JP4725592B2 - Refrigeration cycle equipment - Google Patents

Description

  The present invention relates to a refrigeration cycle apparatus including a two-stage booster compressor.

  Conventionally, Patent Document 1 discloses a refrigeration cycle apparatus including a two-stage booster compressor. The two-stage booster compressor disclosed in Patent Document 1 includes a first compression mechanism that sucks and compresses refrigerant from outside the housing that forms the outer shell of the compressor, and an intermediate-pressure refrigerant that is pressurized by the first compression mechanism ( And a second compression mechanism for further compressing the medium pressure refrigerant) and discharging it to the outside of the housing.

Furthermore, in the two-stage booster compressor disclosed in Patent Document 1, the housing is filled with the high-pressure refrigerant that has been pressurized by the second compression mechanism by filling the housing with the medium-pressure refrigerant that has been pressurized by the first compression mechanism. In contrast, the pressure resistance required for the housing is reduced. As a result, the thickness of the housing is reduced to reduce the weight of the compressor as a whole.
JP-A-9-159297

  By the way, in this type of refrigeration cycle apparatus, in order to improve the coefficient of performance (COP) of the cycle, the high pressure Pd of the high pressure refrigerant from the discharge port of the compressor to the inlet of the decompression means approaches the predetermined target pressure Pt. So that it is controlled. Specifically, a pressure control valve, which is a decompression means, is arranged downstream of the radiator that dissipates the high-pressure refrigerant discharged from the compressor, and means such as adjusting the valve opening of the pressure control valve, The high pressure Pd is brought close to the target pressure Pt.

  However, in the refrigeration cycle apparatus including the two-stage booster compressor as in the refrigeration cycle apparatus of Patent Document 1, even if the high pressure Pd of the high pressure refrigerant is controlled to approach the target pressure Pt during the cycle operation, The medium pressure Pm of the medium pressure refrigerant that has been boosted by the one compression mechanism cannot be controlled.

The reason for this is that the suction pressure Ps and the intermediate pressure Pm in the two-stage booster compressor have a relationship represented by the following formula F1.
Ps × V1κ = Pm × (η1 × V2) κ (F1)
V1 is the discharge volume of the first compression mechanism, V2 is the discharge volume of the second compression mechanism, and η1 is the volumetric efficiency of the first compression mechanism.

That is, the intermediate pressure Pm can be expressed by the following formula F2 from the formula F1.
Pm = Ps × (V1 / (η1 × V2)) κ (F2)
As is apparent from Formula F2, even if the high pressure Pd is controlled to approach the target pressure Pt, if the suction pressure Ps increases, the intermediate pressure Pm also increases.

  Therefore, as shown in FIG. 12, as the suction pressure Ps increases, the intermediate pressure Pm also increases. On the other hand, since the high pressure Pd is controlled so as to approach the target pressure Pt, the pressure difference between the medium pressure Pm and the high pressure Pd decreases as the suction pressure Ps increases. For this reason, when the suction refrigerant Ps rises, the intermediate pressure Pm may rise to the vicinity of the high pressure Pd.

  FIG. 12 is a graph showing the relationship between the suction pressure Ps of the two-stage booster compressor, the intermediate pressure Pm, and the high pressure Pd.

  And in the refrigerating cycle device of patent documents 1, if medium pressure Pm in a housing rises to the neighborhood of high pressure Pd, the pressure demand equivalent to the case where the inside of a housing is filled up with high pressure refrigerant will be demanded for a housing. . Therefore, the weight reduction effect as the whole compressor by reducing the thickness of the housing described above cannot be obtained sufficiently.

  In view of the above points, an object of the present invention is to suppress an unnecessary increase in pressure of a medium-pressure refrigerant that has been pressurized by a first compression mechanism of a two-stage booster compressor.

In order to achieve the above object, the present invention provides a first compression mechanism (19) that sucks and compresses a refrigerant, and a second compression that compresses and discharges a medium-pressure refrigerant pressurized by the first compression mechanism (19). A two-stage booster compressor (14) having a mechanism (20), a radiator (15) that dissipates heat generated by the refrigerant discharged from the two-stage booster compressor (14) to a heat exchange target fluid, and heat radiation Decompression means (16, 30) for decompressing and expanding the refrigerant flowing out of the evaporator (15), an evaporator (17) for evaporating the refrigerant decompressed by the decompression means (16, 30), and a two-stage booster compressor (14) a target pressure determining means (S4) for determining a target pressure (Pt) of the high pressure (Pd) of the high-pressure refrigerant from the discharge port to the pressure reducing means (16, 30) inlet, the high pressure (Pd) being Refrigeration cycle controlled to approach the target pressure (Pt) A location,
Furthermore, it comprises suction pressure detection means (23, 25, 26) for detecting a physical quantity having a correlation with the suction pressure (Ps) of the refrigerant sucked into the two-stage booster compressor (14), and the suction pressure detection means (23 25, 26), when the detected values (Tam, Twi, Te) are equal to or higher than a predetermined set value (TamA, TwiA, TeA), the detected values (Tam, Twi, Te) increase. Along with this, the high pressure (Pd) is reduced so that the medium pressure (Pm) of the medium pressure refrigerant is less than a predetermined pressure .

  According to this, when the detection values (Tam, Twi, Te) detected by the suction pressure detection means (23, 25, 26) become equal to or higher than a predetermined set value (TamA, TwiA, TeA) As the values (Tam, Twi, Te) increase, the high pressure (Pd) is decreased, and by increasing η1 in the above formula F2, the first compression mechanism (Ps) is increased even if the suction pressure (Ps) is increased. The intermediate pressure of the medium pressure refrigerant whose pressure has been increased in 19) can be reliably reduced.

  That is, as described with reference to FIG. 12, even if the pressure difference between the intermediate pressure and the high pressure (Pd) becomes smaller as the suction pressure (Ps) increases, the high pressure (Pd) is reduced. Then, the intermediate pressure can be reliably reduced. As a result, it is possible to prevent the intermediate pressure from increasing unnecessarily as the suction pressure (Ps) of the two-stage booster compressor (14) increases.

  Further, even when the housing accommodating the first compression mechanism (19) and the second compression mechanism (20) is filled with the high-pressure refrigerant or the medium-pressure refrigerant, the high pressure (Pd) accompanying the increase in the suction pressure. Since an unnecessary increase in the intermediate pressure can be suppressed, the pressure resistance of the housing need not be increased more than necessary. As a result, the above-described compressor (11) as a whole can be sufficiently lightened.

  According to a second aspect of the present invention, in the refrigeration cycle apparatus according to the first aspect, a high pressure detecting means (28) for detecting a high pressure (Pd) and a high pressure detecting means for detecting the temperature of the high pressure refrigerant. (24).

  Specifically, when the detection values (Tam, Twi, Te) of the suction pressure detection means (23, 25, 26) are not equal to or higher than a predetermined set value (TamA, TwiA, TeA). The target high pressure (Pt) can be determined based on the detected temperature detected by the high pressure temperature detecting means (24).

  Further, when the detection values (Tam, Twi, Te) of the suction pressure detection means (23, 25, 26) are equal to or higher than a predetermined set value (TamA, TwiA, TeA), the high pressure detection means (28 ), The high pressure (Pd) can be reduced.

  As in the third aspect of the invention, in the refrigeration cycle apparatus of the first or second aspect, the target pressure determining means (S4) is configured such that the detected values (Tam, Twi, Te) are preset values (TamA). , TwiA, TeA) or higher, the target pressure (Pt) may be reduced as the detection values (Tam, Twi, Te) increase. Thereby, the high pressure (Pd) can be specifically reduced.

  Further, as in the invention according to claim 4, in the refrigeration cycle apparatus according to any one of claims 1 to 3, the discharge capacity control for controlling the refrigerant discharge capacity of the two-stage booster compressor (14). Means (22a), and the discharge capacity control means (22a) detects the detected values (Tam, TwiA, TeA) when the detected values (Tam, Twi, Te) are equal to or higher than a predetermined set value (TamA, TwiA, TeA). The refrigerant discharge capacity may be reduced as Twi, Te) increase. As a result, the high pressure (Pd) can be reliably reduced.

  In the present invention, “reducing the refrigerant discharge capacity” does not simply mean reducing the refrigerant discharge capacity and reducing the flow rate of the refrigerant circulating in the cycle, and does not circulate the refrigerant in the cycle. This means that the operation of the cycle is substantially stopped by setting the state, that is, the flow rate of the refrigerant circulating in the cycle to zero.

  Further, as in the invention according to claim 5, in the refrigeration cycle apparatus according to any one of claims 1 to 3, the throttle opening degree control means for controlling the throttle opening degree of the decompression means (16, 30). (22b), and the throttle opening control means (22b) detects the detected value (Tam, TwiA, TeA) when the detected value (Tam, Twi, Te) is equal to or higher than a predetermined set value (TamA, TwiA, TeA). The throttle opening may be increased as Twi, Te) increase. As a result, the high pressure (Pd) can be reliably reduced.

  As in the sixth aspect of the present invention, the suction pressure detecting means includes a fluid temperature sensor (23) that detects the temperature (Twi) of the heat exchange target fluid that exchanges heat with the high-pressure refrigerant in the radiator (15). Or an outside air temperature sensor (25) that detects the outside air temperature (Tam) of the outdoor air that exchanges heat with the low-pressure refrigerant in the evaporator (17), as in the invention described in claim 7. Further, as in the invention described in claim 8, it may be constituted by an evaporation temperature sensor (26) for detecting the evaporation temperature (Te) of the low-pressure refrigerant in the evaporator (17).

  According to this, since the suction pressure detection means (23, 25, 26) can be configured by an inexpensive temperature sensor with respect to the pressure sensor, the manufacturing cost of the refrigeration cycle apparatus can be reduced.

  In a ninth aspect of the present invention, in the refrigeration cycle apparatus according to any one of the first to eighth aspects, the decompression means is a high-speed refrigerant flow ejected from the nozzle portion (30a) that decompresses and expands the high-pressure refrigerant. It is characterized by comprising an ejector (30) which sucks the refrigerant into the interior and mixes the sucked refrigerant and a high-speed refrigerant flow to increase the pressure.

  According to this, the suction pressure (Ps) of the suction refrigerant of the two-stage booster compressor (14) can be increased by the boosting action of the ejector (30), and the driving power of the two-stage booster compressor (14) can be reduced. Therefore, the COP can be improved.

  As in the tenth aspect of the invention, in the refrigeration cycle apparatus according to any one of the first to ninth aspects, the heat exchange target fluid may be water. Thereby, a refrigerating cycle device is applicable to a hot water supply machine.

  As in the invention according to claim 11, in the refrigeration cycle apparatus according to any one of claims 1 to 10, the refrigerant may be carbon dioxide. Thereby, a supercritical refrigeration cycle in which the high-pressure pressure (Pd) exceeds the critical pressure of the refrigerant can be configured and the temperature of the high-pressure refrigerant can be raised, so that heat is radiated to the heat exchange target fluid by the radiator (15). The amount of heat can be increased.

  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.

(First embodiment)
1st Embodiment of this invention is described with FIGS. In the present embodiment, the refrigeration cycle apparatus of the present invention is applied to a heat pump type hot water heater 10, and FIG. 1 is an overall configuration diagram of the heat pump type hot water heater 10 of the present embodiment.

  The heat pump hot water heater 10 includes a hot water storage tank 11 that stores hot water, a water circulation passage 12 that circulates the hot water in the hot water storage tank 11, and a heat pump cycle 13 that heats the hot water. First, the hot water storage tank 11 is a hot water tank that is made of a metal (for example, stainless steel) having excellent corrosion resistance, has a heat insulating structure, and can keep hot hot water for a long time.

  Hot water stored in the hot water storage tank 11 is discharged from a hot water outlet 11a provided in the upper part of the hot water storage tank 11, mixed with cold water from the water at a temperature control valve (not shown), and then adjusted to a temperature in the kitchen or bath. Hot water is supplied. In addition, tap water is supplied from a water supply port 11 b provided in the lower part of the hot water storage tank 11.

  In the water circulation passage 12, an electric water pump 12a for circulating hot water is disposed. The hot water is supplied from the electric water pump 12a to a water passage 15a of a water-refrigerant heat exchanger 15 to be described later → hot water in the upper part of the hot water storage tank 11. It flows in the order of the water inlet 11d → the hot water storage tank 11 → the hot water outlet 11c at the lower part of the hot water storage tank 11 → the electric water pump 12a.

  The heat pump cycle 13 is a refrigeration cycle in which a compressor 14, a water-refrigerant heat exchanger 15, an electric expansion valve 16, an evaporator 17 and the like are sequentially connected by piping. The heat pump cycle 13 employs carbon dioxide as a refrigerant, and constitutes a supercritical refrigeration cycle in which the high-pressure refrigerant discharged from the compressor 14 has a high pressure equal to or higher than the critical pressure of the refrigerant.

  The compressor 14 sucks, compresses and discharges the refrigerant in the heat pump cycle 13. Details of the compressor 14 will be described with reference to FIG. FIG. 2 is a schematic cross-sectional view of the compressor 14 in the axial direction. As shown in FIG. 2, the compressor 14 is an electric two-stage booster compressor that includes a housing 18, a first compression mechanism 19, a second compression mechanism 20, an electric motor 21, and the like.

  The housing 18 is formed in a substantially cylindrical shape and constitutes an outer shell of the compressor 14, and is a metal (for example, stainless steel) pressure-resistant hermetic housing the first compression mechanism 19, the second compression mechanism 20 and the electric motor 21. It is a container.

  More specifically, the housing 18 includes a cup-shaped motor housing 18a and a compression mechanism housing 18b. Further, the opening end portions of the motor housing 18a and the compression mechanism housing 18b are joined to each other by joining means such as welding in a state where the first compression mechanism 19, the second compression mechanism 20, the electric motor 21 and the like are accommodated therein. Is formed.

  The housing 18 (specifically, the compression mechanism housing 18b) has a single inlet 18c for sucking refrigerant from the outside of the housing 18 into the housing 18 and for discharging the refrigerant from the inside of the housing 18 to the outside. One discharge port 18d is provided.

  The first compression mechanism 19 is a known rotary type (rolling piston) type compression mechanism that eccentrically rotates the pump rotor portion 19 b inside a cylinder portion 19 a fixed to the inner wall of the housing 18. The refrigerant suction port 19 c that sucks the refrigerant in the first compression mechanism 19 is connected to the suction port 18 c of the housing 18. Therefore, the refrigerant sucked from the outside to the inside of the housing 18 is sucked into the first compression mechanism 19.

  On the other hand, the refrigerant discharge port 19 d that discharges the refrigerant in the first compression mechanism 19 is open inside the housing 18. Accordingly, the pressure inside the housing 18 becomes the medium pressure of the medium pressure refrigerant.

  Similar to the first compression mechanism 19, the second compression mechanism 20 is a rotary type compression mechanism that includes a cylinder portion 20 a and a pump rotor portion 20 b. In the second compression mechanism 20, the refrigerant suction port 20 c that sucks the refrigerant is open inside the housing 18. Therefore, the intermediate pressure refrigerant whose pressure has been increased by the first compression mechanism is sucked into the second compression mechanism 20.

  On the other hand, the refrigerant discharge portion 20 d that discharges the refrigerant in the second compression mechanism 20 is connected to the discharge port 18 d of the housing 18. Therefore, the high-pressure refrigerant whose pressure has been increased by the second compression mechanism 20 is discharged from the inside of the housing 18 to the outside. That is, in the compressor 14, the refrigerant flows as shown by the thick arrows in FIG.

  Next, the electric motor 21 has a stator coil 21a fixed to the inner wall of the housing 18, a rotor 21b rotating in the stator coil 21a, and a shaft 21c fixed to the rotor 21b and rotated integrally with the rotor 21b. It is a well-known AC type electric motor configured. Of course, a DC motor may be used.

  The shaft 21 c is rotatably supported by a bearing portion 18 e fixed to the housing 18 and is integrally connected to the pump rotor portion 20 a of the second compression mechanism 20 and the pump rotor portion 19 a of the first compression mechanism 19. ing.

  Accordingly, when electric power is supplied to the electric motor 21, the shaft electric motor 21 is rotated together with the rotor 21 b, the rotational driving force is transmitted to the first and second compression mechanisms 19, 20, and the refrigerant is sucked into the suction port of the compressor 14. 18c → first compression mechanism 19 → second compression mechanism 20 → discharge port 18d in the order of suction, compression and discharge. At this time, as described above, the pressure of the refrigerant (carbon dioxide) is increased until it reaches a critical pressure or higher (supercritical state).

  A refrigerant passage 15 b inlet side of the water-refrigerant heat exchanger 15 is connected to the discharge port 18 d of the compressor 14. The water-refrigerant heat exchanger 15 is a heat exchanger configured to include a water passage 15a through which hot water passes and a refrigerant passage 15b through which high-temperature and high-pressure refrigerant discharged from the compressor 14 passes. 14 is a heat radiator that dissipates heat to the hot water supply.

  Therefore, in this embodiment, the hot water is the heat exchange target fluid. In the heat pump cycle 13 of the present embodiment, a supercritical cycle in which the pressure of the refrigerant discharged from the compressor 14 is equal to or higher than the critical pressure constitutes the refrigerant passing through the refrigerant passage 15b of the water-refrigerant heat exchanger 15. The heat is dissipated in a supercritical state without condensing.

  The outlet side of the refrigerant passage 15 b of the water-refrigerant heat exchanger 15 is connected to the inlet side of the electric expansion valve 16. The electric expansion valve 16 is a pressure reducing means for decompressing and expanding the refrigerant flowing out from the refrigerant passage 15b, and a pressure for controlling the high pressure Pd of the refrigerant from the discharge port 18d of the compressor 14 to the inlet side of the electric expansion valve 16. It is also a control means.

  More specifically, the electric expansion valve 16 includes a valve body 16a that adjusts the throttle opening, and an electric actuator 16b that includes a servo motor that variably controls the throttle opening of the valve body 16a. Is done.

  An evaporator 17 is connected to the outlet side of the electric expansion valve 16. The evaporator 17 heat-exchanges the low-pressure refrigerant decompressed by the electric expansion valve 16 and the outside air (outdoor air) blown by the electric blower fan 17a, thereby evaporating the low-pressure refrigerant and exerting an endothermic effect. It is a heat exchanger for heat absorption.

  In the present embodiment, as the evaporator 17, a so-called fin-and-tube type heat configured to include a plate-like fin that promotes heat exchange between the low-pressure refrigerant and the outside air and a tube through which the refrigerant passes. An exchanger is used.

  An accumulator 17 b is connected to the refrigerant outlet side of the evaporator 17. The accumulator 17b is a gas-liquid separator that separates the gas-liquid refrigerant flowing out of the evaporator 17 and stores excess refrigerant. A suction port 18c of the compressor 14 is connected to the gas phase refrigerant outlet of the accumulator 17b.

  Next, an outline of the electric control unit of the present embodiment will be described. The control device 22 shown in FIG. 1 is composed of a microcomputer and its peripheral circuits, and on the output side thereof is an electric water pump 12a, an electric motor 21 of the compressor 14, an electric actuator 16b of the electric expansion valve 16, an electric air blower. A fan 17a and the like are connected to control the operation of these devices.

  Note that the control device 22 is configured integrally with the above-described control means for controlling each actuator. In the present embodiment, in particular, the operation of the electric motor 21 of the compressor 14 in the control device 22 is controlled. The configuration (hardware and software) to be used is the discharge capacity control means 22a, and the configuration to control the operation of the electric actuator 16b of the electric expansion valve 16 is the throttle opening degree control means 22b. Of course, each control means 22a, 22b may be constituted by a separate control device.

  Further, on the input side of the control device 22, a fluid temperature sensor 23 that detects the hot water temperature Twi on the inlet side of the water passage 15 a of the water-refrigerant heat exchanger 15, and the outlet side of the refrigerant passage 15 b of the water-refrigerant heat exchanger 15. High-pressure refrigerant temperature sensor 24 which is a high-pressure temperature detecting means for detecting the refrigerant temperature Tdo of the refrigerant, an outside air temperature sensor 25 for detecting the outside air temperature Tam for exchanging heat with the low-pressure refrigerant downstream of the electric expansion valve 16 in the evaporator 17, and the evaporator 17 is an evaporating temperature sensor 26 for detecting the evaporating temperature Te of the low-pressure refrigerant, and a high-pressure pressure that is a high-pressure detecting means for detecting the high-pressure pressure Pd of the refrigerant from the outlet 18d of the compressor 14 to the inlet side of the electric expansion valve 16. Sensors 28 and the like are connected, and detection signals from these sensor groups are input to the control device 22.

  The evaporating temperature sensor 26 of the present embodiment specifically detects the fin temperature of the evaporator 17, but of course, it may directly detect the temperature of the refrigerant on the outlet side of the evaporator 17, or the evaporator The surface temperature of the refrigerant pipe on the 17 outlet side may be detected.

  In addition, the high pressure sensor 28 of the present embodiment specifically detects the refrigerant pressure discharged from the discharge port 18d of the compressor 14, but the electric expansion valve 16 of the compressor 14 is discharged from the discharge port 18d of the compressor 14. What is necessary is just to detect the high pressure which reaches the inlet side.

  If the high pressure detection means is arranged between the discharge port 18d of the compressor 14 and the water-refrigerant heat exchanger 15, the pressure of the refrigerant discharged from the compressor 14 without being affected by the pressure loss in the refrigerant passage 15b. On the other hand, if it is arranged between the water-refrigerant heat exchanger 15 and the inlet side of the electric expansion valve 3, the water-refrigerant heat exchanger 15 detects the refrigerant pressure after heat dissipation. Therefore, the heat resistance required for the high pressure detection means can be reduced.

  Further, an operation panel 27 is connected to the input side of the control device 22, and an operation signal for hot water heater operation / stop, a hot water supply temperature setting signal for the water heater, and the like are input to the control device 22.

  Next, the operation of the heat pump type water heater 10 of the present embodiment in the above configuration will be described with reference to FIGS. FIG. 3 is a flowchart showing a control process executed by the control device 22. This control process starts when a hot water heater operation signal of the operation panel 27 is input to the control device 22 in a state where power is supplied to the heat pump hot water heater 10 from the outside.

  First, in step S1, flags, timers, and the like are initialized, and in the next step S2, an operation signal of the operation panel 27 and a vehicle environmental state signal, that is, detection signals detected by the sensor groups 23 to 26, etc. are read.

  Next, the process proceeds to step S3, and control states of various actuators are determined based on the operation signal and detection signal read in step S2. In the present embodiment, specifically, control signals (rotational speed control signals) output to the electric water pump 12a, the electric motor 21, and the electric blower fan 17a are determined.

  Next, it progresses to step S4 and the target pressure Pt of the high-pressure refrigerant discharged from the compressor 14 is determined. Therefore, this step S4 constitutes the target pressure determining means of this embodiment. Details of step S4 will be described with reference to FIG.

  First, in step S41, it is determined whether or not the hot water temperature Twi detected by the fluid temperature sensor 23 is equal to or higher than a predetermined set value TwiA. If Twi is greater than or equal to TwiA in step S41, the process proceeds to step S45, and if Twi is not greater than or equal to TwiA, the process proceeds to step S42.

  Next, in step S42, it is determined whether or not the outside air temperature Tam detected by the outside air temperature sensor 25 is equal to or higher than a predetermined set value TamA. In step S41, if Tam is equal to or higher than TamA, the process proceeds to step S45, and if Tam is not equal to or higher than TamA, the process proceeds to step S43.

  Next, in step S43, it is determined whether or not the evaporation temperature Te detected by the evaporation temperature sensor 26 is equal to or higher than a predetermined set value TeA. In step S43, if Te is equal to or higher than TeA, the process proceeds to step S45, and if Te is not equal to or higher than TeA, the process proceeds to step S44.

  That is, in steps S41 to S43, if none of the hot water temperature Twi, the outside air temperature Tam, and the evaporation temperature Te is equal to or higher than the set values TamA, TwiA, and TeA, the process proceeds to step S44. In Steps S41 to S43, when at least one of Twi, Tam, and Te is equal to or higher than TamA, TwiA, and TeA, the process proceeds to Step S45.

  In step S44, referring to the control map stored in advance based on the refrigerant temperature Tdo detected by the high-pressure refrigerant temperature sensor 24, the target pressure Pt (hereinafter, referred to as the coefficient of performance (COP) of the cycle is substantially maximized). The target pressure Pt determined in step S44 is called a normal target pressure Pt.), And the process proceeds to step S5.

  On the other hand, in step S45, a target pressure Pt having a value lower than the value determined in step S44 (hereinafter, the target pressure Pt determined in step S45 is referred to as a target pressure Pt for suppressing the intermediate pressure) is determined. The process proceeds to step S5. Here, details of the determination of the target pressure Pt when suppressing the intermediate pressure in step S45 will be described with reference to FIGS.

  First, in step S41, when the hot water temperature Twi is equal to or higher than TwiA, the high-pressure refrigerant discharged from the compressor 14 is converted into a water-refrigerant heat exchanger when Twi is lower than TwiA. 15 is difficult to dissipate heat, so the high-pressure pressure Pd of the high-pressure refrigerant is likely to increase.

  When the high pressure Pd rises, as will be described later, the throttle opening of the electric expansion valve 16 increases to bring the high pressure Pd closer to the target high pressure Pt. Therefore, the low-pressure pressure (that is, the suction pressure Ps) of the refrigerant from the outlet side of the electric expansion valve 16 to the suction port 18c of the compressor 14 also rises, and as described with reference to FIG. The pressure rises and exceeds the pressure resistance of the housing 18.

  Therefore, the target pressure Pt is determined with reference to the control map as shown by the thick line in FIG. 5 so that the intermediate pressure Pm is less than the withstand pressure of the housing 40 as shown as a reference by the broken line. Specifically, when Twi becomes equal to or higher than TwiA, the target pressure Pt is decreased as Twi increases.

  Note that, by reducing the target pressure Pt, the throttle opening of the electric expansion valve 16 increases and the suction pressure Ps also increases, but the target is within a range where the effect of increasing the volumetric efficiency η1 of the first compression mechanism 19 is greater. If the pressure Pt is determined, an increase in the intermediate pressure Pm can be suppressed.

  In step S42, when the outside air temperature Tam is equal to or higher than TamA, the refrigerant evaporation temperature in the evaporator 17 rises and is sucked into the compressor 14 as compared to the case where Tam is lower than TamA. As the suction pressure Ps of the suction refrigerant increases, the intermediate pressure Pm increases and exceeds the pressure resistance of the housing 18.

  Therefore, the target pressure Pt is determined with reference to the control map as shown by the thick line in FIG. 6 so that the intermediate pressure Pm is less than the withstand pressure of the housing 40 as shown as a reference by the broken line. Specifically, when Tam becomes equal to or higher than TamA, the target pressure Pt is decreased as Tam increases.

  In step S43, when the evaporation temperature Te is equal to or higher than TeA, the refrigerant evaporation pressure in the evaporator 17 increases and the suction pressure Ps increases compared to the case where Te is lower than TeA. Therefore, the intermediate pressure Pm increases and exceeds the pressure resistance of the housing 18.

  Therefore, the target pressure Pt is determined with reference to the control map as shown by the thick line in FIG. 7 so that the intermediate pressure Pm is less than the withstand pressure of the housing 40 as shown as a reference by the broken line. Specifically, when Te is equal to or higher than TeA, the target pressure Pt is decreased as Te increases.

  Further, in step S45, the lowest value among the values obtained by referring to the control maps shown in FIGS. 5 to 7 is set to the target pressure Pt at the time of suppressing the intermediate pressure to suppress the increase in the intermediate pressure Pm. To do. Of course, the arithmetic average value of the values obtained by referring to each map or the weighted average value considering the contribution of the values obtained by referring to each map may be used as the target pressure Pt at the time of suppressing the intermediate pressure. .

  Therefore, in the present embodiment, the outside air temperature Tam, the hot water temperature Twi, and the evaporation temperature Te are physical quantities that are correlated with the suction refrigerant pressure, and the outside air temperature sensor 25, the fluid temperature sensor 23, and the evaporation temperature sensor 26 that detect these physical quantities. Constitutes a suction pressure detecting means.

  Next, in step S5, the throttle opening degree of the electric expansion valve 16 (control to be output to the electric actuator 16b) so that the high pressure Pd detected by the high pressure sensor 28 approaches the target pressure Pt determined in step S4. Signal). Next, it progresses to step S6 and a control signal is output with respect to various actuators 12a, 21, 17a, 16a, etc. from the control apparatus 22 so that the control state determined by step S5 may be obtained.

  In the next step S <b> 7, when a hot water heater stop signal from the operation panel 27 is input to the control device 22, the operation of various actuators is stopped to stop the system. On the other hand, when the hot water heater stop signal is not input, after waiting for a predetermined control period, the process returns to step S2.

  Therefore, when the heat pump type water heater 10 of this embodiment is operated, the high-temperature and high-pressure refrigerant discharged from the compressor 14 flows into the refrigerant passage 15b of the water-refrigerant heat exchanger 15 and is driven by the electric water pump 12a. Heat exchange is performed with hot water flowing into the water passage 15a from the lower side of the hot water storage tank 11. Thereby, the hot water is heated, and the heated hot water is stored above the hot water storage tank 11.

  At this time, in the heat pump cycle 13 of the present embodiment, carbon dioxide is used as the refrigerant and the supercritical refrigeration cycle is configured. Therefore, the temperature of the high-pressure refrigerant is increased as compared with the case where CFC is used as the refrigerant. be able to. As a result, the amount of heat radiated to the hot water in the water-refrigerant heat exchanger 15 can be increased, and the hot water temperature can be increased.

  On the other hand, the high-pressure refrigerant flowing out of the water-refrigerant heat exchanger 15 is decompressed and expanded by the electric expansion valve 16. The refrigerant expanded under reduced pressure by the electric expansion valve 16 flows into the evaporator 5 and absorbs heat from the outside air blown from the electric blower fan 17a to evaporate. Then, the refrigerant that has flowed out of the evaporator 5 flows into the accumulator 17b and is gas-liquid separated. Further, the gas-phase refrigerant that has flowed out of the accumulator 17 b is sucked into the compressor 14.

  In the present embodiment, since the operation is performed as described above, the cycle can be operated at the normal target pressure Pt at which the COP becomes substantially maximum at the normal time when the intermediate pressure Pm does not exceed the withstand pressure of the housing 18. As a result, the heat pump cycle 13 can be operated while exhibiting a high COP.

  On the other hand, when the detected value of the suction pressure detecting means constituted by the outside air temperature sensor 25, the fluid temperature sensor 23, and the evaporation temperature sensor 26 is equal to or higher than a predetermined set value, the value is lower than the target high pressure Pt at normal time. The cycle is operated at the target high pressure Pt when the intermediate pressure is suppressed. As a result, it is possible to reduce the high pressure Pd and prevent the intermediate pressure Pm from exceeding the pressure resistance of the housing 18.

  That is, as shown in the Mollier diagram of FIG. 8, the heat pump cycle 13 can be operated while exhibiting a high COP in a normal state (thick solid line of FIG. 8), and the suction pressure Ps is increased. At the time of suppressing the intermediate pressure that needs to suppress the increase in the pressure Pm (the thick broken line in FIG. 8), the intermediate pressure Pm can be lowered below the withstand pressure of the housing 18.

  Accordingly, even in the heat pump cycle 13 including the two-stage booster compressor in which the housing 18 is filled with the medium-pressure refrigerant boosted by the first compression mechanism 19 as in the present embodiment, the pressure resistance of the housing 18 is increased. It is not necessary to increase the pressure more than necessary.

  In the present embodiment, the suction pressure detecting means is constituted by the outside air temperature sensor 25, the fluid temperature sensor 23, and the evaporation temperature sensor 26. Of course, the suction pressure detecting means is a suction pressure for directly detecting the suction pressure Ps. You may comprise with a sensor. On the other hand, in this embodiment, since the suction pressure detection means is constituted by a temperature sensor that is cheaper than the pressure sensor, the manufacturing cost of the heat pump hot water heater 10 as a whole can be reduced.

(Second Embodiment)
In the first embodiment, when the detected value of the suction pressure detecting means becomes equal to or higher than a predetermined set value, the target pressure Pt is decreased as the detected value increases, thereby decreasing the high pressure Pd. However, in the present embodiment, an example will be described in which the high pressure Pd is reduced without changing the target pressure Pt.

  The overall configuration of the heat pump type water heater 10 of the present embodiment is exactly the same as that of the first embodiment. Hereinafter, the operation of the present embodiment will be described with reference to FIGS. FIG. 9 is a flowchart showing a control process executed by the control device 22. In FIG. 9, 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.

  First, as in the first embodiment, initialization of flags, timers, and the like is performed in step S1, operation signals and detection signals are read in next step S2, and various actuators (electric water pump 12a, electric motor 21 are read in step S3. And the control signal output to the electric blower fan 17a) is determined.

  Next, the process proceeds to step S4 ', and the target pressure Pt of the high-pressure refrigerant discharged from the compressor 14 is determined. Therefore, in this embodiment, this step S4 'target pressure determination means is comprised. Specifically, as in step S44 of step S4 of the first embodiment, the COP is abbreviated with reference to a control map stored in advance based on the refrigerant temperature Tdo detected by the high-pressure refrigerant temperature sensor 24. The maximum target pressure Pt is determined.

  In step S5, the throttle opening degree of the electric expansion valve 16 (control signal output to the electric actuator 16b) is determined so that the high pressure Pd approaches the target pressure Pt determined in step S4 '. Next, it progresses to step S50 and the control signal output to the electric motor 21 of the compressor 14 is corrected. Details of step S50 will be described with reference to FIG.

  First, in step S51, it is determined whether or not the hot water temperature Twi detected by the fluid temperature sensor 23 is equal to or higher than a predetermined set value TwiA. In step S51, if Twi is equal to or greater than TwiA, the process proceeds to step S54. If Twi is not equal to or greater than TwiA, the process proceeds to step S52.

  Next, in step S52, it is determined whether or not the outside air temperature Tam detected by the outside air temperature sensor 25 is equal to or higher than a predetermined set value TamA. In step S52, if Tam is equal to or higher than TamA, the process proceeds to step S54. If Tam is not equal to or higher than TamA, the process proceeds to step S53.

  Next, in step S53, it is determined whether or not the evaporation temperature Te detected by the evaporation temperature sensor 26 is equal to or higher than a predetermined set value TeA. In step S53, if Te is equal to or greater than TeA, the process proceeds to step S54. If Te is not equal to or greater than TeA, the process proceeds to step S6.

  That is, as in steps S41 to S43 of the first embodiment, in steps S51 to S53, none of the outside air temperature Tam, the hot water temperature Twi, and the evaporation temperature Te is higher than the set values TamA, TwiA, and TeA. If yes, go to Step S6. Therefore, in this case, the control signal output to the electric motor 21 of the compressor 14 maintains the value determined in step S3 and is not corrected.

  On the other hand, if at least one of the outside air temperature Tam, the hot water temperature Twi, and the evaporation temperature Te is equal to or higher than the set values TamA, TwiA, TeA, the process proceeds to step S54. In step S54, the control signal output to the electric motor 21 of the compressor 14 is corrected so that the intermediate pressure Pm does not exceed the withstand pressure of the housing 18 with reference to the control map stored in advance.

  Specifically, as in FIGS. 5 to 7 described above, when Tam is equal to or higher than TamA, the refrigerant discharge capacity of the compressor 14 is reduced as Tam increases, and Twi becomes equal to or higher than TwiA. When Te is increased, the refrigerant discharge capacity of the compressor 14 is decreased. When Te is TeA or more, the refrigerant discharge capacity of the compressor 14 is decreased as Te is increased. Correct the control signal.

  Furthermore, in step S54, the value which becomes the lowest refrigerant | coolant discharge capability is made into a control signal among the values obtained with reference to the control map mentioned above. Of course, as in the first embodiment, the arithmetic average value of the values obtained by referring to each map or the weighted average value taking into account the contribution of the values obtained by referring to each map may be used as the control signal. Good.

  Next, in step S55, it is determined whether or not the control signal output to the electric motor 21 corrected in step S54 is equal to or less than a reference value. When the control signal to the electric motor 21 is less than the reference value, a sufficient amount of heat required for hot water supply cannot be obtained, and boiling of the hot water is stopped. That is, the operation of various actuators is stopped to stop the system.

  On the other hand, if the control signal to the electric motor 21 is not below the reference value, the process proceeds to step S6. Step S6 and subsequent steps are the same as in the first embodiment.

  Therefore, even if the heat pump type water heater 10 of this embodiment is operated, the detected value of the suction pressure detecting means constituted by the outside air temperature sensor 25, the fluid temperature sensor 23, and the evaporation temperature sensor 26 is equal to or higher than a predetermined set value. When the detected value increases, the refrigerant discharge capacity of the compressor 14 is reduced and the high pressure Pd is reduced, so that the suction pressure Ps is increased in the same manner as in the first embodiment. The accompanying increase in the intermediate pressure Pm can be suppressed.

  Furthermore, when the amount of heat necessary for hot water supply cannot be obtained, the operation of the heat pump type hot water heater 10 can be stopped.

(Third embodiment)
In the present embodiment, an example in which the electric expansion valve 16 is eliminated and an ejector 30 is employed as shown in the overall configuration diagram of FIG. 11 will be described. The ejector 30 is connected to the outlet side of the refrigerant passage 15b of the water-refrigerant heat exchanger 15 and is a decompression unit that decompresses and expands the high-pressure refrigerant. It is also a refrigerant circulating means for sucking the inside.

  Specifically, the ejector 30 narrows the passage area of the high-pressure refrigerant flowing out from the outlet side of the refrigerant passage 15b of the water-refrigerant heat exchanger 15 to reduce the refrigerant pressure, and the refrigerant injection of the nozzle portion 30a. The refrigerant suction port 30b is disposed so as to communicate with the port and sucks the refrigerant flowing out of the evaporator 17.

  The nozzle portion 30a is a variable nozzle portion configured to be able to change the throttle passage area, and specifically, a needle valve 30c that is disposed inside the nozzle portion 30a and adjusts the throttle opening degree of the nozzle portion 30. It has an electric actuator 30d composed of a servo motor that displaces the needle valve 30c in the axial direction of the nozzle portion 30a.

  Further, on the downstream side of the refrigerant flow of the nozzle portion 30a and the refrigerant suction port 30b, there is a mixing portion 30e that mixes the high-speed refrigerant flow injected from the nozzle portion 30a and the suction refrigerant sucked from the refrigerant suction port 30b. And it has the diffuser part 30f which makes a pressure | voltage rise part in the refrigerant | coolant flow downstream of the mixing part 30e.

  The diffuser portion 30f is formed in a shape that gradually increases the refrigerant passage area, and acts to decelerate the refrigerant flow to increase the refrigerant pressure, that is, to convert the velocity energy of the refrigerant into pressure energy.

  An accumulator 17b is connected to the downstream side of the ejector 30 (specifically, the outlet side of the diffuser portion 30f). The accumulator 17b of the present embodiment has a liquid phase refrigerant outlet, and this liquid phase refrigerant outlet is connected to the refrigerant inlet side of the evaporator 17. Other configurations are the same as those of the first embodiment.

  Next, the operation of this embodiment in the above configuration will be described. Control processing executed by the control device 22 in the present embodiment is basically the same as that in the first embodiment. In the present embodiment, since the ejector 30 is employed as the pressure reducing means, the throttle opening degree control means 20b is different from the electric actuator 16b of the electric expansion valve 16 of the first embodiment with respect to the electric actuator 30d of the ejector 30. A similar control signal is output.

  Therefore, when the heat pump type water heater 10 of the present embodiment is operated, the high-temperature and high-pressure refrigerant discharged from the compressor 14 flows into the refrigerant passage 15b of the water-refrigerant heat exchanger 15 as in the first embodiment. Then, the electric water pump 12a exchanges heat with hot water flowing into the water passage 15a from the lower side of the hot water storage tank 11. Thereby, hot water supply is heated.

  On the other hand, the high-pressure refrigerant that has flowed out of the water-refrigerant heat exchanger 15 flows into the nozzle portion 30a of the ejector 30, and isentropically decompressed and expanded. During the decompression and expansion, the pressure energy of the refrigerant is converted into velocity energy, and the refrigerant is ejected from the refrigerant injection port of the nozzle portion 30a as a high-speed refrigerant flow. At this time, the refrigerant on the downstream side of the evaporator 17 is sucked from the refrigerant suction port 30b by the suction action of the high-speed refrigerant flow.

  Further, in the mixing unit 30e, the refrigerant injected from the refrigerant injection port of the nozzle unit 30a and the suction refrigerant sucked from the refrigerant suction port 30b are mixed and flow into the diffuser unit 30f. In the diffuser portion 30f, the passage area is enlarged, so that the speed (expansion) energy of the refrigerant is converted into pressure energy, so that the pressure of the refrigerant increases.

  The refrigerant flowing out from the diffuser portion 30f of the ejector 30 flows into the accumulator 17b and is separated into a gas phase refrigerant and a liquid phase refrigerant. The liquid-phase refrigerant separated by the accumulator 17b flows into the evaporator 17, absorbs heat from the outside air blown from the electric blower fan 17a, evaporates, and is sucked from the refrigerant suction port 30b of the ejector 30. On the other hand, the gas-phase refrigerant separated by the accumulator 17b is sucked into the compressor 14.

  In this embodiment, since it operates as described above, the detection values of the suction pressure detection means (the outside air temperature sensor 25, the fluid temperature sensor 23, and the evaporation temperature sensor 26) are equal to or greater than a predetermined set value as in the first embodiment. Since the cycle is operated at the target high pressure Pt at the time of suppressing the intermediate pressure, which is lower than the target high pressure Pt at the normal time, the high pressure Pd can be reduced.

  Accordingly, exactly the same as in the first embodiment, it is possible to suppress an increase in the intermediate pressure Pm accompanying an increase in the suction pressure Ps, and thus it is not necessary to increase the pressure resistance of the housing 18 more than necessary.

  Furthermore, in this embodiment, since the ejector 30 is used as the pressure reducing means, the suction refrigerant of the compressor 14 can be raised by the pressure increasing action of the diffuser portion 30f, and the driving power of the compressor 14 can be reduced. . As a result, the normal COP can be further increased.

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

  (1) In the above-described embodiment, the suction pressure detection means is configured by the outside air temperature sensor 25, the fluid temperature sensor 23, and the evaporation temperature sensor 26, but the configuration of the suction pressure detection means is not limited to this. For example, any one of the outside air temperature sensor 25, the fluid temperature sensor 23, and the evaporation temperature sensor 26 may be used.

  According to the study by the present inventors, if the suction pressure detecting means is constituted by a combination of the fluid temperature sensor 23 and the outside air temperature sensor 25 or a combination of the fluid temperature sensor 23 and the evaporation temperature sensor 26, the suction pressure Ps. It is possible to sufficiently detect that the intermediate pressure Pm is equal to or higher than the withstand pressure of the housing 18 due to the rise in the pressure.

  (2) In the first embodiment described above, the throttle opening control means 22b controls the throttle opening of the electric expansion valve 16 (opening of the valve body 16a), thereby bringing the high pressure Pd closer to the target pressure Pt. Although the example has been described, the discharge capacity control means 22a may control the refrigerant discharge capacity (the number of rotations of the electric motor 21) of the compressor 14 so that the high pressure Pd approaches the target pressure Pt.

  Furthermore, in step S5, the throttle opening degree of the electric expansion valve 16 is determined so that the high pressure Pd detected by the high pressure sensor 28 approaches the target pressure Pt. For example, from the temperature of the refrigerant discharged from the compressor 14, The high pressure Pd may be estimated so that the estimated value approaches the target pressure Pt.

  Further, the high pressure Pd is at least the outside temperature Tam, the boiling temperature Two on the outlet side of the water passage 15a of the water-refrigerant heat exchanger 15, the hot water temperature Twi on the inlet side of the water passage 15a, and the rotational speed Nc of the compressor 14. You may estimate using one.

  (3) In the second embodiment described above, the discharge capacity control means 22a allows the refrigerant discharge capacity (specifically, the electric motor 21 of the electric motor 21) so that the intermediate pressure Pm does not exceed the pressure resistance of the housing 18. The example of controlling the rotation speed) has been described, but the throttle opening degree control means 22b controls the throttle opening degree of the electric expansion valve 16 (opening degree of the valve body 16a), so that the intermediate pressure Pm is reduced in the housing 18. The pressure resistance may not be exceeded.

  (4) In the above-described second embodiment, the example in which the electric expansion valve 16 whose operation is electrically controlled is employed as the pressure reducing means has been described. However, the valve opening (throttle opening) is controlled by a mechanical mechanism. A pressure control valve configured to be changeable may be employed.

  Specifically, the pressure control valve has a temperature sensing portion provided on the outlet side of the refrigerant passage 15b of the water-refrigerant heat exchanger 15, and the temperature of the refrigerant on the outlet side of the refrigerant passage 15b is inside the temperature sensing portion. It is possible to employ a configuration in which the valve opening is adjusted by the balance between the internal pressure of the temperature sensing portion and the refrigerant pressure at the outlet side of the refrigerant passage 15b (high pressure Pd).

  In this type of pressure control valve, the target high pressure Pt corresponding to the temperature of the refrigerant on the outlet side of the refrigerant passage 15b can be determined by the type and amount of the temperature-sensitive medium enclosed in the temperature-sensing unit. Therefore, in the refrigeration cycle apparatus that employs the pressure control valve, the temperature sensing unit constitutes the target pressure determining means. Note that the detailed configuration of this type of pressure control valve can be referred to Japanese Patent Laid-Open No. 2000-81157.

  Even in the refrigeration cycle apparatus employing such a pressure control valve, the detection values Tam, Twi, Te detected by the suction pressure detection means 23, 25, 26 are set in advance as in the second embodiment. When the values TamA, TwiA, and TeA are reached, the high pressure Pd can be reduced by reducing the refrigerant discharge capacity of the compressor 14.

  (5) In the above-described third embodiment, the example in which the control device 22 executes the same control process as in the first embodiment has been described. However, in the configuration of the third embodiment, the control device 22 has the second embodiment. The control processing described in the other embodiments may be executed.

  (6) In the above-described embodiment, the normal target pressure Pt is determined based on the refrigerant temperature Tdo detected by the high-pressure refrigerant temperature sensor 24 so that the COP becomes substantially maximum. Even if a high pressure sensor that directly detects the high pressure Pd on the outlet side of the refrigerant passage 15b of the heat exchanger 15 is employed, and the target pressure Pt at which the COP becomes substantially maximum is determined based on the detected value of the high pressure sensor. Good.

  (7) In the above-described embodiment, the example in which the two-stage booster type compressor 14 in which the medium pressure refrigerant boosted by the first compression mechanism 19 is filled in the housing 18 is used has been described. Since the high pressure Pd can be lowered as the suction pressure Ps increases, a two-stage booster compressor in which the housing 18 is filled with the high-pressure refrigerant discharged from the compressor 14 may be employed.

  (8) In the above-described embodiment, the example in which the refrigeration cycle apparatus of the present invention is applied to the heat pump type hot water heater 10 has been described. However, the application of the present invention is not limited to this, and the amount of heat absorbed by the low-pressure refrigerant It can be widely applied to a refrigeration cycle apparatus that radiates heat. For example, the present invention can be applied to a heat pump type floor heating apparatus. In this case, the floor heating LLC is the heat exchange target fluid.

It is a whole lineblock diagram of the heat pump type hot water heater of a 1st embodiment. It is sectional drawing of the two-stage pressure | voltage rise type compressor of 1st Embodiment. It is a flowchart which shows the control processing of the control apparatus of 1st Embodiment. It is a flowchart which shows the principal part of the control processing of the control apparatus of 1st Embodiment. It is explanatory drawing of the control map regarding the external temperature of 1st Embodiment, and a target pressure. It is explanatory drawing of the control map regarding the hot water temperature and target pressure of 1st Embodiment. It is explanatory drawing of the control map regarding the evaporation temperature and target pressure of 1st Embodiment. It is a Mollier diagram which shows the action | operation of the heat pump cycle of 1st Embodiment. It is a flowchart which shows the control processing of the control apparatus of 2nd Embodiment. It is a flowchart which shows the principal part of the control processing of the control apparatus of 2nd Embodiment. It is a whole block diagram of the heat pump type water heater of 3rd Embodiment. It is a graph which shows the relationship between the suction pressure of a 2 step | paragraph pressure | voltage rise type compressor, an intermediate pressure, and a high pressure.

Explanation of symbols

DESCRIPTION OF SYMBOLS 14 Compressor 15 Water-refrigerant heat exchanger 16 Electric expansion valve 17 Evaporator 19 1st compression mechanism 20 2nd compression mechanism 22a Discharge capacity control means 22b Throttle opening degree control means 23 Outside temperature sensor 24 High pressure refrigerant temperature sensor 25 Fluid Temperature sensor 26 Evaporation temperature sensor 28 High pressure sensor 30 Ejector S4 Target pressure determination means

Claims (11)

A two-stage boosting type having a first compression mechanism (19) that sucks and compresses refrigerant and a second compression mechanism (20) that compresses and discharges the medium-pressure refrigerant pressurized by the first compression mechanism (19). A compressor (14);
A radiator (15) for radiating the heat quantity of the refrigerant discharged from the two-stage booster compressor (14) to the heat exchange target fluid;
Decompression means (16, 30) for decompressing and expanding the refrigerant flowing out of the radiator (15);
An evaporator (17) for evaporating the refrigerant decompressed by the decompression means (16, 30);
A target pressure determining means (S4) for determining a target pressure (Pt) of the high pressure (Pd) of the high-pressure refrigerant from the discharge port of the two-stage booster compressor (14) to the inlet of the pressure reducing means (16, 30). Prepared,
The high-pressure pressure (Pd) is a refrigeration cycle apparatus controlled so as to approach the target pressure (Pt),
Further, it comprises suction pressure detection means (23, 25, 26) for detecting a physical quantity having a correlation with the suction pressure (Ps) of the refrigerant sucked into the two-stage booster compressor (14),
When the detected values (Tam, Twi, Te) detected by the suction pressure detecting means (23, 25, 26) are equal to or higher than a predetermined set value (TamA, TwiA, TeA), the detected value (Tam , Twi, Te), the refrigeration cycle is characterized in that the high pressure (Pd) is reduced so that the medium pressure (Pm) of the medium pressure refrigerant is less than a predetermined pressure. apparatus. Furthermore, high pressure detection means (28) for detecting the high pressure (Pd),
The refrigeration cycle apparatus according to claim 1, further comprising high-pressure temperature detecting means (24) for detecting a temperature of the high-pressure refrigerant.   The target pressure determining means (S4) is configured to detect the detected values (Tam, Twi, Te) when the detected values (Tam, Twi, Te) are equal to or higher than a predetermined set value (TamA, TwiA, TeA). The refrigeration cycle apparatus according to claim 1 or 2, wherein the target pressure (Pt) is decreased as the pressure increases. A discharge capacity control means (22a) for controlling the refrigerant discharge capacity of the two-stage booster compressor (14);
The discharge capacity control means (22a) detects the detected value (Tam, Twi, Te) when the detected value (Tam, Twi, Te) is equal to or higher than a predetermined set value (TamA, TwiA, TeA). The refrigeration cycle apparatus according to any one of claims 1 to 3, wherein the refrigerant discharge capacity is reduced as the value of the refrigerant increases. A throttle opening control means (22b) for controlling the throttle opening of the pressure reducing means (16, 30);
When the detected value (Tam, Twi, Te) is equal to or higher than a preset value (TamA, TwiA, TeA), the throttle opening control means (22b) is configured to detect the detected value (Tam, Twi, Te). The refrigeration cycle apparatus according to any one of claims 1 to 3, wherein the throttle opening is increased in accordance with an increase in ().   The suction pressure detection means includes a fluid temperature sensor (23) that detects a temperature (Twi) of the fluid to be heat exchanged to exchange heat with the high-pressure refrigerant in the radiator (15). The refrigeration cycle apparatus according to any one of claims 1 to 5.   The said suction pressure detection means is comprised by the outdoor temperature sensor (25) which detects the outdoor temperature (Tam) of the outdoor air which heat-exchanges with the said low pressure refrigerant | coolant in the said evaporator (17). The refrigeration cycle apparatus according to any one of 1 to 6.   8. The suction pressure detection means comprises an evaporation temperature sensor (26) for detecting an evaporation temperature (Te) of the low-pressure refrigerant in the evaporator (17). The refrigeration cycle apparatus according to one.   The decompression means sucks the refrigerant into the interior by a high-speed refrigerant flow injected from the nozzle part (30a) for decompressing and expanding the high-pressure refrigerant, and mixes the sucked refrigerant and the high-speed refrigerant flow to increase the pressure. The refrigeration cycle apparatus according to any one of claims 1 to 8, wherein the refrigeration cycle apparatus is configured by an ejector (30).   The refrigeration cycle apparatus according to any one of claims 1 to 9, wherein the heat exchange target fluid is water.   The refrigeration cycle apparatus according to any one of claims 1 to 10, wherein the refrigerant is carbon dioxide.

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