JP6381827B2 - Refrigeration cycle apparatus and showcase having the same - Google Patents

Description

  The present invention relates to a refrigeration cycle apparatus and a showcase including the refrigeration cycle apparatus, and more particularly to a refrigeration cycle apparatus in which the temperature of refrigerant on the discharge side of a compressor is controlled and a showcase including the refrigeration cycle apparatus.

  In the refrigeration cycle apparatus, superheat control is known in which the opening degree of the expansion valve is controlled so that the superheat degree of the evaporator becomes a target value. For example, a superheat degree control device for a refrigeration apparatus disclosed in Japanese Patent Application Laid-Open No. 2-17358 (Patent Document 1) includes superheat degree detection means (specifically, an inlet side and an outlet side of an evaporator) for detecting the superheat degree of a refrigerant. The temperature sensor which detects the temperature of the refrigerant | coolant in this is provided. The superheat degree control means controls the opening degree of the expansion valve so that the change amount of the opening degree of the expansion valve becomes larger as the difference between the superheat degree of the refrigerant detected by the superheat degree detection means and the target value is larger.

  In the superheat degree control, by appropriately setting the target value of the superheat degree (for example, by setting it to about several K), the refrigerant on the outlet side of the evaporator becomes close to the saturated gas. That is, since the refrigerant in the evaporator is in a gas-liquid two-phase state, highly efficient heat exchange is realized between the air and the refrigerant.

Japanese Patent Laid-Open No. 2-17358

  Conventionally, refrigerants such as R404C have been generally used in low-temperature refrigeration cycle apparatuses. However, the use of these refrigerants is currently prohibited for the purpose of reducing the environmental burden. And technical development for using a refrigerant | coolant (for example, R410A or R32 etc.) with a lower environmental load is advanced.

  In the refrigeration cycle apparatus, it is required to operate the compressor so that the gas refrigerant is compressed at a high compression ratio. With the adiabatic compression at a high compression ratio in the compressor, the temperature of the gas refrigerant increases. Many refrigerants with a low environmental load have a characteristic that the temperature of the refrigerant on the discharge side of the compressor (hereinafter also referred to as “discharge temperature”) is likely to rise as compared with conventional refrigerants. Therefore, in the refrigeration cycle apparatus in which the superheat degree control is executed, for example, when a refrigerant with a low environmental load is used, the discharge temperature of the refrigerant can rise excessively. As a result, if the temperature of the compressor exceeds the upper limit of the allowable range, an abnormality may occur in the compressor.

  As one of measures for solving this problem, it is conceivable to execute discharge temperature control for controlling the refrigerant discharge temperature to a target value by adjusting the opening of the expansion valve. In the discharge temperature control, an excessive increase in the discharge temperature can be suppressed by appropriately setting the target value of the discharge temperature. Therefore, it is possible to prevent an abnormality of the compressor.

  On the other hand, the discharge temperature control uses the refrigerant discharge temperature as a control target, and does not use the refrigerant superheat degree as a control target parameter unlike the superheat degree control. Therefore, the degree of superheat may not be maintained at an appropriate value, and the degree of superheat may increase excessively.

  The present invention has been made to solve the above-described problems, and an object thereof is to provide a technique capable of suppressing an excessive increase in the degree of superheat in a refrigeration cycle apparatus in which discharge temperature control is executed.

  Another object of the present invention is to provide a technique capable of suppressing an excessive increase in the degree of superheat in a showcase including a refrigeration cycle apparatus in which discharge temperature control is performed.

  A refrigeration cycle apparatus according to an aspect of the present invention includes a refrigerant circuit and a control device. The refrigerant circuit includes a compressor, a condenser, an expansion valve, and an evaporator, and is configured to be able to circulate the refrigerant. The control device controls the temperature of the refrigerant discharged from the compressor by adjusting the opening of the expansion valve. The evaporator includes a refrigerant flow path. At least a part of the flow paths is configured to form a parallel flow in which the refrigerant flows from the upstream side to the downstream side in the air flow direction in the evaporator.

  According to the present invention, the flow paths included in the evaporator are configured to form a parallel flow. Therefore, on the outlet side of the flow path, heat exchange is performed between the refrigerant after circulating through the evaporator and being cooled, and the refrigerant. Furthermore, the refrigerant temperature does not exceed the air temperature during heat exchange. Therefore, the refrigerant temperature at the outlet side of the flow path compared with the case where the flow path is configured to form a counterflow (when the refrigerant flows from the downstream side in the air flow direction toward the upstream side in the evaporator). Is hard to rise. Therefore, an excessive increase in the degree of superheat of the refrigerant can be suppressed.

1 is a block diagram schematically showing a configuration of a showcase including a refrigeration cycle apparatus according to Embodiment 1. FIG. It is a block diagram which shows the structure of the refrigeration cycle apparatus shown in FIG. 1 in detail. It is a functional block diagram of a control device for explaining discharge temperature control. It is a figure which shows the structure of the evaporator contained in the refrigerating-cycle apparatus which concerns on a comparative example. It is a figure for demonstrating the evaporator contained in the refrigerating-cycle apparatus which concerns on Embodiment 1. FIG. 6 is a diagram for explaining an evaporator included in a refrigeration cycle apparatus according to Embodiment 2. FIG. 6 is a diagram for explaining an evaporator included in a refrigeration cycle apparatus according to Embodiment 3. FIG. FIG. 6 is a diagram showing a configuration of an evaporator included in a refrigeration cycle apparatus according to Modification 1 of Embodiment 3. It is a figure which shows the structure of the evaporator contained in the refrigerating-cycle apparatus which concerns on the modification 2 of Embodiment 3. FIG.

  Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings. In the drawings, the same or corresponding parts are denoted by the same reference numerals and description thereof will not be repeated.

  In the following, as an embodiment of the refrigeration cycle apparatus according to the present invention, a configuration in which the refrigeration cycle apparatus is applied to a showcase will be described as an example. However, the use of the refrigeration cycle apparatus is not particularly limited, and the refrigeration cycle apparatus may be applied to, for example, an air conditioner.

[Embodiment 1]
1 is a block diagram schematically showing a configuration of a showcase provided with a refrigeration cycle apparatus according to Embodiment 1. FIG. Referring to FIG. 1, a showcase 1 is a container for displaying commodities such as food. The showcase 1 includes a refrigeration cycle apparatus 100, a case main body 200, and a cooling fan 300. The configuration of the refrigeration cycle apparatus 100 will be described in detail with reference to FIGS.

  The case main body 200 has an opening 210 for taking in and out a product (article) and accommodates the product. The cooling blower 300 sends cool air so as to cover the opening 210 provided in the case main body 200. Thereby, the air curtain CUR is formed. The air curtain CUR prevents the air cooled by the refrigeration cycle apparatus 100 from leaking out of the case body 200 from the opening 210.

  FIG. 2 is a block diagram showing the configuration of the refrigeration cycle apparatus 100 shown in FIG. 1 in more detail. Referring to FIG. 2, the refrigeration cycle apparatus 100 includes a refrigerant circuit 110, blowers 52 and 54, and a control device 70. The refrigerant circuit 110 includes a compressor 10, a condenser 20, an expansion valve 30, an evaporator 40, and a pipe 60.

  A refrigerant is sealed in the refrigerant circuit 110. In the present embodiment, a low environmental load HFC (Hydrofluoro Carbon) refrigerant is employed. Examples of HFC refrigerants include, for example, R32, R134a, R407C, and R410A.

  The compressor 10 is a variable capacity compressor driven by, for example, an inverter (not shown). The gas refrigerant compressed to high temperature and high pressure flows into the condenser 20.

  The condenser 20 is a heat exchanger configured to include, for example, heat transfer tubes and heat radiating fins (not shown). The condenser 20 is provided with a blower 52. In the condenser 20, the gas refrigerant is condensed by dissipating heat to the surrounding outside air to become a liquid refrigerant.

  The expansion valve 30 is, for example, a throttle valve (electronic expansion valve) whose opening degree can be controlled by a stepping motor (not shown). The expansion valve 30 is also used for adjusting the flow rate of the refrigerant. The expansion valve 30 decompresses the liquid refrigerant by expanding the high-pressure liquid refrigerant condensed by the condenser 20. This liquid refrigerant enters a gas-liquid two-phase state and flows into the evaporator 40.

  The evaporator 40 is a heat exchanger configured to include the heat transfer tubes 401 and the heat radiating fins 402 (see FIG. 5) similarly to the condenser 20. The evaporator 40 is provided with a blower 54. In the evaporator 40, the air is cooled by the refrigerant. On the other hand, the refrigerant changes from a gas-liquid two-phase state to a low-pressure gas state. Thereafter, the gas refrigerant returns to the compressor 10 and is compressed again by the compressor 10 and discharged. In this way, the refrigerant circulates in the refrigerant circuit to form a refrigeration cycle.

  The refrigeration cycle apparatus 100 further includes a discharge temperature sensor 81, a condensation temperature sensor 82, and an evaporation temperature sensor 84. The discharge temperature sensor 81 is provided on the discharge side of the compressor 10 and detects the temperature (discharge temperature) Td of the refrigerant discharged from the compressor 10. The condensation temperature sensor 82 is provided in the condenser 20 and detects the temperature of the refrigerant (condensation temperature) in the condenser 20. The evaporation temperature sensor 84 is provided in the evaporator 40 and detects the temperature of the refrigerant (evaporation temperature) in the evaporator 40. Each sensor outputs a signal indicating the detection result to the control device 70.

  Although not shown, the control device 70 includes a CPU (Central Processing Unit), a memory such as a RAM (Random Access Memory) and a ROM (Read Only Memory), and an input / output interface. Based on the detection signals from the above-described sensors, the control device 70 controls each device by causing the CPU to read and execute a program stored in advance in a ROM or the like into the RAM. More specifically, the control device 70 performs discharge temperature control for controlling the refrigerant discharge temperature to a target value by controlling the opening degree of the expansion valve 30 based on detection signals from the sensors. Hereinafter, the discharge temperature control in the present embodiment will be described in more detail.

  FIG. 3 is a functional block diagram of the control device 70 for explaining the discharge temperature control. Each functional block shown in FIG. 3 may be realized by hardware processing such as an electric circuit having a function corresponding to the block, or may be realized by software processing according to a preset program.

  Referring to FIGS. 2 and 3, control device 70 includes a high pressure calculation unit 71, a low pressure calculation unit 72, an intake temperature calculation unit 73, a storage unit 74, a target value calculation unit 75, and an adjustment unit 76. including.

  The high pressure calculator 71 calculates the pressure (high pressure) Pd of the refrigerant discharged from the compressor 10 based on the refrigerant condensation temperature detected by the condensation temperature sensor 82. The high pressure calculation unit 71 outputs the calculated high pressure Pd to the target value calculation unit 75.

  The low pressure calculation unit 72 calculates the pressure (low pressure) Ps of the refrigerant sucked into the compressor 10 based on the refrigerant evaporation temperature detected by the evaporation temperature sensor 84. The low pressure calculation unit 72 outputs the calculated low pressure Ps to the target value calculation unit 75.

  The suction temperature calculation unit 73 calculates the temperature (suction temperature) Ts of the refrigerant sucked into the compressor 10 based on the refrigerant evaporation temperature detected by the evaporation temperature sensor 84. The suction temperature calculation unit 73 outputs the calculated suction temperature Ts to the target value calculation unit 75.

  The storage unit 74 stores n, which is a constant corresponding to the performance of the compressor 10 and the type of refrigerant. The storage unit 74 outputs the constant n to the target value calculation unit 75.

  The target value calculation unit 75 uses the high pressure Pd, the low pressure Ps, the suction temperature Ts, and the constant n to calculate the target value Td * of the refrigerant discharge temperature Td according to the following equation (1). The calculated target value Td * is output to the adjustment unit 76. Equation (1) is derived by approximating the compression process of the compressor 10 with a quasi-static process (polytropic process).

  The adjustment unit 76 adjusts the opening degree of the expansion valve 30 so that the discharge temperature Td detected by the discharge temperature sensor 81 is controlled to the target value Td * of the discharge temperature Td calculated by the target value calculation unit 75. . Thereby, since the excessive raise of discharge temperature Td can be suppressed, it becomes possible to prevent abnormality of the compressor 10. FIG.

  Furthermore, it is desirable that the suction temperature calculation unit 73 use the saturation temperature of the refrigerant in the evaporator 40 as the refrigerant suction temperature Ts. In other words, the target value Td * of the discharge temperature Td is determined so that the refrigerant on the suction side of the compressor 10 is in the state of wet steam where the saturated liquid and saturated gas coexist (the dryness x is 1). Is desirable. Thereby, since the refrigerant | coolant inside the evaporator 40 will be in a gas-liquid two-phase state, compared with the case where the dryness x is set to less than 1, the heat transfer rate from air to a refrigerant | coolant becomes high. As a result, highly efficient heat exchange can be realized between the air and the refrigerant in the evaporator 40.

  Next, the configuration of the evaporator 40 will be described. Here, in order to facilitate understanding of the characteristics of the evaporator 40 in the present embodiment, the configuration of the evaporator included in the refrigeration cycle apparatus according to the comparative example will be described first. The configuration other than the evaporator in the refrigeration cycle apparatus according to the comparative example is the same as the corresponding configuration of the refrigeration cycle apparatus 100 (see FIG. 2) according to the first embodiment.

  FIG. 4 is a diagram for explaining an evaporator included in a refrigeration cycle apparatus according to a comparative example. FIG. 4A shows the configuration of the evaporator 90 included in the refrigeration cycle apparatus according to the comparative example. With reference to FIG. 4 (A), the evaporator 90 is comprised including the heat exchanger tube 901 which is a flow path through which a refrigerant | coolant flows, and the several radiation fin 902 for radiating the heat | fever of a refrigerant | coolant.

  Each of the plurality of heat radiation fins 902 has a flat plate shape extending in the XZ plane in the drawing. Each of the plurality of radiating fins 902 is arranged at a predetermined interval in the Y direction.

  The heat transfer tube 901 extends in a direction (Y direction) penetrating the plurality of heat radiation fins 902. That is, the refrigerant flows in the positive Y direction or the negative Y direction inside the heat transfer tube 901. However, in FIG. 4, only the cross section of the heat transfer tube 901 is shown. The flow of the refrigerant when viewed as the entire evaporator 90 is schematically indicated by an arrow REF. The air sent to the evaporator 90 by the blower 54 (see FIG. 2) is indicated by an arrow AIR.

  The evaporator 90 is a counterflow heat exchanger in which the refrigerant flow direction (negative X direction) is opposite to the air flow direction (positive X direction). That is, the heat transfer tube 901 is configured such that the refrigerant flows from the downstream side in the air flow direction toward the upstream side.

  FIG. 4B is a diagram for explaining the relationship between the air temperature and the refrigerant temperature in the evaporator 90. 4B and FIGS. 5B and 6B described later, the horizontal axis represents a position along the path of the heat transfer tube 901 (401, 411). Regarding the heat transfer tube 901 provided in the evaporator 90, the most upstream position is described as UP and the most downstream position is described as DW. The vertical axis represents temperature. Curve A91 shows the air temperature under certain conditions. Curve R91 shows the refrigerant temperature under the same conditions.

  Referring to FIGS. 4A and 4B, the refrigerant flows into evaporator 90 in a gas-liquid two-phase state. Therefore, even if some refrigerant is gasified by heat exchange between air and the refrigerant, the refrigerant temperature hardly changes. Therefore, the refrigerant temperature from the most upstream UP to the MD in the middle of the path is substantially constant. In the midway MD, the refrigerant is in a gas state (a dryness x is 1). On the downstream side of the refrigerant from the midway MD, all the refrigerant is in a superheated gas state. Therefore, the refrigerant temperature from the midway MD to the most downstream DW increases as going downstream.

  As shown in FIG. 4B, in the evaporator 90 which is a counterflow heat exchanger, a difference between the air temperature and the refrigerant temperature is secured to some extent at all positions on the path of the heat transfer tube 901. Heat exchange takes place. Therefore, high heat exchange efficiency can be realized.

  Here, changes in the air temperature and the refrigerant temperature when the temperature of the air sucked into the evaporator 90 rises will be described. The air temperature in this case is indicated by a curve A92, and the refrigerant temperature is indicated by a curve R92.

  When the suction temperature to the evaporator 90 rises, the air temperature in the most downstream DW of the refrigerant rises significantly compared to before the rise of the suction temperature. Then, since the temperature difference between the air temperature and the refrigerant temperature in the most downstream DW of the refrigerant becomes large, the amount of increase in the refrigerant temperature can also become large. As the refrigerant temperature increases, the degree of superheat can also increase. When the discharge temperature control is executed, the refrigerant discharge temperature Td is controlled to the target value as described in detail with reference to FIG. However, the discharge temperature control does not use the superheat degree of the refrigerant as a direct control target parameter unlike the superheat degree control. Therefore, the degree of superheat may not be maintained at an appropriate value, and the degree of superheat may increase excessively. As a result, the ratio of the gas-liquid two-phase state of the refrigerant flowing through the heat transfer tube 901 may decrease and heat exchange efficiency may decrease as compared to before the suction temperature rises.

  Therefore, according to the present embodiment, in the refrigeration cycle apparatus in which the discharge temperature control is executed, a parallel flow heat exchanger is employed as at least a part of the evaporator 40 as described in detail below.

  FIG. 5 is a diagram for explaining the evaporator 40 included in the refrigeration cycle apparatus 100 according to the first embodiment. FIG. 5A shows the configuration of the evaporator 40. With reference to FIG. 5 (A), the evaporator 40 is comprised including the heat exchanger tube 401 and the radiation fin 402. As shown in FIG.

  The evaporator 40 is a parallel flow heat exchanger in which a refrigerant flow direction (positive X direction) is a forward direction with respect to an air flow direction (positive X direction). That is, the heat transfer tube 401 is configured such that the refrigerant flows from the upstream side to the downstream side in the air flow direction.

  FIG. 5B is a diagram for explaining the relationship between the air temperature and the refrigerant temperature in the evaporator 40. The air temperature under a certain condition is indicated by a curve A1, and the refrigerant temperature is indicated by a curve R1. Further, when the intake temperature of the air into the evaporator 40 is increased, the air temperature after the temperature increase is indicated by a curve A2, and the refrigerant temperature after the temperature increase is indicated by a curve R2.

  With reference to FIGS. 5A and 5B, in the first embodiment, the intake temperature of air into evaporator 40 is the air temperature in the most upstream UP of the refrigerant. That is, in the most downstream DW of the refrigerant, heat exchange is performed between the refrigerant after flowing through the evaporator 40 and being cooled, and the refrigerant. Compared with the comparative example, in the most downstream DW of the refrigerant, the temperature difference between the air temperature and the refrigerant temperature is small, so the amount of increase in the refrigerant temperature is also small. Furthermore, the refrigerant temperature does not exceed the air temperature during heat exchange. Therefore, an excessive increase in the degree of superheat of the refrigerant can be suppressed.

  In the discharge temperature control, feedback control of the opening degree of the expansion valve 30 is executed according to the detection result of the refrigerant discharge temperature Td by the discharge temperature sensor 81. In general, the discharge temperature sensor 81 is provided outside the pipe 60 (see FIG. 2). Therefore, when the discharge temperature Td starts to rise, a delay time due to the heat capacity of the pipe 60 may occur before the temperature rise is detected by the discharge temperature sensor 81. Alternatively, when the discharge temperature Td rises stepwise, such as when the compressor 10 is started, there is a possibility that the discharge temperature cannot be controlled to follow the target value due to a delay in the feedback loop. In such a case, although the discharge temperature control is executed, the discharge temperature Td can rise excessively.

  On the other hand, in Embodiment 1, by adopting a parallel flow heat exchanger as the evaporator 40, the refrigerant temperature in the most downstream DW of the refrigerant is less likely to rise compared to the comparative example. That is, since the temperature of the refrigerant flowing into the compressor 10 is difficult to rise, an excessive increase in the discharge temperature Td is suppressed. Therefore, it is possible to prevent the abnormality of the compressor 10 more reliably.

  Many of the HFC refrigerants have a characteristic that the discharge temperature Td is likely to rise as compared with conventional refrigerants (such as R404C). However, according to the first embodiment, since an excessive increase in the discharge temperature Td can be suppressed, it is possible to more suitably employ the HFC refrigerant. Thus, according to the first embodiment, it is possible to realize a refrigeration cycle apparatus having a low environmental load while preventing an abnormality of the compressor 10.

  Further, as described with reference to FIG. 1, in the showcase 1, the air cooled by the refrigeration cycle apparatus 100 is trapped inside the case body 200 by the air curtain CUR. In general, in an apparatus having the same cooling capacity, the amount of air flowing through the evaporator (air volume) when the air curtain is formed may be smaller than the air volume when the air curtain is not formed. On the other hand, if the air volume is small, the temperature of air sucked into the evaporator tends to rise. Therefore, when a counterflow heat exchanger is employed as an evaporator in the showcase, an excessive increase in the degree of superheat tends to occur as described with reference to FIG. Therefore, it is particularly effective to apply the refrigeration cycle apparatus 100 according to Embodiment 1 to a showcase.

[Embodiment 2]
Embodiment 2 demonstrates the structure which can suppress the excessive raise of a superheat degree more effectively by changing arrangement | positioning of the heat exchanger tube contained in a parallel flow heat exchanger. In addition, since structures other than the evaporator in the refrigeration cycle apparatus according to Embodiment 2 and Embodiment 3 described later are the same as the corresponding structures of the refrigeration cycle apparatus 100 according to Embodiment 1 (see FIG. 2). Detailed description will not be repeated.

  FIG. 6 is a diagram for explaining a configuration of an evaporator included in the refrigeration cycle apparatus according to Embodiment 2. FIG. 6A shows a configuration of an evaporator included in the refrigeration cycle apparatus according to Embodiment 2. With reference to FIG. 6 (A), the evaporator 41 is a parallel flow heat exchanger similarly to the evaporator 40 (refer FIG. 5 (A)) which concerns on Embodiment 1, and the heat exchanger tube 411, a radiation fin 412.

  In the evaporator 41, the heat transfer tubes 411 form a plurality of flow path rows ROW arranged in a direction (X direction) along the air flow direction. Furthermore, the heat transfer tube 411 forms a plurality of flow path stages CLM arranged in a direction (Z direction) intersecting with the air flow direction. Hereinafter, the number of the plurality of flow path rows ROW is also referred to as “row number” of the heat transfer tubes 411, and the number of the plurality of flow path stages CLM is also referred to as “stage number” of the heat transfer tubes 411. In the example shown in FIG. 6A, the number of columns is 7, and the number of stages is 3. That is, the number of columns is larger than the number of stages.

  In the example shown in FIG. 5A as Embodiment 1, the number of columns is 3, and the number of stages is 7. That is, in the first embodiment, the number of columns is smaller than the number of stages.

  FIG. 6B is a diagram for explaining the relationship between the air temperature and the refrigerant temperature in the evaporator 41. 6 (A) and 6 (B), since the number of columns is larger than the number of stages in the second embodiment, the flow path length of air in the evaporator 41 is larger than that in the first embodiment. long. As a result, the air sucked into the evaporator 41 is sufficiently cooled, and the amount of decrease in the air temperature is larger than that in the first embodiment. Therefore, in the most downstream DW of the refrigerant, the temperature difference between the air temperature and the refrigerant temperature becomes small, so the refrigerant temperature becomes low. Therefore, according to Embodiment 2, an excessive increase in the degree of superheat can be more effectively suppressed.

[Embodiment 3]
In the first and second embodiments, the configuration in which the entire heat transfer tube included in the evaporator forms a parallel flow has been described as an example. In Embodiment 3, a configuration in which only a part of the heat transfer tubes forms a parallel flow will be described.

  FIG. 7 is a diagram showing a configuration of an evaporator included in the refrigeration cycle apparatus according to Embodiment 3. Referring to FIG. 7, evaporator 42 includes heat exchange units 42A and 42B. The heat exchange part 42A is a counterflow heat exchanger, and the heat exchange part 42B is a parallel flow heat exchanger.

  The heat exchange part 42B is provided on the downstream side of the air flow and the downstream side of the refrigerant flow with respect to the heat exchange part 42A. As a result, as in the first and second embodiments, the most downstream position of the air flow coincides with the most downstream position of the refrigerant flow. Therefore, an increase in the refrigerant temperature at the most downstream side of the refrigerant is suppressed. Therefore, an excessive increase in the degree of superheat can be suppressed.

  Further, in general, the heat exchange efficiency of the counter flow heat exchanger is higher than the heat exchange efficiency of the parallel flow heat exchanger. Therefore, the heat exchange efficiency of the evaporator 42 is higher than the heat exchange efficiency of the evaporator 41 (see FIG. 6A), which is a parallel flow heat exchanger as a whole. As described above, according to the third embodiment, it is possible to achieve a relatively high heat exchange efficiency while suppressing an excessive increase in the degree of superheat.

[Modification 1 of Embodiment 3]
In the first and second modifications of the third embodiment, a configuration for reducing the power consumption of the refrigeration cycle apparatus by changing the shape of the heat transfer tube of the evaporator will be described.

  FIG. 8 is a diagram illustrating a configuration of an evaporator included in the refrigeration cycle apparatus according to the first modification of the third embodiment. Referring to FIG. 8, evaporator 43 includes heat exchange units 43A and 43B. As with the evaporator 42 (see FIG. 7) in the third embodiment, the heat exchange unit 43A is a counterflow heat exchanger, and the heat exchange unit 43B is a parallel flow heat exchanger. The heat exchange unit 43A includes a heat transfer tube 431A. The heat exchange unit 43B includes a heat transfer tube 431B.

  Inside the evaporator 43, the refrigerant is mainly in a gas-liquid two-phase state. On the upstream side of the refrigerant flow, the liquid phase ratio is higher than the gas phase ratio. The proportion of the gas phase gradually increases as it goes downstream of the refrigerant flow. As the gas phase ratio increases, the density of the refrigerant decreases, so the flow rate of the refrigerant increases. As the flow rate of the refrigerant increases, the pressure loss (mainly loss due to frictional resistance) in the heat transfer tubes 431A and 431B increases. Thus, the pressure loss is greater on the downstream side than on the upstream side of the refrigerant flow. The heat exchange unit 43B is provided on the downstream side of the refrigerant flow with respect to the heat exchange unit 43A. For this reason, the pressure loss in the heat exchange part 43B tends to become relatively large.

  In the first modification, the inner diameter of the heat transfer tube 431B included in the heat exchange unit 43B is larger than the inner diameter of the heat transfer tube 431A included in the heat exchange unit 43A. That is, the cross-sectional area of the heat transfer tube 431B is larger than the cross-sectional area of the heat transfer tube 431B. Therefore, when the flow rate of the refrigerant is equal, when comparing the evaporator 42 in the third embodiment and the evaporator 43 in the first modification, the flow rate of the refrigerant flowing through the heat exchange unit 43B is the same as that of the refrigerant flowing through the heat exchange unit 42B. Slower than the flow rate. Thereby, the pressure loss in a parallel flow heat exchanger is reduced. Therefore, the power consumption of the refrigeration cycle apparatus can be reduced.

[Modification 2 of Embodiment]
FIG. 9 is a diagram illustrating a configuration of an evaporator included in the refrigeration cycle apparatus according to the second modification of the third embodiment. Referring to FIG. 9, evaporator 44 includes heat exchange units 44A and 44B. The heat exchange unit 44A is a counterflow heat exchanger, and the heat exchange unit 44B is a parallel flow heat exchanger. The heat exchange unit 44A includes a heat transfer tube 441A. The heat exchange unit 44B includes a heat transfer tube 441B.

  In the heat exchange part 44A, the heat transfer tube 441B is not branched and is formed as a single flow path. On the other hand, in the heat exchanging part 44B, the heat transfer tube 441B is branched, and a plurality of (two in FIG. 9) flow paths are formed. The two flow paths merge again on the downstream side of the heat exchange unit 44B.

  Thus, in the modification 2, the number of branches of the heat transfer tube 441B is larger than the number of branches of the heat transfer tube 441A. Therefore, as in Modification 1, when the flow rate of the refrigerant is the same, the flow rate of the refrigerant flowing through the heat exchange unit 44B is slower than the flow rate of the refrigerant flowing through the heat exchange unit 42B. Thereby, the pressure loss in a parallel flow heat exchanger is reduced. Therefore, the power consumption of the refrigeration cycle apparatus can be reduced.

  It is also possible to combine Modification 1 and Modification 2 of Embodiment 3. That is, a configuration in which the cross-sectional area of the flow path of the heat transfer tube in the parallel flow heat exchanger is large and the number of branches may be adopted as compared with the heat transfer tube in the counter flow heat exchanger. In this case, since the pressure loss is further reduced as compared with the configuration in which only one of them is adopted, the power consumption can be further reduced.

  In the first to third embodiments, the configuration in which the heat transfer tube is employed as the refrigerant flow path has been described as an example. However, the present invention can also be applied to a configuration including a plate heat exchanger as an evaporator. . In a plate heat exchanger, a plurality of plates are arranged at predetermined intervals in a direction (generally orthogonal direction) that intersects the air flow direction. In this case, the number of rows corresponds to the length of each of the plurality of plates (the length in the direction along the air flow direction). On the other hand, the number of stages corresponds to the width of the plurality of plates in the arrangement direction (the sum of the intervals between the plurality of plates).

  The embodiment disclosed this time should be considered as illustrative in all points and not restrictive. The scope of the present invention is defined by the terms of the claims, rather than the description above, and is intended to include any modifications within the scope and meaning equivalent to the terms of the claims.

  DESCRIPTION OF SYMBOLS 1 Showcase, 10 Compressor, 20 Condenser, 30 Expansion valve, 40-44,90 Evaporator, 42A, 42B, 43A, 43B, 44A, 44B Heat exchange part, 52, 54 Blower, 60 Piping, 70 Control device 71 High pressure calculation unit 72 Low pressure calculation unit 73 Saturation temperature calculation unit 74 Storage unit 75 Target value calculation unit 76 Control unit 81 Discharge temperature sensor 82 Condensation temperature sensor 84 Evaporation temperature sensor 100 Freezing Cycle device, 110 refrigerant circuit, 200 case main body, 210 opening, 300 cooling fan, 401, 411, 431A, 431B, 441A, 441B, 901 heat transfer tube, 402, 412, 902 radiating fins.

Claims (6)

A refrigerant circuit including a compressor, a condenser, an expansion valve, and an evaporator, and configured to circulate the refrigerant;
A controller for controlling the temperature of the refrigerant discharged from the compressor by adjusting the opening of the expansion valve;
The evaporator includes a flow path for the refrigerant,
At least a part of the flow path is configured to form a parallel flow in which the refrigerant flows from the upstream side to the downstream side in the air flow direction in the evaporator ,
The other part of the flow path is configured to form a counter flow in which the refrigerant flows from the downstream side in the flow direction toward the upstream side,
The flow path that forms the parallel flow is a refrigeration cycle device that is provided on the downstream side in the flow direction and on the downstream side of the refrigerant that flows through the evaporator as compared to the flow path that forms the counterflow . The flow path is
A plurality of flow path rows arranged in a direction along the flow direction;
Forming a plurality of flow path stages arranged in a direction intersecting the flow direction,
The refrigeration cycle apparatus according to claim 1, wherein the number of the plurality of flow path rows is larger than the number of the plurality of flow path stages. The cross-sectional area of the formed channel to form a parallel flow is larger than the cross-sectional area of the formed channel to form said counterflow refrigeration cycle apparatus according to claim 1. The number of branches composed channel to form a parallel flow, the greater than the number of branches formed channel to form a counter flow, refrigeration cycle apparatus according to claim 1. The refrigeration cycle apparatus according to any one of claims 1 to 4 , wherein the refrigerant includes an HFC (Hydrofluoro Carbon) refrigerant. The refrigeration cycle apparatus according to any one of claims 1 to 5 ,
A case body having an opening and containing an article cooled by the refrigeration cycle apparatus;
A showcase comprising a blower for sending air so that an air curtain is formed in the opening.

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