WO2017061233A1

WO2017061233A1 - Refrigeration cycle device

Info

Publication number: WO2017061233A1
Application number: PCT/JP2016/076611
Authority: WO
Inventors: 昌宏高津
Original assignee: 株式会社デンソー
Priority date: 2015-10-05
Filing date: 2016-09-09
Publication date: 2017-04-13
Also published as: JPWO2017061233A1; JP6477908B2; DE112016004544T5

Abstract

This refrigeration cycle device is provided with a compression mechanism (21), a heat radiator (23), a first pressure reducer (25), an evaporator (26), a second pressure reducer (27), an internal heat exchanger (24), a temperature detection unit (31), and a control device (30). The compression mechanism (21) discharges a high-pressure refrigerant in a supercritical state, and introduces an intermediate-pressure refrigerant having an intermediate pressure between a low pressure and a high pressure during a process in which a refrigerant is compressed from the low pressure to the high pressure. The temperature detection unit detects the temperature of a heat exchange medium flowing into the heat radiator. When the temperature detected by the temperature detection unit is higher than the critical temperature of the refrigerant, the control device adjusts the pressure of the high-pressure refrigerant so that the higher the detected temperature, the higher the pressure of the high-pressure refrigerant becomes, and also adjusts the flow rate of the intermediate-pressure refrigerant so that the higher the detected temperature, the greater the flow rate of the intermediate-pressure refrigerant becomes.

Description

Refrigeration cycle equipment

Cross-reference to related applications

This application is based on Japanese Patent Application No. 2015-197893 filed on October 5, 2015, the description of which is incorporated herein by reference.

The present disclosure relates to a refrigeration cycle apparatus.

As a conventional refrigeration cycle apparatus, there is a conventional refrigeration cycle apparatus that constitutes a two-stage compression / one-stage expansion cycle and a high-pressure refrigerant discharged from a two-stage compression type compression mechanism is operated in a supercritical state as in Patent Document 1. In such a refrigeration cycle apparatus, after the high-pressure refrigerant discharged from the compression mechanism is radiated by the radiator, a part of the high-pressure refrigerant passes through the internal heat exchanger, is depressurized by the first decompressor, and is absorbed by the evaporator. Then, it is sucked into the compression mechanism. The remainder of the high-pressure refrigerant radiated by the radiator is decompressed by the second decompressor to become an intermediate-pressure refrigerant, heated by heat exchange with the high-pressure refrigerant by the internal heat exchanger, and injected into the compression mechanism (that is, introduced). Is done.

JP 2014-159950 A

By the way, in the refrigeration cycle apparatus described above, the temperature of the heat exchange medium that exchanges heat with the high-pressure refrigerant in the radiator (hereinafter referred to as the heat exchange medium temperature) may be higher than the critical temperature of the refrigerant (hereinafter referred to as the critical temperature). is there. In this case, as compared with the case where the heat exchange medium temperature is lower than the critical temperature, the amount of decrease in the enthalpy difference in the radiator due to the increase in the heat exchange medium temperature is increased. The heat exchange medium temperature is a temperature before heat exchange with the refrigerant. The enthalpy difference in the radiator is the enthalpy difference between the refrigerant on the inlet side and the outlet side of the radiator. For this reason, when the heat exchange medium temperature is higher than the critical temperature, the heat dissipation performance of the radiator decreases as the heat exchange medium temperature increases. Therefore, in the case where the radiator is used for heating, when the heat exchange medium temperature becomes higher than the critical temperature, the heating capacity is lowered.

In addition, when the heat exchange medium temperature is higher than the critical temperature, when the heat exchange medium temperature rises, the inlet temperature of the internal heat exchanger rises, and the amount of heat that can be exchanged on the high pressure side increases, resulting in a drastic difference in enthalpy. Increase. The enthalpy difference on the high pressure side of the internal heat exchanger is the enthalpy difference between the refrigerant on the inlet side and the outlet side of the high pressure side passage of the internal heat exchanger.

Therefore, when the heat exchange medium temperature is higher than the critical temperature, it is conceivable that the higher the heat exchange medium temperature, the higher the pressure of the high-pressure refrigerant. Thereby, the reduction | decrease of the enthalpy difference in a radiator can be suppressed, and the fall of the heating capability of a radiator can be suppressed.

However, it is assumed that when the heat exchange medium temperature rises, only the pressure of the high-pressure refrigerant is increased, and the injection flow rate of the intermediate-pressure refrigerant injected into the compression mechanism is the same as that before the heat exchange medium temperature rises. . That is, it is assumed that the injection flow rate is constant regardless of the heat exchange medium temperature. In this case, as a result of an increase in the enthalpy difference on the high pressure side of the internal heat exchanger, the amount of heat received by the intermediate pressure refrigerant from the high pressure refrigerant in the internal heat exchanger increases, and the enthalpy of the intermediate pressure refrigerant injected into the compression mechanism increases. growing. As a result, the discharge refrigerant temperature of the compression mechanism increases. At this time, depending on the temperature of the heat exchange medium on the inlet side of the internal heat exchanger, the discharge refrigerant temperature becomes too high, and the reliability of the compression mechanism is lowered.

This disclosure aims to provide a refrigeration cycle apparatus that can improve the reliability of a compression mechanism while suppressing a decrease in heating performance of a radiator.

According to one aspect of the present disclosure,
Refrigeration cycle equipment
The refrigerant is compressed from a low pressure to a high pressure higher than the low pressure, and the high-pressure refrigerant in a supercritical state is discharged. A compression mechanism for introducing a refrigerant;
A radiator that dissipates the high-pressure refrigerant by heat exchange between the high-pressure refrigerant discharged from the compression mechanism and the heat exchange medium;
A first pressure reducer that depressurizes a part of the high-pressure refrigerant flowing out of the radiator to a low pressure to form a low-pressure refrigerant;
An evaporator that evaporates the low-pressure refrigerant and sucks the low-pressure refrigerant after evaporation into the compression mechanism;
A second pressure reducer that reduces the other part of the high-pressure refrigerant flowing out of the radiator to an intermediate pressure to obtain an intermediate-pressure refrigerant;
An internal heat exchanger that exchanges heat between the high-pressure refrigerant that flows out of the radiator and flows toward the first decompressor, and the intermediate-pressure refrigerant that flows out of the second decompressor and flows toward the compression mechanism;
A temperature detector that detects the temperature of the heat exchange medium flowing into the radiator;
Based on the detected temperature of the temperature detector, the operation of the refrigeration cycle including the compression mechanism, the radiator, the first decompressor, the evaporator, the second decompressor, and the internal heat exchanger is controlled to increase the pressure. A control device for adjusting the pressure of the refrigerant and the flow rate of the intermediate pressure refrigerant,
When the detected temperature is higher than the critical temperature of the refrigerant, the control device adjusts the pressure of the high-pressure refrigerant so that the higher the detected temperature, the higher the pressure of the high-pressure refrigerant. The flow rate of the intermediate pressure refrigerant is adjusted so that the flow rate increases.

According to this, when the heat exchange medium temperature is higher than the critical temperature, the higher the heat exchange medium temperature, the higher the pressure of the high-pressure refrigerant can be suppressed, thereby suppressing the heating capacity of the radiator.

Furthermore, when the heat exchange medium temperature is higher than the critical temperature, as the heat exchange medium temperature rises, the enthalpy difference on the high pressure side of the internal heat exchanger increases rapidly, and the higher the heat exchange medium temperature, the higher the heat exchange medium temperature. The flow rate of the intermediate pressure refrigerant to be injected is increased. Thereby, the raise of the enthalpy of the intermediate pressure refrigerant | coolant injected can be suppressed. For this reason, the temperature rise of the discharge refrigerant | coolant of a compression mechanism can be suppressed, and the reliability of a compression mechanism can be improved.

It is a figure showing the whole hot-water supply heater composition in a 1st embodiment. It is a flowchart which shows the control processing of the control apparatus in 1st Embodiment. It is a figure which shows the relationship between the target pressure determined by step S2 of FIG. 2, and hot-water supply water temperature. FIG. 3 is a Mollier diagram of carbon dioxide refrigerant for explaining the target pressure determined in step S <b> 2 of FIG. 2. It is a figure which shows the relationship between the injection flow volume relevant to the passage opening degree of the 2nd expansion valve determined by step S4 of FIG. 2, and hot-water supply water temperature. It is a figure for demonstrating the subject which this indication solves, Comprising: It is a Mollier diagram which shows cycle balance when the hot-water supply water temperature of the inlet side of the water refrigerant | coolant heat exchanger in the refrigeration cycle apparatus shown in FIG. is there. It is a figure for demonstrating the subject which this indication solves, Comprising: It is a Mollier diagram which shows cycle balance when the hot-water supply water temperature of the inlet side of the water refrigerant | coolant heat exchanger in the refrigeration cycle apparatus shown in FIG. is there. It is a figure for demonstrating the subject which this indication solves, Comprising: It is the Mollier diagram which shows cycle balance when the hot-water-heating-water temperature of the inlet side of the water refrigerant | coolant heat exchanger in the refrigeration cycle apparatus shown in FIG. is there. It is a figure for demonstrating the subject which this indication solves, Comprising: It is the Mollier diagram which shows cycle balance when the hot-water supply water temperature of the inlet side of the water refrigerant | coolant heat exchanger in the refrigeration cycle apparatus shown in FIG. is there. It is a figure for demonstrating the subject which this indication solves, Comprising: The hot_water | molten_metal water temperature of the inlet side of the water refrigerant | coolant heat exchanger in the refrigeration cycle apparatus shown in FIG. It is a Mollier diagram which shows the cycle balance when making it higher than time. It is a figure for demonstrating the effect of 1st Embodiment, Comprising: It is the Mollier diagram which shows cycle balance when the hot-water supply water temperature of the inlet side of the water refrigerant | coolant heat exchanger in the refrigerating-cycle apparatus shown in FIG. is there. It is a flowchart which shows the control processing of the control apparatus in 2nd Embodiment. FIG. 13 is a diagram for explaining step S6-1 in FIG. 12, and is a Mollier diagram showing cycle balance when the hot water temperature on the inlet side of the water refrigerant heat exchanger in the refrigeration cycle apparatus shown in FIG. 1 is 37 ° C. It is. It is a flowchart which shows the control processing of the control apparatus in 3rd Embodiment. FIG. 15 is a diagram for explaining step S6-2 in FIG. 14 and is a Mollier diagram showing the cycle balance when the hot water temperature on the inlet side of the water refrigerant heat exchanger in the refrigeration cycle apparatus shown in FIG. 1 is 37 ° C. It is.

Hereinafter, embodiments of the present disclosure will be described with reference to the drawings. In the following embodiments, parts that are the same or equivalent to each other will be described with the same reference numerals.

(First embodiment)
In the present embodiment, the refrigeration cycle apparatus of the present disclosure is applied to a hot water heater. As shown in FIG. 1, the hot water heater 1 according to the present embodiment includes a hot water supply circuit 10 that circulates hot water in a hot water storage tank and a refrigeration cycle device 20 that circulates refrigerant.

The hot water supply circuit 10 includes a hot water storage tank 11 that stores hot water, a water pipe 12 that connects the water refrigerant heat exchanger 23 and the hot water storage tank 11 of the refrigeration cycle apparatus 20, and the water refrigerant heat exchanger 23 and the hot water storage tank 11. And a water circulation pump 13 for circulating water between them.

The hot water storage tank 11 is connected to the water passage 23 a of the water refrigerant heat exchanger 23. Hot water stored in the hot water storage tank 11 is heated by the water refrigerant heat exchanger 23. The heated hot water is supplied to a kitchen, a bath, or the like, or supplied to a heating device that heats the room using hot water. The water circulation pump 13 is an electric pump.

The refrigeration cycle apparatus 20 mainly includes a first compressor 21a, a second compressor 21b, a water-refrigerant heat exchanger 23, an internal heat exchanger 24, a first expansion valve 25, an outdoor heat exchanger 26, and a second expansion valve 27. It is provided as a major component. Each component is connected by refrigerant piping. These components constitute a two-stage compression / single-stage expansion cycle. The refrigeration cycle apparatus 20 uses carbon dioxide having a critical temperature of 31 ° C. as a refrigerant.

The first compressor 21a is a low-stage compressor that compresses the sucked-in low-pressure refrigerant and discharges an intermediate-pressure refrigerant having an intermediate pressure higher than that of the low-pressure refrigerant. The second compressor 21b is a high-stage compressor that sucks and compresses the intermediate-pressure refrigerant discharged from the first compressor 21a and discharges high-pressure high-pressure refrigerant having a pressure higher than that of the intermediate-pressure refrigerant. The pressure of the high-pressure refrigerant at this time is a pressure exceeding the critical pressure of the refrigerant, that is, a pressure at which the refrigerant becomes a supercritical state. The first compressor 21a and the second compressor 21b are each an electric compressor driven by an electric motor. In the present embodiment, the two-stage compression type compression mechanism 21 is configured using two single-stage compressors, ie, a first compressor 21a and a second compressor 21b. The two-stage compression type compression mechanism 21 compresses the refrigerant from a low pressure to a high pressure higher than the low pressure, and also converts an intermediate pressure refrigerant that is an intermediate pressure between the low pressure and the high pressure in the middle of the compression process of the refrigerant from the low pressure to the high pressure. Introduce.

The water refrigerant heat exchanger 23 has a water passage 23a through which hot water flows and a refrigerant passage 23b through which high-pressure refrigerant flows. The inlet side of the refrigerant passage 23b is connected to the discharge port side of the second compressor 21b. The water-refrigerant heat exchanger 23 is a radiator that radiates the high-pressure refrigerant and heats the hot-water supply by heat exchange between the high-pressure refrigerant discharged from the second compressor 21 b and the hot-water supply water in the hot water storage tank 11. Therefore, in the present embodiment, the hot water supply constitutes a heat exchange medium that exchanges heat with the high-pressure refrigerant. In other words, the hot water supply constitutes a cooling medium that cools the high-pressure refrigerant.

The internal heat exchanger 24 has a high-pressure side passage 24a through which high-pressure refrigerant flows and an intermediate-pressure side passage 24b through which intermediate-pressure refrigerant flows. The inlet side of the high-pressure side passage 24a is connected to the refrigerant passage 23b. The outlet side of the intermediate pressure side passage 24b is connected between the first compressor 21a and the second compressor 21b. The internal heat exchanger 24 is a heat exchanger that exchanges heat between a part of the high-pressure refrigerant flowing out of the water refrigerant heat exchanger 23 and the intermediate-pressure refrigerant flowing out of the second expansion valve 27.

The inlet side of the first expansion valve 25 is connected to the outlet side of the high-pressure side passage 24a. The first expansion valve 25 is a first pressure reducer that depressurizes the high-pressure refrigerant that has flowed out of the high-pressure side passage 24a into a low-pressure refrigerant. The first expansion valve 25 is an electric expansion valve that is configured so that the passage opening is variable and the passage opening is electrically adjusted.

The inlet side of the outdoor heat exchanger 26 is connected to the outlet side of the first expansion valve 25. The outdoor heat exchanger 26 is an evaporator that evaporates the low-pressure refrigerant by heat exchange between the low-pressure refrigerant decompressed by the first expansion valve 25 and the outside air. The outlet side of the outdoor heat exchanger 26 is connected to the suction side of the first compressor 21a.

The second expansion valve 27 is a second pressure reducer that reduces the other part of the high-pressure refrigerant flowing out of the water-refrigerant heat exchanger 23 to an intermediate pressure to obtain an intermediate-pressure refrigerant. The second expansion valve 27 is an electric expansion valve similar to the first expansion valve 25. The inlet side of the second expansion valve 27 is connected to a branch point 28 provided in the middle of the refrigerant passage between the refrigerant passage 23 b of the water refrigerant heat exchanger 23 and the high-pressure side passage 24 a of the internal heat exchanger 24.

In other words, the refrigeration cycle apparatus 20 includes an injection circuit (that is, an injection passage) 29 that is a refrigerant passage extending from the branch point 28 between the first compressor 21a and the second compressor 21b. In the middle of the injection circuit 29, the second expansion valve 27 and the intermediate pressure side passage 24b of the internal heat exchanger 24 are arranged.

In the refrigeration cycle apparatus 20 configured in this way, the low-pressure refrigerant is compressed in the order of the first compressor 21a and the second compressor 21b to become a high-pressure refrigerant. The high-pressure refrigerant is radiated by the water refrigerant heat exchanger 23 and then branches at the branch point 28. One of the branched high-pressure refrigerants is cooled by the internal heat exchanger 24 and then depressurized by the first expansion valve 25 to become a low-pressure refrigerant. The low-pressure refrigerant is heated by the outdoor heat exchanger 26 and evaporated, and then sucked into the first compressor 21a. The other high-pressure refrigerant branched at the branch point 28 is decompressed by the second expansion valve 27 and becomes an intermediate-pressure refrigerant. The intermediate pressure refrigerant is heated by the internal heat exchanger 24 and then injected (that is, introduced) between the discharge port of the first compressor 21a and the suction port of the second compressor 21b. In other words, the intermediate pressure refrigerant joins the refrigerant in the middle of the compression process from the low pressure to the high pressure in the first compressor 21a and the second compressor 21b.

At this time, the flow rate of the refrigerant injected between the first compressor 21a and the second compressor 21b is adjusted by adjusting the passage opening degree of the second expansion valve 27. Hereinafter, the refrigerant flow rate of the intermediate pressure refrigerant to be injected is referred to as an injection flow rate.

Thus, the intermediate pressure refrigerant is injected between the first compressor 21a and the second compressor 21b. As a result, an increase in the refrigerant temperature discharged from the second compressor 21b is suppressed, the heat circulation capacity is increased by increasing the refrigerant circulation amount of the water refrigerant heat exchanger 23, or the inlet refrigerant enthalpy of the outdoor heat exchanger 26 is lowered. The effect of increasing the cooling capacity can be obtained.

The refrigeration cycle apparatus 20 includes a control device 30. The control device 30 includes a microcomputer and its peripheral circuits.

A water temperature sensor 31, a first refrigerant temperature sensor 32, a second refrigerant temperature sensor 33, and a refrigerant pressure sensor 34 are connected to the input side of the control device 30. The water temperature sensor 31 is provided on the inlet side of the water passage 23 a of the water refrigerant heat exchanger 23. The water temperature sensor 31 is a temperature detection unit that detects the temperature of hot water flowing into the water passage 23a. The first refrigerant temperature sensor 32 is a first refrigerant temperature detector that detects the temperature of the intermediate pressure refrigerant on the inlet side of the intermediate pressure side passage 24b of the internal heat exchanger 24 (that is, a point A8 in FIG. 6 and the like described later). . The second refrigerant temperature sensor 33 is a second refrigerant temperature detection unit that detects the temperature of the intermediate pressure refrigerant on the outlet side of the intermediate pressure side passage 24b of the internal heat exchanger 24 (that is, a point A9 in FIG. 6 and the like described later). . The refrigerant pressure sensor 34 is a pressure detection unit that detects the pressure of the high-pressure refrigerant. The refrigerant pressure sensor 34 is provided for the refrigerant passage between the high-pressure side passage 24 a of the internal heat exchanger 24 and the first expansion valve 25. Sensor signals of these

sensors

31, 32, 33, and 34 are input to the control device 30.

The output side of the control device 30 is connected to refrigeration cycle components such as the first compressor 21a, the second compressor 21b, the first expansion valve 25, and the second expansion valve 27. The control device 30 controls the operation of the refrigeration cycle by controlling the first compressor 21a, the second compressor 21b, the first expansion valve 25, the second expansion valve 27, and the like.

Specifically, the control device 30 starts the operation of the refrigeration cycle by starting the operation of the first compressor 21a and the second compressor 21b. At this time, each rotation speed of the first compressor 21a and the second compressor 21b is set to a predetermined rotation speed.

Then, as shown in FIG. 2, the control device 30 controls the operation of the refrigeration cycle and appropriately adjusts the pressure of the high-pressure refrigerant (hereinafter referred to as high-pressure pressure) and the injection flow rate. Each step shown in FIG. 2 constitutes a functional unit that realizes various functions of the control device 40.

In step S1, the sensor signal of the water temperature sensor 31 is read. Thereby, the detected temperature of the water temperature sensor 31, that is, the hot water temperature on the inlet side of the water refrigerant heat exchanger 23 is read.

Subsequently, in step S2, the target pressure of the high pressure is determined based on the temperature detected by the water temperature sensor 31. At this time, the target pressure Px is determined so as to satisfy the relationship shown in FIG. That is, as shown in FIG. 3, when the hot water temperature is equal to or lower than the critical temperature of the refrigerant, the target pressure Px is made constant at a predetermined value P1 regardless of the hot water temperature. When the hot water temperature is higher than the critical temperature, the target pressure Px is set higher than the predetermined value P1, and the target pressure Px is increased as the hot water temperature is higher. In FIG. 3, when the hot water temperature is higher than the critical temperature, the target pressure Px increases linearly as the hot water temperature rises, but may increase in a curved line.

In addition, as shown in FIG. 4, the target pressure when the hot water temperature is higher than the critical temperature is based on the pressure at the intersection of the 600 kg / m ³ isodensity line and the isothermal line in the Mollier diagram of the carbon dioxide refrigerant. Is preferably high. The isotherm here is an isotherm having the same temperature as the temperature detected by the water temperature sensor 31 in the refrigerant isotherm in the Mollier diagram. If the pressure is lower than this, the enthalpy rise of the refrigerant at the outlet of the water refrigerant heat exchanger 23 becomes large, the difference in the enthalpy of the refrigerant at the inlet and outlet of the water refrigerant heat exchanger 23 becomes small, and the reduction in heating capacity is suppressed. It is because it becomes impossible.

Further, as shown in FIG. 4, the target pressure in this case is preferably lower than the pressure at the intersection of the 700 kg / m ³ isodensity line and the isotherm. If the pressure is higher than this, the increase in the work amount of the

compressors

21a and 21b becomes larger than the enthalpy difference increment in the water-refrigerant heat exchanger 23 and the efficiency is lowered, which is not desirable.

Therefore, the target pressure in this case is particularly preferably the pressure at the intersection of the 650 kg / m ³ isodensity line and the isotherm, as shown in FIG. In the present embodiment, step S2 constitutes a pressure determining unit that determines the target pressure of the high-pressure refrigerant.

In step S3, the sensor signal of the refrigerant pressure sensor 34 is read. Thereby, the detected pressure of the refrigerant pressure sensor 34, that is, the pressure of the high-pressure refrigerant is read. Hereinafter, the pressure of the high-pressure refrigerant is also referred to as high-pressure.

In step S4, the passage opening degree of the first expansion valve 25 is controlled based on the detected pressure of the refrigerant pressure sensor 34 so that the actual high pressure becomes the target pressure. Specifically, if the detected pressure is higher than the target pressure, the passage opening degree of the first expansion valve 25 is increased so that the actual high pressure is lowered. If the detected pressure is lower than the target pressure, the passage opening degree of the first expansion valve 25 is decreased and adjusted so that the actual high pressure is increased. In this way, the actual high pressure is brought close to the target pressure. In the present embodiment, step S4 constitutes an opening degree control unit that controls the passage opening degree of the first pressure reducer.

In step S5, the passage opening degree of the second expansion valve 27 is determined based on the temperature detected by the water temperature sensor 31. At this time, the passage opening degree of the second expansion valve 27 is determined so as to satisfy the relationship shown in FIG. That is, as shown in FIG. 5, the injection flow rate is increased as the hot-water supply temperature is higher, both in the case where the hot-water supply temperature is lower than the critical temperature of the refrigerant and in the case where the hot-water supply temperature is higher than the critical temperature of the refrigerant. The rate of increase in the injection flow rate is made larger when the hot water temperature is higher than the critical temperature than when it is lower than the critical temperature. The increase rate of the injection flow rate is the ratio of the increase amount of the injection flow rate to the increase amount of the hot water temperature.

Therefore, the passage opening degree of the second expansion valve 27 is larger as the hot water temperature is higher, and the passage opening degree is higher when the hot water temperature is higher than the critical temperature than when the hot water temperature is lower than the critical temperature. The increase rate is determined to be large. The increase rate of the passage opening is the ratio of the increase amount of the passage opening to the increase amount of the hot water temperature.

In step S6, the passage opening of the second expansion valve 27 is controlled so as to be the passage opening determined in step S5. Specifically, when the hot water temperature becomes high, the passage opening degree of the second expansion valve 27 is increased to increase the injection flow rate. On the other hand, when the hot water temperature is lowered, the passage opening degree of the second expansion valve 27 is reduced to reduce the injection flow rate. In this embodiment, Step S6 constitutes an opening degree control unit that controls the passage opening degree of the second pressure reducer.

Thus, when the detected hot water temperature is higher than the critical temperature of the refrigerant, the control device 30 increases the high pressure as the hot water temperature increases. Further, the control device 40 increases the injection flow rate as the hot water supply water temperature is higher, and increases the increase rate of the injection flow rate than when the hot water supply temperature is lower than the critical temperature of the refrigerant.

By the way, in the refrigeration cycle apparatus 20 having the configuration shown in FIG. 1, hot water flowing into the water-refrigerant heat exchanger 23 is generally within the temperature range of 5 ° C. or more and within 70 ° C., and is the critical temperature of the refrigerant 31. Temperature changes across ℃.

As shown in FIGS. 6-9, when the hot water temperature on the inlet side of the water refrigerant heat exchanger 23 is higher than the critical temperature, the water refrigerant is lower than when the hot water temperature is lower than the critical temperature. The decreasing rate of the enthalpy difference in the heat exchanger 23 increases. The decreasing rate of the enthalpy difference in the water refrigerant heat exchanger 23 is a ratio of the decreasing amount of the enthalpy difference of the refrigerant in the water refrigerant heat exchanger 23 to the increasing amount of the hot water temperature. Also, as shown in FIGS. 6-9, when the hot water temperature on the inlet side of the water-refrigerant heat exchanger 23 is higher than the critical temperature, the internal heat is higher than when the hot water temperature is lower than the critical temperature. The increasing rate of the enthalpy difference in the high-pressure side passage 24a of the exchanger 24 becomes large. The increase rate of the enthalpy difference in the high-pressure side passage 24a is the ratio of the increase amount of the enthalpy difference of the refrigerant in the high-pressure side passage 24a to the increase amount of the hot water temperature.

6, 7, 8, and 9 show examples of cycle balance when the hot water temperature on the inlet side of the water-refrigerant heat exchanger 23 is 25 ° C., 30 ° C., 32 ° C., and 37 ° C., respectively. 6 and 7 show cases where the hot water temperature is lower than the critical temperature. 8 and 9 show the case where the hot water temperature is higher than the critical temperature.

Further, points A1 to A9 in each figure indicate the state of the refrigerant at each position of the refrigeration cycle apparatus 20 shown in FIG. Point A1 indicates the state of the refrigerant on the suction side of the first compressor 21a. Point A2 shows the state of the refrigerant on the discharge side of the first compressor 21a. Point A3 indicates the state of the refrigerant on the suction side of the second compressor 21b and after the intermediate pressure refrigerant merges. Point A4 indicates the state of the refrigerant on the discharge side of the second compressor 21b. Point A5 is the outlet side of the refrigerant passage 23b of the water refrigerant heat exchanger 23, and shows the state of the refrigerant on the inlet side of the high-pressure side passage 24a of the internal heat exchanger 24 and on the inlet side of the second expansion valve 27. Show. Point A6 indicates the state of the refrigerant on the outlet side of the high-pressure side passage 24a of the internal heat exchanger 24. Point A <b> 7 indicates the state of the refrigerant on the outlet side of the first expansion valve 25. Point A8 indicates the state of the refrigerant on the outlet side of the second expansion valve 27 and on the inlet side of the intermediate pressure side passage 24b of the internal heat exchanger 24. Point A9 indicates the state of the refrigerant on the outlet side of the intermediate pressure side passage 24b of the internal heat exchanger 24 and before joining the refrigerant on the suction side of the second compressor 21b.

6-9, the high pressure (that is, the pressure at points A4, A5, and A6) is the same pressure. 6-9, the saturation temperature of the injected intermediate pressure refrigerant (that is, the temperature at the point A8) is set to the same temperature, specifically 15 ° C.

Theoretically, the refrigerant temperature on the outlet side of the water refrigerant heat exchanger 23 is equal to the hot water temperature on the inlet side of the water refrigerant heat exchanger 23. Therefore, in the following description, in FIGS. 6-9, it is assumed that the temperature at the point A5 is equal to the hot water temperature on the inlet side of the water-refrigerant heat exchanger 23. Similarly, theoretically, the temperature of the high-pressure refrigerant on the outlet side of the internal heat exchanger 24 is equal to the temperature of the intermediate-pressure refrigerant on the inlet side. Therefore, in the following description, in FIGS. 6-9, it is assumed that the temperature at point A6 is equal to the temperature at point A8.

The broken line in FIG. 7 indicates the cycle balance of FIG. 6 when the hot water temperature is 25 ° C. The arrow D1a shown in FIG. 7 has shown the enthalpy difference in the water refrigerant | coolant heat exchanger 23 when hot-water supply water temperature is 25 degreeC. The longer the arrow marked D1, the greater the amount of heat exchange between the refrigerant and hot water in the water / refrigerant heat exchanger 23. The same applies to other drawings. The arrow D2a shown in FIG. 7 has shown the enthalpy difference in the high voltage | pressure side of the internal heat exchanger 24 when hot-water supply water temperature is 25 degreeC. The larger the arrow marked D2, the greater the amount of heat exchange in the internal heat exchanger 24. The same applies to other drawings.

The thick solid line in FIG. 7 indicates the cycle balance when the hot water temperature is 30 ° C. An arrow D1b in FIG. 7 indicates the enthalpy difference in the water refrigerant heat exchanger 23 at this time. An arrow D2b in FIG. 7 indicates the enthalpy difference on the high pressure side of the internal heat exchanger 24 at this time.

The broken line in FIG. 9 shows the cycle balance of FIG. 8 when the hot water temperature is 32 ° C. The arrow D1c shown in FIG. 9 has shown the enthalpy difference in the water refrigerant | coolant heat exchanger 23 at this time. An arrow D2c shown in FIG. 9 indicates the enthalpy difference on the high-pressure side of the internal heat exchanger 24 at this time.

The thick solid line in FIG. 9 indicates the cycle balance when the hot water temperature is 37 ° C. An arrow D1d in FIG. 9 indicates the enthalpy difference in the water refrigerant heat exchanger 23 at this time. An arrow D2d in FIG. 7 indicates the enthalpy difference on the high-pressure side of the internal heat exchanger 24 at this time.

As can be seen by comparing the arrows D1a and D1b in FIG. 7, when the hot water temperature rises from 25 ° C. to 30 ° C., the enthalpy difference in the water-refrigerant heat exchanger 23 decreases. Similarly, as can be seen by comparing the arrows D1c and D1c in FIG. 9, when the hot water temperature rises from 32 ° C. to 37 ° C., the enthalpy difference in the water-refrigerant heat exchanger 23 decreases.

Then, comparing the difference in length between the arrows D1a and D1b in FIG. 7 and the difference in length between the arrows D1c and D1d in FIG. 9, the following can be understood. The amount of decrease in the enthalpy difference in the water-refrigerant heat exchanger 23 when the amount of rising hot water temperature is the same at 5 ° C. is higher than the critical temperature when the hot water temperature rises in a temperature range higher than the critical temperature. Greater than when it rises in the lower temperature range. In this way, when the hot water temperature is higher than the critical temperature, when the hot water temperature rises, the enthalpy difference in the water refrigerant heat exchanger 23 is greatly reduced as compared with the case where the hot water temperature is lower than the critical temperature. To do. This means that when the hot water temperature rises, the amount of heat released from the refrigerant in the water / refrigerant heat exchanger 23 is greatly reduced. For this reason, the heating capability of the water-refrigerant heat exchanger 23 is reduced, and hot water is difficult to boil.

Further, as can be seen by comparing the arrows D2a and D2b in FIG. 7, when the hot water temperature rises from 25 ° C. to 30 ° C., the enthalpy difference on the high pressure side of the internal heat exchanger 24 increases. Similarly, as can be seen by comparing the arrows D2c and D2d in FIG. 9, when the hot water temperature rises from 32 ° C. to 37 ° C., the enthalpy difference on the high pressure side of the internal heat exchanger 24 increases.

Then, comparing the difference in length between the arrows D2a and D2b in FIG. 7 and the difference in length between the arrows D2c and D2d in FIG. 9 reveals the following. The amount of increase in the enthalpy difference on the high pressure side of the internal heat exchanger 24 when the amount of the hot water temperature rises is the same 5 ° C. is higher when the hot water temperature rises in a temperature range higher than the critical temperature. It is larger than when it rises in a lower temperature range. As described above, when the hot water temperature is higher than the critical temperature, when the hot water temperature rises, the internal heat is increased as compared with the case where the hot water temperature is lower than the critical temperature, contrary to the water refrigerant heat exchanger 23. The enthalpy difference in the high-pressure side passage 24a of the exchanger 24 greatly increases.

Therefore, in order to suppress a decrease in the heating capacity of the water-refrigerant heat exchanger 23, when the hot-water supply temperature is higher than the critical temperature, it is conceivable that the higher the hot-water supply temperature, the higher the high-pressure pressure. An example of the cycle balance at this time is shown in FIG.

The thick solid line in FIG. 10 indicates the cycle balance when the hot water temperature is 37 ° C. and the high pressure is higher than that in FIG. The broken line in FIG. 10 shows the cycle balance of FIG. 8 when the hot water temperature is 32 ° C. In FIG. 10, the injection flow rate is the same as when the hot water temperature is 32 ° C. An arrow D1e in FIG. 10 indicates the enthalpy difference in the water refrigerant heat exchanger 23 at this time. An arrow D2e in FIG. 10 indicates the enthalpy difference on the high pressure side of the internal heat exchanger 24 at this time.

However, in the case where the hot water temperature is higher than the critical temperature, when the hot water temperature rises, if the high pressure is increased without changing the injection flow rate, as shown in FIG. The rapid increase of the enthalpy difference in the 24 high-pressure side passages 24a cannot be suppressed.

For this reason, the amount of heat that the intermediate pressure refrigerant receives from the high pressure refrigerant in the internal heat exchanger 24 increases, and the enthalpy of the intermediate pressure refrigerant injected between the first and

second compressors

21a and 21b increases. In other words, the temperature of the intermediate pressure refrigerant to be injected (that is, the temperature at point A9 in FIG. 10) increases. As a result, the injection effect is weakened, and the discharge refrigerant temperature of the second compressor 21b (that is, the temperature at point A4 in FIG. 10) increases. At this time, depending on the hot water temperature on the inlet side of the internal heat exchanger 24, the discharge refrigerant temperature becomes too high, and the reliability of the second compressor 21b is lowered.

Therefore, in this embodiment, when the hot water temperature is higher than the critical temperature of the refrigerant, the higher the hot water temperature, the higher the high pressure, and the higher the hot water temperature, the greater the injection flow rate. More specifically, regarding the injection flow rate, the rate of increase in the injection flow rate is greater when the hot water temperature is higher than the critical temperature of the refrigerant than when the hot water temperature is lower than the critical temperature of the refrigerant. I have to. An example of the cycle balance at this time is shown in FIG.

The thick solid line in FIG. 11 shows the cycle balance when the hot water temperature is 37 ° C. and the high pressure is made higher than in FIG. 9 and the injection flow rate is increased more than in FIG. ing. The broken line in FIG. 11 shows the cycle balance of FIG. 8 when the hot water temperature is 32 ° C. An arrow D1f in FIG. 10 indicates the enthalpy difference in the water refrigerant heat exchanger 23 at this time. An arrow D2f in FIG. 10 indicates the enthalpy difference on the high pressure side of the internal heat exchanger 24 at this time.

According to this, as shown in FIG. 11, the higher the hot water temperature, the higher the high pressure, so that the decrease in the enthalpy difference in the water refrigerant heat exchanger 23 can be suppressed. Therefore, it is possible to suppress a decrease in the heating capacity of the water refrigerant heat exchanger 23.

Further, when the hot water temperature is higher than the critical temperature of the refrigerant, the enthalpy difference in the high-pressure side passage 24a of the internal heat exchanger 24 increases sharply as compared with the case where the hot water temperature is equal to or lower than the critical temperature of the refrigerant. To do.

Therefore, in this embodiment, when the hot water temperature is higher than the critical temperature of the refrigerant, the injection flow rate is increased as the hot water temperature is higher. More specifically, when the hot water temperature is higher than the critical temperature of the refrigerant, the rate of increase of the injection flow rate is increased compared to the case where the hot water temperature is lower than the critical temperature of the refrigerant. This is because when the hot water temperature is higher than the critical temperature of the refrigerant, the rate of increase in the enthalpy difference in the high-pressure side passage 24a of the internal heat exchanger 24 is compared with the case where the hot water temperature is lower than the critical temperature of the refrigerant. Because is big.

Thereby, as shown at a point A9 in FIG. 11, it is possible to suppress an increase in the enthalpy of the injected intermediate pressure refrigerant. For this reason, the discharge temperature rise suppression effect by injection can be maintained.

Therefore, according to the refrigeration cycle apparatus 20 of the present embodiment, it is possible to suppress a decrease in the heating capacity of the water-refrigerant heat exchanger 23 and to improve the reliability of the second compressor 21b.

In this embodiment, as shown in FIG. 3, when the hot water temperature is equal to or lower than the critical temperature of the refrigerant, the target pressure is constant at a predetermined value regardless of the hot water temperature. This is because the influence on the heating capacity, COP, and discharge temperature rise is small even if it is constant.

Also, unlike the present embodiment, even when the hot water temperature is lower than the critical temperature of the refrigerant, the target pressure may be increased as the hot water temperature is higher. At this time, the rate of increase of the target pressure is set to be larger when the hot water temperature is higher than the critical temperature than when it is lower than the critical temperature. The increase rate of the target pressure is the ratio of the increase amount of the target pressure to the increase amount of the hot water temperature.

Incidentally, in FIG. 3, the rate of increase of the target pressure when the hot water temperature is below the critical temperature of the refrigerant is zero. Therefore, the rate of increase of the target pressure is greater when the hot water temperature is higher than the critical temperature than when it is lower than the critical temperature.

Further, in the present embodiment, as shown in FIG. 5, even when the hot water temperature is lower than the critical temperature of the refrigerant, the injection flow rate is increased as the hot water temperature is higher. The flow rate may be constant. At this time, the increase rate of the injection flow rate when the hot water temperature is equal to or lower than the critical temperature of the refrigerant is zero. Accordingly, even at this time, the increase rate of the injection flow rate is larger when the hot water temperature is higher than the critical temperature than when it is lower than the critical temperature.

(Second Embodiment)
In the present embodiment, the opening degree control of the second expansion valve 27 is different from the first embodiment. In the present embodiment, the control device 30 appropriately adjusts the pressure of the high-pressure refrigerant and the injection flow rate by executing the steps shown in FIG. 12 and controlling the operation of the refrigeration cycle. Steps S1 to S4 in FIG. 12 are the same as steps S1 to S4 in FIG.

In step S5-1, sensor signals of the first refrigerant temperature sensor 32 and the second refrigerant temperature sensor 33 are read. Thereby, the temperature of the intermediate pressure refrigerant on the inlet side of the internal heat exchanger 24 (namely, point A8 in FIG. 13) and the temperature of the intermediate pressure refrigerant on the outlet side of the internal heat exchanger 24 (namely, point A9 in FIG. 13). ) Temperature and read.

Subsequently, in step S5-2, based on the detected temperature of the first refrigerant temperature sensor 32 and the detected temperature of the second refrigerant temperature sensor 33, an intermediate is injected between the first compressor 21a and the second compressor 21b. The enthalpy of the pressure refrigerant (that is, the point A9 in FIG. 13) is calculated. For example, the control device 30 has a map that shows the relationship between the detected temperature of the first refrigerant temperature sensor 32, the detected temperature of the second refrigerant temperature sensor 33, and the enthalpy of the intermediate pressure refrigerant. Calculate the enthalpy of the pressurized refrigerant. Note that the control device 30 calculates the pressure at points A8 and A9 in FIG. 13 from the detected temperature of the first refrigerant temperature sensor 32, and calculates the pressure at the point A9 in FIG. 13 from the calculated pressure and the detected temperature of the second refrigerant temperature sensor 33. Enthalpy may be calculated. In the present embodiment, step S5-2 constitutes an enthalpy calculation unit.

Thus, the temperature of the intermediate pressure refrigerant on each of the inlet side and the outlet side of the internal heat exchanger 24 is a physical quantity of the intermediate pressure refrigerant related to the enthalpy of the intermediate pressure refrigerant. Therefore, the first refrigerant temperature sensor 32 and the second refrigerant temperature sensor 33 constitute a physical quantity detection unit that detects the physical quantity of the intermediate pressure refrigerant related to the enthalpy of the intermediate pressure refrigerant. The temperature detected by the first refrigerant temperature sensor 32 and the temperature detected by the second refrigerant temperature sensor 33 are detected physical quantities detected by the physical quantity detector.

Subsequently, in step S6-1, based on the calculated value calculated in step S5-2, as shown in FIG. 13, the enthalpy at the position of point A9 becomes a predetermined target value E1. The passage opening degree of the second expansion valve 27 is controlled. This target value E1 is a fixed value. Therefore, the control device 30 adjusts the injection flow rate so that the enthalpy becomes constant even when the hot water temperature changes. The target value E1 is determined by experiment, experience, or the like. In the present embodiment, step S6-1 constitutes an opening degree control unit that controls the passage opening degree of the second pressure reducer.

Here, as explained in the first embodiment, when the injection flow rate is kept constant, the enthalpy of the intermediate pressure refrigerant to be injected becomes larger as the enthalpy difference on the high pressure side of the internal heat exchanger 24 becomes larger. Therefore, keeping the enthalpy of the injected intermediate pressure refrigerant constant means that the injection flow rate is adjusted according to the change in the enthalpy difference on the high pressure side of the internal heat exchanger 24.

Further, in the first embodiment, as shown in FIG. 5, the injection flow rate is increased as the hot water temperature is higher. Further, when the hot water temperature is higher than the critical temperature of the refrigerant, the rate of increase of the injection flow rate is increased compared to the case where the hot water temperature is lower than the critical temperature of the refrigerant. Thereby, as shown by a point A9 in FIG. 11, an increase in the enthalpy of the injected intermediate pressure refrigerant is suppressed.

Therefore, adjusting the injection flow rate so as to keep the enthalpy of the intermediate pressure refrigerant to be injected constant is the same as adjusting the injection flow rate so as to satisfy the relationship shown in FIG. Therefore, also in this embodiment, there exists an effect similar to 1st Embodiment.

(Third embodiment)
In the present embodiment, the opening degree control of the second expansion valve 27 is different from the first embodiment. In the present embodiment, the control device 30 appropriately adjusts the pressure of the high-pressure refrigerant and the injection flow rate by executing the steps shown in FIG. 14 and controlling the operation of the refrigeration cycle. Steps S1 to S4 in FIG. 14 are the same as steps S1 to S4 in FIG.

In step S5-3, the sensor signals of the first refrigerant temperature sensor 32 and the second refrigerant temperature sensor 33 are read in the same manner as in step S5-1 of FIG. 12 of the second embodiment.

Subsequently, in step S5-4, based on the detected temperature of the first refrigerant temperature sensor 32 and the detected temperature of the second refrigerant temperature sensor 33, an intermediate is injected between the first compressor 21a and the second compressor 21b. The degree of superheat of the pressure refrigerant (that is, point A9 in FIG. 15) is calculated. The degree of superheat of the intermediate pressure refrigerant is a difference Tsh between the temperature of the saturated gas line of the intermediate pressure refrigerant and the temperature at point A9 in the Mollier diagram shown in FIG. Therefore, the degree of superheat of the intermediate pressure refrigerant is calculated by obtaining the difference between the temperature detected by the second refrigerant temperature sensor 33 and the temperature detected by the first refrigerant temperature sensor 32. In the present embodiment, step S5-4 constitutes a superheat degree calculation unit that calculates the superheat degree.

Thus, the temperature of the saturated gas line of the intermediate pressure refrigerant, that is, the temperature of the intermediate pressure refrigerant on the inlet side of the internal heat exchanger 24 and the temperature of the intermediate pressure refrigerant on the outlet side of the internal heat exchanger 24 are the intermediate pressure. It is a physical quantity of intermediate pressure refrigerant related to the degree of superheat of the refrigerant. Therefore, the 1st refrigerant | coolant temperature sensor 32 and the 2nd refrigerant | coolant temperature sensor 33 comprise the physical quantity detection part which detects the physical quantity of the intermediate pressure refrigerant | coolant relevant to the superheat degree of an intermediate pressure refrigerant | coolant. The detected temperature of the first refrigerant temperature sensor 32 and the detected temperature of the second refrigerant temperature sensor 33 are detected physical quantities detected by the physical quantity detector.

Subsequently, in step S6-2, based on the calculated value calculated in step S5-4, as shown in FIG. 15, the superheat degree Tsh at point A9 is set to a predetermined target value Tsh1. The passage opening degree of the second expansion valve 27 is controlled. This target value Tsh1 is a fixed value. Therefore, the control device 30 adjusts the injection flow rate so that the degree of superheat Tsh is constant even when the hot water temperature changes. The target value Tsh1 is determined by experiment, experience, or the like. In the present embodiment, step S6-2 constitutes an opening degree control unit that controls the passage opening degree of the second pressure reducer.

Here, in the first embodiment, as shown in FIG. 5, the injection flow rate is increased as the hot water temperature is higher. Further, when the hot water temperature is higher than the critical temperature of the refrigerant, the rate of increase of the injection flow rate is increased compared to the case where the hot water temperature is lower than the critical temperature of the refrigerant. Thereby, as shown by a point A9 in FIG. 11, an increase in the enthalpy of the injected intermediate pressure refrigerant is suppressed. Suppressing the increase in enthalpy is equivalent to suppressing the degree of superheat.

Therefore, adjusting the injection flow rate so as to keep the superheat degree Tsh of the intermediate pressure refrigerant to be injected constant is the same as adjusting the injection flow rate so as to satisfy the relationship shown in FIG. Therefore, also in this embodiment, there exists an effect similar to 1st Embodiment.

Further, the superheat degree of the intermediate pressure refrigerant to be injected can be easily obtained by the first and second

refrigerant temperature sensors

32 and 33. Therefore, according to the present embodiment, the injection flow rate can be adjusted with simple control.

(Other embodiments)
The present disclosure is not limited to the above-described embodiment, and can be appropriately changed within the scope described in the claims as follows.

(1) In each of the above embodiments, the control device 30 controls the passage opening of the first expansion valve 25 to control the operation of the refrigeration cycle and adjust the pressure of the high-pressure refrigerant. You may go. For example, the control device 30 may control the operation of the refrigeration cycle by controlling the rotation speeds of the first and

second compressors

21a and 21b to adjust the pressure of the high-pressure refrigerant.

(2) In each of the above embodiments, the two-stage compression mechanism 21 is configured using two single-

stage compressors

21a and 21b. However, a two-stage compression mechanism having another structure is employed. May be. For example, a single compressor of two-stage compression type in which two compression units are accommodated in one container may be used. Further, a scroll type compressor that has an intermediate pressure port and injects high-pressure refrigerant from the intermediate pressure port into the refrigerant in the middle of the compression process may be used.

(3) In each of the above embodiments, the water stored in the hot water storage tank 11 is heated by the water refrigerant heat exchanger 23. However, the water is heated without being stored in the hot water storage tank 11, and the heated water is used for hot water supply or heating. Also good.

(4) In each of the above embodiments, the heat radiation from the radiator of the refrigeration cycle apparatus 20 is used for heating water used for hot water supply or heating, but may be used for other heating applications. For example, heat dissipation by the radiator may be performed on air instead of water.

(5) In each of the above embodiments, the refrigeration cycle apparatus 20 is used for heating, but may be used for cooling.

(6) In the above embodiments, carbon dioxide is used as the refrigerant of the refrigeration cycle apparatus 20, but other refrigerants in which the pressure of the high-pressure refrigerant discharged from the compression mechanism becomes a supercritical pressure may be used.

(7) In each of the above-described embodiments, each function unit such as the enthalpy calculation unit, the superheat degree calculation unit, and the pressure determination unit is realized by the function of the control device 30, but at least a part of these function units is provided. You may implement | achieve by the control part (hardware etc.) different from the control apparatus 30. FIG. In this case, the control device 30 and another control unit constitute a control device that controls the operation of the refrigeration cycle and adjusts the pressure of the high-pressure refrigerant and the flow rate of the intermediate-pressure refrigerant.

(8) The above embodiments are not irrelevant to each other, and can be combined as appropriate unless the combination is clearly impossible. In each of the above-described embodiments, it is needless to say that elements constituting the embodiment are not necessarily essential unless explicitly stated as essential and clearly considered essential in principle. Yes.

Claims

A refrigeration cycle apparatus,
The refrigerant is compressed from a low pressure to a high pressure higher than the low pressure, and a high-pressure refrigerant in a supercritical state is discharged. A compression mechanism (21) for introducing an intermediate pressure refrigerant that is a pressure;
A radiator (23) for radiating the high-pressure refrigerant by heat exchange between the high-pressure refrigerant discharged from the compression mechanism and a heat exchange medium;
A first pressure reducer (25) that depressurizes a part of the high-pressure refrigerant flowing out of the radiator to the low pressure to form a low-pressure refrigerant;
An evaporator (26) for evaporating the low-pressure refrigerant and sucking the low-pressure refrigerant after evaporation into the compression mechanism;
A second pressure reducer (27) that reduces the other part of the high-pressure refrigerant flowing out of the radiator to the intermediate pressure to obtain an intermediate-pressure refrigerant;
An internal heat exchanger that exchanges heat between the high-pressure refrigerant that flows out from the radiator and flows toward the first pressure reducer, and the intermediate-pressure refrigerant that flows out from the second pressure reducer and flows toward the compression mechanism (24) and
A temperature detector (31) for detecting the temperature of the heat exchange medium flowing into the radiator;
Based on the temperature detected by the temperature detector, the refrigeration cycle is configured to include the compression mechanism, the radiator, the first decompressor, the evaporator, the second decompressor, and the internal heat exchanger. A controller (30) for controlling the operation to adjust the pressure of the high-pressure refrigerant and the flow rate of the intermediate-pressure refrigerant;
When the detected temperature is higher than the critical temperature of the refrigerant, the control device adjusts the pressure of the high-pressure refrigerant so that the higher the detected temperature, the higher the pressure of the high-pressure refrigerant, and the detected temperature A refrigeration cycle apparatus that adjusts the flow rate of the intermediate pressure refrigerant so that the flow rate of the intermediate pressure refrigerant increases as the flow rate increases.
When the detected temperature is higher than the critical temperature of the refrigerant, the control device increases the flow rate of the intermediate pressure refrigerant as the detected temperature is higher, and the detected temperature is lower than the critical temperature. 2. The refrigeration cycle apparatus according to claim 1, wherein the flow rate of the intermediate pressure refrigerant is adjusted so that the ratio of the increase amount of the flow rate of the intermediate pressure refrigerant to the increase amount of the detected temperature is increased.
The second decompressor is configured such that the passage opening is adjusted,
The refrigeration cycle apparatus according to claim 1 or 2, wherein the control device adjusts a flow rate of the intermediate pressure refrigerant by controlling a passage opening degree of the second decompressor.
A physical quantity detector (32, 33) for detecting a physical quantity of the intermediate pressure refrigerant associated with the enthalpy of the intermediate pressure refrigerant that has flowed out of the internal heat exchanger;
The control device includes an enthalpy calculating unit (S5-2) that calculates the enthalpy based on a physical quantity detected by the physical quantity detecting unit, and a target in which the enthalpy is determined based on a calculated value of the enthalpy calculating unit. The refrigeration cycle apparatus according to claim 3, further comprising an opening degree control unit (S6-1) for controlling a passage opening degree of the second pressure reducer so as to be a value.
A physical quantity detector (32, 33) for detecting a physical quantity of the intermediate pressure refrigerant related to the degree of superheat of the intermediate pressure refrigerant that has flowed out of the internal heat exchanger;
The control device includes a superheat degree calculation unit (S5-4) that calculates the superheat degree based on the physical quantity detected by the physical quantity detection unit, and the superheat degree is calculated in advance based on the calculated value of the superheat degree calculation unit. The refrigeration cycle apparatus according to claim 3, further comprising an opening degree control unit (S6-2) for controlling a passage opening degree of the second pressure reducer so as to be a predetermined target value.
A pressure detector (34) for detecting the pressure of the high-pressure refrigerant;
The first pressure reducer is configured such that the passage opening is adjusted,
The control device is configured to determine a target pressure of the high-pressure refrigerant based on the detected temperature, and to determine whether the pressure of the high-pressure refrigerant is equal to the target pressure based on the detected pressure of the pressure detection unit. An opening degree control unit (S4) for controlling the passage opening degree of the first pressure reducer,
The said pressure determination part determines the said target pressure so that the pressure of the said high pressure refrigerant may become so high that the said detection temperature is high, when the said detection temperature is higher than the critical temperature of the said refrigerant | coolant. The refrigeration cycle apparatus according to any one of the above.
The refrigerant is carbon dioxide;
The target pressure is higher than a pressure at an intersection of an isodensity line of 600 kg / m 3 on the Mollier diagram of the refrigerant and an isothermal line having the same temperature as the detected temperature, and 700 kg on the Mollier diagram of the refrigerant. The refrigeration cycle apparatus according to claim 6, wherein the refrigeration cycle apparatus is lower than a pressure at an intersection of an isodensity line of / m 3 and an isothermal line having the same temperature as the detected temperature.
The refrigeration cycle apparatus according to any one of claims 1 to 7, wherein the heat exchange medium is hot water used for hot water supply or heating.