JP2009115415A - Supercritical refrigerating cycle - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a supercritical refrigerating cycle in which the electric power of a compressor can be reduced by the improvement of a coefficient of performance in a cycle stabilized period and quick cooling performance can be attained by the improvement of transit performance in a cool down transit period. <P>SOLUTION: The supercritical refrigerating cycle connects the compressor 1, a gas cooler 2, an expansion valve 3, an evaporator 4 and an accumulator 5 sequentially in an annular form. A main high temperature side heat exchanging part 6a and a sub high temperature side heat exchanging part 7a are set in series along a refrigerant path from an outlet of a gas cooler 2 to an inlet of the expansion valve 3, and a main low temperature side heat exchanging part 6b and a sub low temperature side heat exchanging part 7b are set on both sides of the accumulator 5 in a refrigerant path from an outlet of the expansion valve 3 to an inlet of the compressor 1. An inner heat exchanger is composed of a main inner heat exchanger 6 combining the main high temperature side heat exchanging part 6a and the main low temperature side heat exchanging part 6b, and a sub inner heat exchanger 7 combining the sub high temperature side heat exchanging part 7a and the sub low temperature side heat exchanging part 7b. <P>COPYRIGHT: (C)2009,JPO&INPIT

Description

The present invention relates to a supercritical refrigeration cycle that uses carbon dioxide refrigerant (CO ₂ refrigerant), which is a natural refrigerant, and whose high pressure exceeds the critical pressure at high outside air temperature.

  Conventionally, as a supercritical refrigeration cycle, a compressor, a gas cooler, an expansion valve, an evaporator, and an accumulator are sequentially connected in an annular manner, and between a high-temperature / high-pressure refrigerant that has exited the gas cooler and a low-temperature / low-pressure refrigerant that has exited the accumulator. The thing provided with the internal heat exchanger which heat-exchanges by is known (for example, refer patent document 1).

In the supercritical refrigeration cycle using this CO ₂ refrigerant, the high pressure exceeds the critical pressure at high outside air temperature, and liquid refrigerant does not accumulate on the high pressure side, so the amount of refrigerant is adjusted with an accumulator. The internal heat exchanger exchanges heat between the high-temperature / high-pressure refrigerant that exits the gas cooler and the low-temperature / low-pressure refrigerant that exits the accumulator, and lowers the enthalpy at the evaporator inlet to reduce the coefficient of performance COP (Coefficient Of Performance). ).
JP 11-193967 A

However, since the conventional supercritical refrigeration cycle has a configuration in which only one internal heat exchanger is installed for the purpose of improving the coefficient of performance COP, it is listed below due to the physical properties of the CO ₂ refrigerant. There was a serious problem.

  (1) At high load, there is a limit to the improvement of the coefficient of performance COP, and the compressor power becomes excessive. In other words, the internal heat exchanger reduces the expansion valve inlet refrigerant temperature Tex and improves the coefficient of performance COP.However, the lower the expansion valve inlet refrigerant temperature Tex, the higher the compressor discharge temperature Td. There is a limit to improving the coefficient of performance COP from the viewpoint. Note that the compressor discharge temperature Td needs to be kept below a limit temperature (for example, 140 ° C.).

  (2) When the load fluctuates due to a change in the fan air volume or the like, the superheat degree SH (superheat) at the evaporator outlet fluctuates, leading to deterioration of the evaporator temperature distribution. In other words, since the evaporator outlet dryness X is controlled to be X = 1.0 in the accumulator, the superheat degree SH is temporarily removed when the load fluctuates due to a change in the fan air volume and the temperature distribution of the evaporator deteriorates.

  These measures include increasing the size of the internal heat exchanger. However, when the size is increased, the compressor discharge temperature Td rises and exceeds the limit temperature, so there is a limit even if the size of the internal heat exchanger is increased. Moreover, the internal diameter of the oil bleed opened to the refrigerant discharge pipe of the accumulator is increased to reduce the evaporator outlet dryness. However, there are many demerits that the refrigerant amount adjustment is inappropriate and the oil circulation rate is increased.

  Here, the “coefficient of performance COP” refers to a coefficient obtained by the equation of evaporator performance / compressor power. “Superheat degree SH” refers to the difference between the temperature of superheated steam and the temperature of saturated steam (boiling point). “Dryness X (also referred to as dryness)” refers to the ratio of X to 1 kg of wet steam, when X is included in 1 kg of wet steam. Therefore, saturated steam becomes X = 1 and saturated liquid becomes X = 0.

  The present invention has been made paying attention to the above-mentioned problem, and it is possible to achieve a reduction in compressor power by improving the coefficient of performance in the cycle stable period, and to improve the cooling performance by improving the transient performance in the cool-down transition period. The object is to provide a supercritical refrigeration cycle that can be achieved.

In order to achieve the above object, in the present invention, an internal heat exchanger in which a compressor, a gas cooler, an expansion valve, an evaporator, and an accumulator are sequentially connected in an annular manner to reduce the inlet refrigerant temperature of the expansion valve by internal heat exchange. In the supercritical refrigeration cycle with
A main high temperature side heat exchange part and a sub high temperature side heat exchange part are set in series along a refrigerant path from the outlet of the gas cooler to the inlet of the expansion valve, and the refrigerant path from the outlet of the expansion valve to the inlet of the compressor The main low temperature side heat exchange part and the sub low temperature side heat exchange part are respectively set at positions on both sides of the accumulator,
The internal heat exchanger is a combination of a main internal heat exchanger that combines the main high temperature side heat exchange section and the main low temperature side heat exchange section, and a combination of the sub high temperature side heat exchange section and the sub low temperature side heat exchange section. It is characterized by comprising a sub-internal heat exchanger.

Therefore, in the supercritical refrigeration cycle of the present invention, the compressor discharge temperature Td is first protected by setting the main low temperature side heat exchange section and the sub low temperature side heat exchange section at positions on both sides of the accumulator, respectively. The main internal heat exchanger and the sub internal heat exchanger can be installed in the cycle while keeping below the temperature.
And by reducing the expansion valve inlet refrigerant temperature Tex by the sub internal heat exchanger added to the main internal heat exchanger, the liquefaction of the refrigerant can be promoted even at a critical pressure or higher, and the high pressure can be lowered. It becomes possible. Thereby, in the cycle stable period, the compressor compression ratio is lowered, and the coefficient of performance COP can be improved by reducing the compression enthalpy in the compressor.
During the cool-down transition period, for example, the sub internal heat exchanger lowers the expansion valve inlet refrigerant temperature Tex instead of the main internal heat exchanger, even if the internal heat exchange performance by the main internal heat exchanger is difficult. Can improve the cool-down performance.
As a result, a reduction in compressor power can be achieved by improving the coefficient of performance during the cycle stabilization period, and quick cooling can be achieved by improving the transient performance during the cool-down transition period.

  Hereinafter, the best mode for realizing the supercritical refrigeration cycle of the present invention will be described based on Examples 1 to 3 shown in the drawings.

First, the configuration will be described.
FIG. 1 is a cycle system diagram showing a CO ₂ refrigeration cycle (supercritical refrigeration cycle) of Example 1, (a) shows a schematic configuration of the cycle, and (b) shows a schematic configuration of the cycle arranged on the Mollier diagram. Show. 2 is a cross-sectional view showing an example of an accumulator used in the CO ₂ refrigeration cycle of Example 1. FIG.

The CO ₂ refrigeration cycle of Example 1 using R744 (CO ₂ refrigerant), which is a natural refrigerant, includes a compressor 1, a gas cooler 2, an expansion valve 3, an evaporator 4, and an accumulator 5, as shown in FIG. A main internal heat exchanger 6 and a sub internal heat exchanger 7 that are sequentially connected in an annular shape and reduce the inlet refrigerant temperature of the expansion valve 3 (hereinafter referred to as “expansion valve inlet refrigerant temperature Tex”) by internal heat exchange are provided. Consists of.

The compressor 1 is driven by an engine, a motor, or the like, and compresses the gas refrigerant from the accumulator 5 to obtain a high-temperature / high-pressure gas refrigerant. In the first embodiment, an external variable displacement control type having a differential pressure control ECV (External Control Valve) for controlling a differential pressure between a high pressure and a low pressure is employed. The saturated gas of R744 (CO ₂ refrigerant) used as a refrigerant is 7 times the density of HFC134a (fluorine refrigerant) and 1.2 times the latent heat of evaporation (per unit mass), so the cooling capacity per unit volume is about It becomes 8 times. For this reason, if the discharge capacity of the compressor 1 is about 15 to 30 cc, sufficient performance can be exhibited.

  The gas cooler 2 is a condenser that exchanges heat between the high-temperature and high-pressure gas refrigerant from the compressor 1 with the outside air to obtain a low-temperature and high-pressure gas refrigerant. The gas cooler 2 includes a pair of left and right header tanks arranged in parallel with each other at intervals, a heat exchange tube arranged in a large number in parallel with both ends connected to the header tank, and adjacent heat tanks. And a fin disposed in the air circulation gap of the exchange tube. Then, the interior of the pair of header tanks is partitioned in the lateral direction by the partitioning means, so that at least two refrigerant passages by the heat exchange tubes are provided, such as an inlet side passage group, an intermediate passage group, and an outlet side passage group. It is divided into passage groups.

  The expansion valve 3 is installed in the engine room, and the pressure of the high-pressure gas refrigerant from the gas cooler 2 is a low-pressure gas-liquid two-phase refrigerant. In the case of Example 1, a control type expansion valve that controls the expansion valve opening degree (orifice opening degree) so as to keep the superheat degree SH of the refrigerant constant based on the outlet refrigerant temperature and the outlet refrigerant pressure of the gas cooler 2 is adopted. is doing.

The evaporator 4 is a heat exchanger that is disposed together with a blower or the like in a vehicle air conditioning unit 8 that performs air conditioning in the vehicle interior. By circulating the low-temperature and low-pressure gas-liquid two-phase refrigerant from the expansion valve 3, heat is taken from the surrounding air, the temperature of the refrigerant is increased, and gasification is promoted. In the refrigeration cycle using R744 (CO ₂ refrigerant), the equilibrium pressure at high load is as high as 7 MPa (about 70 bar) or higher, so refrigerant leakage into the passenger compartment is more reliable than HFC134a (fluorine refrigerant). It is necessary to ensure sex. For this reason, the evaporator 4 has a structure in which the core, piping, and flange are integrated, thereby preventing refrigerant leakage into the vehicle compartment including slow leaks such as O-ring seals.

The accumulator 5 separates gas-liquid from the gas-liquid two-phase refrigerant introduced from the evaporator 4, supplies gas refrigerant to the compressor 1, and stores excess liquid refrigerant in the CO ₂ refrigeration cycle inside the main body. In the refrigeration cycle using R744 (CO ₂ refrigerant), liquid refrigerant does not accumulate on the high pressure side when the high pressure exceeds the critical pressure. For this reason, a general liquid tank cannot be adopted in a refrigeration cycle using HFC134a (fluorine refrigerant), and an appropriate amount of refrigerant is managed using the accumulator 5.

  The structure of the accumulator 5 will be described with reference to FIG. In the accumulator 5, a desiccant layer 54 that defines an upper layer 52 and a lower layer 53 is set in an intermediate portion of the internal space of the container body 51. The head portion 55 is provided with a refrigerant suction pipe 56 for introducing the gas-liquid two-phase refrigerant from the evaporator 4 and a U-shaped refrigerant discharge pipe 57 for discharging the gas refrigerant to the compressor 1. In the U-shaped refrigerant discharge pipe 57, a gas refrigerant suction port 58 is opened in the upper layer 52, and an oil bleed 59 is opened in the lowest U-shaped curved position immersed in the liquid refrigerant in the lower layer 53. .

  In this accumulator 5, the gas-liquid two-phase refrigerant introduced into the container main body 51 is passed through the desiccant layer 54 from the upper layer 52 to the lower layer 53, thereby separating the gas and liquid, stabilizing the liquid refrigerant liquid level, and Remove moisture. Then, the gas refrigerant separated from the gas and liquid and having passed through the desiccant layer 54 and transferred to the upper layer 52 is sucked through the gas refrigerant suction port 58 set in the upper part of the refrigerant discharge pipe 57 and returned to the compressor 1. Also, the lubricating oil that has settled below the liquid refrigerant separated from the gas and liquid is sucked from the oil bleed 59 and returned to the compressor using the flow of the refrigerant. Further, excess liquid refrigerant in the refrigeration cycle is stored in the container body 51.

  The accumulator 5 uses the oil / refrigerant compatibility characteristics to control the oil circulation rate and the outlet refrigerant state with an oil bleed diameter of about 1 mm in diameter. The accumulator outlet refrigerant has a dryness X of X = 0.9. The oil circulation rate OCR is controlled to OCR = 2-7%.

  In Embodiments 1 to 3 of the present invention, the main high temperature side heat exchange unit 6a and the sub high temperature side heat exchange unit 7a are set in series along the refrigerant path from the outlet of the gas cooler 2 to the inlet of the expansion valve 3. Yes. In addition, a main low temperature side heat exchanging portion 6b and a sub low temperature side heat exchanging portion 7b are set at positions on both sides of the refrigerant path from the outlet of the expansion valve 3 to the inlet of the compressor 1 with the accumulator 5 interposed therebetween. is doing.

The internal heat exchanger in the CO ₂ refrigeration cycle includes a main internal heat exchanger 6 in which the main high temperature side heat exchange unit 6a and the main low temperature side heat exchange unit 6b are combined, and the sub high temperature side heat exchange unit 7a. And the sub internal heat exchanger 7 in which the sub low temperature side heat exchanging portion 7b is combined.

  In Example 1, as shown in FIGS. 1 (a) and 1 (b), the main internal heat exchanger 6 is connected to a main high temperature side heat exchanger 6a set in the refrigerant path immediately after the outlet of the gas cooler 2, It is configured by a combination with a main low temperature side heat exchange section 6b set in a refrigerant path from the outlet of the accumulator 5 to the inlet of the compressor 1. This main internal heat exchanger 6 keeps the compressor discharge temperature Td below the limit temperature (about 140 ° C in the Td protection line in Fig. 1 (b)), and the expansion valve inlet exceeding the critical temperature (= 31.1 ° C). Internal heat exchange is performed to lower the refrigerant temperature Tex to near the critical temperature of the gas region and the liquid region (31.1 ° C. line in FIG. 1B).

  A sub high temperature side heat exchanging part 7a in which the sub internal heat exchanger 7 is set as a refrigerant path downstream from the main high temperature side heat exchanging part 6a, and a refrigerant path from the outlet of the evaporator 4 to the inlet of the accumulator 5 And a sub-low temperature side heat exchanging part 7b set in the above. This sub internal heat exchanger 7 has an internal valve that lowers the expansion valve inlet refrigerant temperature Tex, which has been reduced by internal heat exchange in the main internal heat exchanger 6, until it enters the liquid region across the critical temperature (= 31.1 ° C.). Perform heat exchange.

Next, the operation will be described.
First, the “supercritical refrigeration cycle technology” will be described, and then the operation in the supercritical refrigeration cycle of Example 1 will be compared with the conventional supercritical refrigeration cycle by comparing “compressor power reduction effect”, “cool down”. The description will be divided into “quick cooling action in the transition period” and “improvement in the stability of the evaporator outlet dryness and the action of increasing the appropriate refrigerant width”.

[Supercritical refrigeration cycle technology]
In order to respond to the increasing global trend to prevent global warming, the development of an air conditioning system using a CO ₂ refrigeration cycle using R744 (CO ₂ refrigerant), a natural refrigerant, is proceeding at a rapid pace.

The current refrigeration cycle uses HFC134a (fluorine refrigerant). When comparing the current refrigeration cycle with the CO ₂ refrigeration cycle, the CO ₂ refrigeration cycle has a higher operating pressure than the current refrigeration cycle, and CO 2 _{The 2} refrigeration cycle is about 7 to 10 times the operating pressure of the current refrigeration cycle. For this reason, it is indispensable to give pressure resistance to the components, but it is necessary to suppress an increase in weight at the same time as the pressure resistance design, and further, control to maintain a safe pressure becomes important.

The characteristics of R744 (CO ₂ refrigerant) are that the critical temperature is as low as 31.1 ° C, and the high pressure exceeds the critical pressure (critical pressure 7.4MPa or higher and temperature 31.1 ° C or higher) when the outside air temperature is about 30 ° C or higher. It is called the critical region). For this reason, the CO ₂ refrigeration cycle is called a supercritical refrigeration cycle.

  As shown in FIG. 3, the current supercritical refrigeration cycle includes one internal heat exchanger for exchanging heat between the high-temperature / high-pressure refrigerant discharged from the gas cooler and the low-temperature / low-pressure refrigerant discharged from the accumulator. It is known to improve the coefficient of performance COP by lowering the enthalpy at the entrance.

However, in the current supercritical refrigeration cycle, only one internal heat exchanger is installed for the purpose of improving the coefficient of performance (COP). This is due to the physical properties of the CO ₂ refrigerant. There is a limit to the improvement of the coefficient COP, and the compressor power becomes excessive. Further, when the load fluctuates due to a change in the fan air volume or the like, the degree of superheat SH at the evaporator outlet fluctuates, leading to deterioration in the evaporator temperature distribution.

  These measures include increasing the size of the internal heat exchanger. However, since the compressor discharge temperature Td rises and exceeds the limit temperature when the size is increased, there is a limit even if the size of the internal heat exchanger is increased. Moreover, the internal diameter of the oil bleed opened to the refrigerant discharge pipe of the accumulator is increased to reduce the evaporator outlet dryness. However, the refrigerant amount adjustment is inappropriate, and further, the oil circulation rate is increased.

Furthermore, supercritical refrigeration cycles have been proposed in Japanese Patent Application Laid-Open No. 2007-155229, Japanese Patent Application Laid-Open No. 2007-51841, Japanese Patent Application Laid-Open No. 11-351680, etc., all of which increase the coefficient of performance COP at high loads. Compressor power cannot be reduced. That is, as a CO ₂ refrigeration cycle, reduction of compressor power at high load remains as a major problem to be solved.

  The present inventor can set a low temperature side heat exchanging portion at a position from the expansion valve outlet to the accumulator inlet in the low pressure side refrigerant path from the expansion valve outlet to the compressor inlet, and is a gas-liquid two-phase refrigerant. We paid attention to the fact that the cooling energy for internal heat exchange is higher than that of the gas refrigerant on the outlet side of the accumulator.

  In accordance with this point of interest, a main high temperature side heat exchange section and a sub high temperature side heat exchange section are set in series along the refrigerant path from the gas cooler outlet to the expansion valve inlet, and the internal heat exchanger is connected to the main high temperature side heat exchange section. And a main internal heat exchanger with a combination of the main low temperature side heat exchange section and a sub internal heat exchanger with a combination of the sub high temperature side heat exchange section and the sub low temperature side heat exchange section.

  By adopting this configuration, it is possible to achieve a reduction in compressor power by improving the coefficient of performance during the cycle stabilization period, and it is possible to achieve rapid cooling by improving the transient performance during the cool-down transition period. I did it.

[Compressor power reduction action]
First, in the CO ₂ refrigeration cycle of the first embodiment, as shown in FIG. 1 (b), a main low temperature side heat exchange section 6b and a sub low temperature side heat exchange section 7b are set at positions on both sides of the accumulator 5, respectively. ing. For this reason, by making the main low temperature side heat exchange section 6b equal to the size of the low temperature side heat exchange section in the current CO ₂ refrigeration cycle, the compressor discharge temperature Td can be kept below the protection temperature (see the Td protection line in FIG. 1 (b)). Or less). Then, a main high temperature side heat exchange section 6a and a sub high temperature side heat exchange section 7a are set in series along the refrigerant path from the outlet of the gas cooler 2 to the inlet of the expansion valve 3 in the CO ₂ refrigeration cycle, and the main low temperature side heat is set. The main internal heat exchanger 6 and the sub internal heat exchanger 7 can be installed by combining the exchange unit 6b and the sub low temperature side heat exchange unit 7b.

Then, the evaporator performance in cycles plateau at CO ₂ refrigeration cycle of Example 1 based on FIG. 4 will be described. FIG. 4 is a comparison diagram of the Mollier diagram of the cycle with sub-internal heat of Example 1 and the conventional cycle shown in FIG. 3 in the cycle stabilization period.

  In the conventional cycle, as shown in the dotted Mollier diagram of FIG. 4, the enthalpy increase amount a 'due to the high pressure side internal heat exchange and the enthalpy increase amount b' due to the low pressure side internal heat exchange cancel each other. Therefore, the actual enthalpy used for heat exchange in the evaporator is A '.

  On the other hand, in the cycle with sub-internal heat of the first embodiment, as shown in the solid Mollier diagram of FIG. 4, the enthalpy increase amount a by two high-pressure side internal heat exchanges and the enthalpy increase by two low-pressure side internal heat exchanges. Mass b cancels. Therefore, the actual enthalpy used for heat exchange in the evaporator 4 is A.

As is clear from the above comparison, the sub-internally heated cycle cancels out the increase in enthalpy by high-low pressure internal heat exchange as in the conventional cycle. Therefore, the actual enthalpy becomes A′≈A, and the evaporator performance in the cycle stable period in the CO ₂ refrigeration cycle of Example 1 is equivalent to the evaporator performance in the cycle stable period in the conventional cycle.

Next, compressor power in the cycle stable period in the CO ₂ refrigeration cycle of Example 1 will be described with reference to FIG.
In the cycle stabilization period in the CO ₂ refrigeration cycle of the first embodiment, the expansion valve inlet refrigerant temperature Tex is lowered until the gas refrigerant is liquefied by the sub internal heat exchanger 7 added to the main internal heat exchanger 6. Can do. Thereby, as shown in FIG. 4, the high pressure (Pd) in the cycle stable period in the sub-heated cycle can be reduced by ΔPd with respect to the high pressure in the cycle stable period in the conventional cycle.

  As a result, the compressor compression ratio is lowered by ΔPd in the cycle stable period in the sub-heated cycle, and the compression enthalpy E in the compressor 1 is reduced to the compression enthalpy in the cycle stable period in the conventional cycle as shown in FIG. Reduced compared to E ′.

  As a result, the compressor power to which the compressor 1 gives the compression energy can be reduced, and furthermore, the same evaporator performance can be obtained as described above, so that the coefficient of performance COP obtained by the equation of evaporator performance / compressor power can be improved. it can.

  Incidentally, FIG. 5 is a diagram showing a table in which each measured value according to the presence or absence of the sub internal heat exchanger, the compressor power, and the coefficient of performance COP are compared. In the example of the table shown in FIG. 5, the compressor power without the sub internal heat exchanger is 2.74 KW, whereas the compressor power with the sub internal heat exchanger is 2.55 KW, so that the compressor power can be reduced. confirmed. In addition, it was confirmed that the coefficient of performance COP can be improved such that the coefficient of performance COP without the sub internal heat exchanger is 1.77, whereas the coefficient of performance COP with the sub internal heat exchanger is 1.90.

[Cool-down transition rapid cooling]
First, the evaporator performance in the cool-down transition period in the CO ₂ refrigeration cycle of Example 1 will be described with reference to FIG. FIG. 6 is a comparison diagram of the Mollier diagram of the sub-internally heated cycle of Example 1 and the conventional cycle shown in FIG. 3 in the cool-down transition period (for example, when 10 minutes have elapsed).

  During the cool-down transition period, the gas refrigerant continues in a dry state at the evaporator outlet, so that both the internal heat exchanger of the conventional system and the main internal heat exchanger 6 of the system with sub-internal heat exchange heat almost. Absent. That is, since both the internal heat exchanger of the conventional system and the main internal heat exchanger 6 of the system with sub-internal heat exchange internal heat in a gas refrigerant state with a superheat degree SH, as shown in e ′ and e of FIG. In addition, both enthalpies become smaller.

  Thus, in the cool-down transition period, the internal heat exchange performance by the main internal heat exchanger 6 is difficult to be obtained. However, in the case of a system with sub internal heat, the sub internal heat exchanger 7 is the main internal heat exchanger. Instead of 6, the expansion valve inlet refrigerant temperature Tex can be greatly reduced.

  Therefore, compared with the evaporator enthalpy F ′ of the conventional system, the evaporator enthalpy F of the system with sub-internal heat increases, the evaporator 4 cools quickly during the cool-down transition period, and effectively takes heat from the high-temperature wind passing through the evaporator 4. As a result, the cool-down performance is improved.

Incidentally, FIG. 7 is a diagram showing a comparison characteristic of the expansion valve inlet refrigerant temperature Tex and a comparison of the direct temperature of the evaporator (= the temperature immediately after the evaporator) with the passage of time from the start of cool-down due to the presence or absence of the sub internal heat exchanger.
As for the expansion valve inlet refrigerant temperature Tex, when the sub-internal heat exchanger is provided, as shown by the solid line characteristic in FIG. 7, the expansion valve inlet refrigerant temperature immediately after the start of the air conditioner is compared with the dotted line characteristic without the sub-internal heat exchanger. The effect of lowering Tex is great, for example, a difference of about 9 ° C appears after 7 minutes. Therefore, the difference between the enthalpy of the evaporator inlet and the evaporator outlet can be increased at an early stage immediately after the air conditioner is started. As for the direct temperature of the evaporator, when the sub internal heat exchanger is present, the direct temperature of the evaporator is lower than the dotted line characteristic without the sub internal heat exchanger, as shown by the solid line characteristics in FIG. After a minute, there will be a difference of 2 ° C or more.

Next, the compressor power in the cool-down transition period in the CO ₂ refrigeration cycle of Example 1 will be described.
When the sub internal heat exchanger 7 exhibits heat exchange performance, cycle formation is quick, and accordingly, the refrigerant flow rate increases and the compressor power increases. Thus, during the cool-down transition period, the compressor power at the same time has increased. However, since the coefficient of performance COP does not change, only the compressor power increases as the cooling capacity of the evaporator 4 increases.

[Improvement of evaporator outlet dryness and increase of proper refrigerant width]
In Example 1, the sub low temperature side heat exchanging part 7 b is set in the refrigerant path from the outlet of the evaporator 4 to the inlet of the accumulator 5. For this reason, as described below, the improvement in the stability of the evaporator outlet dryness and the effect of increasing the appropriate refrigerant width are exhibited.

FIG. 8 is a characteristic diagram showing the dryness of the outlet and the dryness of the inlet with respect to the amount of refrigerant charged in the conventional cycle.
In the conventional cycle, since the evaporator outlet is connected to the accumulator as shown in FIG. 3, the dryness of the evaporator outlet is stabilized at about 1.0 in the range of the proper refrigerant width 200 g as shown in FIG.

FIG. 9 is a characteristic diagram showing the dryness of the evaporator outlet and the dryness of the evaporator inlet with respect to the refrigerant charging amount when the oil bleed diameter is increased in the conventional cycle.
When the oil bleed diameter is increased in the conventional cycle, as shown in FIG. 9, the dryness of the evaporator outlet with respect to the refrigerant charging amount gradually decreases as the refrigerant charging amount increases. That is, there is no stable range of dryness at the outlet of the evaporator, and, for example, it changes in summer or winter.

FIG. 10 is a characteristic diagram showing the degree of dryness of the outlet, the degree of dryness of the inlet, and the degree of dryness of the sub-internal heat outlet with respect to the amount of refrigerant filled in the cycle with sub-internal heat.
As in the first embodiment, in the case of the cycle with sub-internal heat, the sub-internal heat exchanger 7 performs a constant heat exchange. Therefore, as shown in FIG. In the range, X can be stably lowered so as to maintain a value of about 0.8. Furthermore, as is clear from the comparison with Fig. 8, the appropriate refrigerant width has been increased from 200g for the conventional system to 250g for the system with sub-internal heat, improving stability against large load fluctuations. Can be made.

By the way, comparing the load fluctuation resistance of the current refrigeration cycle using HFC134a (fluorine refrigerant) and the CO ₂ refrigeration cycle using R744 (CO ₂ refrigerant), the amount of heat held by the evaporator (0 ° C) is HFC134a ( R744 (CO ₂ refrigerant) is 25 kJ, while fluorine refrigerant) is 140 kJ. Therefore, R744 (CO ₂ refrigerant) has a heat quantity smaller than that of HFC134a (fluorine refrigerant) and exhibits a characteristic that is not stable against load fluctuations.

Next, as an example of load fluctuation, when the fan air volume is switched from low (Lo) to high (Hi), the stability is improved in the case of a CO ₂ refrigeration cycle with sub-internal heat as shown in FIGS. Based on

FIG. 11 is a diagram showing the evaporator outlet dryness characteristics without sub-internal heat and the evaporator outlet dryness characteristics with sub-internal heat (200 mm × 3 stages) when the fan air flow is switched from low (Lo) to high (Hi). It is.
As for the evaporator outlet dryness, when there is no sub-internal heat, as shown by the dotted line characteristics in FIG. 11, when the fan airflow is low, the airflow changes slightly below the dryness X = 1.0. When switched to Hi, the dryness increases to a value slightly greater than 1.0 = 1.0, and after about 2 minutes 30 seconds, the dryness converges to a value of 1.0 or less.
On the other hand, when there is sub-internal heat, as shown by the solid line characteristics in FIG. 11, when the fan air volume is low, the dryness value X is smaller than 1.0 (about X = 0.7). When the air volume is switched to Hi, the dryness momentarily becomes a value slightly larger than X = 1.0, but the dryness converges to a value of 1.0 or less from the point before 2 minutes have passed.
That is, since it is in a gas-liquid two-phase state from the entrance to the exit of the evaporator, the temperature distribution in the evaporator is good. Note that the evaporator outlet dryness is 1.0 or less in the gas-liquid two-phase saturation state, and 1.0 or more is the superheat range (temperature rises).

Figure 12 shows the evaporator outlet superheat characteristics without sub-internal heat and the evaporator outlet superheat characteristics with sub-internal heat (200mm x 3 stages) when the fan air flow is switched from low (Lo) to high (Hi). It is.
As for the evaporator outlet superheat degree, when there is no sub-internal heat, as shown in the dotted line characteristic of FIG. 12, when the fan air volume is switched from low (Lo) to high (Hi), the time when the evaporator outlet dryness is 1.0 or more Since it continues for a long time, the superheat time lasts more than 2 minutes.
On the other hand, when there is sub-internal heat, as shown in the solid line characteristics in FIG. 12, when the fan air volume is switched from low (Lo) to high (Hi), the evaporator dryness is 1.0 or more. As a result, the superheat time is within 2 minutes.
That is, when there is sub-internal heat, the value of the superheat degree immediately after the fan air volume change and the time in the superheat degree region can be greatly reduced.

FIG. 13 is a graph showing the temperature characteristics immediately after the evaporator without sub-internal heat and the temperature characteristics immediately after the evaporator with sub-internal heat (200 mm × 3 stages) when the fan air volume is switched from low (Lo) to high (Hi). .
With regard to the temperature immediately after the evaporator, when there is no sub-internal heat, as shown in the dotted line characteristics of FIG. 13, when the fan air flow is switched from low (Lo) to high (Hi), the temperature immediately after the evaporator varies greatly by about 10 ° C. Is seen.
On the other hand, when there is sub-internal heat, as shown in the solid line characteristics in FIG. 13, even if the fan air volume is switched from low (Lo) to high (Hi), the temperature immediately after the evaporator fluctuates by about 5 ° C. Can be suppressed.
As a result, the temperature distribution (maximum temperature-minimum temperature) of the evaporator is reduced, and the cooling performance can be improved.

Next, the effect will be described.
In the supercritical refrigeration cycle of Example 1, the effects listed below can be obtained.

  (1) The compressor 1, the gas cooler 2, the expansion valve 3, the evaporator 4, and the accumulator 5 are sequentially connected in an annular manner, and an internal heat exchanger that lowers the inlet refrigerant temperature of the expansion valve 3 by internal heat exchange is provided. In the supercritical refrigeration cycle, a main high temperature side heat exchange section 6a and a sub high temperature side heat exchange section 7a are set in series along a refrigerant path from the outlet of the gas cooler 2 to the inlet of the expansion valve 3, and the expansion valve Main low temperature side heat exchanging part 6b and sub low temperature side heat exchanging part 7b in the refrigerant path from the outlet 3 to the inlet of the compressor 1 with the accumulator 5 in between. The main internal heat exchanger 6 combining the main high temperature side heat exchange unit 6a and the main low temperature side heat exchange unit 6b, the sub high temperature side heat exchange unit 7a, and the sub low temperature Since the sub-internal heat exchanger 7 is combined with the side heat exchanger 7b, it is possible to achieve a reduction in compressor power by improving the coefficient of performance during the cycle stabilization period, and at the same time by improving the transient performance during the cool-down transition period. Coolness can be achieved.

  (2) The main internal heat exchanger 6 performs internal heat exchange to lower the expansion valve inlet refrigerant temperature Tex exceeding the critical temperature to near the critical temperature of the gas region and the liquid region, and the sub internal heat exchanger 6 7 performs an internal heat exchange in which the expansion valve inlet refrigerant temperature Tex, which has been lowered by the internal heat exchange in the main internal heat exchanger 6, is lowered until it enters the liquid region across the critical temperature, and both the internal heat exchangers 6 and 7, the heat exchange amount by the main low temperature side heat exchange section 6b provided on the outlet side of the accumulator 5 is set to an amount that keeps the compressor discharge temperature Td below the limit temperature. By reducing the valve inlet refrigerant temperature Tex until it is liquefied, the high pressure (Pd) can be lowered, and the compressor power can be reliably reduced.

(3) The main internal heat exchanger 6 is set to a main high temperature side heat exchanging portion 6a set as a refrigerant path immediately after the outlet of the gas cooler 2, and a refrigerant path from the outlet of the accumulator 5 to the inlet of the compressor 1. The sub internal heat exchanger 7 includes a sub high temperature side heat exchange unit 7a set in a refrigerant path downstream from the main high temperature side heat exchange unit 6a. The sub-internal heat exchanger 7 performs a constant heat exchange because the sub-internal heat exchanger 7 performs a constant heat exchange because the sub-internal heat exchanger 7 is configured to be combined with the sub-low temperature side heat exchanging portion 7b set in the refrigerant path from the outlet of the evaporator 4 to the inlet of the accumulator 5. The outlet dryness of No. 4 can be stably reduced, and a CO ₂ refrigeration cycle strong against sudden load fluctuations can be obtained. In addition, the appropriate refrigerant width is increased, and the stability of the CO ₂ refrigeration cycle can be improved against large load fluctuations such as changes in fan airflow.

  The second embodiment is an example in which the sub low temperature side heat exchanging portion 7b is set at a position different from the first embodiment.

First, the configuration will be described.
FIG. 14 is a cycle system diagram showing the CO ₂ refrigeration cycle (supercritical refrigeration cycle) of Example 2. (a) shows a schematic configuration of the cycle, and (b) shows a schematic configuration of the cycle arranged on the Mollier diagram. Show.

As shown in FIG. 14, the CO ₂ refrigeration cycle using the natural refrigerant R744 (CO ₂ refrigerant) includes a compressor 1, a gas cooler 2, an expansion valve 3, an evaporator 4, and an accumulator 5. The main internal heat exchanger 6 and the sub internal heat exchanger 7 that are sequentially connected in an annular shape and reduce the refrigerant temperature at the inlet of the expansion valve 3 by internal heat exchange are provided.

  In Example 2, as shown in FIGS. 14 (a) and 14 (b), the main internal heat exchanger 6 has a main high temperature side heat exchange section 6a set in the refrigerant path immediately after the outlet of the gas cooler 2, It is configured by a combination with a main low temperature side heat exchange section 6b set in a refrigerant path from the outlet of the accumulator 5 to the inlet of the compressor 1. The main internal heat exchanger 6 keeps the compressor discharge temperature Td below the limit temperature, and supplies the expansion valve inlet refrigerant temperature Tex exceeding the critical temperature (= 31.1 ° C.) near the critical temperature of the gas region and the liquid region ( The internal heat exchange is reduced to 31.1 ° C. in FIG.

  A sub high temperature side heat exchanger 7a in which the sub internal heat exchanger 7 is set as a refrigerant path downstream from the main high temperature side heat exchanger 6a, and refrigerant from the outlet of the expansion valve 3 to the inlet of the evaporator 4 It is configured by a combination with the sub low temperature side heat exchange section 7b set in the path. This sub internal heat exchanger 7 has an internal valve that lowers the expansion valve inlet refrigerant temperature Tex, which has been reduced by internal heat exchange in the main internal heat exchanger 6, until it enters the liquid region across the critical temperature (= 31.1 ° C.). Perform heat exchange. Since other configurations are the same as those of the first embodiment, the corresponding configurations are denoted by reference numerals and description thereof is omitted.

  Next, regarding the operation, in the second embodiment, the sub low temperature side heat exchanging portion 7b is set as a refrigerant path from the outlet of the expansion valve 3 to the inlet of the evaporator 4 different from the first embodiment. For this reason, the evaporator outlet dryness is 1.0, and the effect of improving the evaporator temperature distribution with respect to the fan air volume change or the like cannot be obtained. However, since the internal heat exchanger is composed of the main internal heat exchanger 6 and the sub internal heat exchanger 7, the “compressor power reduction action” and the “cooling action in the cool-down transition period” are performed as in the first embodiment. Obtainable.

Next, the effect will be described.
In the vehicle supercritical refrigeration cycle of the second embodiment, in addition to the effects (1) and (2) of the first embodiment, the following effects can be obtained.

  (4) The main internal heat exchanger 6 is set to a main high temperature side heat exchanging part 6a set as a refrigerant path immediately after the outlet of the gas cooler 2, and a refrigerant path from the outlet of the accumulator 5 to the inlet of the compressor 1. The sub internal heat exchanger 7 includes a sub high temperature side heat exchange unit 7a set in a refrigerant path downstream from the main high temperature side heat exchange unit 6a. And the sub low temperature side heat exchanging part 7b set in the refrigerant path from the outlet of the expansion valve 3 to the inlet of the evaporator 4, the sub internal heat exchanger 7 has an inlet and an outlet of the expansion valve 3. Due to this temperature difference, internal heat exchange can be performed effectively until the refrigerant at the inlet of the expansion valve 3 becomes liquid refrigerant.

  Example 3 is a setting in which the positions of the main low temperature side heat exchange unit 6b and the sub low temperature side heat exchange unit 7b are changed with respect to the setting of the main low temperature side heat exchange unit 6b and the sub low temperature side heat exchange unit 7b of Example 1. It is an example.

First, the configuration will be described.
FIG. 15 is a cycle system diagram showing the CO ₂ refrigeration cycle (supercritical refrigeration cycle) of Example 3, where (a) shows a schematic cycle configuration, and (b) shows a cycle schematic configuration arranged on the Mollier diagram. Show.

As shown in FIG. 15, the CO ₂ refrigeration cycle of Example 3 using R744 (CO ₂ refrigerant), which is a natural refrigerant, includes a compressor 1, a gas cooler 2, an expansion valve 3, an evaporator 4, and an accumulator 5. The main internal heat exchanger 6 and the sub internal heat exchanger 7 that are sequentially connected in an annular shape and reduce the refrigerant temperature at the inlet of the expansion valve 3 by internal heat exchange are provided.

  In Example 3, as shown in FIGS. 15 (a) and 15 (b), the main internal heat exchanger 6 has a main high temperature side heat exchange section 6a set in a refrigerant path immediately after the outlet of the gas cooler 2, It is configured by a combination with a main low temperature side heat exchange section 6b set in the refrigerant path from the outlet of the evaporator 4 to the inlet of the accumulator 5. This main internal heat exchanger 6 reduces the expansion valve inlet refrigerant temperature Tex, which exceeds the critical temperature (= 31.1 ° C), to near the critical temperature of the gas region and the liquid region (31.1 ° C line in Fig. 14 (b)). Perform internal heat exchange.

  A sub high temperature side heat exchanging part 7a in which the sub internal heat exchanger 7 is set as a refrigerant path downstream of the main high temperature side heat exchanging part 6a, and a refrigerant path from the outlet of the accumulator 5 to the inlet of the compressor 1 And a sub-low temperature side heat exchanging part 7b set in the above. The sub-internal heat exchanger 7 reduces the expansion valve inlet refrigerant temperature Tex, which has decreased due to the internal heat exchange in the main internal heat exchanger 6, while keeping the compressor discharge temperature Td below the limit temperature. ) To reduce the internal heat exchange until it enters the liquid region. Since other configurations are the same as those of the first embodiment, the corresponding configurations are denoted by reference numerals and description thereof is omitted.

  Next, in the third embodiment, the main low temperature side heat exchanging portion 6b and the sub low temperature side heat exchanging portion 7b are set to be different from those in the first embodiment. Therefore, by controlling the evaporator outlet dryness by the main low temperature side heat exchanging section 6b, as in the first embodiment, "compressor power reduction action", "cooling action during cool-down transition period", "evaporator outlet dryness" Improvement in degree of stability and an effect of increasing the appropriate refrigerant width ”.

Next, the effect will be described.
In the vehicle supercritical refrigeration cycle of the third embodiment, the following effects can be obtained in addition to the effects (1) and (2) of the first embodiment.

(5) The main internal heat exchanger 6 is set to a main high temperature side heat exchanging portion 6a set as a refrigerant path immediately after the outlet of the gas cooler 2, and a refrigerant path from the outlet of the evaporator 4 to the inlet of the accumulator 5. The sub internal heat exchanger 7 includes a sub high temperature side heat exchange unit 7a set in a refrigerant path downstream from the main high temperature side heat exchange unit 6a. Since the main internal heat exchanger 6 performs a constant heat exchange because it is configured by a combination with the sub low temperature side heat exchange section 7b set in the refrigerant path from the outlet of the accumulator 5 to the inlet of the compressor 1, the evaporator The outlet dryness of No. 4 can be stably reduced, and a CO ₂ refrigeration cycle strong against sudden load fluctuations can be obtained. In addition, the appropriate refrigerant width is increased, and the stability of the CO ₂ refrigeration cycle can be improved against large load fluctuations such as changes in fan airflow.

  As described above, the supercritical refrigeration cycle of the present invention has been described based on Examples 1 to 3. However, the specific configuration is not limited to these examples, and each claim of the claims Design changes and additions are permitted without departing from the spirit of the invention.

  In Examples 1-3, the example which sets each of the main internal heat exchanger 6 and the sub internal heat exchanger 7 independently was shown. However, for example, the main and sub high-temperature side heat exchanging portions may be configured integrally, and in appearance, it may be configured as if having one internal heat exchanger.

  In short, the main high temperature side heat exchange section and the sub high temperature side heat exchange section are set in series along the refrigerant path from the outlet of the gas cooler to the inlet of the expansion valve, and this is the refrigerant path from the outlet of the expansion valve to the inlet of the compressor. The main low temperature side heat exchange section and the sub low temperature side heat exchange section are set on both sides of the accumulator, and the internal heat exchanger is combined with the main high temperature side heat exchange section and the main low temperature side heat exchange section. The configuration is not limited to the first to third embodiments as long as the main internal heat exchanger and the sub internal heat exchanger in which the sub high temperature side heat exchange unit and the sub low temperature side heat exchange unit are combined.

  In Examples 1 to 3, an example of a supercritical refrigeration cycle applied to a vehicle air conditioner system is shown. However, as a supercritical refrigeration cycle other than a vehicle, for example, a home air conditioner system or a factory or office air conditioner system. Is also applicable. In short, if it is a supercritical refrigeration cycle equipped with an internal heat exchanger that sequentially connects the compressor, gas cooler, expansion valve, evaporator, and accumulator in an annular fashion and reduces the inlet refrigerant temperature of the expansion valve by internal heat exchange Applicable.

A cycle system diagram showing a CO ₂ refrigeration cycle of Example 1 (supercritical refrigerating cycle), (a) shows a cycle schematic structure shows the (b) cycle schematic configuration disposed in Mollier chart. 3 is a cross-sectional view showing an example of an accumulator used in the CO ₂ refrigeration cycle of Example 1. FIG. It is a figure which shows the cycle schematic structure which has arrange | positioned the present supercritical refrigerating cycle (conventional cycle) on the Mollier diagram. It is a comparison figure of the Mollier diagram of the cycle with the sub internal heat of Example 1 in the cycle stable period, and the conventional cycle shown in FIG. It is a figure which shows the table | surface which contrasted each measured value by the presence or absence of a sub internal heat exchanger, compressor power, and a coefficient of performance COP. FIG. 4 is a comparison diagram of a Mollier diagram of a sub-internally heated cycle of Example 1 and a conventional cycle shown in FIG. 3 in a cool-down transition period (for example, when 10 minutes have elapsed). It is a figure which shows the comparative characteristic of the expansion valve inlet_port | entrance refrigerant | coolant temperature Tex by the time passage from the cooldown start by the presence or absence of a sub internal heat exchanger, and the evaporative temperature (= temperature immediately after an evaporator). It is a characteristic view which shows the degree of dryness of the evaporating outlet and the evaporating degree of the evaporating inlet with respect to the refrigerant filling amount in the conventional cycle. It is a characteristic view which shows the dryness of the evaporating exit and the evaporating rate of the evaporating entrance with respect to the refrigerant filling amount when the oil bleed diameter is increased in the conventional cycle. It is a characteristic view which shows the dryness of an evaporator outlet, the dryness of an evaporator inlet, and the dryness of an internal heat outlet of a refrigerant with respect to the refrigerant filling amount in a cycle with sub-internal heat. It is a figure which shows the evaporator outlet dryness characteristic without sub internal heat, and the evaporator outlet dryness characteristic with sub internal heat (200 mm × 3 stages) when the fan air volume is switched from low (Lo) to high (Hi). It is a figure which shows the evaporator outlet superheat degree characteristic without sub internal heat, and the evaporator outlet superheat degree characteristic with sub internal heat (200 mm x 3 stages) when the fan air volume is switched from low (Lo) to high (Hi). It is a figure which shows the temperature characteristic immediately after the evaporator without sub internal heat, and the temperature characteristic immediately after the evaporator with sub internal heat (200 mm × 3 stages) when the fan air volume is switched from low (Lo) to high (Hi). A cycle system diagram showing a CO ₂ refrigeration cycle of Example 2 (supercritical refrigerating cycle), (a) shows a cycle schematic structure shows the (b) cycle schematic configuration disposed in Mollier chart. A cycle system diagram showing a CO ₂ refrigeration cycle of Example 3 (supercritical refrigeration cycle), (a) shows a cycle schematic structure shows the (b) cycle schematic configuration disposed in Mollier chart.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Compressor 2 Gas cooler 3 Expansion valve 4 Evaporator 5 Accumulator 6 Main internal heat exchanger 6a Main high temperature side heat exchange part 6b Main low temperature side heat exchange part 7 Sub internal heat exchanger 7a Sub high temperature side heat exchange part 7b Sub low temperature side heat exchange Part 8 Vehicle air conditioning unit

Claims (5)

In a supercritical refrigeration cycle having an internal heat exchanger that sequentially connects a compressor, a gas cooler, an expansion valve, an evaporator, and an accumulator in an annular manner, and reduces the inlet refrigerant temperature of the expansion valve by internal heat exchange.
A main high temperature side heat exchange part and a sub high temperature side heat exchange part are set in series along a refrigerant path from the outlet of the gas cooler to the inlet of the expansion valve, and the refrigerant path from the outlet of the expansion valve to the inlet of the compressor The main low temperature side heat exchange part and the sub low temperature side heat exchange part are respectively set at positions on both sides of the accumulator,
The internal heat exchanger is a combination of a main internal heat exchanger that combines the main high temperature side heat exchange section and the main low temperature side heat exchange section, and a combination of the sub high temperature side heat exchange section and the sub low temperature side heat exchange section. A supercritical refrigeration cycle comprising a sub-internal heat exchanger. In the supercritical refrigeration cycle according to claim 1,
The main internal heat exchanger performs internal heat exchange for reducing the expansion valve inlet refrigerant temperature exceeding the critical temperature to near the critical temperature of the gas region and the liquid region,
The sub-internal heat exchanger performs an internal heat exchange in which the expansion valve inlet refrigerant temperature lowered by the internal heat exchange in the main internal heat exchanger is lowered until entering the liquid region across the critical temperature,
Of the internal heat exchangers, the supercritical refrigeration characterized in that the heat exchange amount by the low temperature side heat exchange section provided on the outlet side of the accumulator is set to an amount that keeps the compressor discharge temperature below the limit temperature. cycle. In the supercritical refrigeration cycle according to claim 1 or 2,
The main internal heat exchanger includes: a main high temperature side heat exchange unit set in a refrigerant path immediately after the gas cooler outlet; a main low temperature side heat exchange unit set in a refrigerant path from the outlet of the accumulator to the inlet of the compressor; , A combination of
The sub internal heat exchanger includes a sub high temperature side heat exchange section set in a refrigerant path downstream from the main high temperature side heat exchange section, and a sub low temperature set in a refrigerant path from the outlet of the evaporator to the inlet of the accumulator. A supercritical refrigeration cycle comprising a combination with a side heat exchanger. In the supercritical refrigeration cycle according to claim 1 or 2,
The main internal heat exchanger includes: a main high temperature side heat exchange unit set in a refrigerant path immediately after the gas cooler outlet; a main low temperature side heat exchange unit set in a refrigerant path from the outlet of the accumulator to the inlet of the compressor; , A combination of
The sub internal heat exchanger includes a sub high temperature side heat exchange section set in a refrigerant path downstream from the main high temperature side heat exchange section, and a sub set in a refrigerant path from the outlet of the expansion valve to the inlet of the evaporator. A supercritical refrigeration cycle comprising a combination with a low temperature side heat exchange section. In the supercritical refrigeration cycle according to claim 1 or 2,
The main internal heat exchanger includes a main high temperature side heat exchange unit set in a refrigerant path immediately after the gas cooler outlet, and a main low temperature side heat exchange unit set in a refrigerant path from the evaporator outlet to the accumulator inlet. , A combination of
The sub internal heat exchanger includes a sub high temperature side heat exchange unit set in a refrigerant path downstream from the main high temperature side heat exchange unit, and a sub low temperature set in a refrigerant path from an outlet of the accumulator to an inlet of the compressor. A supercritical refrigeration cycle comprising a combination with a side heat exchanger.

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