WO2018216187A1

WO2018216187A1 - Refrigeration cycle device

Info

Publication number: WO2018216187A1
Application number: PCT/JP2017/019664
Authority: WO
Inventors: 智隆石川; 悠介有井; 久登森田
Original assignee: 三菱電機株式会社
Priority date: 2017-05-26
Filing date: 2017-05-26
Publication date: 2018-11-29
Also published as: JPWO2018216187A1; CN110678707A

Abstract

This refrigeration cycle device is provided with: a refrigerant circuit to which a compressor, a condenser, a pressure reducer, and an evaporator are connected and in which a non-azeotropic refrigerant mixture circulates; and a temperature control means for controlling the saturated liquid temperature of the refrigerant so as to be equal to or higher than the temperature of a cooling fluid in the condenser.

Description

Refrigeration cycle equipment

The present invention relates to a refrigeration cycle apparatus using a non-azeotropic refrigerant mixture.

Conventionally, a technique using a non-azeotropic refrigerant mixture in a refrigeration cycle apparatus such as a refrigerator or an air conditioner has been disclosed (for example, see Patent Document 1). Patent Document 1 discloses that heat exchange efficiency is improved by making the heat exchange fluids of the cooling fluid and the refrigerant to be subjected to heat exchange counter-flow in the refrigerant condensation process. Yes. For example, in the case of air cooling, the cooling fluid is ambient air.

JP 2000-320917 A

However, even if the non-azeotropic refrigerant mixture and the cooling fluid in which the temperature gradient occurs in the refrigerant condensing process are opposed to each other, if the low load operation is maintained or the heat exchange performance in the condenser continues to be improved, The saturated liquid temperature becomes lower than the fluid temperature. At this time, since the refrigerant has a temperature substantially equal to that of the cooling fluid in the course of the condensation process, no further heat exchange is possible, and the refrigerant flows out of the condenser without becoming a saturated liquid. As a result, sufficient heating capacity or cooling capacity cannot be exhibited, resulting in a decrease in refrigeration performance. Although the heat exchange performance is improved, the performance may decrease.

The present invention has been made to solve the above-described problems, and provides a refrigeration cycle apparatus that can suppress a decrease in operating performance in a refrigeration cycle using a non-azeotropic refrigerant mixture.

The refrigeration cycle apparatus according to the present invention includes a refrigerant circuit in which a compressor, a condenser, a decompression device, and an evaporator are connected, in which a non-azeotropic refrigerant is circulated, and a saturated fluid temperature of the refrigerant is a cooling fluid in the condenser Temperature control means for making the temperature equal to or higher than this temperature.

According to the present invention, since the temperature control means for making the saturated liquid temperature of the refrigerant equal to or higher than the temperature of the cooling fluid in the condenser is provided, even if the refrigerant is a non-azeotropic refrigerant, the temperature is higher than the temperature of the cooling fluid. Saturated liquid temperature is maintained and condensation is promoted. For this reason, the fall of driving performance can be controlled.

It is a figure which shows the example of 1 structure of the refrigerating-cycle apparatus in Embodiment 1 of this invention. 6 is a Ph diagram illustrating a condensation process for a single refrigerant or an azeotropic refrigerant mixture as Comparative Example 1. FIG. 6 is a Ph diagram illustrating a condensation process for a non-azeotropic refrigerant mixture as Comparative Example 2. FIG. It is a figure for demonstrating the structure of the temperature control means in Embodiment 1 of this invention. FIG. 5 is a Ph diagram illustrating a condensation process in the condenser according to the first embodiment of the present invention. It is a figure which shows the example of 1 structure of the refrigerating-cycle apparatus in Embodiment 2 of this invention. It is a figure which shows the example of 1 structure of the refrigerating-cycle apparatus in Embodiment 3 of this invention. It is a figure which shows the example of 1 structure of the refrigerating-cycle apparatus in Embodiment 4 of this invention. It is a figure which shows the other structural example of the refrigeration cycle apparatus in Embodiment 4 of this invention. It is a figure which shows the example of 1 structure of the refrigerating-cycle apparatus in Embodiment 5 of this invention. It is a figure which shows the example of 1 structure of the refrigerating-cycle apparatus in Embodiment 6 of this invention. It is sectional drawing which shows the case where the condenser shown in FIG. 1 is a flat tube heat exchanger.

Embodiment 1 FIG.
The configuration of the refrigeration cycle apparatus according to the first embodiment will be described. FIG. 1 is a diagram illustrating a configuration example of a refrigeration cycle apparatus according to Embodiment 1 of the present invention. As shown in FIG. 1, the refrigeration cycle apparatus 1 includes a compressor 2, a condenser 3, a fan 6, a decompression device 4, and an evaporator 5. The compressor 2, the condenser 3, the decompression device 4, and the evaporator 5 are connected in order by refrigerant piping, and the refrigerant circuit 10 in which a refrigerant circulates is comprised.

The compressor 2 compresses the refrigerant to be sucked and discharges the refrigerant in a high temperature and high pressure state. The compressor 2 is a variable capacity compressor of a type that can control the number of revolutions by an inverter circuit or the like and adjust the discharge amount of the refrigerant.

The condenser 3 is an air heat exchanger that exchanges heat between the refrigerant and air. The condenser 3 cools the gas refrigerant discharged from the compressor 2 by heat exchange with outdoor air (outside air), water, brine, or the like as a heat source. In this Embodiment 1, the condenser 3 consists of a heat source side heat exchanger, and the heat source (cooling fluid) which the condenser 3 cools a refrigerant | coolant shall be outside air. The fan 6 serves to supply outside air to the condenser 3 and to promote heat exchange to the condenser 3. The fan 6 can adjust the air volume.

The decompression device 4 decompresses and expands the refrigerant flowing through the refrigerant circuit 10. The decompression device 4 is an expansion valve, a throttle device, or the like. The decompression device 4 includes, for example, a flow rate control unit such as an electronic expansion valve, a capillary tube (capillary tube), and a refrigerant flow rate control unit such as a temperature-sensitive expansion valve.

The evaporator 5 evaporates the refrigerant flowing through the refrigerant circuit 10 to be a gaseous refrigerant in order to maintain the set temperature by heat exchange between the object to be cooled and the refrigerant. The evaporator 5 evaporates and gasifies the refrigerant. In the evaporator 5, the object to be cooled is cooled directly or indirectly by heat exchange with the refrigerant. The evaporator 5 is composed of a use side heat exchanger, and the cooling target is, for example, air in the air conditioning target space.

In the refrigeration cycle apparatus 1 of Embodiment 1, the refrigerant circulating in the refrigerant circuit 10 is a non-azeotropic refrigerant mixture. Non-azeotropic refrigerant mixtures are, for example, R407C and R448A. The non-azeotropic refrigerant mixture is a refrigerant mixture of R32, R125, R134a, R1234yf, and CO ₂ , wherein the ratio of R32 XR32 (wt%) is 33 <XR32 <39, and the ratio of R125 XR125 (Wt%) is 27 <XR125 <33, R134a ratio XR134a (wt%) is 11 <XR134a <17, and R1234yf ratio XR1234yf (wt%) is 11 <XR1234yf <17. and conditions, the ratio of CO ₂ _XCO ₂ (wt%) is 3 <XR125 <and conditions is 9, a refrigerant satisfying all the conditions sum of XR32 and XR125 and XR134a and XR1234yf and XCO ₂ is 100, the Also good. In the first embodiment, it is assumed that the refrigerant is R448A.

Since most of the energy consumed by the refrigeration cycle apparatus 1 is occupied by the compression power of the compressor 2, the operation performance improves when the compressor power is reduced. One way to reduce compressor power is to reduce the compression ratio. In order to reduce the compression ratio, the suction pressure of the compressor 2 may be increased or the discharge pressure may be decreased. Here, attention is focused on lowering the discharge pressure in order to improve the operation performance, that is, lowering the outlet pressure of the compressor 2.

On the other hand, the refrigeration effect corresponding to the refrigerant latent heat of the refrigeration cycle apparatus 1 has a sufficient effect by completely converting the refrigerant into a condensate in the condenser 3 or completely evaporating the refrigerant in the evaporator 5. Obtainable. Here, attention is paid to the fact that the refrigerant is completely condensed by the condenser 3 in order to improve the operation performance.

In order to achieve the above-described improvement in operating performance, when the refrigerant is a single refrigerant or an azeotropic refrigerant mixture, the heat transfer performance of the condenser 3 can be improved to reduce the discharge pressure and promote the condensation of the refrigerant. Conceivable. When a non-azeotropic refrigerant mixture is used as in the first embodiment, a temperature gradient occurs even if the condensation pressure is constant. Therefore, if the heat transfer performance of the condenser 3 is improved as in the conventional case, the operation performance is remarkably increased. It may be damaged. Below, this performance fall is demonstrated using a comparative example.

As Comparative Example 1, the case where the refrigerant is a single refrigerant or an azeotropic refrigerant mixture will be described. FIG. 2 is a Ph diagram showing a condensation process for a single refrigerant or an azeotropic refrigerant mixture as Comparative Example 1. As shown in FIG. 2, when the refrigerant is a single refrigerant or an azeotropic refrigerant mixture, if the heat transfer performance of the condenser is improved, the refrigerant condensing temperature and the outside air temperature approach each other as much as possible, and the refrigerant accelerates condensation. As a result, the liquid is liquefied and the discharge pressure is also reduced. Therefore, the driving performance can be improved.

Next, as Comparative Example 2, the case where the refrigerant is a non-azeotropic refrigerant mixture will be described. FIG. 3 is a Ph diagram showing a condensation process for a non-azeotropic refrigerant mixture as Comparative Example 2. When the refrigerant is a non-azeotropic refrigerant mixture, the mixed refrigerants have different boiling points, so that the refrigerant with a low boiling point evaporates first, and the refrigerant with a high boiling point evaporates later. As a result, as shown in FIG. 3, the isotherm is a line having a temperature gradient that descends to the right. When the refrigerant is a non-azeotropic refrigerant, the refrigerant pressure decreases when the heat transfer performance of the condenser is improved and the refrigerant condensing temperature and the outside air temperature are brought as close as possible. May be lower. Since it is impossible to cool the refrigerant below the outside air temperature, the refrigerant cannot exchange heat with the outside air during the condensation process. As a result, the refrigerant cannot be sufficiently liquefied and the refrigeration effect is significantly impaired. Specifically, when the refrigerant is a conventional non-azeotropic refrigerant mixture such as R407C, when the saturated liquid temperature is lowered by 1 ° C. from the outside air temperature, the refrigeration effect is reduced by about 20%, whereas the compressor power is 3 % Does not decrease.

Therefore, when a refrigeration cycle is performed using a non-azeotropic refrigerant mixture, the saturated liquid temperature needs to be higher than the outside air temperature. In the first embodiment, when the refrigerant is R448A, the refrigeration cycle apparatus 1 is provided with temperature control means for making the saturated liquid temperature of the refrigerant equal to or higher than the temperature of the cooling fluid in the condenser. An example of the temperature control means will be described.

In the condenser 3, the temperature control means 20 of the first embodiment is such that the refrigerant flow and the cooling fluid flow are counterflows, and the saturated liquid temperature of the refrigerant flows into the condenser 3. It is the structure which makes it more than the temperature of the entrance side. FIG. 4 is a diagram for explaining the configuration of the temperature control means in Embodiment 1 of the present invention. FIG. 5 is a Ph diagram showing a condensation process in the condenser according to Embodiment 1 of the present invention.

As shown in FIG. 4, the flow of the refrigerant flowing through the refrigerant flow path 13 of the condenser 3 and the flow of outside air supplied to the condenser 3 by the fan 6 are counterflows. As shown in FIG. 5, in the condensation process in the condenser 3 of the first embodiment, the condensation temperature of the refrigerant and the outside air temperature approach each other as much as possible, the refrigerant is condensed and liquefied, and the discharge pressure drop is also promoted. . By the temperature control means 20, it is possible to avoid significantly impairing the refrigeration effect and improve driving performance.

Furthermore, in the example of FIG. 4, the flow of the refrigerant and the flow of the air are opposite flows, and the refrigerant flows toward the upstream of the air flow while changing its direction a plurality of times. For this reason, the condensation temperature of the refrigerant and the outside air temperature approach each other as much as possible, and the effect that the refrigerant is accelerated to condense and liquefy, and the discharge pressure drop is further promoted becomes more remarkable.

In the first embodiment, in the condenser 3, since the refrigerant flow and the external air flow are counterflows, the saturated liquid temperature only needs to be higher than the temperature on the suction side of the external air, and the temperature of the external air is high. There is no need to consider the higher outlet side. Therefore, it is not necessary to raise the saturated liquid temperature more than necessary, and it is not necessary to raise the discharge pressure of the compressor 2. Even if the refrigerant is a non-azeotropic refrigerant, condensation is promoted in the condenser 3 by maintaining the saturated liquid temperature above the temperature of the cooling fluid. As a result, it is possible to suppress a decrease in operation performance and improve the operation performance of the entire apparatus by improving the refrigeration effect.

The compressor 2 of the first embodiment is driven by an inverter circuit. Since the compressor 2 can control rotation speed by an inverter circuit, it can perform low load operation. During low-load operation, the temperature difference between the outside air temperature and the condensation temperature is the smallest. Further, since the condenser 3 is an air-cooled condenser, an air temperature difference D _air that is a difference between the temperature on the suction side and the temperature on the blow-out side of the outside _air becomes large. This is because the air-cooled condenser has a smaller heat capacity than, for example, a water-cooled condenser. This is because gaseous air has a small density and a small specific heat, so that it is difficult to obtain a mass flow rate unless means such as a large-diameter fan is used, and the heat capacity becomes small. Therefore, the air temperature difference D _air inevitably increases. The temperature difference between the outside air temperature and the condensation temperature cannot be condensed unless the air temperature difference D _{air is} not less than the air temperature difference.

The refrigeration cycle apparatus 1 according to the first embodiment includes the compressor 2 provided with the inverter circuit and the air-cooled condenser 3, so that a refrigerant having a large temperature difference between the saturated liquid temperature and the saturated gas temperature is used. And the effect which can suppress the fall of driving performance becomes remarkable.

For example, in a small refrigerator or the like, the temperature difference between the outside air temperature and the condensation temperature is about 2 ° C. during operation at a low load. If the temperature gradient of the non-azeotropic refrigerant mixture is 2 ° C. or higher, the saturated liquid temperature may be lower than the outside air temperature. Therefore, the temperature control means described above is highly effective for refrigerants having a temperature gradient of 2 ° C. or higher, and is an effective means for a refrigeration cycle apparatus for non-azeotropic refrigerant mixture.

Embodiment 2. FIG.
In the second embodiment, the pressure measuring means is provided in the refrigerant pipe on the downstream side of the condenser 3. In the second embodiment, as in the first embodiment, the case where the refrigerant is R448A will be described. In the second embodiment, detailed description of the same configuration as that of the first embodiment is omitted.

FIG. 6 is a diagram illustrating a configuration example of the refrigeration cycle apparatus according to Embodiment 2 of the present invention. As shown in FIG. 6, the refrigeration cycle apparatus 1a further includes a pressure measurement unit 11, and includes a control unit 7 as a temperature control unit 20a. The control unit 7 is connected to the pressure measuring means 11 and the compressor 2 via a signal line. The control unit 7 is, for example, a microcomputer. The compressor 2 may have an inverter circuit as described in the first embodiment.

The pressure measuring means 11 is provided in the refrigerant pipe between the condenser 3 and the decompression device 4. The pressure measuring unit 11 measures the pressure of the refrigerant between the condenser 3 and the decompression device 4 and outputs the measurement result to the control unit 7. The control unit 7 calculates the saturated liquid temperature of the refrigerant using the measurement result of the pressure measuring unit 11. Then, the control unit 7 adjusts the discharge pressure by controlling the capacity of the compressor 2 so that the calculated saturated liquid temperature becomes equal to or higher than the outside air temperature.

According to the second embodiment, since the control unit 7 can accurately and accurately measure the saturated liquid temperature from the measurement result of the pressure measuring means 11, it is not necessary to unnecessarily increase the discharge pressure of the compressor 2. Therefore, driving performance is improved.

Here, consider the case where pressure measuring means is installed upstream of the condenser 3. In this case, it is necessary to correct the numerical deviation caused by the pressure loss of the condenser 3. Consider a case where a temperature measuring means is installed instead of the pressure measuring means. In this case, the temperature measuring means does not know whether there is supercooling, and the control unit 7 may make an erroneous determination. Therefore, it is most desirable to install the pressure measuring means on the downstream side of the condenser 3 as in the second embodiment.

Embodiment 3 FIG.
In Embodiment 3, as the temperature control means, the opening degree of the decompression device 4 is controlled, and the saturation temperature of the refrigerant is maintained at a temperature equal to or higher than the outside air temperature. In the third embodiment, the decompression device 4 is the expansion valve 4a. In the third embodiment, as in the first embodiment, the refrigerant is R448A. In the third embodiment, detailed description of the same configuration as in the first and second embodiments is omitted.

FIG. 7 is a diagram illustrating a configuration example of a refrigeration cycle apparatus according to Embodiment 3 of the present invention. As shown in FIG. 7, the refrigeration cycle apparatus 1b has a temperature control means 20b. The temperature control means 20b has the control part 7 which controls the saturated liquid temperature of a refrigerant | coolant by adjusting the opening degree of the expansion valve 4a. The control part 7 adjusts the opening degree of the expansion valve 4a so that the saturated liquid temperature of the refrigerant | coolant in the condenser 3 may maintain the state more than external temperature.

It is desirable for the expansion valve 4a to be a variable throttle so that it can cope with various operating conditions. If the expansion valve 4a is throttled, the low-pressure side refrigerant moves to the high-pressure side, so that the pressure can be increased, but at the same time, the pressure of the low-pressure side refrigerant decreases, so the amount of refrigerant filled needs to be appropriate. If the control unit 7 controls the opening degree of the expansion valve 4a to make the saturated liquid temperature higher than the outside air temperature, the supercooling can be automatically taken. On the other hand, even if the supercooling is not achieved, the saturated liquid temperature may be higher than the outside air temperature. In this case, if the controller 7 sets the saturated liquid temperature as a control target, the supercooling can be controlled with higher accuracy.

In the third embodiment, by controlling the saturated liquid temperature in the high pressure control by the expansion valve 4a, the refrigeration cycle apparatus 1b using R448A which is a kind of non-azeotropic refrigerant mixture has a refrigerating effect remarkably. Improve driving performance by avoiding damage.

Embodiment 4 FIG.
In the fourth embodiment, a liquid receiver is provided between the condenser 3 and the decompression device 4 as temperature control means. In the fourth embodiment, similarly to the first embodiment, the refrigerant is R448A. In the fourth embodiment, detailed description of the same configuration as in the first to third embodiments is omitted.

FIG. 8 is a diagram illustrating a configuration example of a refrigeration cycle apparatus according to Embodiment 4 of the present invention. As shown in FIG. 8, the refrigeration cycle apparatus 1c has a temperature control means 20c. The temperature control means 20 c has a liquid receiver 12. As shown in FIG. 8, the liquid receiver 12 is provided between the condenser 3 and the decompression device 4. The liquid receiver 12 serves to temporarily store the liquid refrigerant condensed by the condenser 3.

Even if the refrigerant is R448A, the saturated liquid temperature naturally exceeds the outside air temperature and the outlet of the condenser 3 is maintained in the saturated liquid state as long as the refrigerant filling amount is such that the liquid refrigerant is stored in the liquid receiver 12. The At this time, the outlet of the condenser 3 is surely in a saturated liquid state. Therefore, temperature measuring means may be installed at the outlet of the condenser 3. FIG. 9 is a diagram illustrating another configuration example of the refrigeration cycle apparatus according to Embodiment 4 of the present invention.

The refrigeration cycle apparatus 1d shown in FIG. 9 further includes a temperature measurement unit 16, and includes a liquid receiver 12 and a control unit 7 as the temperature control unit 20d. As described above, in the fourth embodiment, since the outlet of the condenser 3 is surely in the saturated liquid state, the temperature measuring unit 16 can measure the saturated liquid temperature with sufficient accuracy. When the discharge pressure is changed as a result of the determination as to whether or not the discharge pressure of the compressor 2 should be changed based on the measurement result of the temperature measurement unit 16, the control unit 7 performs the same control as in the second embodiment. Compared with the refrigeration cycle apparatus 1a shown in FIG. 6, the refrigeration cycle apparatus 1d shown in FIG. 9 uses a temperature measurement means instead of the pressure measurement means, so that the cost can be reduced.

Embodiment 5 FIG.
In the fifth embodiment, as the temperature control means, the saturated liquid temperature is controlled by adjusting the heat transfer performance of the condenser 3, and the saturated liquid temperature is maintained at a temperature equal to or higher than the outside air temperature. In the fifth embodiment, similarly to the first embodiment, the refrigerant is R448A. In the fifth embodiment, detailed description of the same configuration as in the first to fourth embodiments is omitted.

FIG. 10 is a diagram illustrating a configuration example of a refrigeration cycle apparatus according to Embodiment 5 of the present invention. As shown in FIG. 10, the refrigeration cycle apparatus 1e further includes a fan 6, and includes a control unit 7 as temperature control means 20d. The fan 6 supplies outside air to the condenser 3. The control unit 7 is connected to the fan 6 via a signal line. The controller 7 adjusts the heat exchange performance of the condenser 3 by controlling the rotational speed of the fan 6.

The fan 6 is a fan whose rotation speed is variable. Therefore, the control part 7 can control the heat transfer performance of the condenser 3 by controlling the rotation speed of the fan 6 and adjusting the air volume. For example, if the target value of the saturated liquid temperature is “outside air temperature + predetermined temperature difference”, the control unit 7 can maintain an appropriate fan speed that is not excessive or insufficient. Further, even in the refrigeration cycle apparatus 1e using R448A, which is a kind of non-azeotropic refrigerant mixture, it is possible to avoid significantly impairing the refrigeration effect and improve operational performance.

Embodiment 6 FIG.
In the sixth embodiment, as the temperature control means, the saturated liquid temperature is controlled by adjusting the capacity of the compressor 2, and the saturated liquid temperature is maintained higher than the outside air temperature. In the sixth embodiment, as in the first embodiment, the refrigerant is R448A. In the sixth embodiment, detailed description of the same configuration as in the first to fifth embodiments is omitted.

FIG. 11 is a diagram illustrating a configuration example of a refrigeration cycle apparatus according to Embodiment 6 of the present invention. As shown in FIG. 11, the refrigeration cycle apparatus 1f has a temperature control means 20f. The temperature control unit 20 f includes a control unit 7 that controls the capacity of the compressor 2. The control unit 7 is connected to the compressor 2 via a signal line.

Compressor 2 is a compressor with a variable rotation speed. For example, as described in the first embodiment, the compressor 2 includes an inverter circuit. In this case, the control unit 7 may execute the low load operation of the compressor 2 by controlling the rotation speed of the inverter circuit. Therefore, the control unit 7 can freely control the capacity by controlling the rotational speed of the compressor 2. Increasing the number of revolutions of the compressor 2 can increase the discharge pressure, but the pressure loss of the condenser 3 also increases, so the increase in the saturated liquid temperature is reduced. It is necessary to make the pressure loss of the condenser 3 appropriate. In the sixth embodiment, even in the refrigeration cycle apparatus 1f using R448A which is a kind of non-azeotropic refrigerant mixture, it is possible to avoid significantly impairing the refrigeration effect and to improve the operation performance.

Although specific examples of the temperature control means have been described in the first to sixth embodiments, the temperature control means is not limited to the means described in each embodiment, and two or more embodiments may be combined.

When the pressure loss of the condenser 3 is large, the saturated liquid temperature is also lowered and may be lower than the outside air temperature. Therefore, the pressure loss of the condenser 3 needs to be appropriate. The narrower each refrigerant flow path of the heat exchanger, the larger the pressure loss. A diameter of 9.52 mm is used in a general small refrigerator used in a conventional non-azeotropic refrigerant R407C or the like. Hereinafter, the diameter of 9.52 mm is expressed as φ9.52. If it is less than φ9.52, the problem of refrigerant distribution also increases, and it is assumed that the pressure loss will increase more than before. Therefore, the probability that the saturated liquid temperature falls below the outside air temperature increases.

FIG. 12 is a cross-sectional view showing a case where the condenser shown in FIG. 1 is a flat tube heat exchanger. FIG. 12 shows one section when the condenser 3 is cut perpendicularly to the direction in which the refrigerant flows. As shown in FIG. 12, the condenser 3 has a flat tube 14 that is a flat heat transfer tube. The flat tube 14 may employ a flat multi-hole tube formed so that a plurality of refrigerant flow paths are partitioned, but may be one in which one refrigerant flow path is formed. The flat tube 14 of the flat multi-hole tube of FIG. 12 has ten flow paths. The flat tube 14 passes through the plurality of fins 15. Although the flat tubes 14 are provided in multiple stages with a certain interval, FIG. 12 shows the flat tubes 14 at the uppermost stage and the lowermost stage. In FIG. 12, the direction in which the outside air supplied from the fan 6 flows is indicated by broken-line arrows.

The effectiveness of the temperature control means described above is very high for the condenser 3 which has a flow path with a pipe having a flat tube shape and an equivalent diameter of less than φ9.52, and is a non-azeotropic refrigerant refrigerating cycle apparatus. Required. When the refrigerant is a non-azeotropic refrigerant mixture, the heat exchange efficiency can be improved by using a flat tube heat exchanger for the condenser 3.

As described above, when the pressure loss of the condenser 3 is large, the saturated liquid temperature is also lowered and may be lower than the outside air temperature. The pressure loss increases or decreases depending on the refrigerant flow rate, and the refrigerant flow rate is required more for the refrigerant having a smaller latent heat when the same capacity is exhibited. If the latent heat is smaller than R407C of the conventional non-azeotropic refrigerant mixture, it is assumed that the refrigerant flow rate is larger than that of the conventional refrigerant and the pressure loss is increased, and the probability that the saturated liquid temperature is also lower than the outside air temperature is increased. Therefore, the effectiveness of the temperature control means described above is very high for a refrigerant having a smaller latent heat than R407C, and is essential in a refrigeration cycle apparatus for a non-azeotropic refrigerant mixture. The latent heat of R407C with respect to the saturation temperature of 40 ° C. is 164 kJ / kg, and the temperature control means described above is particularly effective for R448A, which has a lower latent heat than R407C.

As described in Embodiment 1, in a general small refrigerator, the temperature difference between the outside air temperature and the condensing temperature is about 2 ° C. in the operation at a low load. At this time, if the temperature gradient of the non-azeotropic refrigerant mixture is 2 ° C. or higher, the saturated liquid temperature may be lower than the outside air temperature. Therefore, the temperature control means described in Embodiments 1 to 6 described above is highly effective for refrigerants having a temperature gradient of 2 ° C. or higher, and is effective means for the refrigeration cycle apparatus for non-azeotropic refrigerant mixture.

The refrigeration cycle apparatuses described in Embodiments 1 to 6 are particularly effective in refrigeration cycle apparatuses such as refrigeration apparatuses, air conditioning apparatuses, and hot water supply apparatuses using non-azeotropic refrigerant mixtures. Furthermore, since the problems of the present invention with respect to low outside air operation in a cold region with a low temperature apparatus with a wide evaporation temperature range and low load operation performance become significant, the refrigeration cycle apparatus described in Embodiments 1 to 6 is , Especially effective.

1, 1a to 1f refrigeration cycle device, 2 compressor, 3 condenser, 4 decompressor, 4a expansion valve, 5 evaporator, 6 fan, 7 control unit, 10 refrigerant circuit, 11 pressure measuring means, 12 receiver, 13 refrigerant flow path, 14 flat tube, 15 fin, 16 temperature measuring means, 20, 20a-20f temperature control means.

Claims

A refrigerant circuit in which a compressor, a condenser, a pressure reducing device, and an evaporator are connected, and a non-azeotropic refrigerant is circulated;
Temperature control means for setting the saturated liquid temperature of the refrigerant to be equal to or higher than the temperature of the cooling fluid in the condenser;
A refrigeration cycle apparatus having
In the condenser, the flow of the refrigerant and the flow of the cooling fluid are counterflows, and the saturated liquid temperature is equal to or higher than the temperature of the cooling fluid flowing into the condenser on the inlet side of the condenser. Item 2. The refrigeration cycle apparatus according to Item 1.
A pressure measuring means provided between the condenser and the pressure reducing device for measuring the pressure of the refrigerant;
The temperature control means includes
The refrigeration cycle apparatus according to claim 1 or 2, further comprising a control unit that controls a discharge pressure of the compressor using a measurement result of the pressure measuring unit.
The temperature control means includes
The refrigeration cycle apparatus according to any one of claims 1 to 3, further comprising a control unit that controls the temperature of the saturated liquid by adjusting an opening degree of the decompression device.
The temperature control means includes
The refrigeration cycle apparatus according to any one of claims 1 to 4, further comprising a liquid receiver provided between the condenser and the decompression device.
A temperature measuring means provided between the condenser and the liquid receiver for measuring the temperature of the refrigerant;
The temperature control means includes
The refrigeration cycle apparatus according to claim 5, further comprising a control unit that controls a discharge pressure of the refrigerant of the compressor using a measurement result of the temperature measuring unit.
A fan for supplying outside air to the condenser;
The temperature control means includes
The refrigeration cycle apparatus according to any one of claims 1 to 6, further comprising a control unit that adjusts a heat exchange performance of the condenser by controlling a rotation speed of the fan.
The temperature control means includes
The refrigeration cycle apparatus according to any one of claims 1 to 7, further comprising a control unit that controls a capacity of the compressor.
The refrigeration cycle apparatus according to any one of claims 1 to 8, wherein an equivalent diameter of a refrigerant flow path in the condenser is less than 9.52 mm.
The refrigeration cycle apparatus according to claim 9, wherein the condenser has a flat tube in which the refrigerant flow path is formed and is formed in a flat shape.
The refrigeration cycle apparatus according to any one of claims 1 to 10, wherein the refrigerant is a refrigerant in which the latent heat of the refrigerant is smaller than the latent heat of the refrigerant R407C.
The compressor is a variable capacity compressor capable of adjusting the discharge amount of the refrigerant,
The condenser is an air heat exchanger that exchanges heat between the refrigerant and air.
The refrigeration cycle apparatus according to any one of claims 1 to 11.
The refrigeration cycle apparatus according to any one of claims 1 to 12, wherein a temperature difference between the saturated liquid temperature and the saturated gas temperature under a constant pressure is 2 ° C or more.