressurized by the low stage side expansion device 16, the refrigerant evaporation pressure in the second evaporator 17. The (refrigerant evaporation temperature) is lower than the refrigerant evaporation pressure (refrigerant evaporation temperature) in the first evaporator 15. Moreover, the 2nd ventilation fan 17a is an electric blower by which rotation speed (blowing air amount) is controlled by the control voltage output from a control apparatus.

  Next, the inlet side of the nozzle portion 18 a of the ejector 18 is connected to the refrigerant outlet side of the first evaporator 15. The ejector 18 functions as a decompression unit that depressurizes the refrigerant on the downstream side of the first evaporator 15, and also circulates in the cycle by sucking (transporting) the refrigerant by the suction action of the jetted refrigerant injected at a high speed. It functions as a means (refrigerant transport means).

  The detailed configuration of the ejector 18 will be described with reference to FIG. The ejector 18 includes a nozzle portion 18a and a body portion 18b. First, the nozzle portion 18a is formed of a substantially cylindrical metal (for example, a stainless alloy) that gradually tapers in the refrigerant flow direction, and the refrigerant is formed in the refrigerant passage (throttle passage) formed therein. Is expanded under reduced pressure in an isentropic manner.

  The refrigerant passage formed in the nozzle portion 18a is provided with a throat portion (minimum passage area portion) having the smallest refrigerant passage area. Further, the refrigerant is directed from the throat portion toward the refrigerant injection port 18c for injecting the refrigerant. A divergent section where the passage area gradually increases is provided. That is, the nozzle portion 18a of the present embodiment is configured as a so-called Laval nozzle.

  Further, in the present embodiment, as the nozzle portion 18a, during the normal operation of the ejector refrigeration cycle 10, the injection refrigerant injected from the refrigerant injection port 18c is in a gas-liquid two-phase state and is further injected from the refrigerant injection port 18c. A refrigerant having a flow velocity of the immediately preceding refrigerant that is equal to or higher than the two-phase sonic speed αh (supersonic state) described in Formula F1 is employed.

  Next, the body portion 18b is formed of a substantially cylindrical metal (for example, aluminum) or resin, and functions as a fixing member for supporting and fixing the nozzle portion 18a therein, and forms an outer shell of the ejector 18. Is. More specifically, the nozzle portion 18a is fixed by press-fitting or the like so as to be housed inside the one end side in the longitudinal direction of the body portion 18b.

  Also, a refrigerant suction port 18d provided in a portion corresponding to the outer peripheral side of the nozzle portion 18a in the outer peripheral side surface of the body portion 18b so as to penetrate the inside and the outside and communicate with the refrigerant injection port 18c of the nozzle portion 18a. Is formed. The refrigerant suction port 18d is a through hole that sucks the refrigerant that has flowed out of the second evaporator 17 due to the suction action of the refrigerant injected from the refrigerant injection port 18c of the nozzle portion 18a into the ejector 18.

  Further, inside the body portion 18b, the mixing refrigerant 18e that mixes the refrigerant injected from the refrigerant injection port 18c and the suction refrigerant sucked from the refrigerant suction port 18d, and the suction refrigerant sucked from the refrigerant suction port 18d. A suction passage 18f that leads to the mixing unit 18e and a diffuser unit 18g as a boosting unit that pressurizes the mixed refrigerant mixed in the mixing unit 18e are formed.

  The suction passage 18f is formed by a space between the outer peripheral side around the tapered tip of the nozzle portion 18a and the inner peripheral side of the body portion 18b. The refrigerant passage area of the suction passage 18f is in the refrigerant flow direction. It is gradually shrinking. Thereby, the flow rate of the suction refrigerant flowing through the suction passage 18f is gradually increased, and the energy loss (mixing loss) when the suction refrigerant and the injection refrigerant are mixed in the mixing unit 18e is reduced.

  The mixing portion 18e is formed by a space in a range from the refrigerant injection port 18c of the nozzle portion 18a to the inlet portion 18h of the diffuser portion 18g in the axial section of the nozzle portion 18a in the internal space of the body portion 18b. Furthermore, the distance La in the axial direction of the nozzle portion 18a from the refrigerant injection port 18c to the inlet portion 18h in the mixing portion 18e is determined so that the flow velocity of the refrigerant flowing into the inlet portion 18h is equal to or lower than the two-phase sound velocity αh. .

Specifically, the total value of the circular opening cross-sectional area of the refrigerant injection port 18c and the annular refrigerant passage cross-sectional area of the suction passage 18f in the axially vertical cross section of the nozzle portion 18a including the refrigerant injection port 18c is converted into a circle. When the equivalent diameter is φDa, the distance La is determined so as to satisfy the following formula F3.
La / φDa ≦ 1 (F3)
In the present embodiment, specifically, La / φDa = 1 is set.

  Furthermore, the mixing portion 18e of the present embodiment is formed in a shape that gradually reduces the refrigerant passage area toward the downstream side of the refrigerant flow. More specifically, it is formed in a shape that combines a truncated cone shape that gradually reduces the refrigerant passage area toward the downstream side of the refrigerant flow and a cylindrical shape that makes the refrigerant passage area constant. Further, the refrigerant passage area of the inlet portion 18h of the diffuser portion 18g is formed to be smaller than the refrigerant passage area of the refrigerant injection port 18c.

As shown in FIG. 2, the axial length of the nozzle portion 18a of the cylindrical portion of the mixing portion 18e is Lb, and the diameter of the cylindrical portion (corresponding to the diameter of the inlet portion 18h of the diffuser portion 18g). Is set to φDb, the distance Lb is determined so as to satisfy the following formula F4.
Lb / φDb ≦ 1 (F4)
In the present embodiment, specifically, Lb / φDb = 1.

  The diffuser portion 18g is disposed so as to be continuous with the outlet of the mixing portion 18e, and is formed so that the refrigerant passage area gradually increases. Thereby, the diffuser part 18g fulfills the function of converting the velocity energy of the mixed refrigerant flowing out from the mixing unit 18e into pressure energy, that is, the function of increasing the pressure of the mixed refrigerant by decelerating the flow rate of the mixed refrigerant.

  More specifically, the wall surface shape of the inner peripheral wall surface of the body portion 18b that forms the diffuser portion 18g of the present embodiment is formed by combining a plurality of curves as shown in FIG. And since the degree of spread of the refrigerant passage cross-sectional area of the diffuser portion 18g gradually increases in the refrigerant flow direction and then decreases again, the refrigerant can be increased in an isentropic manner.

  The suction port of the compressor 11 is connected to the refrigerant outlet side of the diffuser portion 18 g of the ejector 18.

  Next, the electric control unit of this embodiment will be described. A control device (not shown) is composed of a well-known microcomputer including a CPU, ROM, RAM, etc. and its peripheral circuits, performs various calculations and processing based on a control program stored in the ROM, and is connected to the output side. The operation of various control target devices 11, 12a, 15a, 17a, etc. is controlled.

  In addition, the control device includes an inside air temperature sensor that detects the interior temperature of the vehicle, an outside air temperature sensor that detects the outside air temperature, a solar radiation sensor that detects the amount of solar radiation in the interior of the vehicle, and the air temperature (evaporator temperature) of the first evaporator 15 ), A second evaporator temperature sensor for detecting the blown air temperature (evaporator temperature) of the second evaporator 17, and an outlet side temperature sensor for detecting the temperature of the refrigerant on the outlet side of the radiator 12. Sensor groups such as an outlet side pressure sensor for detecting the pressure of the refrigerant on the outlet side of the radiator 12 and an internal temperature sensor for detecting the internal temperature of the cool box are connected, and the detection values of these sensor groups are input.

  Furthermore, an operation panel (not shown) disposed near the instrument panel in the front part of the vehicle interior is connected to the input side of the control device, and operation signals from various operation switches provided on the operation panel are input to the control device. The As various operation switches provided on the operation panel, there are provided an air conditioning operation switch for requesting air conditioning in the vehicle interior, a vehicle interior temperature setting switch for setting the vehicle interior temperature, and the like.

  Note that the control device of the present embodiment is configured integrally with control means for controlling the operation of various control target devices connected to the output side of the control device. The configuration (hardware and software) for controlling the operation constitutes the control means of each control target device. For example, in the present embodiment, the configuration (hardware and software) that controls the operation of the compressor 11 constitutes the discharge capacity control means.

  Next, the operation of the present embodiment in the above configuration will be described using the Mollier diagram of FIG. First, when the operation switch of the operation panel is turned on (ON), the control device operates the electric motor of the compressor 11, the cooling fan 12a, the first blower fan 15a, the second blower fan 17a, and the like. Thereby, the compressor 11 sucks the refrigerant, compresses it, and discharges it.

  The high-temperature and high-pressure gas-phase refrigerant discharged from the compressor 11 (point a3 in FIG. 3) flows into the radiator 12, exchanges heat with the blown air (outside air) blown from the cooling fan 12a, and dissipates heat. Condensates (point a3 → b3 in FIG. 3).

  The refrigerant that has flowed out of the radiator 12 flows into the high-stage expansion device 13 and is decompressed in an enthalpy manner (b3 point → c3 point in FIG. 3). At this time, the throttle opening degree of the high stage side expansion device 13 is adjusted so that the degree of superheat of the refrigerant on the outlet side of the first evaporator 15 (point d3 in FIG. 3) is within a predetermined range.

  The refrigerant flow depressurized by the high-stage expansion device 13 is branched by the branching section 14. One refrigerant branched in the branching section 14 flows into the first evaporator 15 and evaporates by absorbing heat from the indoor air blown by the first blower fan 15a (point c3 → d3 in FIG. 3). point). Thereby, the indoor blowing air is cooled.

  The other refrigerant branched at the branching portion 14 flows into the low-stage expansion device 16 and is further decompressed in an isenthalpy manner (point c3 → point e3 in FIG. 3). The refrigerant depressurized by the low stage side expansion device 16 flows into the second evaporator 17 and evaporates by absorbing heat from the internal blown air circulated by the second blower fan 17a (e3 in FIG. 3). Point → f3 point). As a result, the internal blown air is cooled.

  Further, the superheated gas phase refrigerant flowing out of the first evaporator 15 flows into the nozzle portion 18a of the ejector 18 and is isentropically decompressed and injected (point d3 → point g3 in FIG. 3). The refrigerant flowing out of the second evaporator 17 is sucked from the refrigerant suction port 18d of the ejector 18 by the suction action of the injection refrigerant.

  The injection refrigerant injected from the nozzle portion 18a and the suction refrigerant sucked from the refrigerant suction port 18d are mixed in the mixing portion 18e of the ejector 18 and flow into the diffuser portion 18g (g3 → h3 point in FIG. 3, f3 point → h3 point).

  In the diffuser part 18g, the velocity energy of the refrigerant is converted into pressure energy by expanding the refrigerant passage area. As a result, the pressure of the mixed refrigerant of the injection refrigerant and the suction refrigerant increases (point h3 → point i3 in FIG. 3). The refrigerant flowing out from the diffuser portion 18g is sucked into the compressor 11 and compressed again (point i3 → point a3 in FIG. 3).

  The ejector-type refrigeration cycle 10 of the present embodiment operates as described above, and can cool indoor blast air blown into the vehicle interior and internal blast air circulated into the cool box. At this time, the refrigerant evaporation pressure (refrigerant evaporation temperature) of the second evaporator 17 is lower than the refrigerant evaporation pressure (refrigerant evaporation temperature) of the first evaporator 15, so that different temperatures are provided in the vehicle interior and the interior of the cool box. Can be cooled with a belt.

  Further, in the ejector refrigeration cycle 10, since the refrigerant whose pressure has been increased by the diffuser portion 18g of the ejector 18 is sucked into the compressor 11, the power consumption of the compressor 11 is reduced and the coefficient of performance (COP) of the cycle is improved. Can be made.

  Here, as in the ejector refrigeration cycle 10 of the present embodiment, in the configuration in which the superheated gas phase refrigerant that has flowed out of the first evaporator 15 flows into the nozzle portion 18a of the ejector 18, the mixed refrigerant in the mixing portion 18e. The degree of dryness x tends to be relatively high (for example, the degree of dryness x is 0.8 or more).

  If the mixed refrigerant becomes a gas-liquid two-phase refrigerant having a relatively high dryness x as described above, the refrigerant pressure-increasing performance in the diffuser portion 18g becomes unstable as described with reference to FIGS. End up.

  On the other hand, according to the ejector 18 of the present embodiment, the axial distance La of the nozzle portion 18a from the refrigerant injection port 18c of the nozzle portion 18a to the inlet portion 18h of the diffuser portion 18g in the mixing portion 18e is the inlet portion. The flow rate of the refrigerant flowing into 18h is determined to be equal to or lower than the two-phase sound velocity αh. Thereby, the shock wave generated when the mixed refrigerant shifts from the supersonic state to the subsonic state can be generated in the mixing unit 18e.

  Therefore, it can suppress that a shock wave generate | occur | produces in the diffuser part 18g, and can suppress that the flow rate of the mixed refrigerant | coolant which distribute | circulates the diffuser part 18g by the effect | action of a shock wave becomes unstable. As a result, even in the ejector 18 that causes the refrigerant that has flowed out of the first evaporator 15 to flow into the nozzle portion 18a, the refrigerant pressure increase performance in the diffuser portion 18g can be stabilized, and the refrigerant pressure increase performance of the ejector 18 can be reduced. Can be suppressed.

  Further, by determining the distance La so as to satisfy the above formula F3, not only the shock wave generated when the mixed refrigerant shifts from the supersonic state to the subsonic state is generated in the mixing unit 18e, but also the ejector 18 It can suppress that the axial direction length of increases unnecessarily.

  Moreover, in the ejector 18 of this embodiment, the shape of the mixing part 18e is formed in a shape that gradually reduces the refrigerant passage area toward the downstream side of the refrigerant flow. Further, the refrigerant passage area of the inlet portion 18h of the diffuser portion 18g is set smaller than the refrigerant passage area of the refrigerant injection port 18c of the nozzle portion 18ac.

  Thereby, in the mixing unit 18e of the present embodiment, the flow rate of the mixed refrigerant is reduced to the two-phase sonic speed αh or less until it reaches the inlet 18h of the diffuser unit 18g so as to effectively decelerate the flow rate of the mixed refrigerant. It is trying to become.

  Further, according to the study by the present inventors, the shape of the mixing portion 18e is a combination of a truncated cone shape that gradually reduces the refrigerant passage area toward the downstream side of the refrigerant flow and a cylindrical shape that makes the refrigerant passage area constant. It has been found that the flow rate of the mixed refrigerant can be effectively decelerated by determining the distance Lb so as to satisfy the above formula F4.

  Therefore, according to the ejector 18 of the present embodiment, as shown in FIG. 4, the energy conversion efficiency (ejector efficiency ηej) in the ejector 18 can be greatly improved over the prior art. As a result, in the ejector type refrigeration cycle 10 of the present embodiment, the COP improvement effect due to the provision of the ejector 18 can be sufficiently obtained.

The ejector efficiency ηej is defined by the following formula F5.
ηej = {Δhd × (Gn + Ge)} / (Δiej × Gn) (F5)
Here, Gn is a flow rate of the refrigerant injected from the nozzle portion 18 a of the ejector 18 and is a flow rate of the refrigerant flowing through the first evaporator 15. Ge is a flow rate of the suction refrigerant sucked from the refrigerant suction port 18d of the ejector 18 and is a flow rate of the refrigerant flowing through the second evaporator 17.

  Further, Δhd is the amount of increase in enthalpy when the refrigerant is isentropically boosted in the diffuser portion 18g of the ejector 18 as shown in FIG. 3, and Δiej is the ejector 18 as shown in FIG. The amount of enthalpy is reduced when the pressure is reduced isentropically by the nozzle portion 18a.

(Second Embodiment)
This embodiment demonstrates the example which changed the structure of the ejector 18 with respect to 1st Embodiment, as shown in FIG. In FIG. 5, the same or equivalent parts as those in the first embodiment are denoted by the same reference numerals. The same applies to the following drawings.

  Specifically, in the ejector 18 of the present embodiment, a tapered portion 18i that gradually reduces the refrigerant passage area toward the refrigerant injection port 18c is formed as a refrigerant passage formed in the nozzle portion 18a. That is, the nozzle portion 18a of the present embodiment is configured as a so-called tapered nozzle. Furthermore, the injection part 18j is formed in the most downstream side of the refrigerant path formed in the nozzle part 18a of this embodiment.

  The injection unit 18j is a space that guides the refrigerant from the most downstream portion of the tapered portion 18i toward the refrigerant injection port 18c. Therefore, the spraying shape or the spreading direction of the jetted refrigerant injected from the refrigerant jet port 18c can be changed by the spreading angle θn in the axial section of the nozzle part 18a of the jetting part 18j. That is, the injection unit 18j can also be expressed as a space that defines the injection direction of the refrigerant injected from the refrigerant injection port 18c.

  Furthermore, in the present embodiment, the spread angle θn of the injection portion 18j in the axial section of the nozzle portion 18a is set to 0 °. That is, the injection part 18j of the present embodiment is formed by a cylindrical space extending in the axial direction of the nozzle part 18a and having a constant refrigerant passage area. In FIG. 5, the angle θn is illustrated as a slight value (about 1 °) in order to clarify the angle θn.

Further, as shown in FIG. 5, the axial length in which the injection portion 18j is formed in the refrigerant passage formed in the nozzle portion 18a is Lc, and the equivalent diameter of the opening area of the refrigerant injection port 18c is φDc. Sometimes, the distance Lc is determined so as to satisfy the following formula F6.
Lc / φDc ≦ 1 (F6)
In the present embodiment, specifically, Lc / φDc = 0.67. Lc / φDc = 1 may be set.

  In the nozzle part 18a of this embodiment, the refrigerant path formed inside is formed as described above, so that the refrigerant injected from the refrigerant injection port 18c to the mixing unit 18e is freely expanded.

  Other configurations and operations of the ejector 18 and the ejector refrigeration cycle 10 are the same as those in the first embodiment. Therefore, when the ejector-type refrigeration cycle 10 of this embodiment is operated, similarly to the first embodiment, the indoor blast air blown into the vehicle interior and the internal blast air circulated into the cool box are cooled. can do.

  Here, as in the ejector refrigeration cycle 10 of the present embodiment, in the configuration in which the superheated gas phase refrigerant that has flowed out of the first evaporator 15 flows into the nozzle portion 18a of the ejector 18, the refrigerant is injected from the refrigerant injection port. In some cases, the flow rate of the refrigerant immediately before the discharge tends to increase, and the diffuser portion 18g of the ejector 18 may not be able to exhibit the desired refrigerant pressure increase performance.

  More specifically, in a general ejector, the loss of kinetic energy when the refrigerant is decompressed at the nozzle portion is recovered by sucking the refrigerant from the refrigerant suction port by the suction action of the injected refrigerant. At this time, the amount of energy recovered (that is, the amount of decrease in the enthalpy indicated by Δiej in FIG. 3) increases when the pressure of the refrigerant flowing into the nozzle unit is constant. It increases with.

Moreover, the maximum value of the flow velocity V of the injection refrigerant immediately after being injected from the refrigerant injection port of the nozzle portion is expressed by the following formula F7.
V = V0 + (2 × Δiej) ^0.5 (F7)
V0 is the initial speed of the refrigerant flowing into the nozzle portion.

  Therefore, when a gas phase refrigerant having a higher enthalpy than the gas-liquid two-phase refrigerant is flowed into the nozzle portion, the flow velocity V of the injected refrigerant is likely to increase, and the wall friction between the refrigerant and the refrigerant passage formed in the nozzle portion also increases. It becomes easy.

  Further, the gas-phase refrigerant flowing at high speed in the refrigerant passage formed in the nozzle portion is condensed, and the gas-liquid two-phase refrigerant having a high gas-liquid density ratio (for example, the gas-liquid two-phase refrigerant having a gas-liquid density ratio of 200 or more). If it becomes, the wall surface friction of a refrigerant | coolant and a refrigerant channel will increase greatly, and the loss of the kinetic energy which a refrigerant | coolant has will be caused. Such loss of kinetic energy lowers the flow rate of the injected refrigerant and lowers the refrigerant pressurization performance in the diffuser section.

  On the other hand, according to the ejector 18 of this embodiment, the injection part 18j is provided in the nozzle part 18a comprised as a taper nozzle, and the mixed refrigerant | coolant injected from the refrigerant | coolant injection port 18c to the mixing part 18e is free. Since the expansion is performed, it is possible to accelerate the injected refrigerant in the mixing unit 18e without providing a divergent portion unlike a Laval nozzle. That is, the refrigerant can be accelerated without causing wall friction between the refrigerant and the refrigerant passage that occurs when the refrigerant is supersonically accelerated at the divergent portion of the Laval nozzle.

  Therefore, the wall friction between the refrigerant and the refrigerant passage can be reduced, the loss of kinetic energy of the refrigerant flowing through the refrigerant passage can be suppressed, and the flow rate of the injected refrigerant can be suppressed from decreasing. As a result, even in the ejector 18 that causes the refrigerant that has flowed out of the first evaporator 15 to flow into the nozzle portion 18a, the energy loss of the refrigerant in the nozzle portion 18a can be reduced, and a decrease in the refrigerant pressure increase performance of the ejector 18 can be suppressed. .

  Further, according to the ejector 18 of the present embodiment, similarly to the first embodiment, the refrigerant pressure increasing performance in the diffuser portion 18g can be stabilized, and the ejector efficiency ηej in the ejector 18 can be improved. Therefore, in the ejector type refrigeration cycle 10 of the present embodiment, the COP improvement effect due to the provision of the ejector 18 can be sufficiently obtained.

  In this embodiment, the example in which the expansion angle θn of the injection portion 18j in the axial section of the nozzle portion 18a is 0 ° has been described. However, if the refrigerant injected from the refrigerant injection port 18c can be freely expanded, The spread angle θn may be set larger than 0 °. That is, the injection unit 18j may be formed by a truncated cone space in which the refrigerant passage cross-sectional area gradually increases in the refrigerant flow direction.

(Third embodiment)
This embodiment demonstrates the example which changed the structure of the ejector 18 with respect to 1st Embodiment, as shown in FIG. 6, FIG. Specifically, in the ejector 18 of the present embodiment, the refrigerant that has flowed in from the refrigerant inlet 18l is disposed upstream of the throat portion (minimum passage area portion) in the refrigerant passage formed in the nozzle portion 18a. A swirl space 18k that swirls around the axis of the nozzle portion 18a is provided.

  More specifically, the swirling space 18k is formed in a cylindrical portion 18m provided on the upstream side of the refrigerant flow of the nozzle portion 18a. Accordingly, the cylindrical portion 18m constitutes the swirling space forming member described in the claims, and in this embodiment, the swirling space forming member and the nozzle portion are integrally formed. .

  The swirling space 18k is formed in a rotating body shape, and its central axis extends coaxially with the nozzle portion 18a. The rotating body shape is a three-dimensional shape formed when a plane figure is rotated around one straight line (central axis) on the same plane. More specifically, the swirl space 18k of the present embodiment is formed in a substantially cylindrical shape.

  Further, the refrigerant inflow passage 18n that connects the refrigerant inlet 18l and the swirl space 18k is, as seen from the central axis direction of the swirl space 18k, in the tangential direction of the inner wall surface of the swirl space 18k, as shown in FIG. It extends. As a result, the refrigerant flowing into the swirl space 18k from the refrigerant inlet 18l flows along the inner wall surface of the swirl space 18k and swirls in the swirl space 18k.

  Here, since centrifugal force acts on the refrigerant swirling in the swirling space 18k, the refrigerant pressure on the central axis side is lower than the refrigerant pressure on the outer peripheral side in the swirling space 18k. Therefore, in the present embodiment, during normal operation, the refrigerant on the central axis side in the swirl space 18k is closer to the gas-liquid two-phase side than the saturated gas line, that is, the refrigerant on the central axis side in the swirl space 18k is The pressure of the refrigerant on the central axis side in the swirling space 18k is reduced so as to start the condensation.

  Such adjustment of the refrigerant pressure on the central axis side in the swirling space 18k can be realized by adjusting the swirling flow velocity of the refrigerant swirling in the swirling space 18k. Further, the swirling flow velocity is adjusted by adjusting the ratio of the flow passage cross-sectional area between the passage cross-sectional area of the refrigerant inflow passage 18n and the axial cross-sectional area of the swirling space 18k, or on the upstream side of the nozzle portion 18a. This can be done by adjusting the aperture of the high-stage expansion device 13 arranged.

  Other configurations and operations of the ejector 18 and the ejector refrigeration cycle 10 are the same as those in the first embodiment. Therefore, when the ejector-type refrigeration cycle 10 of this embodiment is operated, similarly to the first embodiment, the indoor blast air blown into the vehicle interior and the internal blast air circulated into the cool box are cooled. can do.

  Here, as in the ejector refrigeration cycle 10 of the present embodiment, in the configuration in which the superheated gas phase refrigerant flowing out from the first evaporator 15 flows into the nozzle portion 18a of the ejector 18, the ejector 18 as described above. The refrigerant is condensed and accelerated while reducing the pressure in the refrigerant passage formed in the nozzle portion 18a.

  In such an ejector 18, as described above, not only energy loss occurs due to wall friction between the refrigerant and the refrigerant passage, but also when the gas-phase refrigerant flowing in the refrigerant passage formed in the nozzle portion 18 a is condensed. Condensation delay may occur, and the pressure increase performance of the refrigerant in the diffuser portion 18g is deteriorated.

  On the other hand, in the ejector 18 of the present embodiment, the refrigerant is swirled in the swirling space 18k, whereby the refrigerant on the swiveling center axis side in the swirling space 18k is depressurized to start condensation, and condensed nuclei are generated. The gas-liquid two-phase refrigerant can flow into the nozzle portion 18a. Therefore, it is possible to suppress the occurrence of a condensation delay in the refrigerant at the nozzle portion 18a.

  As a result, as shown in FIG. 8, the nozzle efficiency ηnoz in the nozzle portion 18a can be greatly improved as compared with the prior art, and the refrigerant is condensed while reducing the pressure in the refrigerant passage formed in the nozzle portion 18a. Even if the ejector 18 is accelerated, it is possible to suppress a decrease in the refrigerant pressure increase performance in the diffuser portion 18g. The nozzle efficiency ηnoz is energy conversion efficiency when the pressure energy of the refrigerant is converted into kinetic energy in the nozzle portion 18a.

  Further, according to the ejector 18 of the present embodiment, similarly to the first embodiment, the refrigerant pressure increasing performance in the diffuser portion 18g can be stabilized, and the ejector efficiency ηej in the ejector 18 can be improved. Therefore, in the ejector type refrigeration cycle 10 of the present embodiment, the COP improvement effect due to the provision of the ejector 18 can be sufficiently obtained.

  Further, according to the ejector 18 of the present embodiment, even when the refrigerant flowing into the swirling space 18k is a gas-liquid two-phase refrigerant, the refrigerant pressure on the center side in the swirling space 18k is reduced to reduce the nozzle. Since the boiling of the refrigerant flowing into the throat portion (minimum passage area portion) of the portion 18a can be promoted, the nozzle efficiency ηnoz can be improved.

(Fourth embodiment)
This embodiment demonstrates the example which changed the structure of the ejector-type refrigerating cycle with respect to 1st Embodiment.

  Specifically, in the ejector refrigeration cycle 10a of the present embodiment, as shown in FIG. 9, the branch portion 14 is disposed on the outlet side of the radiator 12, and one of the refrigerants branched at the branch portion 14 is made high. The pressure is reduced until it becomes a low-pressure refrigerant by the stage side expansion device 13, and flows into the refrigerant inlet side of the first evaporator 15. Further, the other refrigerant branched by the branching section 14 is decompressed by the low stage side expansion device 16 until it becomes a low-pressure refrigerant, and flows into the refrigerant inlet side of the second evaporator 17.

  Furthermore, in this embodiment, the throttle opening of the low stage side throttle device 16 is set to be smaller than the throttle opening of the high stage side throttle device 13, which is lower than the pressure reduction amount in the high stage side throttle device 13. The amount of pressure reduction in the expansion device 16 is large. Therefore, the refrigerant evaporation pressure (refrigerant evaporation temperature) in the second evaporator 17 is lower than the refrigerant evaporation pressure (refrigerant evaporation temperature) in the first evaporator 15. Other configurations are the same as those of the first embodiment.

  Therefore, when the ejector refrigeration cycle 10a of this embodiment is operated, as shown in the Mollier diagram of FIG. 10, the high-temperature and high-pressure gas-phase refrigerant discharged from the compressor 11 (point a10 in FIG. 10) As in the first embodiment, heat is dissipated by the radiator 12 and condensed (a10 point → b10 point in FIG. 10).

  The flow of the refrigerant flowing out of the radiator 12 is branched at the branching section 14. One refrigerant branched by the branching section 14 is decompressed by the high stage side expansion device 13 (b10 point → c10 point in FIG. 10) and flows into the first evaporator 15. The other refrigerant branched by the branching section 14 is decompressed by the low-stage expansion device 16 (b10 point → e10 point in FIG. 10) and flows into the second evaporator 17. Subsequent operations are the same as those in the first embodiment.

  Therefore, when the ejector-type refrigeration cycle 10a of this embodiment is operated, the indoor blast air blown into the vehicle interior and the internal blast air circulated into the cool box are cooled as in the first embodiment. can do.

  Furthermore, also in the ejector type refrigeration cycle 10a of the present embodiment, the ejector 18 exhibits the same effect as that of the first embodiment, so that the COP improvement effect due to the provision of the ejector 18 can be sufficiently obtained. Moreover, you may apply the ejector 18 disclosed by 2nd, 3rd, 8th, 9th embodiment to the ejector-type refrigerating cycle 10a of this embodiment.

(Fifth embodiment)
This embodiment demonstrates the example which changed the structure of the ejector-type refrigerating cycle with respect to 1st Embodiment.

  Specifically, in the ejector refrigeration cycle 10b of the present embodiment, as shown in FIG. 11, a fixed throttle with a fixed throttle opening is adopted as the high stage side throttle device 13, and the low stage side throttle device 16 is used. A temperature expansion valve is used. Further, a liquid storage tank (liquid storage part) 19 for storing excess refrigerant in the cycle is disposed between the refrigerant outlet side of the first evaporator 15 and the inlet side of the nozzle part 18 a of the ejector 18.

  The detailed configuration of the liquid storage tank 19 will be described with reference to FIG. In addition, the up and down arrows in FIG. 12 indicate the up and down directions in a state where the liquid storage tank 19 is mounted on the vehicle.

  The liquid storage tank 19 includes a main body 19a formed of a cylindrical member extending in the vertical direction and closed at both ends, a refrigerant inflow port 19b for allowing the refrigerant flowing out of the first evaporator 15 to flow into the main body 19a, A refrigerant outflow port 19c for allowing the gas-liquid two-phase refrigerant to flow out from the main body 19a to the nozzle 18a side of the ejector 18 is provided.

  The refrigerant inflow port 19b is connected to the cylindrical side surface of the main body portion 19a and is configured by a refrigerant pipe extending in the tangential direction of the cylindrical side surface of the main body portion 19a. The refrigerant outflow port 19c is connected to the lower end surface (bottom surface) in the axial direction of the main body 19a, and is constituted by a refrigerant pipe extending coaxially with the main body 19a over the inside and outside of the main body 19a.

  Furthermore, the upper end part of the refrigerant | coolant outflow port 19c is extended to the upper side rather than the connection site | part of the refrigerant | coolant inflow port 19b. Further, a liquid phase refrigerant introduction hole 19d through which the liquid phase refrigerant stored in the main body portion 19a flows into the refrigerant outflow port 19c is formed below the refrigerant outflow port 19c.

  Accordingly, under the operating conditions in which the flow rate of the circulating refrigerant circulating in the cycle decreases and the gas-liquid two-phase refrigerant flows out from the first evaporator 15, the refrigerant flowing into the main body 19a from the refrigerant inflow port 19b is the cylinder of the main body 19a. The refrigerant gas and liquid are separated by the action of centrifugal force generated by the swirling flow.

  The separated liquid phase refrigerant falls downward due to the action of gravity, and is stored in the main body 19a as an excess refrigerant. On the other hand, the separated gas-phase refrigerant is mixed with the liquid-phase refrigerant flowing into the refrigerant outflow port 19c from the liquid-phase refrigerant introduction hole 19d when flowing out to the inlet side of the nozzle portion 18a via the refrigerant outflow port 19c. And flows out as a gas-liquid two-phase refrigerant.

  Also, under the operating conditions in which the flow rate of the circulating refrigerant circulating in the cycle increases and the gas phase refrigerant flows out from the first evaporator 15, the gas phase refrigerant flowing in from the refrigerant inflow port 19b is not separated into gas and liquid, and the refrigerant flows out. It flows out to the inlet side of the nozzle part 18a through the port 19c. At this time, the gas-phase refrigerant that has flowed into the refrigerant outflow port 19c is mixed with the liquid-phase refrigerant that has flowed into the refrigerant outflow port 19c from the liquid-phase refrigerant introduction hole 19d and flows out as a gas-liquid two-phase refrigerant.

  That is, the liquid storage tank 19 of the present embodiment constitutes a two-phase unit that causes the refrigerant flowing out from the first evaporator 15 to flow out into the gas-liquid two-phase state to the inlet side of the nozzle portion 18a. More specifically, the liquid storage tank 19 mixes the liquid-phase refrigerant stored in the main body 19a and the refrigerant that has flowed out of the first evaporator 15, and flows it out to the inlet side of the nozzle portion 18a.

  Other configurations and operations of the ejector 18 and the ejector refrigeration cycle 10b are the same as those in the first embodiment. Therefore, when the ejector-type refrigeration cycle 10 of this embodiment is operated, similarly to the first embodiment, the indoor blast air blown into the vehicle interior and the internal blast air circulated into the cool box are cooled. can do.

  Here, in the ejector-type refrigeration cycle configured to flow the gas-phase refrigerant into the nozzle portion 18a of the ejector 18, the dryness x of the mixed refrigerant obtained by mixing the injection refrigerant and the suction refrigerant in the mixing section 18e is also a relatively high value. (For example, the dryness x is 0.8 or more).

  In such an ejector-type refrigeration cycle, as described with reference to FIG. 25, a condensation delay may occur, and the refrigerant pressure increase performance in the diffuser portion 18g may be reduced. Moreover, as demonstrated using FIG. 20, FIG. 21, the refrigerant | coolant pressurization performance in the diffuser part 18g may become unstable.

  Furthermore, when the dryness x of the mixed refrigerant increases to become a gas-liquid two-phase refrigerant having a high dryness of 0.995 or more, as described with reference to FIGS. According to the above examination, the diffuser portion 18g of the ejector 18 may not be able to exhibit the desired refrigerant boosting performance, and the flow rate of the suctioned refrigerant sucked from the refrigerant suction port 18d of the ejector 18 may decrease.

  On the other hand, since the ejector refrigeration cycle 10b of the present embodiment includes the liquid storage tank 19 as the two-phase conversion means, the gas-liquid two-phase refrigerant can surely flow into the nozzle portion 18a of the ejector 18. Can do. Therefore, it is possible to reliably suppress the occurrence of the condensation delay.

  Furthermore, by causing the gas-liquid two-phase refrigerant to flow into the nozzle portion 18a and reducing the pressure in an isentropic manner, the injected refrigerant that is injected from the refrigerant injection port 18c is also surely converted into the gas-liquid two-phase refrigerant. It can suppress that dryness x raises. Therefore, it is possible to suppress the refrigerant pressure increase performance in the diffuser portion 18g from becoming unstable and the suction refrigerant flow rate sucked from the refrigerant suction port 18d from being lowered.

  In addition, since the dryness x of the injected refrigerant can be reduced to reduce the two-phase sonic speed αh of the mixed refrigerant, it occurs when the flow speed of the two-phase refrigerant changes from the supersonic state to the subsonic state. The shock wave can be a gas-weak shock wave. Therefore, it is possible to effectively suppress the refrigerant pressure increase performance in the diffuser portion 18g from becoming unstable.

  As a result, according to the ejector refrigeration cycle 10 of the present embodiment, the COP can be sufficiently improved even in the cycle configuration in which the refrigerant on the downstream side of the first evaporator 15 flows into the nozzle portion 18a of the ejector 18. .

  Further, in this embodiment, since the two-phase unit is constituted by the liquid storage tank 19, the gas liquid is reliably supplied to the nozzle portion 18a of the ejector 18 with a very simple configuration without causing a complicated cycle configuration. Two-phase refrigerant can be introduced.

  Further, in the ejector type refrigeration cycle 10 of the present embodiment, a temperature type expansion valve that is a variable throttle mechanism is adopted as the low stage side throttle device 16, and the refrigerant flowing out of the second evaporator 17 is within a predetermined reference range. It is trying to become. In other words, the throttle opening degree of the low stage side expansion device 16 of the present embodiment is adjusted so that the refrigerant flowing out from the second evaporator 17 is equal to or less than a predetermined reference superheat degree.

  Therefore, by appropriately setting the reference superheat degree, the mixed refrigerant obtained by mixing the jet refrigerant that is in the gas-liquid two-phase state and the suction refrigerant that is in the gas phase state below the reference superheat degree is dried. It is possible to reliably suppress the degree x from increasing. Furthermore, the throttle opening degree of the low stage side expansion device 16 may be adjusted so that the refrigerant flowing out from the second evaporator 17 becomes a saturated gas phase refrigerant or a gas-liquid two-phase refrigerant.

  Further, according to the ejector 18 of the present embodiment, similarly to the first embodiment, the refrigerant pressure increasing performance in the diffuser portion 18g can be stabilized, and the ejector efficiency ηej in the ejector 18 can be improved. As a result, according to the ejector refrigeration cycle 10b of the present embodiment, the COP improvement effect by providing the ejector 18 can be sufficiently obtained.

  Moreover, you may apply the ejector 18 disclosed by 2nd, 3rd, 8th, 9th embodiment to the ejector-type refrigerating cycle 10a of this embodiment.

(Sixth embodiment)
In the present embodiment, an example in which the configuration of the ejector refrigeration cycle is changed as shown in FIG. 13 with respect to the fifth embodiment will be described.

  Specifically, in the ejector type refrigeration cycle 10b of the present embodiment, a discharge refrigerant passage 20a that guides the gas-phase refrigerant discharged from the compressor 11 into the liquid storage tank 19 is added. It is desirable that a throttle means for preventing the refrigerant pressure in the liquid storage tank 19 from increasing is provided in the discharge refrigerant passage 20a. Therefore, in the present embodiment, the discharge refrigerant passage 20a is constituted by a capillary tube.

  Therefore, the liquid storage tank 19 which is the two-phase unit of the present embodiment mixes the liquid phase refrigerant stored in the liquid storage tank 19 with the gas phase refrigerant discharged from the compressor 11 to thereby form the nozzle 18a. It is configured to flow out to the inlet side. Other configurations and operations are the same as those of the fifth embodiment. Even if the biphasic means is configured as in this embodiment, the same effect as in the fifth embodiment can be obtained.

  Moreover, you may apply the ejector 18 disclosed by 2nd, 3rd, 8th, 9th embodiment to the ejector-type refrigerating cycle 10a of this embodiment.

(Seventh embodiment)
In the present embodiment, an example in which the configuration of the ejector refrigeration cycle is changed as shown in FIG. 14 with respect to the fifth embodiment will be described.

  Specifically, in the ejector refrigeration cycle 10b of the present embodiment, a condensed refrigerant passage 20b that guides the liquid-phase refrigerant that has flowed out of the radiator 12 into the liquid storage tank 19 is added. In the condensed refrigerant passage 20b, it is desirable to provide a throttle means for preventing the refrigerant pressure in the liquid storage tank 19 from increasing. Therefore, in the present embodiment, the condensed refrigerant passage 20b is constituted by a capillary tube.

  Therefore, the liquid storage tank 19 that is the two-phase unit of the present embodiment mixes the liquid-phase refrigerant that has flowed out of the radiator 12 and the gas-phase refrigerant that has flowed out of the first evaporator 15 to the inlet side of the nozzle portion 18a. It is configured to flow into Other configurations and operations are the same as those of the fifth embodiment. Even if the biphasic means is configured as in this embodiment, the same effect as in the fifth embodiment can be obtained.

  Moreover, you may apply the ejector 18 disclosed by 2nd, 3rd, 8th, 9th embodiment to the ejector-type refrigerating cycle 10a of this embodiment.

(Eighth embodiment)
In the present embodiment, as shown in FIG. 15, the ejector 18 of the second embodiment has a cylindrical portion 18m provided on the upstream side of the refrigerant flow of the nozzle portion 18a as in the third embodiment. A swirling space 18k for swirling the refrigerant flowing in from the refrigerant inlet 18l is provided. Other configurations and operations of the ejector 18 and the ejector refrigeration cycle 10 are the same as those in the second embodiment.

  Therefore, when the ejector type refrigeration cycle 10 of this embodiment is operated, the indoor blast air blown into the vehicle compartment and the blast air circulated into the cool box are cooled as in the second embodiment. can do.

  Further, in the ejector 18 of the present embodiment, similarly to the third embodiment, the refrigerant is swirled in the swirling space 18k so that the gas-liquid two-phase refrigerant in which condensed nuclei are generated flows into the nozzle portion 18a. As a result, the nozzle efficiency ηnoz can be improved. Therefore, it is possible to suppress a decrease in the refrigerant pressure increase performance in the diffuser portion 18g.

  Moreover, since the refrigerant | coolant injected from the refrigerant | coolant injection port 18c of the nozzle part 18a is freely expanded similarly to 2nd Embodiment, it can suppress that wall surface friction increases. Therefore, the energy loss of the refrigerant in the nozzle portion 18a can be reduced, and the decrease in the refrigerant pressure increase performance of the ejector 18 can be suppressed.

  Further, similarly to the first embodiment, the refrigerant pressure-increasing performance in the diffuser portion 18g can be stabilized, and the ejector efficiency ηej in the ejector 18 can be improved. Therefore, in the ejector type refrigeration cycle 10 of the present embodiment, the COP improvement effect due to the provision of the ejector 18 can be sufficiently obtained.

(Ninth embodiment)
In the eighth embodiment, an example in which a fixed nozzle in which the refrigerant passage area of the minimum passage area portion formed at the inlet portion of the injection portion 18j is fixed is used as the nozzle portion 18a of the ejector 18 is described. Now, as shown in FIG. 16, an example will be described in which a variable nozzle configured to change the refrigerant passage area of the minimum passage area portion is adopted.

  Specifically, the ejector 18 of the present embodiment includes a needle valve 18y as a valve body that changes the refrigerant passage area of the nozzle portion 18a, and a stepping motor 18x as a drive means that displaces the needle valve 18y. Has been.

  The needle valve 18y is formed in the shape of a needle whose central axis is arranged coaxially with the central axis of the nozzle portion 18a. More specifically, the needle valve 18y is formed in a shape that tapers toward the downstream side of the refrigerant flow, and the taper tip portion on the most downstream side is further on the downstream side of the refrigerant flow than the refrigerant injection port 18c of the nozzle portion 18a. It arrange | positions so that it may protrude toward. That is, the nozzle portion 18a of the present embodiment is configured as a so-called plug nozzle.

  The stepping motor 18x is disposed on the refrigerant inlet 18l side of the nozzle portion 18a, and displaces the needle valve 18y in the axial direction of the nozzle portion 18a. Thereby, the area of the refrigerant passage having an annular cross section formed between the inner peripheral wall surface of the nozzle portion 18a and the outer peripheral wall surface of the needle valve 18y is changed. The operation of the stepping motor 18x is controlled by a control signal output from the control device.

  Other configurations and operations of the ejector 18 and the ejector refrigeration cycle 10 are the same as those in the eighth embodiment. Therefore, also in the ejector-type refrigeration cycle 10 and the ejector 18 of the present embodiment, the same effects as in the eighth embodiment can be obtained.

  Moreover, according to the ejector 18 of this embodiment, since the nozzle part 18a is comprised as a variable nozzle, the refrigerant | coolant flow volume according to the load of the ejector type refrigeration cycle 10 can be flowed into the nozzle part 18a of the ejector 18. .

  Moreover, since the nozzle part 18a of this embodiment is comprised as a plug nozzle, an injection refrigerant | coolant can be injected to the mixing part 18e along the outer surface of the needle valve 18y from the refrigerant | coolant injection port 18c. Therefore, even if the flow rate of the refrigerant flowing into the nozzle portion 18a changes, the injected refrigerant can be easily expanded freely, and the kinetic energy of the refrigerant flowing through the refrigerant passage is reduced by reducing the wall friction between the refrigerant and the refrigerant passage. Loss can be suppressed.

  Further, as shown in FIG. 16, the needle valve 18y of the present embodiment is disposed so as to penetrate through the swirling space 18k, so that friction between the refrigerant swirling in the swirling space 18k and the inner wall of the nozzle portion 18a occurs. It is easy to produce condensed nuclei.

  In addition, in the nozzle part 18a shown in FIG. 16, the needle valve 18y employs a shape that tapers toward the downstream side of the refrigerant flow. However, as in the modification shown in FIG. 17, from the diffuser part 18g side. Alternatively, a taper shape that tapers toward the upstream side of the refrigerant flow may be employed. In this case, it suffices to arrange the taper tip on the most upstream side so that it protrudes toward the taper 18i rather than the injection part 18j.

(10th Embodiment)
This embodiment demonstrates the example which changed the structure of the ejector-type refrigerating cycle 10a with respect to 4th Embodiment. Specifically, in the ejector-type refrigeration cycle 10a of the present embodiment, as shown in FIG. 18, a high-stage ejector 131 is employed as the first pressure reducing means instead of the high-stage expansion device 13.

  The basic configuration of the high-stage ejector 131 is the same as that of the ejector 18 described above. Therefore, similarly to the ejector 18, the high-stage ejector 131 also has the high-stage nozzle portion 131 a that depressurizes the refrigerant, the high-stage refrigerant suction port 131 d that sucks the refrigerant flowing out of the first evaporator 15, and the mixed refrigerant. The high-stage side diffuser section (high-stage-side boost section) 131g to be boosted has a high-stage body section 131b in which is formed.

  Here, since the liquid-phase refrigerant condensed by the radiator 12 can be caused to flow into the high-stage nozzle portion 131a of the high-stage ejector 131 of the present embodiment, the high-stage ejector 131 has a high-stage side. The gas-liquid two-phase refrigerant having a high degree of dryness does not flow into the nozzle portion 131a so that the high stage side diffuser portion 131g cannot exhibit the desired boosting performance.

  Therefore, in the present embodiment, the high-stage ejector 131 is not of the same configuration as the ejector 18 described above, and the ejector-type refrigeration cycle 10a is introduced when the liquid-phase refrigerant is introduced into the high-stage nozzle portion 131a. The thing set so that high COP can be demonstrated as a whole is adopted.

  A gas-liquid separator 21 that separates the gas-liquid refrigerant flowing out from the high-stage diffuser portion 131g of the high-stage ejector 131 is connected to the outlet side of the high-stage side diffuser portion 131g of the high-stage ejector 131.

  A refrigerant inlet of the first evaporator 15 is connected to the liquid-phase refrigerant outlet of the gas-liquid separator 21 via a fixed throttle 22, and a high-stage ejector is connected to the refrigerant outlet of the first evaporator 15. A refrigerant suction port 131 is connected. On the other hand, the gas-phase refrigerant outlet of the gas-liquid separator 21 is connected to the inlet side of the nozzle portion 18 a of the ejector 18. Other configurations are the same as those of the fourth embodiment.

  Therefore, when the ejector refrigeration cycle 10 a of this embodiment is operated, the flow of the liquid-phase refrigerant that has flowed out of the radiator 12 is branched at the branching section 14. One refrigerant branched by the branching portion 14 flows into the high-stage nozzle portion 131a of the high-stage ejector 131 and is isentropically decompressed and injected.

  The refrigerant flowing out of the first evaporator 15 is sucked from the high stage side refrigerant suction port 131d of the high stage side ejector 131 by the suction action of the jet refrigerant. The mixed refrigerant of the refrigerant injected from the high stage side nozzle part 131a and the suction refrigerant sucked from the high stage side refrigerant suction port 131d flows into the high stage side diffuser part 131g and is pressurized.

  The refrigerant that has flowed out of the high stage side diffuser portion 131g flows into the gas-liquid separator 21 and is gas-liquid separated. Then, the liquid phase refrigerant separated by the gas-liquid separator 21 flows into the first evaporator 15 via the fixed throttle 22. On the other hand, the gas-phase refrigerant separated by the gas-liquid separator 21 flows into the nozzle portion 18 a of the ejector 18. Other operations are the same as those in the fourth embodiment.

  Therefore, according to the ejector type refrigeration cycle 10a of the present embodiment, not only the same effects as in the fourth embodiment can be obtained, but also the power consumption of the compressor 11 can be reduced by the boosting action of the high stage ejector 131, and the cycle The COP as a whole can be further improved.

  The ejector-type refrigeration cycle 10a that employs the high-stage ejector 131 as the first pressure reducing means is not limited to the cycle configuration shown in FIG. 18, but may be configured as shown in FIG. 19, for example.

  Specifically, in the ejector-type refrigeration cycle 10a shown in FIG. 19, the refrigerant inlet side of the first evaporator 15 is connected to the outlet side of the high stage side diffuser part 131g of the high stage side ejector 131, and further, the branch part ( A second branch portion 14 a that further branches the refrigerant flow is connected to the other refrigerant outlet of the first branch portion 14.

  Then, the refrigerant inlet of the third evaporator 23 is connected to one refrigerant outlet of the second branch portion 14a via the fixed throttle 132, and the high stage ejector 131 of the third evaporator 23 is connected to the refrigerant outlet of the third evaporator 23. The high stage side refrigerant suction port 131d is connected. The third evaporator 23 performs heat exchange between the low-pressure refrigerant decompressed by the fixed throttle 132 and the blown air blown from the third blower fan 23a, thereby evaporating the low-pressure refrigerant and exerting an endothermic effect. It is a heat exchanger.

  In addition, the refrigerant inlet of the second evaporator 17 is connected to the other refrigerant outlet of the second branch portion 14 a via the low-stage expansion device 16. Other configurations are the same as those of the fourth embodiment. Even with such a cycle configuration, the COP of the entire cycle can be further improved by the boosting action of the high-stage ejector 131.

(Other embodiments)
The present invention is not limited to the above-described embodiment, and can be variously modified as follows without departing from the spirit of the present invention.

  (1) In the above-described embodiment, the ejector refrigeration cycle 10, 10 a, 10 b including the ejector 18 is applied to a refrigeration cycle apparatus for a vehicle, the indoor blowing air is cooled by the first evaporator 15, and the second evaporation Although the example which cooled the ventilation air for storage in the container 17 was demonstrated, application of the ejector-type refrigerating cycle 10-10b is not limited to this.

  For example, when applied to a vehicle refrigeration cycle apparatus, the front-seat air blown to the front seat side of the vehicle is cooled by the first evaporator 15, and the air is blown to the rear seat side of the vehicle by the second evaporator 17. The rear-seat blown air may be cooled.

  Further, for example, when applied to a refrigeration refrigerator, the first evaporator 15 blows air to a refrigeration room that refrigerates and stores food / drinking water at a low temperature (specifically, 0 ° C. to 10 ° C.). The refrigeration room air is cooled, and the second evaporator 17 blows air to the freezer room where the food is stored frozen at a very low temperature (specifically, −20 ° C. to −10 ° C.). The air may be cooled.

  (2) In the above-described embodiment, the example in which the ejector 18 is applied to the ejector refrigeration cycles 10 to 10b has been described. However, the applicable cycle configuration of the ejector 18 is not limited thereto.

  For example, in the ejector refrigeration cycles 10 to 10b, the gas-liquid refrigerant flowing out from the diffuser portion 18g is separated between the outlet side of the diffuser portion 18g of the ejector 18 and the suction port side of the compressor 11 and separated. An accumulator that allows the gas-phase refrigerant to flow out to the suction port side of the compressor 11 may be disposed.

  Moreover, you may arrange | position the beneficiary device which isolate | separates the gas-liquid of the refrigerant | coolant which flowed out from the heat radiator 12, and flows out a liquid phase refrigerant | coolant downstream in the refrigerant | coolant exit side of the heat radiator 12. FIG. Further, an internal heat exchanger for exchanging heat between the high-temperature refrigerant flowing out from the radiator 12 and the low-temperature refrigerant sucked into the compressor 11 may be arranged. Furthermore, an auxiliary pump for refrigerant pressure feeding may be provided between the refrigerant outlet side of the second evaporator 17 and the refrigerant suction port 18 d of the ejector 18.

  (3) In the above-described embodiment, the example in which the temperature type expansion valve, the fixed throttle, and the high stage side ejector are employed as the high stage side throttle device 13 and the low stage side throttle device 16 has been described. 13. As the low-stage side throttle device 16, an electric type constituted by including a valve body configured to be able to change the throttle opening degree and an electric actuator composed of a stepping motor that changes the throttle opening degree of the valve body. The variable aperture mechanism may be used.

  Moreover, although the above-mentioned embodiment demonstrated the example which employ | adopted what consists of the heat exchange part which heat-exchanges the compressor 11 discharge refrigerant | coolant and external air as the heat radiator 12, as the heat radiator 12, the compressor 11 discharge refrigerant | coolant. Heat exchange between the refrigerant and the outside air to condense the refrigerant discharged from the compressor 11, a modulator for separating the gas and liquid of the refrigerant flowing out from the condensing unit, and heat exchange between the liquid phase refrigerant flowing out from the modulator and the outside air A so-called subcool type condenser having a supercooling unit for supercooling the liquid phase refrigerant may be employed.

  Moreover, although the above-mentioned embodiment demonstrated the example which formed structural members, such as the body part 18b of the ejector 18, with the metal, if the function of each structural member can be exhibited, a material will not be limited. That is, you may form these structural members with resin.

  (4) In the above-described embodiment, the example in which the refrigerant passage area of the inlet portion 18h of the diffuser portion 18g is set smaller than the refrigerant passage area of the refrigerant injection port 18c of the nozzle portion 18ac has been described. The opening diameter of the refrigerant injection port 18c may be set smaller than the opening diameter of the inlet portion 18h.

  Further, when the opening diameter of the inlet portion 18h is set larger than the opening diameter of the refrigerant injection port 18c, by providing the inlet portion 18h with a protruding portion that protrudes into the refrigerant passage, the refrigerant passage area of the inlet portion 18h. May be made smaller than the refrigerant passage area of the refrigerant injection port 18c.

  (5) In the above-described ninth embodiment, the example in which the refrigerant passage area of the minimum passage area portion of the refrigerant passage formed in the nozzle portion 18a by the valve body (needle valve 18y) can be changed has been described. As a body, a conical shape extending from the refrigerant passage formed in the nozzle portion 18a to the inside of the diffuser portion 18g is adopted, and the refrigerant passage area of the diffuser portion 18g is changed simultaneously with the minimum passage area portion of the nozzle portion 18a. It is good also as a structure.

  (6) In the above-described embodiment, the example in which R134a is adopted as the refrigerant has been described, but the refrigerant is not limited to this. For example, R600a, R1234yf, R410A, R404A, R32, R1234yfxf, R407C, etc. can be employed. Or you may employ | adopt the mixed refrigerant | coolant etc. which mixed multiple types among these refrigerant | coolants.

  (7) The means disclosed in each of the above embodiments may be appropriately combined within a practicable range. For example, the biphasic means described in the fifth to seventh embodiments may be applied to the ejector refrigeration cycle 10a described in the fourth embodiment. For example, the ejector 18 disclosed in the second, third, eighth, and ninth embodiments may be applied as the ejector 18 of the ejector refrigeration cycle 10a described in the tenth embodiment.

  (8) In the above-described embodiment, the radiator 12 is an outdoor heat exchanger that exchanges heat between the refrigerant and the outside air, and the first and second evaporators 15 and 17 are user-side heat exchangers that cool the blown air. On the contrary, the first and second evaporators 15 and 17 are configured as outdoor heat exchangers that absorb heat from a heat source such as outside air, and the radiator 12 heats a fluid to be heated such as air or water. The present invention may be applied to a heat pump cycle configured as an indoor heat exchanger.

10, 10a, 10 Ejector refrigeration cycle 15, 17 First and second evaporators 18 Ejector 18a Nozzle part 18b Body part 18c Refrigerant injection port 18d Refrigerant suction port 18e Mixing unit 18g Diffuser unit 18h Inlet unit

Claims (14)

A compressor (11) for compressing and discharging the refrigerant;
A radiator (12) for radiating the refrigerant discharged from the compressor (11);
A first decompression means (13) and a second decompression means (16) for decompressing the refrigerant on the downstream side of the radiator (12);
A first evaporator (15) for evaporating the refrigerant depressurized by the first depressurization means (13);
A second evaporator (17) for evaporating the refrigerant decompressed by the second decompression means (16);
The second evaporator (17) downstream refrigerant from the refrigerant suction port (18d) by the suction action of the high-speed jet refrigerant injected from the nozzle part (18a) for depressurizing the first refrigerant (15) downstream refrigerant. And an ejector (18) having a pressure increasing part (18g) for mixing and increasing the pressure of the injected refrigerant and the suction refrigerant sucked from the refrigerant suction port (18d),
The ejector-type refrigeration cycle further comprises a two-phase means (19) for causing the refrigerant flowing out from the first evaporator (15) to flow into the nozzle part (18a) in a gas-liquid two-phase state.   The biphasic means has a liquid storage part (19) for storing excess refrigerant in the cycle, and flows out of the liquid phase refrigerant stored in the liquid storage part (19) and the first evaporator (15). The ejector-type refrigeration cycle according to claim 1, wherein the refrigerant is mixed and discharged to the inlet side of the nozzle portion (18a).   The biphasic means has a liquid storage section (19) for storing excess refrigerant in the cycle, and is discharged from the liquid phase refrigerant stored in the liquid storage section (19) and the compressor (11). The ejector refrigeration cycle according to claim 1, wherein the ejector refrigeration cycle is mixed with a refrigerant and allowed to flow out to the inlet side of the nozzle portion (18a).   The biphasic means mixes the refrigerant that has flowed out of the radiator (12) and the refrigerant that has flowed out of the first evaporator (15) and causes the refrigerant to flow out to the inlet side of the nozzle portion (18a). The ejector type refrigeration cycle according to claim 1. The second pressure reducing means (16) is composed of a variable throttle mechanism configured to be able to change the throttle opening.
The throttle opening degree of the second decompression means (16) is adjusted so that the refrigerant flowing out of the second evaporator (17) is less than a gas-liquid two-phase refrigerant or a predetermined reference superheat degree. The ejector type refrigeration cycle according to any one of claims 1 to 4. The ejector (18) has a body part (18b) that forms the refrigerant suction port (18d) and the pressure increasing part (18g),
In the internal space of the body part (18b), the range of the refrigerant from the refrigerant injection port (18c) of the nozzle part (18a) to the inlet part (18h) of the pressure increasing part (18g) A mixing part (18e) for mixing the refrigerant is formed,
The distance (La) from the refrigerant injection port (18c) to the inlet part (18h) in the mixing part (18e) is such that the flow rate of the refrigerant flowing into the inlet part (18h) is equal to or lower than the two-phase sound speed. The ejector refrigeration cycle according to any one of claims 1 to 5, wherein the ejector refrigeration cycle is determined. The distance from the refrigerant injection port (18c) in the mixing unit (18e) to the inlet unit (18h) is La,
The cross-sectional area of the opening of the refrigerant injection port (18c) and the cross-sectional area of the suction passage (18f) through which the suction refrigerant flows in the axially vertical cross section of the nozzle portion (18a) including the refrigerant injection port (18c). When the diameter when the total value is converted into a circle is Da,
La / Da ≦ 1
The ejector refrigeration cycle according to claim 6, wherein: The ejector (18) has a body part (18b) that forms the refrigerant suction port (18d) and the pressure increasing part (18g),
In the internal space of the body part (18b), the range of the refrigerant from the refrigerant injection port (18c) of the nozzle part (18a) to the inlet part (18h) of the pressure increasing part (18g) A mixing part (18e) for mixing the refrigerant is formed,
As the refrigerant passage formed in the nozzle portion (18a), a tapered portion (18i) that gradually reduces the refrigerant passage area, and an injection portion that guides the refrigerant from the tapered portion (18i) to the refrigerant injection port (18c). (18j) is provided,
The nozzle part (18a) is configured to allow the jet refrigerant to be jetted to the mixing part (18e) when the spread angle (θn) in the axial section of the jet part (18j) is 0 ° or more. 8. The ejector refrigeration cycle according to claim 1, wherein the ejector refrigeration cycle is formed so as to freely expand.   The ejector refrigeration cycle according to any one of claims 6 to 8, wherein the mixing section (18e) is formed in a shape in which a refrigerant passage area decreases toward a refrigerant flow downstream side.   The said mixing part (18e) is formed in the shape which combined the truncated cone shape and cylinder shape which a refrigerant path area reduces gradually toward a refrigerant | coolant flow downstream, The shape of Claim 9 characterized by the above-mentioned. Ejector refrigeration cycle. When the axial length of the cylindrical portion of the mixing portion (18e) is Lb and the diameter of the cylindrical portion is Db,
Lb / Db ≦ 1
The ejector refrigeration cycle according to claim 10, wherein:   The ejector type according to any one of claims 6 to 11, wherein a refrigerant passage area of the inlet portion (18h) is set smaller than a refrigerant passage area of the refrigerant injection port (18c). Refrigeration cycle.   The ejector (18) includes a swirling space forming member (18m) that forms a swirling space (18k) for swirling the refrigerant flowing into the nozzle portion (18a) around the axis of the nozzle portion (18a). The ejector-type refrigeration cycle according to any one of claims 1 to 12.   The ejector refrigeration cycle according to any one of claims 1 to 13, wherein the ejector (18) includes a valve body (18y) that changes a refrigerant passage area of the nozzle portion (18a).

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