JP2009002649A - Ejector type refrigerating cycle - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To operate an ejector type refrigerating cycle while bringing a flow rate ratio η into a value closely near to the optimum flow rate ratio ηmax and while exhibiting a high cooling capacity as the whole cycle, in the ejector type refrigerating cycle provided with a plurality of evaporators including the evaporator arranged in a branch passage branched from an ejector upstream. <P>SOLUTION: A refrigerant passage area of a nozzle part 15a of an ejector 15 and a throttle opening of a fixed throttle 19 are appropriately set so that when a first evaporator 16 outlet side refrigerant superheat degree becomes a prescribed value, the flow rate ratio η that is the ratio of a refrigerant flow rate Gnoz of the nozzle part 15a and a refrigerant flow rate Ge sucked by a refrigerant suction port 15b exhibits a high cooling capacity as the whole cycle. Further, the refrigerant flow rate of the whole cycle is controlled so that the first evaporator 16 outlet side refrigerant superheat degree becomes the prescribed value by a variable throttle mechanism 31 arranged in a refrigerant passage from a radiator 12 outlet side to a branch part Z inlet side. <P>COPYRIGHT: (C)2009,JPO&INPIT

Description

  The present invention relates to an ejector-type refrigeration cycle having an ejector serving as a refrigerant decompression unit and a refrigerant circulation unit.

  Conventionally, an ejector-type refrigeration cycle including a plurality of evaporators is known from Patent Document 1 and the like.

  In Patent Document 1, as shown in FIG. 6, a first evaporator 16 is connected to the downstream side of the ejector 15, and an accumulator 32 that forms a gas-liquid separator is arranged on the downstream side of the refrigerant of the first evaporator 16. In addition, a cycle is disclosed in which the second evaporator 20 is disposed between the liquid-phase refrigerant outlet side of the accumulator 32 and the refrigerant suction port 15b side of the ejector 15, and the two evaporators are operated simultaneously.

  In this cycle, the refrigerant flowing out of the second evaporator 20 is sucked using the pressure drop caused by the high-speed flow of the refrigerant when the refrigerant is decompressed and expanded by the ejector 15 and the speed of the refrigerant at the time of expansion is increased. The energy is converted into pressure energy by the diffuser unit 15d (pressure increase unit) to increase the refrigerant pressure (the suction pressure of the compressor 11). Thereby, since the driving power of the compressor 11 can be reduced, the operating efficiency of the cycle can be improved.

  Further, in this cycle, the two evaporators, the first evaporator 16 and the second evaporator 20, can exert an endothermic (cooling) action from different spaces or from the same space in the two evaporators. And it is also possible to cool the room with two evaporators (see paragraph 0192 of Patent Document 1).

Further, in Patent Document 1, in an ejector-type refrigeration cycle in which an evaporator (corresponding to the second evaporator) is disposed only on the refrigerant suction port 15b side of the ejector 15, a machine is disposed upstream of the ejector or upstream of the evaporator. It is described that a control valve of an electric type or an electric type is provided. And it is described that these control valves perform superheat degree control etc. of an evaporator exit refrigerant | coolant by the opening degree adjustment (refer FIG. 27, 29, 31-33 of patent document 1).
Japanese Patent No. 3322263

  By the way, in this prior art cycle, the entire amount of refrigerant flowing out of the radiator 12 passes through the nozzle portion 15a of the ejector 15. Here, if the refrigerant flow rate that passes through the nozzle portion 15a of the ejector 15 is Gnoz, this Gnoz is such a flow rate that the refrigerant on the outlet side of the first evaporator 16 is below a predetermined dryness.

  Further, the refrigerant decompressed by the nozzle portion 15 a is mixed with the refrigerant sucked from the refrigerant suction port 15 b of the ejector 15 and flows into the first evaporator 16. Then, the refrigerant flowing out of the first evaporator 16 is separated into a gas phase refrigerant and a liquid phase refrigerant by the accumulator 32.

  On the other hand, the second evaporator 20 is supplied with the liquid phase refrigerant separated by the accumulator 32 when the refrigerant suction port 15b of the ejector 15 is reduced in pressure and exerts a suction action. Here, the refrigerant flow rate sucked from the refrigerant suction port 15b is Ge.

  Since the liquid-phase refrigerant that has flowed into the second evaporator 20 evaporates in the second evaporator 20, most or all of the refrigerant sucked from the refrigerant suction port 15b is a gas-phase refrigerant. Therefore, since the refrigerant flow rate contributing to the cooling capacity of the first evaporator 16 is substantially the flow rate of the liquid phase refrigerant of Gnoz, the cooling capacity of the first evaporator 16 is affected by Gnoz.

  Therefore, in the prior art cycle, if the refrigerant flow rate Ge sucked into the refrigerant suction port 15b of the ejector 15 is increased and the flow rate of the liquid-phase refrigerant flowing into the second evaporator 20 is increased, the first evaporator Since the cooling capacity of the second evaporator 20 can be increased without reducing the cooling capacity of 16, the cooling capacity of the entire cycle is also increased.

  Here, the cooling capacity of the evaporator is, for example, an increase in the enthalpy of the refrigerant when the refrigerant absorbs heat from the air in the evaporator. This increase in enthalpy is the increase in specific enthalpy per unit weight of the refrigerant multiplied by the refrigerant flow rate. Further, the cooling capacity of the entire cycle is the total value Qer of the enthalpy increase of the refrigerant in the first and second evaporators 16 and 20. Of course, the coefficient of performance COP obtained by dividing Qer by the power consumed by the compressor 11 may be used.

  Therefore, in the prior art cycle, as shown in FIG. 8, the flow rate ratio η (which is the ratio of the refrigerant flow rate Gnoz passing through the nozzle portion 15 a of the ejector 15 and the refrigerant flow rate Ge sucked into the refrigerant suction port 15 b of the ejector 15. As (η = Ge / Gnoz) increases, the cooling capacity Qer of the entire cycle also increases.

  However, when the cycle heat load is low, the difference between the high and low pressures of the refrigerant in the cycle is small, so the input to the ejector 15 is small. In this case, in the prior art cycle, the refrigerant flow rate Ge depends only on the refrigerant suction ability of the ejector 15, so the input of the ejector 15 decreases → the refrigerant suction capacity of the ejector 15 decreases → the liquid phase flowing into the second evaporator 20. A decrease in the flow rate of the refrigerant → a decrease in the flow rate ratio η occurs, and the cooling capacity Qer decreases.

  Therefore, the present applicant has previously proposed a cycle as shown in FIG. 7 in Japanese Patent Application No. 2004-290120 (Japanese Patent Laid-Open No. 2005-308380: hereinafter referred to as a prior application example). In this prior application cycle, a branch passage 18 is provided between the discharge side of the radiator 12 and the ejector 15 refrigerant inlet, and the throttle mechanism 42 and the second evaporator 20 for adjusting the pressure and flow rate of the refrigerant in the branch passage 18. The outlet side of the second evaporator 20 is connected to the refrigerant suction port 15b side of the ejector 15.

  That is, in the cycle of the prior application example, the refrigerant flow is branched at the upstream portion of the ejector 15, and the branched refrigerant is sucked into the refrigerant suction port 15 b via the branch passage 18, so that the branch passage 18 is parallel to the ejector 15. Connection. For this reason, the refrigerant can be supplied to the branch passage 18 by utilizing not only the refrigerant suction capability of the ejector 15 but also the refrigerant suction / discharge capability of the compressor 11.

  Therefore, even if the phenomenon that the input of the ejector 15 is reduced and the refrigerant suction capacity of the ejector 15 is reduced occurs, the degree of decrease in the refrigerant flow rate Ge sucked into the refrigerant suction port 15b of the ejector 15 can be made smaller than that of the prior art cycle. .

  In the cycle of the prior application example, since the refrigerant flow is branched at the upstream portion of the ejector 15, the refrigerant flow rate Gn flowing out of the radiator 12 is equal to the refrigerant flow rate Gnoz passing through the nozzle portion 15 a of the ejector 15. It becomes equal to the sum of the refrigerant flow rate flowing into the evaporator 20. The refrigerant flow rate flowing into the second evaporator 20 is equal to the refrigerant flow rate Ge sucked into the refrigerant suction port 15 b of the ejector 15.

  Therefore, in the cycle of the prior application example, the relationship of Gn = Gnoz + Ge is established, and if Gnoz decreases, Ge increases, and conversely, if Gnoz increases, Ge decreases, so the cooling capacity of the first evaporator 16 decreases. However, the cooling capacity of the second evaporator 20 increases, and conversely, even if the cooling capacity of the second evaporator 20 decreases, the cooling capacity of the first evaporator 16 increases.

  Therefore, as shown in FIG. 8, there is an advantage that the change in the cooling capacity Qer with respect to the change in the flow rate ratio η is less than the cycle of the prior art, and further, there is a peak in the optimum flow rate ratio ηmax. That is, in the cycle of the prior application example, in order to operate the cycle while exhibiting the high cooling capacity Qer, it is necessary to operate so that the flow rate ratio η approaches the optimum flow rate ratio ηmax.

  In view of the above points, the present invention is an ejector-type refrigeration cycle including a plurality of evaporators including an evaporator disposed in a branch passage branched from the upstream side of the ejector as in the cycle of the prior application. The object is to operate while approaching the ratio ηmax and exhibiting a high cooling capacity as a whole cycle.

In order to achieve the above object, according to the first aspect of the present invention, a compressor (11) that sucks and compresses a refrigerant, and a radiator (12) that radiates high-pressure refrigerant discharged from the compressor (11). And a branch part (Z) for branching the flow of the refrigerant flowing out from the radiator (12), and a high speed at which one of the refrigerants branched at the branch part (Z) is injected from the nozzle part (15a) for decompression and expansion. The refrigerant flow is sucked from the refrigerant suction port (15b), the suction refrigerant from the refrigerant suction port (15b) and the high-speed refrigerant flow are mixed, and the mixed refrigerant flow is decelerated to reduce the refrigerant flow. An ejector (15) for increasing the pressure, a first evaporator (16) for evaporating the refrigerant flowing out from the ejector (15), and the other refrigerant branched at the branching section (Z) as the refrigerant suction port (15b) Branch passage (18) leading to Includes an arrangement has been throttle means for flow rate adjustment by reducing the pressure of the refrigerant (19), stop means (19) is disposed on the downstream side second evaporator for evaporating the refrigerant and (20),
Furthermore, the variable throttle means (31, 35, 38, 40) is disposed in the refrigerant passage from the radiator (12) outlet side to the branch (Z) inlet side, and decompresses and expands the refrigerant flowing out from the radiator (12). Ejector refrigeration cycle comprising

  According to this, since the variable throttle means (31, 35, 38, 40) is arranged in the refrigerant passage from the radiator (12) outlet side to the branching section (Z) inlet side, the nozzle of the ejector (15) By appropriately setting the refrigerant passage area of the section (15a) and the throttle opening (refrigerant passage area) of the throttle means (19), the flow rate ratio (η) can be brought close to the optimum flow rate ratio (ηmax). . As a result, the ejector refrigeration cycle can be operated while exhibiting a high cooling capacity as a whole cycle.

  According to a second aspect of the present invention, in the ejector refrigeration cycle according to the first aspect, the variable throttle means (31, 35, 38, 40) includes the refrigerant state inside the cycle, the first evaporator (16) and the first evaporator. The refrigerant flow rate of the entire cycle is adjusted based on a physical quantity related to at least one of the cooling target space temperature of the two evaporators (20) and the ambient temperature of the cooling target space.

  According to this, since the refrigerant flow rate of the entire cycle is adjusted based on the physical quantity related to the refrigerant state or the like inside the cycle, the refrigerant passage of the nozzle portion (15a) of the ejector (15) at the adjusted refrigerant flow rate. The area, the size of the part responsible for the function of mixing and boosting the refrigerant sucked from the refrigerant suction port (15b) and the high-speed refrigerant flow decompressed and expanded by the nozzle part (15a), and the throttle means (19) By appropriately setting the throttle opening (refrigerant passage area) and the like in advance, the ejector refrigeration cycle can be operated while exhibiting a high cooling capacity as a whole cycle.

  Here, the physical quantity related to the refrigerant state in the cycle in the present invention is, for example, a physical quantity related to the temperature, pressure, and flow rate of the refrigerant, and the degree of superheat, the degree of supercooling, and the flow rate at a specific location in the cycle. And the ratio of the flow rate at another location. In addition, the temperature of the cooling target space is, for example, the inside air temperature, and the ambient temperature of the cooling target space is, for example, the outside air temperature, and includes a temperature that can indirectly estimate the refrigerant state. It is.

  In order to exert the effect of the above feature, specifically, as in the invention described in claim 3, the physical quantity is a physical quantity related to the degree of superheat of the refrigerant on the outlet side of the first evaporator (16), and the variable throttle The means (31) may adjust the refrigerant flow rate of the entire cycle so that the degree of superheat of the refrigerant on the outlet side of the first evaporator (16) approaches a predetermined value.

  Further, as in the invention described in claim 4, the physical quantity is a physical quantity related to the degree of superheat of the refrigerant on the outlet side of the second evaporator (20), and the variable throttle means (35) includes the second evaporator (20 ) The refrigerant flow rate of the entire cycle may be adjusted so that the degree of superheat of the outlet side refrigerant approaches a predetermined value.

  Further, as in the invention described in claim 5, the physical quantity is a physical quantity related to the degree of supercooling of the refrigerant on the outlet side of the radiator (12), and the variable throttle means (38) is provided on the outlet side of the radiator (12). The refrigerant flow rate of the entire cycle may be adjusted so that the degree of supercooling of the refrigerant approaches a predetermined value.

  Further, as in the invention described in claim 6, the high-pressure refrigerant is boosted to a critical pressure or higher, and the physical quantity is the temperature and pressure of the refrigerant on the outlet side of the radiator (12), and the variable throttle means (38) may adjust the refrigerant flow rate of the entire cycle so that the pressure of the refrigerant on the radiator (12) outlet side approaches a predetermined value.

  Further, as in the seventh aspect of the invention, the physical quantity is a physical quantity related to the refrigerant discharge flow rate of the compressor (11), and the variable throttle means (40) has a predetermined value for the refrigerant flow rate of the entire cycle. The refrigerant flow rate of the entire cycle may be adjusted so that

  If the physical quantity is a physical quantity related to the degree of superheat of the refrigerant on the outlet side of the first evaporator (16), high cooling capacity can be exhibited as a whole cycle, and at the same time, the refrigerant on the outlet side of the first evaporator (16) is superheated. Therefore, it is possible to prevent the liquid refrigerant from being sucked into the compressor (11). As a result, malfunction of the compressor (11) can be prevented, and cycle stability can be improved.

  If the physical quantity is the temperature and pressure of the refrigerant on the outlet side of the radiator (12), the above-described effects can be exhibited even when the high-pressure refrigerant is boosted to a critical pressure or higher and the degree of supercooling cannot be detected.

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

(Form which is the premise of the present invention)
First, the form which becomes the premise of this invention is demonstrated. FIG. 1 shows an example in which an ejector refrigeration cycle 10 which is a premise of the present invention is applied to a vehicular refrigeration apparatus. The vehicular refrigeration apparatus of this embodiment cools the internal temperature to, for example, an extremely low temperature around −20 ° C.

  First, in the ejector refrigeration cycle, the compressor 11 sucks, compresses and discharges refrigerant, and is driven to rotate by a vehicle travel engine (not shown) through an electromagnetic clutch 11a and a belt. In this embodiment, a swash plate type variable displacement compressor capable of continuously variably controlling the discharge capacity by an external control signal is used.

  Specifically, the pressure in the swash plate chamber (not shown) is controlled using the discharge pressure and the suction pressure of the compressor 11, the piston stroke is changed by changing the inclination angle of the swash plate, The discharge capacity is continuously changed in a range of approximately 0% to 100%. The refrigerant discharge capacity can be adjusted by changing the discharge capacity.

  Here, the discharge capacity is the geometric volume of the working space where the refrigerant is sucked and compressed, and is the cylinder volume between the top dead center and the bottom dead center of the piston stroke.

  Further, the control of the pressure in the swash plate chamber will be described. The compressor 11 includes an electromagnetic capacity control valve 11b. The electromagnetic capacity control valve 11b generates a force F1 due to the low-pressure refrigerant pressure on the suction side of the compressor 11. A pressure responsive mechanism (not shown) that generates and an electromagnetic mechanism (not shown) that generates an electromagnetic force F2 that opposes the force F1 generated by the low-pressure refrigerant pressure Ps are incorporated.

  The electromagnetic force F2 of this electromagnetic mechanism is determined by a control current In output from the air conditioning control device 21 described later. The ratio of introducing the high-pressure refrigerant and the low-pressure refrigerant into the swash plate chamber is changed by a valve body (not shown) that is displaced according to the force F1 and the electromagnetic force F2 according to the low-pressure refrigerant pressure Ps. The pressure is changed.

  In the compressor 11, the discharge capacity can be continuously changed from 100% to about 0% by adjusting the pressure in the swash plate chamber. Therefore, the compressor 11 can be reduced by reducing the discharge capacity to about 0%. Can be substantially deactivated. Therefore, it is good also as a clutchless structure which always connects the rotating shaft of the compressor 11 to a vehicle engine via a pulley and the belt V. FIG.

  The radiator 12 is connected to the refrigerant discharge side of the compressor 11, and performs heat exchange between the high-pressure refrigerant discharged from the compressor 11 and the outside air (air outside the vehicle compartment) blown by the radiator fan 12a. And a heat exchanger for cooling the high-pressure refrigerant.

  The radiator fan 12a is driven by a driving electric motor 12b, and the driving electric motor 12b is driven to rotate when an applied voltage V1 is output from the air conditioning control device 21. Therefore, since the rotation speed of the drive electric motor 12b can be changed when the air conditioning control device 21 changes the applied voltage V1, the amount of air blown by the radiator fan 12a can be changed.

  In the present embodiment, since a normal chlorofluorocarbon refrigerant is used as the circulating refrigerant in the cycle, the ejector refrigeration cycle 10 constitutes a subcritical cycle in which the high pressure does not exceed the critical pressure. Therefore, the radiator 12 acts as a condenser that condenses the refrigerant.

  A liquid receiver 13 is disposed as a gas-liquid separator that separates the gas-liquid of the refrigerant and accumulates the liquid-phase refrigerant, and the liquid-phase refrigerant is led downstream from the liquid receiver 13. The A variable throttle mechanism 14 is connected to the refrigerant downstream side of the liquid receiver 13.

  Specifically, the variable throttle mechanism 14 is a well-known temperature expansion valve, and functions to reduce the high-pressure liquid-phase refrigerant from the liquid receiver 13 to a gas-liquid two-phase intermediate-pressure refrigerant.

  This temperature type expansion valve adjusts the opening degree of a valve body part (not shown) according to the superheat degree of the 1st evaporator 16 outlet side refrigerant | coolant mentioned later, and, thereby, the refrigerant | coolant which passes the variable throttle mechanism 14 The flow rate is adjusted so that the degree of superheat of the refrigerant on the outlet side of the first evaporator 16 approaches a predetermined value. That is, in this embodiment, the valve body portion of the temperature type expansion valve is a means for adjusting the flow rate ratio η.

  A diaphragm mechanism 14a serving as a pressure responsive means is coupled to the valve body of the temperature type expansion valve. ) And the first evaporator 16 outlet side refrigerant pressure introduced by the pressure equalizing pipe 14c, the valve body is displaced to adjust the opening degree of the valve body. That is, in this embodiment, the temperature sensing cylinder 14b and the pressure equalizing pipe 14c serve as means for detecting a physical quantity related to the refrigerant state in the cycle.

  An ejector 15 is connected to the outlet side of the variable throttle mechanism 14. The ejector 15 is a pressure reducing means for reducing the pressure of the refrigerant, and is also a refrigerant circulating means for transporting the fluid to circulate the refrigerant by a suction action (winding action) of the refrigerant flow ejected at a high speed.

  The ejector 15 includes a nozzle portion 15a for reducing the passage area of the intermediate-pressure refrigerant flowing through the variable throttle mechanism 14 to a small size and isentropically decompressing and expanding the intermediate-pressure refrigerant, and a refrigerant outlet of the nozzle portion 15a. A refrigerant suction port 15b that is disposed in the same space and sucks a gas-phase refrigerant from the second evaporator 20 described later is provided.

  Further, on the downstream side of the nozzle portion 15a and the refrigerant suction port 15b, a mixing unit 15c that mixes the high-speed refrigerant flow from the nozzle portion 15a and the suction refrigerant from the refrigerant suction port 15b is provided. And the diffuser part 15d which makes | forms a pressure | voltage rise part is arrange | positioned downstream of the mixing part 15c.

  The diffuser portion 15d is formed in a shape that gradually increases the refrigerant passage area, and has the function of decelerating the refrigerant flow to increase the refrigerant pressure, that is, the function of converting the velocity energy of the refrigerant into pressure energy.

  A first evaporator 16 is connected to the downstream side of the diffuser portion 15 d of the ejector 15. The first evaporator 16 is a heat absorber that exchanges heat between the air blown from the evaporator blower 16a and the refrigerant, evaporates the refrigerant, and exhibits an endothermic effect.

  The evaporator blower 16a is driven by a driving electric motor 16b, and the driving electric motor 16b is driven to rotate when an applied voltage V2 is output from the air conditioning control device 21. Therefore, since the rotation speed of the drive electric motor 16b can be changed when the air conditioning control device 21 changes the applied voltage V2, the amount of air blown by the evaporator blower 16a can be changed.

  The refrigerant flow downstream side of the first evaporator 16 is connected to the internal heat exchanger 17, and the refrigerant outlet side of the internal heat exchanger 17 is connected to the suction side of the compressor 11.

  Next, the branch passage 18 is a refrigerant passage that connects between the liquid receiver 13 and the variable throttle mechanism 14 and the refrigerant suction port 15 b of the ejector 15. Z indicates a branch portion of the branch passage 18. In the branch passage 18, the internal heat exchanger 17 described above is disposed, a fixed throttle 19 is disposed on the downstream side of the internal heat exchanger 17, and the second evaporator 20 is disposed on the downstream side of the fixed throttle 19. Is arranged.

  Here, the internal heat exchanger 17 performs heat exchange between the high-temperature high-pressure refrigerant passing through the branch passage 18 and the low-temperature low-pressure refrigerant downstream of the first evaporator 16. The refrigerant passing through the branch passage 18 is cooled by heat exchange between the refrigerants in the internal heat exchanger 17, so that the enthalpy of refrigerant between the refrigerant inlet and outlet in the first evaporator 16 and the second evaporator 20. The difference (cooling capacity) can be increased.

  The fixed throttle 19 adjusts the flow rate of the refrigerant flowing into the second evaporator 20 and depressurizes it. Specifically, the fixed throttle 19 can be constituted by a fixed throttle such as a capillary tube or an orifice.

  Furthermore, the throttle opening of the fixed throttle 19 in this embodiment passes through the variable throttle mechanism 14 when the superheat degree of the refrigerant on the outlet side of the first evaporator 16 reaches a predetermined value, and further passes through the nozzle portion 15a of the ejector 15. The flow rate ratio η between the refrigerant flow rate Gnoz and the refrigerant flow rate Ge sucked into the refrigerant suction port 15b of the ejector 15 is designed in advance so as to be the optimum flow rate ratio ηmax shown in FIG. Here, η = Ge / Gnoz, and the optimum flow rate ratio ηmax is a flow rate ratio at which the cooling capacity Qer of the entire system is maximized.

  In such a design, the refrigerant passage area of the nozzle portion 15a of the ejector 15 and the mixing portion with respect to the throttle opening degree of the variable throttle mechanism 14 when the superheat degree of the refrigerant on the outlet side of the first evaporator 16 reaches a predetermined value. This can be realized by designing the dimensions of 15c and the diffuser portion 15d and the throttle opening of the fixed throttle 19 to appropriate values. This design also takes into account the pressure loss of the passage through which the refrigerant passing through the variable throttle mechanism 14 and the refrigerant passage (branch passage 18) through which the refrigerant passes through the fixed throttle 19 flow.

  The second evaporator 20 is a heat absorber that evaporates the refrigerant and exerts an endothermic effect. Furthermore, in this embodiment, the first evaporator 16 and the second evaporator 20 are assembled into an integral structure. Specifically, the constituent parts of the first evaporator 16 and the second evaporator 20 are made of aluminum and joined into an integral structure by brazing.

  Therefore, the air blown by the above-described evaporator blower 16a flows as indicated by an arrow A, and is first cooled by the first evaporator 16 and then cooled by the second evaporator 20. ing. That is, the same cooling target space is cooled by the first evaporator 16 and the second evaporator 20.

  The air conditioning control device 21 includes a known microcomputer including a CPU, a ROM, a RAM, and the like and peripheral circuits thereof. The air conditioning control device 21 performs various calculations and processes based on the control program stored in the ROM to control the operations of the various devices 11a, 11b, 12b, 16b and the like.

  Further, the air conditioning controller 21 receives detection signals from various sensor groups and various operation signals from an operation panel (not shown). Specifically, an outside air sensor that detects an outside air temperature (a temperature outside the passenger compartment) is provided as the sensor group. The operation panel is provided with a temperature setting switch for setting the cooling temperature of the space to be cooled.

  Next, the operation of this embodiment will be described with the above configuration. When the electromagnetic clutch 11a is energized by the control output of the air conditioning control device 21 and the electromagnetic clutch 11a is in the connected state, the rotational driving force is transmitted to the compressor 11 from the vehicle travel engine. When the control current In is output from the air conditioning controller 21 to the electromagnetic capacity control valve 11b based on the control program, the compressor 11 sucks and compresses the gas-phase refrigerant and discharges it.

  The high-temperature and high-pressure gas-phase refrigerant compressed and discharged from the compressor 11 flows into the radiator 12. In the radiator 12, the high-temperature and high-pressure refrigerant is cooled and condensed by the outside air. The high-pressure refrigerant after heat dissipation that has flowed out of the radiator 12 is separated into a gas-phase refrigerant and a liquid-phase refrigerant in the liquid receiver 13, and the liquid-phase refrigerant that has flowed out of the liquid receiver 13 is variable in the branch portion Z. The refrigerant flow is divided into a refrigerant flow toward the throttle mechanism 14 and a refrigerant flow toward the branch passage 18.

  The refrigerant flow toward the variable throttle mechanism 14 is decompressed and adjusted in flow rate by the variable throttle mechanism 14 and flows into the ejector 15. Here, the variable throttle mechanism 14 adjusts the flow rate of refrigerant passing through the variable throttle mechanism 14 so that the degree of superheat of the refrigerant on the outlet side of the first evaporator 16 approaches a predetermined value. That is, the refrigerant flow rate Gnoz is adjusted.

  The refrigerant flow flowing into the ejector 15 is further decompressed and expanded by the nozzle portion 15a. Accordingly, the pressure energy of the refrigerant is converted into velocity energy at the nozzle portion 15a, and the refrigerant is ejected at a high velocity from the nozzle outlet of the nozzle portion 15a. The refrigerant (gas phase refrigerant) after passing through the second evaporator 20 in the branch passage 18 is sucked from the refrigerant suction port 15b by the refrigerant suction action at this time.

  The refrigerant ejected from the nozzle portion 15a and the refrigerant sucked into the refrigerant suction port 15b are mixed in the mixing portion 15c on the downstream side of the nozzle portion 15a and flow into the diffuser portion 15d. In the diffuser portion 15d, the speed (expansion) energy of the refrigerant is converted into pressure energy due to the expansion of the passage area, so that the pressure of the refrigerant rises.

  Then, the refrigerant that has flowed out of the diffuser portion 15 d of the ejector 15 flows into the first evaporator 16. In the first evaporator 16, the low-temperature low-pressure refrigerant absorbs heat from the blown air of the evaporator blower 16a and evaporates. The gas-phase refrigerant after passing through the first evaporator 16 flows into the internal heat exchanger 17 and exchanges heat with the high-temperature and high-pressure refrigerant that flows into the branch passage 18 at the branch portion Z. And the gaseous-phase refrigerant | coolant which flowed out from the internal heat exchanger 17 is suck | inhaled by the compressor 11, and is compressed again.

  On the other hand, the refrigerant flow that has flowed into the branch passage 18 flows into the internal heat exchanger 17 and performs heat exchange with the low-temperature and low-pressure gas-phase refrigerant that has flowed out of the first evaporator 16 as described above. Then, the refrigerant cooled by the internal exchanger 19 is decompressed by the fixed throttle 19 to become a low-pressure refrigerant, and this low-pressure refrigerant flows into the second evaporator 20.

  In the second evaporator 20, the low-pressure refrigerant that has flowed in absorbs heat from the blown air cooled by the first evaporator 16 and evaporates. The gas-phase refrigerant after passing through the second evaporator 20 is sucked into the ejector 15 from the refrigerant suction port 15b. Here, the refrigerant flow rate Ge sucked into the refrigerant suction port 15b of the ejector 15 is designed so that the throttle opening degree of the fixed throttle 19 is previously set to a predetermined amount as described above, so that the flow rate ratio η with Gnoz is the optimum flow rate. The flow rate is close to the ratio ηmax.

  The vapor-phase refrigerant evaporated in the second evaporator 20 is sucked from the refrigerant suction port 15b of the ejector 15 and mixed with the liquid-phase refrigerant that has passed through the nozzle portion 15a in the mixing unit 15c. Inflow.

  As described above, in this embodiment, the refrigerant on the downstream side of the diffuser portion 15d of the ejector 15 can be supplied to the first evaporator 16, and the refrigerant on the branch passage 18 side can also be supplied to the second evaporator 20 through the fixed throttle 19. The first evaporator 16 and the second evaporator 20 can exhibit a cooling action at the same time.

  At that time, the refrigerant evaporation pressure of the first evaporator 16 is the pressure after the pressure is increased by the diffuser portion 15 d, while the outlet side of the second evaporator 20 is connected to the refrigerant suction port 15 b of the ejector 15. The lowest pressure immediately after the pressure reduction at the nozzle portion 15a can be applied to the second evaporator 20. Thereby, the refrigerant evaporation pressure (refrigerant evaporation temperature) of the second evaporator 20 can be made lower than the refrigerant evaporation pressure (refrigerant evaporation temperature) of the first evaporator 16.

  Further, the amount of compression work of the compressor 11 can be reduced by the amount by which the suction pressure of the compressor 11 can be increased by the pressure increasing action at the diffuser portion 15d of the ejector 15, and a power saving effect can be exhibited.

  Furthermore, the variable throttle mechanism 14 in this embodiment adjusts the refrigerant flow rate Gnoz so that the degree of superheat of the refrigerant on the outlet side of the first evaporator 16 approaches a predetermined value. As a result, the flow rate ratio η is adjusted so as to approach the optimum flow rate ratio ηmax that increases the cooling capacity of the entire system, so that the entire cycle can be operated while exhibiting a high cooling capacity.

  Moreover, since the superheat degree control of the refrigerant is performed on the outlet side of the first evaporator 16, it is possible to prevent the liquid phase refrigerant from returning to the compressor 11 and to ensure the stability of the cycle.

(First embodiment)
In the above-described embodiment of the present invention, the variable throttle mechanism 14 is disposed between the branching portion Z and the ejector 15. However, in this embodiment, the variable throttle mechanism 14 is eliminated as shown in FIG. Thus, the variable throttle mechanism 31 is provided between the liquid receiver 13 and the branch portion Z. In other words, the variable throttle mechanism 31 of the present embodiment is disposed in the refrigerant passage from the radiator 12 outlet side to the branching section Z inlet side. In FIG. 2, the same or equivalent parts as those in the above-mentioned premise are denoted by the same reference numerals. The same applies to the following drawings.

  The variable throttle mechanism 31 is a temperature expansion valve that adjusts the refrigerant flow rate so that the degree of superheat of the refrigerant on the outlet side of the first evaporator 16 approaches a predetermined value, and the configuration of the temperature expansion valve is a premise of the present invention. It is the same as the form. That is, the valve body of the variable throttle mechanism 31 is means for adjusting the refrigerant flow rate of the entire cycle, and the temperature sensing cylinder and the pressure equalizing pipe of the variable throttle mechanism 31 are the first evaporator 16 as physical quantities related to the refrigerant state in the cycle. It is means for detecting a physical quantity related to the degree of superheat of the outlet side refrigerant.

  Further, in the present embodiment, the flow rate ratio η becomes the optimum flow rate ratio ηmax at the refrigerant flow rate passing through the variable throttle mechanism 31 when the superheat degree of the refrigerant on the outlet side of the first evaporator 16 reaches a predetermined value. The refrigerant passage area of the nozzle portion 15a of the ejector 15 and the throttle opening of the fixed throttle 19 are designed in advance to a predetermined amount. The other cycle configurations are the same as the preconditions of the present invention.

  Therefore, when the cycle of the present embodiment is operated, the variable throttle mechanism 31 adjusts the flow rate of the refrigerant passing through the variable throttle mechanism 31 so that the degree of superheat of the refrigerant on the outlet side of the first evaporator 16 approaches a predetermined value. As a result, since the flow rate ratio η is adjusted so as to approach the optimum flow rate ratio ηmax, the cycle as a whole can be operated while exhibiting a high cooling capacity, as in the premise of the present invention. Furthermore, since the superheat degree control of the refrigerant is performed on the outlet side of the first evaporator 16, it is possible to prevent the liquid phase refrigerant from returning to the compressor 11 and to ensure the stability of the cycle.

(Second Embodiment)
In the first embodiment, the example using the variable throttle mechanism 31 has been described. However, in the present embodiment, as shown in FIG. 3, the liquid receiver 13 and the variable throttle mechanism 31 are abolished and the first evaporator 16 is removed. An accumulator 32 that separates the liquid-phase refrigerant and the gas-phase refrigerant is provided on the downstream side, and a variable throttle mechanism 35 is provided between the radiator 12 and the branch portion Z.

  The variable throttle mechanism 35 is a temperature type expansion valve that adjusts the refrigerant flow rate so that the degree of superheat of the refrigerant on the outlet side of the second evaporator 20 approaches a predetermined value. The configuration of the temperature type expansion valve is the same as that of the first embodiment. In this embodiment, the temperature sensing cylinder and the pressure equalizing pipe of the variable throttle mechanism 35 are variable according to the temperature and pressure of the refrigerant on the outlet side of the second evaporator 20. The valve body of the throttle mechanism 35 is displaced. That is, the valve body of the variable throttle mechanism 35 is a means for adjusting the refrigerant flow rate of the entire cycle, and the temperature sensing cylinder and the pressure equalizing pipe of the variable throttle mechanism 35 are the second evaporator 20 as physical quantities related to the refrigerant state in the cycle. It is means for detecting a physical quantity related to the degree of superheat of the outlet side refrigerant.

  Further, in the present embodiment, the flow rate ratio η is set to the optimum flow rate ratio ηmax at the flow rate of the refrigerant passing through the variable throttle mechanism 35 when the superheat degree of the refrigerant on the outlet side of the second evaporator 20 reaches a predetermined value. The refrigerant passage area of the nozzle portion 15a of the ejector 15 and the throttle opening of the fixed throttle 19 are designed in advance to a predetermined amount. Other cycle configurations are the same as those in the first embodiment.

  Therefore, when the cycle of this embodiment is operated, the variable throttle mechanism 35 adjusts the flow rate of the refrigerant passing through the variable throttle mechanism 35 so that the degree of superheat of the refrigerant on the outlet side of the second evaporator 20 approaches a predetermined value. As a result, the flow rate ratio η is adjusted so as to approach the optimum flow rate ratio ηmax, so that the entire cycle can be operated while exhibiting a high cooling capacity in exactly the same manner as in the first embodiment. Further, since the accumulator 32 is provided on the outlet side of the first evaporator 16, it is possible to prevent the liquid phase refrigerant from returning to the compressor 11 and to ensure the stability of the cycle.

(Third embodiment)
In the second embodiment, an example in which the variable throttle mechanism 35 is employed has been described. However, in this embodiment, as shown in FIG. 4, the variable throttle mechanism 35 is eliminated and the temperature Tc of the outlet side refrigerant of the radiator 12 is eliminated. A temperature sensor 55 for detecting the pressure and a pressure sensor 56 for detecting the pressure Pc are provided. Further, an electric variable throttle mechanism 38 is provided between the radiator 12 and the branch portion Z and downstream of the temperature sensor 55 and the pressure sensor 56. Is provided.

  The electric variable throttle mechanism 38 has a valve mechanism that adjusts the refrigerant passage area and a stepping motor that is rotationally driven by a control signal (pulse signal) output from the air conditioning controller 21. When the stepping motor rotates, the valve mechanism This is a flow rate adjusting valve capable of continuously adjusting the refrigerant passage area by displacing the valve body. The air conditioning control device 21 calculates the degree of supercooling of the refrigerant on the outlet side of the radiator 12 based on the detection values of the temperature sensor 55 and the pressure sensor 56, and the degree of subcooling of the refrigerant on the outlet side of the radiator 12 approaches a predetermined value. In this way, the control signal (pulse signal) is output to adjust the refrigerant passage area of the electric variable throttle mechanism 38.

  That is, the electric variable throttle mechanism 38 in the present embodiment is a means for adjusting the refrigerant flow rate of the entire cycle, and the temperature sensor 55 and the pressure sensor 56 are refrigerants at the outlet side of the radiator 12 as physical quantities related to the refrigerant state in the cycle. It is means for detecting a physical quantity related to the degree of supercooling.

  Further, in the present embodiment, the flow rate ratio η becomes the optimum flow rate ratio ηmax in the refrigerant flow rate that passes through the electric variable throttle mechanism 38 when the degree of supercooling of the refrigerant on the outlet side of the radiator 12 reaches a predetermined value. In addition, the refrigerant passage area of the nozzle portion 15a of the ejector 15 and the throttle opening of the fixed throttle 19 are designed in advance to a predetermined amount. Other cycle configurations are the same as those of the second embodiment.

  Therefore, when the cycle of the present embodiment is operated, the electric variable throttle mechanism 38 causes the refrigerant flow rate that passes through the electric variable throttle mechanism 38 so that the degree of supercooling of the refrigerant on the outlet side of the radiator 12 approaches a predetermined value. Control. As a result, the flow rate ratio η is adjusted so as to approach the optimum flow rate ratio ηmax, so that the same effect as in the second embodiment can be obtained.

(Fourth embodiment)
In the third embodiment, the temperature sensor 55 and the pressure sensor 56 are arranged on the outlet side of the radiator 12, but in this embodiment, the temperature sensor 55 and the pressure sensor 56 are eliminated as shown in FIG. A temperature sensor 57 for detecting the temperature Tsi of the refrigerant on the suction side of the compressor 11 between the internal heat exchanger 17 connected to the outlet side of the evaporator 16 and the compressor 11 and a pressure sensor 58 for detecting the pressure Psi are provided. 11 is provided with a tachometer 11c. Furthermore, the electric variable throttle mechanism 38 is abolished, and the electric variable throttle mechanism 40 is provided between the radiator 12 and the branch point Z.

  The tachometer 11c detects the rotation speed Nc of the compressor 11, and is a magnetic rotation speed sensor that detects a magnetic flux change due to rotation of the compressor 11 by a Hall element or an MRE element. Alternatively, the engine speed Ne may be detected by the engine ECU to calculate the speed Nc.

  The detected values of the temperature sensor 57, the pressure sensor 58, and the tachometer 11c are input to the air conditioning control device 21, and the air conditioning control device 21 calculates the compressor 11 intake refrigerant density based on Tsi and Psi. The refrigerant discharge flow rate of the compressor 11 is calculated from the suction refrigerant density, the rotation speed Nc, and the control current In. The air conditioning controller 21 stores in advance a refrigerant passage area corresponding to the output (pulse count number) of a control signal (pulse signal) output to the electric variable throttle mechanism 40.

  In the present embodiment, the refrigerant passage area of the nozzle portion 15a of the ejector 15 is set so that the flow rate ratio η becomes the optimum flow rate ratio ηmax when the flow rate of the refrigerant passing through the electric variable throttle mechanism 40 reaches a predetermined value. The apertures of the equal and fixed throttles 19 are designed in advance to a predetermined amount. Other cycle configurations are the same as those of the third embodiment.

  That is, in this embodiment, the electric variable throttle mechanism 40 is a means for adjusting the refrigerant flow rate of the entire cycle, and the temperature sensor 57, the pressure sensor 58, and the tachometer 11c are refrigerants as physical quantities related to the refrigerant state in the cycle. It is a means for detecting a physical quantity related to the discharge flow rate.

  Therefore, when the cycle of the present embodiment is operated, the air conditioning control device 21 calculates the refrigerant discharge flow rate of the compressor 11, and the control signal of the electric variable throttle mechanism 40 stored in advance and the calculated refrigerant discharge flow rate ( Based on the refrigerant passage area according to the output (pulse count number) of the pulse signal), the refrigerant passage area of the electric variable throttle mechanism 40 is set so that the refrigerant flow rate passing through the electric variable throttle mechanism 40 becomes a predetermined value. adjust. As a result, the flow rate ratio η approaches the optimal flow rate ratio ηmax, so that the same effect as in the third embodiment can be obtained.

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

  (1) In the above embodiment, the present invention is applied to a vehicular refrigeration apparatus, and the cooling target space of the first evaporator 16 and the second evaporator 20 is the same. And the cooling target space of the second evaporator 20 may be different.

  For example, the first evaporator 16 may use a front seat side region in the vehicle interior as a cooling target space, and the second evaporator 20 may use the rear seat side region in the vehicle interior as a cooling target space. As described above, when the space to be cooled is different, a dedicated evaporator blower may be provided for each evaporator, and the air flow rate of each evaporator blower may be individually controlled. Accordingly, the flow rate ratio η can be adjusted by adjusting the refrigerant flow rate Gnoz that substantially passes through the nozzle portion 15a of the ejector 15 and the refrigerant flow rate Ge sucked into the refrigerant suction port 15b of the ejector 15.

  (2) In the first to fourth embodiments, a variable displacement compressor is used as the compressor 11, and the electric capacity control valve 11 b is controlled by the air conditioning control device 21 to control the refrigerant discharge capacity of the compressor 11. However, a fixed displacement compressor is used, and the ratio of the operation state and the non-operation state (operation rate) of the fixed displacement compressor is controlled by an electromagnetic clutch to control the refrigerant discharge capacity of the compressor. It may be.

  An electric compressor may be used as the compressor. In this case, the refrigerant discharge capacity can be controlled by controlling the rotational speed of the electric compressor 11.

  (3) In the first embodiment, the degree of superheat of the refrigerant on the outlet side of the first evaporator 16 is detected by the temperature sensing cylinder and the pressure equalizing pipe of the temperature type expansion valve. However, the detecting means is not limited to this.

  For example, it can be estimated from the first evaporator 16 refrigerant evaporation temperature or pressure and the first evaporator 16 outlet side refrigerant temperature.

  Moreover, it can also estimate from the 1st evaporator 16 refrigerant | coolant evaporation temperature or pressure, and the 1st evaporator 16 blowing air temperature. This is because the cooling capacity of the first evaporator 16 decreases due to the increase in the degree of superheat, and the temperature of the air blown from the first evaporator 16 increases.

  The degree of superheat can also be estimated from the intake temperature of the first evaporator 16 and the refrigerant temperature on the outlet side of the first evaporator 16. The degree of superheat can also be estimated from the first evaporator 16 intake air temperature and the first evaporator 16 blown air temperature.

  Further, the degree of superheat can be estimated only with the first evaporator 16 intake air temperature. This is because when the air in the cooling target space (freezer) is circulated and cooled by an evaporator like a refrigerator, the first evaporator 16 blowout temperature rises when the superheat degree becomes excessive, and as a result, the inside of the freezer This is because the intake air temperature of the first evaporator 16 also rises because the temperature rises.

  Therefore, the degree of superheat of the refrigerant on the outlet side of the first evaporator 16 can be detected even using the means for detecting the physical quantity.

  (4) In the second embodiment, the degree of superheat of the refrigerant on the outlet side of the second evaporator 20 is detected by the temperature sensing cylinder and the pressure equalizing pipe of the temperature type expansion valve. However, the detecting means is not limited to this.

  For example, it can be estimated from the second evaporator 20 refrigerant evaporation temperature or pressure and the second evaporator 20 outlet side refrigerant temperature, similarly to the degree of superheat of the first evaporator 16 outlet refrigerant. Furthermore, you may estimate with the 2nd evaporator 20 refrigerant | coolant evaporation temperature or pressure, and the 2nd evaporator 20 blowing air temperature.

  In addition, a combination of the second evaporator 20 suction air temperature and the second evaporator 20 outlet side refrigerant temperature, a combination of the second evaporator 20 suction air temperature and the second evaporator 20 blown air temperature, the second The degree of superheat can be estimated only with the evaporator 20 intake air temperature.

  Therefore, the degree of superheat of the refrigerant on the outlet side of the second evaporator 20 can be detected using the means for detecting the physical quantity.

  (5) In the third embodiment, the degree of supercooling of the radiator 12 outlet side refrigerant is detected by the temperature sensors 55 and 56, but the means for detecting the degree of supercooling of the radiator 12 outlet side refrigerant is limited to this. It is not something.

  For example, the combination of the refrigerant condensing temperature (refrigerant pressure) in the radiator 12 and the radiator 12 outlet side refrigerant temperature, the combination of the radiator 12 intake air temperature and the radiator 12 outlet refrigerant temperature, and the radiator 12 outlet side refrigerant. It can be estimated from the dryness of.

  Therefore, even if the means for detecting the physical quantity is used, the degree of supercooling of the refrigerant on the outlet side of the radiator 12 can be detected.

  (6) In the above embodiment, temperature type expansion valves are used as the variable throttle mechanisms 31 and 35, and stepping motor-driven flow rate adjusting valves are used as the electric variable throttle mechanisms 38 and 40. However, a plurality of fixed throttles having different characteristics are used. You may use the variable aperture mechanism which switches and uses.

  Further, the above-described variable aperture mechanism, electric variable aperture mechanism, and fixed aperture of each of the above embodiments may be used in combination.

  (7) In the above embodiment, the two evaporators, the first evaporator 16 and the second evaporator 20, are used, but three or more evaporators may be used by further increasing the number of evaporators. .

  For example, in the configuration of the first embodiment, a second branch passage that connects the internal heat exchanger 17 of the branch passage 18 and the fixed throttle 19 to the outlet side of the first evaporator 16 is provided, and the second branch passage is provided. A fixed throttle and a third evaporator may be provided.

  In this case, when the superheat degree of the refrigerant on the outlet side of the first evaporator 16 reaches a predetermined value, the refrigerant flow rate that passes through the variable throttle mechanism 14, the refrigerant flow rate that passes through the fixed throttle 19, and the second branch passage are arranged. The throttle opening degree of the fixed throttles arranged in the fixed throttle 19 and the second branch passage may be set so that the flow rate of the refrigerant passing through the fixed throttle increases the cooling capacity Qer of the entire system.

  (8) Although the refrigeration cycle of the above embodiment shows an example of a subcritical cycle in which the high pressure does not exceed the critical pressure of the refrigerant, the refrigeration cycle may be applied to a supercritical cycle in which the high pressure exceeds the critical pressure of the refrigerant.

  For example, in the configuration of the third embodiment, when the compressor 11 increases the refrigerant until it reaches the supercritical pressure, the air conditioning control device 21 cannot calculate the degree of supercooling. In this case, the air conditioning controller 21 controls the electric variable throttle mechanism 38 so that the pressure of the refrigerant on the outlet side of the radiator 12 becomes a predetermined value.

  Then, when the pressure of the refrigerant on the outlet side of the radiator 12 reaches a predetermined value, the refrigerant flow rate Gnoz passing through the electric variable throttle mechanism 38 and further passing through the nozzle portion 15a of the ejector 15, and the refrigerant suction port 15b of the ejector 15 By designing the throttle opening of the fixed throttle 19 to a predetermined amount in advance so that the flow rate ratio η with respect to the refrigerant flow rate Ge sucked into Even if it exists, the completely same effect as 3rd Embodiment can be acquired.

  (9) In the above-described embodiment, the ejector-type refrigeration cycle according to the present invention is applied to a vehicle refrigeration apparatus. However, the ejector-type refrigeration cycle is applied to a vapor compression cycle such as a stationary refrigerator, a stationary freezer, a cooling apparatus, and a heat pump cycle for a water heater. You may apply.

  (10) In the above embodiment, a chlorofluorocarbon refrigerant is used as the refrigerant, but a CO2 refrigerant and an HC refrigerant may be used. Freon is a general term for organic compounds composed of carbon, fluorine, chlorine and hydrogen, and is widely used as a refrigerant.

  Fluorocarbon refrigerants include HCFC (hydro-chloro-fluoro-carbon) refrigerants, HFC (hydro-fluoro-carbon) refrigerants, etc. These are refrigerants called substitute chlorofluorocarbons because they do not destroy the ozone layer. is there.

  The HC (hydrocarbon) refrigerant is a refrigerant substance that contains hydrogen and carbon and exists in nature. Examples of the HC refrigerant include R600a (isobutane) and R290 (propane).

  (11) In the above-described forms and embodiments, the example in which the fixed throttle 19 is used as the throttle means disposed in the branch passage 18 has been described. Of course, a variable throttle mechanism may be used. As this variable throttle mechanism, for example, a temperature type expansion valve that adjusts the refrigerant flow rate so that the superheat degree of the refrigerant on the outlet side of the first and second evaporators 16 and 20 approaches a predetermined value may be adopted. Further, for example, an electric variable throttle mechanism that adjusts the refrigerant flow rate so that the degree of supercooling of the refrigerant on the outlet side of the radiator 12 approaches a predetermined value may be employed.

It is a cycle block diagram which shows the ejector-type refrigerating cycle of the form used as the premise of this invention. It is a cycle lineblock diagram showing the ejector type refrigerating cycle of a 1st embodiment. It is a cycle block diagram which shows the ejector type refrigeration cycle of 2nd Embodiment. It is a cycle block diagram which shows the ejector type refrigeration cycle of 3rd Embodiment. It is a cycle lineblock diagram showing the ejector type freezing cycle of a 4th embodiment. It is a cycle block diagram which shows the ejector-type refrigerating cycle of a prior art. It is a cycle block diagram which shows the ejector type refrigeration cycle of a prior application example. It is explanatory drawing which shows the relationship between the flow rate ratio (eta) and the cooling capacity Qer in a prior art and a prior application example.

Explanation of symbols

DESCRIPTION OF SYMBOLS 11 Compressor 12 Radiator 15 Ejector 15a Nozzle part 15b Refrigerant suction port 16 First evaporator 18 Branch passage 19 Fixed throttle 20 Second evaporator 31, 35 Variable throttle mechanism 38, 40 Electric variable throttle mechanism

Claims (7)

A compressor (11) for sucking and compressing refrigerant;
A radiator (12) that radiates heat of the high-pressure refrigerant discharged from the compressor (11);
A branch part (Z) for branching the flow of the refrigerant flowing out of the radiator (12);
The refrigerant is sucked from the refrigerant suction port (15b) by the high-speed refrigerant flow injected from the nozzle portion (15a) that decompresses and expands one of the refrigerant branched at the branch portion (Z), and the refrigerant suction port ( An ejector (15) that mixes the suction refrigerant from 15b) with the high-speed refrigerant flow, decelerates the mixed refrigerant flow and increases the pressure of the refrigerant flow;
A first evaporator (16) for evaporating the refrigerant flowing out of the ejector (15);
A branch passage (18) for guiding the other refrigerant branched at the branch portion (Z) to the refrigerant suction port (15b);
Throttle means (19) disposed in the branch passage (18) for reducing the pressure of the refrigerant to adjust the flow rate;
A second evaporator (20) disposed downstream of the throttle means (19) and evaporating the refrigerant,
Furthermore, the variable throttle means (31, 35,) disposed in the refrigerant passage from the radiator (12) outlet side to the branch portion (Z) inlet side to decompress and expand the refrigerant flowing out from the radiator (12). 38, 40). An ejector refrigeration cycle comprising:   The variable throttling means (31, 35, 38, 40) includes a refrigerant state in a cycle, a cooling target space temperature of the first evaporator (16) and the second evaporator (20), and a cooling target space. The ejector refrigeration cycle according to claim 1, wherein the refrigerant flow rate of the entire cycle is adjusted based on a physical quantity related to at least one of the ambient temperatures. The physical quantity is a physical quantity related to the degree of superheat of the refrigerant on the outlet side of the first evaporator (16),
The variable throttle means (31) adjusts the refrigerant flow rate of the entire cycle so that the degree of superheat of the refrigerant on the outlet side of the first evaporator (16) approaches a predetermined value. The ejector refrigeration cycle described. The physical quantity is a physical quantity related to the degree of superheat of the refrigerant on the outlet side of the second evaporator (20),
The variable throttle means (35) adjusts the refrigerant flow rate of the entire cycle so that the degree of superheat of the refrigerant on the outlet side of the second evaporator (20) approaches a predetermined value. The ejector refrigeration cycle described. The physical quantity is a physical quantity related to the degree of supercooling of the refrigerant on the outlet side of the radiator (12),
The said variable throttle means (38) adjusts the refrigerant | coolant flow rate of the whole cycle so that the subcooling degree of the refrigerant | coolant (12) exit side refrigerant | coolant may approach the predetermined value. Ejector type refrigeration cycle. The high-pressure refrigerant is designed to be pressurized above the critical pressure,
The physical quantity is the temperature and pressure of the refrigerant on the outlet side of the radiator (12),
The ejector according to claim 2, wherein the variable throttle means (38) adjusts the refrigerant flow rate of the entire cycle so that the pressure of the refrigerant on the outlet side of the radiator (12) approaches a predetermined value. Refrigeration cycle. The physical quantity is a physical quantity related to the refrigerant discharge flow rate of the compressor (11),
The ejector refrigeration cycle according to claim 2, wherein the variable throttle means (40) adjusts the refrigerant flow rate of the entire cycle so that the refrigerant flow rate of the entire cycle approaches a predetermined value.

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