JP6715918B2 - Refrigeration cycle equipment - Google Patents

Description

The present invention relates to a refrigeration cycle device used for applications such as air conditioning and hot water supply.

Conventionally, as in Patent Document 1, an evaporator, a compressor (corresponding to the compressor of the present invention), a condenser, and a capillary tube (corresponding to the pressure reducing device of the present invention) in which refrigerant circulates There has been known a refrigeration cycle device that is connected to each other and circulates a refrigerant.

Further, a non-azeotropic mixed refrigerant in which a plurality of kinds of refrigerants are mixed may be used as the refrigerant used in the refrigeration cycle device. The temperature of the non-azeotropic mixed refrigerant changes due to the phase change under the same pressure, and the temperature becomes lower on the downstream side than on the upstream side in the evaporation process.If the non-azeotropic mixed refrigerant is used as the refrigerant of the refrigeration cycle device, the outlet side of the evaporator Has a higher temperature than the inlet side of the evaporator. Further, the lower the temperature of the refrigerant sent to the evaporator is, the more the cooling performance of the evaporator is improved, but if the temperature of the refrigerant sent to the evaporator is lower than 0° C., the problem of frost formation on the evaporator occurs. Therefore, when a non-azeotropic mixed refrigerant is used in the refrigeration cycle device, the temperature of the refrigerant flowing to the evaporator is high and the cooling performance may be insufficient.

On the other hand, in Patent Document 1, the evaporation in which the pressure loss in the pipeline in the evaporation process is increased and the pressure in the pipeline in the evaporation process is gradually decreased so that the temperature of the refrigerant in the evaporator becomes constant. A refrigeration cycle apparatus is described. In such a refrigeration cycle apparatus, a rise in temperature during the evaporation process is suppressed, and both improvement of cooling performance and prevention of frost formation can be achieved.

JP, 9-133433, A

However, if the pressure loss in the pipeline in the evaporation process, that is, the pressure loss of the evaporator is increased, the pressure of the refrigerant in the evaporator is decreased, and the heat exchange performance of the evaporator is deteriorated. Therefore, depending on the pressure loss of the evaporator, the performance deterioration due to the increase of the pressure loss becomes larger than the performance improvement due to the suppression of the temperature rise in the evaporation process, and the entire refrigeration cycle apparatus. There is a problem in that the driving performance of In particular, when a low-temperature device with a low evaporation temperature, hot water supply/heating operation in a cold region, or a refrigerant with a low operating pressure is used, the decrease in the refrigerant pressure in the evaporator greatly affects the performance of the refrigeration cycle device, This problem becomes apparent.

The present invention has been made in view of the above problems, and an object of the present invention is to provide a refrigeration cycle apparatus that prevents a decrease in the operating performance of the entire refrigeration cycle apparatus.

の式で表される関数Ｔ_{ｓａｔ＿ｕ}（ｕ）の最小値をＴ_{ｓａｔ＿ｕ＿ｍｉｎ}として、前記制御装置は、前記流速センサの測定値が、前記関数Ｔ_{ｓａｔ＿ｕ}（ｕ）がＴ_{ｓａｔ＿ｕ＿ｍｉｎ}＋０．３（℃）未満となるような前記平均流速ｕに近づくように、前記冷媒が前記蒸発器を流れるように、前記圧縮機の回転数を制御することを特徴としている。 The refrigeration cycle apparatus of the present invention includes a compressor, a condenser, a decompression device, an evaporator, and a refrigeration cycle in which a single refrigerant is circulated, and a control device, and the evaporator includes , A flow rate sensor is attached , the heat exchange amount of the evaporator is Q(W), the refrigerant flow path diameter of the evaporator is d(m), and the heat transfer area inside the tube of the evaporator is A _i (m ^{2 ).} ), the average thermal conductivity of the refrigerant is k (W/mK), the average density of the refrigerant is ρ (kg/m ³ ), the average viscosity of the refrigerant is μ (Pa·s), and the Prandtl number of the refrigerant is Pr, the average flow velocity of the refrigerant flowing through the evaporator is u (m/s), and a function of the average flow velocity u that is defined by the type of the refrigerant and represents the refrigerant saturation temperature difference between the inlet and the outlet of the evaporator is T _sat. (U),

The minimum of the function _{T Sat_u} represented by the formula (u) as _{T Sat_u_min,} wherein the control device, measurement of the flow velocity sensor, before Symbol function _{T Sat_u} (u) is _{T sat_u_min} +0.3 (℃) closer to the average flow velocity u that is less than said refrigerant to flow through said evaporator, is characterized by controlling the rotation speed of the compressor.

In the refrigeration cycle of the present invention, the refrigerant flows through the evaporator at an average flow rate u (m/s) at which the function T _{sat_u} (u) is less than T _{sat_u_min} +0.3 (°C), so the performance resulting from the increase in pressure loss is Of the refrigeration cycle apparatus can be prevented, and the deterioration of the operating performance of the entire refrigeration cycle apparatus can be prevented.

FIG. 1 is a schematic diagram showing the overall configuration of the refrigeration cycle device according to the first embodiment. FIG. 3 is a schematic diagram showing a configuration of an evaporator of the refrigeration cycle device according to the first embodiment. It is sectional drawing of the AA cross section shown in FIG. 2 of the evaporator of the refrigerating-cycle apparatus which concerns on Embodiment 1. 7 is a graph of a refrigerant saturation temperature difference ΔT _sat between an inlet and an outlet of an evaporator of a refrigeration cycle device according to a second embodiment and an average flow rate u of a refrigerant flowing through the evaporator. 7 is a graph of a temperature difference generated between a cooling target and a refrigerant due to the heat transfer performance of the refrigeration cycle device according to the second embodiment, and an average flow rate u of the refrigerant flowing through the evaporator. ₇ is a graph of a refrigerant saturation temperature T _{sat_out at} the outlet of the evaporator 4 of the refrigeration cycle device according to the second embodiment and an average flow rate u of the refrigerant flowing through the evaporator.

Embodiment 1.
FIG. 1 is a schematic diagram showing the configuration of the entire refrigeration cycle apparatus according to the first embodiment. FIG. 2 is a schematic diagram showing the configuration of the evaporator of the refrigeration cycle device according to the first embodiment. FIG. 3 is a cross-sectional view of the evaporator of the refrigeration cycle apparatus according to Embodiment 1 taken along the line AA shown in FIG. 2. In each figure, the same parts or corresponding parts are designated by the same reference numerals. In the refrigeration cycle apparatus 100 according to the first embodiment, as shown in FIG. 1, a compressor 1, a condenser 2, an expansion valve 3, and an evaporator 4 are connected by a refrigerant pipe 5, respectively, and the compressor 1, the condenser A refrigeration cycle in which a refrigerant circulates inside the vessel 2, the expansion valve 3, the evaporator 4, and the compressor 1 in this order. Of the refrigerant pipes 5, one connecting the compressor 1 and the condenser 2 is connected to the refrigerant pipe 5a, one connecting the condenser 2 and the expansion valve 3 is connected to the refrigerant pipe 5b, and the expansion valve 3 and the evaporator 4 are connected to each other. The connected one is called a refrigerant pipe 5c, and the one connected between the evaporator 4 and the compressor 1 is called a refrigerant pipe 5d. The refrigerant that circulates in the refrigeration cycle device 100 according to Embodiment 1 is not particularly limited, but is determined as any one refrigerant according to the application of the refrigeration cycle device 100 and the like.

The compressor 1 draws in the refrigerant, compresses it, and discharges it into a high-temperature, high-pressure gas state. The compressor 1 may be configured to have a rotation speed controlled by, for example, an inverter circuit, and the discharge amount of the refrigerant can be adjusted by controlling the rotation speed.

The refrigerant compressed by the compressor 1 into a high temperature and high pressure gas state flows into the condenser 2, and heat is exchanged between the refrigerant and the heat source to cool the refrigerant to a low temperature and high pressure liquid state. Examples of the heat source include air, water, brine, and the like. In the first embodiment, the heat source of the condenser 2 is the outside air which is the outdoor air, and the condenser 2 performs heat exchange between the outside air and the refrigerant. Further, in the first embodiment, in order to promote heat exchange of the condenser 2, the condenser blower 6 that blows the outside air to the condenser 2 while the refrigerant is circulating in the refrigeration cycle device 100 is provided. The condenser blower 6 is preferably configured by one capable of adjusting the air volume.

The low-temperature high-pressure liquid-state refrigerant cooled by the condenser 2 flows into the expansion valve 3 and decompresses and expands the refrigerant to a low-temperature low-pressure liquid state. As the expansion valve 3, for example, a refrigerant flow rate control means such as an electronic expansion valve or a temperature-sensitive expansion valve, a capillary tube (capillary tube), or the like is configured. The expansion valve 3 corresponds to the pressure reducing device of the present invention.

In the evaporator 4, the low-temperature low-pressure liquid-state refrigerant that has been decompressed and expanded by the expansion valve 3 flows in, heat is exchanged between the refrigerant and the cooling target, and the heat of the cooling target is absorbed by the refrigerant to cool the cooling target. To cool. When cooling an object to be cooled, the refrigerant is vaporized into a high-temperature low-pressure gas state. In the first embodiment, indoor air is used as the cooling target, and the evaporator 4 performs heat exchange between the indoor air and the refrigerant. Further, in the first embodiment, in order to promote heat exchange of the evaporator 4, the evaporator blower 7 that blows indoor air to the evaporator 4 while the refrigerant circulates in the refrigeration cycle device 100 is provided. .. The evaporator blower 7 is preferably configured by one that can adjust the air volume.

A specific configuration of the evaporator 4 will be described based on FIGS. 2 and 3. The evaporator 4 is a plate fin tube heat exchanger configured by a plurality of heat transfer tubes 41, a plurality of fins 42, a refrigerant distributor 43, and a header 44. In FIG. 2, the number of the heat transfer tubes 41 is 5 and the number of the fins 42 is 28, but the number and the number of these are examples, and the present invention is not limited to this number and the number.

The heat transfer tube 41 is a tube that constitutes a flow path through which the refrigerant flows, and is made of a metal having a high thermal conductivity such as aluminum or copper. The plurality of heat transfer tubes 41 are arranged in parallel, and are connected to the refrigerant distributor 43 and the header 44. The flow path 41a of the heat transfer tube 41 of the first embodiment has a circular cross section as shown in FIG.

The fins 42 are thin plates made of metal having high thermal conductivity such as aluminum and copper. The plurality of fins 42 are arranged perpendicularly to the axial direction of the heat transfer tube 41 (direction of arrow X in FIG. 2) with a predetermined interval. Further, the heat transfer tube 41 and the fin 42 are joined by a method such as brazing so that heat is transferred from the heat transfer tube 41 to the fin 42.

The refrigerant distributor 43 has one inflow port and a plurality of outflow ports, and is a device that distributes the refrigerant flowing from the inflow port to the plurality of outflow ports and causes the refrigerant to flow out. The inlet of the refrigerant distributor 43 is connected to the expansion valve 3 via the refrigerant pipe 5c, and the plurality of outlets of the refrigerant distributor 43 are connected to the heat transfer tubes 41, respectively.

The header 44 includes a plurality of inflow ports and one outflow port, and is a device that combines the refrigerants that have flowed in from the inflow ports and causes them to flow out from the one outflow port. The inlet of the header 44 is connected to the heat transfer pipe 41, and the outlet of the header 44 is connected to the compressor 1 via the refrigerant pipe 5d.

Next, the flow of the refrigerant in the evaporator 4 will be described. First, the low-temperature low-pressure liquid-state refrigerant decompressed and expanded by the expansion valve 3 flows in from the inlet of the refrigerant distributor 43. The refrigerant that has flowed in from the inlet of the refrigerant distributor 43 is distributed and flows to the heat transfer tubes 41 from the respective outlets of the refrigerant distributor 43. The refrigerant flowing into the heat transfer tube 41 flows along the axial direction of the heat transfer tube 41 (direction of arrow X in FIG. 2 ). On the surfaces of the heat transfer tubes 41 and the fins 42, indoor air to be cooled is blown by the evaporator blower 7, and the refrigerant flowing through the heat transfer tubes 41 exchanges heat with the indoor air in contact with the heat transfer tubes 41 and the fins 42. And absorbs the heat of the indoor air. The refrigerant that has exchanged heat with indoor air in the heat transfer tube 41 flows in from the inlet of the header 44, merges with the header 44, and flows from the outlet of the header 44 to the compressor 1.

Next, the refrigerant flow path length L(m), the refrigerant flow path diameter d(m), the pipe outside heat transfer area A _o (m ² ) and the pipe inside heat transfer area A of the evaporator 4 used in describing the present invention will be described. Define _i (m ² ). In the description of the first embodiment, the contents in parentheses after each symbol indicate the unit of the symbol. As shown in FIG. 2, the refrigerant flow path length L varies from the joint portion 41b between the heat transfer tube 41 and the fin 42a located on the most upstream side to the joint portion 41c between the heat transfer tube 41 and the fin 42b located on the most downstream side. Up to. The refrigerant flow path diameter d is the inner diameter of the heat transfer tube 41, as shown in FIG. The tube outside heat transfer area A _o is the sum of the surface areas of the plurality of heat transfer tubes 41 on the outer side and the surface areas of the plurality of fins 42 from the joint portion 41 b to the joint portion 41 c. The tube inner heat transfer area A _i is the sum of the inner areas of the plurality of heat transfer tubes 41 from the joint portion 41b to the joint portion 41c. Further, the refrigerant flow path length L, the refrigerant flow path diameter d, the pipe outside heat transfer area A _o, and the pipe inside heat transfer area A _i are numerical values determined when the refrigeration cycle apparatus 100 is configured, and thus are constants.

Next, the operation performance of the refrigeration cycle apparatus 100 will be described. Since most of the energy consumed by the refrigeration cycle apparatus 100 is occupied by the compression power of the compressor 1, it is effective to reduce the compression power of the compressor 1 in order to improve the operating performance. As a method of reducing the compression power of the compressor 1, there is a reduction of the compression ratio. If the suction pressure of the compressor 1 is raised or the discharge pressure is lowered, the compression ratio is reduced. The present invention focuses on increasing the suction pressure of the compressor 1.

Increasing the suction pressure of the compressor 1 is, in other words, increasing the outlet pressure of the evaporator 4. Further, the pressure of the refrigerant can be indicated by the refrigerant saturation temperature, and in the following description, the refrigerant pressure will be indicated by the refrigerant saturation temperature. When a refrigerant that circulates in the refrigeration cycle apparatus 100, such as a non-azeotropic mixed refrigerant, whose temperature changes due to a phase change under the same pressure, the refrigerant saturation temperature changes with respect to the enthalpy in the evaporation process, so that the refrigerant saturation occurs. The temperature difference is the refrigerant saturation temperature difference in the pressure difference at isenthalpy.

From the above description, by raising the refrigerant saturation temperature at the outlet of the evaporator 4, the outlet pressure of the evaporator 4 rises, the suction pressure of the compressor 1 rises, the compression ratio of the compressor 1 decreases, and the compressor 1 It can be seen that the compression power is reduced, and finally the operation performance of the refrigeration cycle apparatus 100 is improved. That is, it is understood that the operation performance of the refrigeration cycle apparatus 100 is maximized by maximizing the refrigerant saturation temperature at the outlet of the evaporator 4.

Next, the refrigerant saturation temperature at the outlet of the evaporator 4 will be described. The refrigerant saturation temperature T _{sat_out at} the outlet of the evaporator 4 is obtained by subtracting the refrigerant saturation temperature difference ΔT _sat at the outlet of the evaporator 4 from the refrigerant saturation temperature T _{sat_in at} the inlet of the evaporator 4 as shown in Formula 2. ..

Here, the refrigerant saturation temperature T _{sat_in at} the inlet of the evaporator 4 will be described. First, the heat exchange amount Q(W) of the evaporator 4 can be expressed as in Equation 3 using the refrigerant saturation temperature T _{sat_in} at the inlet of the evaporator 4. When Equation 3 is transformed into an equation for deriving the refrigerant saturation temperature T _{sat_in} at the inlet of the evaporator 4, Equation 4 is obtained, and the refrigerant saturation temperature T _{sat_in at} the inlet of the evaporator 4 is _calculated based on Equation 4 below. Can be derived. Here, the symbols used in Equations 3 and 4 are defined. T _air is the temperature (° C.) of the object to be cooled by the evaporator 4, and in the case of the first embodiment, the temperature of indoor air corresponds. α _o is the heat transfer coefficient (W/m ² K) of the object to be cooled by the evaporator 4, and in the case of the first embodiment, the heat transfer coefficient of indoor air corresponds. α _i is the heat transfer coefficient (W/m ² K) of the refrigerant. Here, the temperature T _air of the cooling target, the heat exchange amount Q of the evaporator 4, and the heat transfer coefficient α _o of the cooling target are values that are determined if the state of the cooling target and the configuration of the evaporator 4 are determined. In the present invention, it is treated as a constant. In addition, regarding A _o and A _i in Formula 3 and Formula 4, since they are the pipe outside heat transfer area A _o (m ² ) and the pipe inside heat transfer area A _i (m ² ) defined above, respectively, Omit the definition of.

The heat transfer coefficient α _i of the refrigerant can be derived based on Equation 5 below. Here, the symbols used in Equation 5 are defined. k is the average thermal conductivity (W/mK) of the refrigerant, and is obtained by taking the average value of the thermal conductivity of the saturated gas state refrigerant and the saturated liquid state refrigerant. Re is the Reynolds number of the refrigerant and is a dimensionless number. Pr is the Prandtl number of the refrigerant and is a dimensionless number. Since the thermal conductivity k of the refrigerant and the Prandtl number Pr of the refrigerant are values that are determined when the type of the refrigerant is determined, they are treated as constants in the present invention. It should be noted that d in Expression 5 is the refrigerant flow path diameter d(m) defined above, and therefore the definition here is omitted.

The Reynolds number Re can be derived based on Equation 6 below. Here, the symbols used in Equation 6 are defined. ρ is an average density (kg/m ³ ) of the refrigerant, and is obtained by taking an average value of the density of the refrigerant in the saturated gas state and the density of the refrigerant in the saturated liquid state. u is the average flow velocity (m/s) of the refrigerant flowing through the evaporator 4. Further, hereinafter, when simply referred to as the average flow velocity u, it means the average flow velocity u (m/s) of the refrigerant flowing through the evaporator 4. μ is the average viscosity (Pa·s) of the refrigerant, and is obtained by taking the average value of the viscosity of the refrigerant in the saturated gas state and the viscosity of the refrigerant in the saturated liquid state. Since the average density ρ of the refrigerant and the average viscosity μ of the refrigerant are values that are determined when the type of the refrigerant is determined, they are treated as constants in the present invention. Further, since the average flow velocity u is a value that can be easily changed by adjusting the rotation speed of the compressor 1 or the opening degree of the expansion valve 3, it is treated as a variable in the present invention. It should be noted that d in Formula 6 is the refrigerant flow path diameter d(m) defined above, and therefore the definition here is omitted.

By substituting Equation 5 into Equation 4 and rearranging, Equation 7 below is obtained, and further substituting Equation 6 into Equation 7 and rearranging yields Equation 8. Here, all symbols other than the average flow velocity u in the formula 7 are treated as constants, the first term and the second term in the formula 8 are also treated as constants, and the third term in the formula 8 is 0.8 power of the average flow velocity u. Inversely proportional to. Therefore, it can be seen that the smaller the value of the average flow velocity u, the larger the value of the third term of Expression 8 and the smaller the refrigerant saturation temperature T _{sat_in at} the inlet of the evaporator 4.

Next, the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 will be described. The refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 is determined by the type of refrigerant and the pressure loss ΔP (Pa) of the evaporator 4. Particularly, as the relationship between the pressure loss ΔP of the evaporator 4 and the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4, the refrigerant saturation temperature difference ΔT between the inlet and the outlet of the evaporator 4 increases as the pressure loss of the evaporator 4 increases. _sat becomes large. Note that the detailed relational expression between the pressure loss ΔP of the evaporator 4 and the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 varies depending on the type of refrigerant, but if the type of refrigerant is determined, the detailed evaporator 4, the relational expression between the pressure loss ΔP and the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 is uniquely determined, and the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 is determined by Equation 9 as follows. The pressure loss ΔP of 4 can be expressed as a function T _sat (ΔP). Further, no matter what kind of refrigerant is used, the relationship in which the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 increases as the pressure loss ΔP of the evaporator 4 increases, does not change.

The pressure loss ΔP of the evaporator 4 can be derived based on the following formula 10. Here, the symbols used in Expression 10 are defined. Φ is an average correction factor for the two-phase flow pressure loss, which is a dimensionless number. Further, the average correction coefficient Φ of the two-phase flow pressure loss can be generally derived by the Lockhart-Martinelli calculation method. λ is a friction loss coefficient of the refrigerant and is a dimensionless number. Further, the average correction coefficient Φ of the two-phase flow pressure loss is a value that is determined if the type of refrigerant used is determined, and is therefore treated as a constant. Regarding L, d, ρ, and u in the mathematical expression 10, the refrigerant channel length L(m) defined above, the refrigerant channel diameter d(m), and the average density ρ of the refrigerant (kg/m ³ ) And the average flow velocity u (m/s), the definition here is omitted.

The friction loss coefficient λ of the refrigerant can be derived based on the following formula 11. Re in Equation 11 is a Reynolds number, and Equation 12 is obtained by substituting Equation 6 in Reynolds number Re in Equation 11.

By substituting the equation 12 into the friction loss coefficient λ of the refrigerant in the equation 10, the following equation 13 is obtained. Here, all symbols other than the average flow velocity u in the formula 13 are treated as constants. From the formula 13, the friction loss coefficient λ of the refrigerant is proportional to the 1.75th power of the average flow velocity u. It can be seen that the friction loss coefficient λ increases.

As described above, if the type of the refrigerant is determined, the detailed relational expression of the pressure loss ΔP of the evaporator 4 and the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 is also uniquely determined, and the equation 13 is other than the average flow velocity u. Since all symbols are constants, the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 can be expressed as a function T _sat (u) of the average flow velocity u as shown in Formula 14 when the type of refrigerant is determined. it can. Further, there is a relationship that the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 increases as the pressure loss of the evaporator 4 increases. As shown in Formula 13, the faster the average flow velocity u is, the friction loss coefficient of the refrigerant is increased. λ has a relationship of increasing. From these two relationships, the value of the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4, that is, the value of the function T _sat (u) increases as the value of the average flow velocity u increases.

Then, by substituting the equation 8 into the refrigerant saturation temperature T _{sat_in} at the inlet of the evaporator 4 of the equation 2 and by substituting the equation 14 into the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4, the equation 15 is obtained. In Expression 15, all the symbols other than the average flow velocity u are treated as constants, so the first term and the second term of Expression 15 are constants. Further, the values of the third term and the fourth term of the formula 15 differ depending on the value of the average flow velocity u, and the average flow velocity for deriving the temperature decrease amount in the evaporator 4 due to the average flow velocity u as shown in the formula 16. It can be expressed as a function T _{sat —} u (u) of u, and by substituting Equation 16 into Equation 15, Equation 17 is obtained. From Expression 17, it is understood that the function T _{sat_u} (u) shown in Expression 16 may be minimized in order to maximize the refrigerant saturation temperature T _{sat_out} at the outlet of the evaporator 4. Further, it can be seen from Equation 16 that the average flow velocity u at which the refrigerant saturation temperature at the outlet of the evaporator 4 has the maximum value T _{sat_out_max} is the average flow velocity u at which the function T _{sat_u} (u) has the minimum value T _{sat_u_min} . Therefore, in the refrigeration cycle device 100 of the first embodiment, the refrigerant flows through the evaporator 4 at the average flow rate u such that the function T _{sat_u} (u) has the minimum value T _{sat_u_min} .

Here, whether or not the function T _{sat —} u (u) shown in Expression 16 can be minimized will be described. It can be seen that the first term of Expression 16 is inversely proportional to the 0.8th power of the average flow velocity u, and increases as the average flow velocity u decreases. In addition, the second term of the mathematical expression 16 is that the average flow velocity u becomes large as the value of the function T _sat (u) indicating the relationship between the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 and the average flow velocity u as described above. It can be seen that the larger the average flow velocity u, the larger. Therefore, the function T _{sat_u} (u) in Expression 16 is a sum of a term that increases as the average flow velocity u _decreases and a term that increases as the average flow velocity u increases, and therefore the function T _{sat_u} (u) has a minimum value. It can be _seen that there is a value T _{sat_u_min and} that the function T _{sat_u} (u) can be minimized. Further, when the average flow velocity u is 0, the first term of Expression 16 diverges to infinity, so the minimum value T _{sat_u_min} of the function T _{sat_u} (u) exists when the average flow velocity u is a value larger than 0. I know what to do.

As described above, in the refrigeration cycle device 100 of the first embodiment, the refrigerant flows through the evaporator 4 at the average flow velocity u such that the function T _{sat_u} (u) becomes the minimum value T _{sat_u_min} , so that the refrigerant at the outlet of the evaporator. The saturation temperature can be maximized, and as a result, the operating performance of the refrigeration cycle apparatus can be maximized, and the deterioration of the operating performance of the entire refrigeration cycle apparatus can be prevented. Further, in a low-temperature device having a low evaporation temperature, a hot water supply/heating operation in a cold region, a refrigerant having a low operating pressure, and the like, the suction pressure of the compressor has a great influence on the operation performance, and is particularly effective.

In order to prevent the deterioration of the operation performance of the entire refrigeration cycle apparatus 100, it is not always necessary that the average flow velocity u be the minimum value T _{sat_u_min of the} function T _{sat_u} (u), and the refrigeration cycle apparatus 100 The average flow velocity u may have a range to such an extent that the driving performance is not extremely deteriorated. For example, in a commonly used refrigerant, it is known that if the refrigerant saturation temperature T _{sat_out at} the outlet of the evaporator 4 decreases by 0.3° C. or more, the operating performance decreases by 1%, which has a _{nonnegligible} effect. Therefore, if the refrigerant flows through the evaporator at an average flow rate u such that the function T _{sat_u} (u) shown in Formula 16 is less than T _{sat_u_min} +0.3 (° C.), the refrigeration cycle device may have a refrigeration cycle apparatus. The effect of preventing the deterioration of the driving performance of the entire 100 is exhibited.

Further, if the average flow velocity u can be controlled, a control device that acquires the average flow velocity u and controls the average flow velocity u based on the acquired average flow velocity u may be provided.

The average flow velocity u (m/s) of the refrigerant flowing through the evaporator 4 is represented by the following formula 18. Here, Gr used in Expression 18 is the mass flow rate (kg/s) of the refrigerant. The mass flow rate of the refrigerant can be controlled by the rotation speed of the compressor 1, the opening degree of the expansion valve 3, and the like. Note that ρ and Ai in Formula 18 are defined here because the average density ρ (kg/m ³ ) of the refrigerant defined above and the heat transfer area inside the tube of the evaporator are A _i (m ² ). Omit.

From Equation 18, it can be seen that the mass flow rate Gr of the refrigerant may be controlled to control the average flow velocity u of the refrigerant flowing through the evaporator 4. That is, examples of means for controlling the average flow velocity u include means for controlling the number of revolutions of the compressor 1 and controlling the opening degree of the expansion valve 3. In the compressor 1, the average flow velocity u can be increased by increasing the rotation speed, and the average flow velocity u can be decreased by decreasing the rotation speed. Further, in the expansion valve 3, the average flow velocity u can be increased by decreasing the opening amount, and the average flow velocity u can be decreased by increasing the opening amount.

As a method of acquiring the average flow velocity u of the refrigerant flowing through the evaporator 4, there is a method of directly measuring the average flow velocity u of the refrigerant flowing through the evaporator 4 using a flow velocity sensor. Furthermore, since the mass flow rate Gr of the refrigerant and the average flow rate u of the refrigerant are in a proportional relationship from the formula 18, the mass flow rate Gr flowing through the evaporator 4 measured using a flow meter or the set value of the rotation speed of the compressor The average flow velocity u of the refrigerant can be acquired based on the two set values including the set value of the opening degree of the expansion valve.

The control device sets, for example, the measured value of the flow velocity sensor attached to the heat transfer tube 41 of the evaporator 4 and the set value of the average flow velocity u at which the preset function T _{sat_u} (u) shown in Formula 16 becomes the minimum value T _{sat_u_min.} The average flow velocity u may be controlled so as to make a comparison and bring the measured value close to the set value. Further, when the refrigerant saturation temperature T _{sat_out} at the outlet of the evaporator 4 is used as the state of the refrigerant, the control device predetermines the rotation speed of the compressor or the opening degree of the evaporator when the operation of the refrigeration cycle device 100 is started. If it is controlled to operate at the rotational speed of the compressor or the opening degree of the evaporator at which the refrigerant saturation temperature T _{sat_out at} the outlet of the evaporator 4 becomes the minimum value T _{sat_u_min by} a method such as a dichotomy method by changing it within a predetermined range. Good.

Moreover, the structure of the refrigeration cycle apparatus 100 of Embodiment 1 is an example, and the structure of the refrigeration cycle apparatus of the present invention is not limited to the structure of Embodiment 1. For example, although the cross section of the flow path of the heat transfer tube 41 of the first embodiment is circular, the cross section of the heat transfer tube is not limited to this and may be elliptical, flat, or rectangular. When the cross section of the flow path of the heat transfer tube is elliptical, flat or rectangular, the equivalent diameter may be used as the refrigerant flow path diameter d(m).

Embodiment 2.
Refrigeration cycle apparatus 100 according to Embodiment 2 differs from refrigeration cycle apparatus 100 according to Embodiment 1 in that the composition of the refrigerant circulating in refrigeration cycle apparatus 100 is limited. Other than that, the configuration of refrigeration cycle apparatus 100 itself is similar to that of the first embodiment, and therefore the description thereof is omitted.

The refrigerant that circulates in the refrigeration cycle apparatus 100 of the second embodiment is a mixed refrigerant of R32, R1234yf, and R125, and the ratio XR32 of _R32 is 67 (wt%) and the ratio XR1234yf of _XR1234yf is 26 (wt%). ), and the ratio of _R125 X _R125 is 7 (wt %). The following effects can be obtained by using such a refrigerant.

First, each single refrigerant has physical properties that are advantages or disadvantages, but by mixing a plurality of refrigerants, the disadvantages can be reduced and the advantages can be increased. As the physical properties of R32 are high operating pressure, it is possible to reduce the influence of performance deterioration due to pressure loss, and it is possible to improve the refrigerating capacity even at low temperatures such as in supermarket showcases. As for the physical properties of R1234yf, since the global warming potential is 0, environmental impact can be reduced. Since the physical properties of R125 are nonflammable, the flammability of R32 and R1234yf can be reduced, and the safety can be improved. Therefore, the above-mentioned mixed refrigerant has less influence on the global environment and can improve safety and performance at the same time.

Further, the mixed refrigerant is a non-azeotropic mixed refrigerant, and the temperature of the non-azeotropic mixed refrigerant changes due to a phase change under the same pressure, and the temperature of the downstream side becomes higher than that of the upstream side in the evaporation process, and the refrigerant of the refrigeration cycle apparatus 100. When a non-azeotropic mixed refrigerant is used for the evaporator 4, the temperature on the outlet side of the evaporator 4 becomes higher than that on the inlet side of the evaporator 4. In particular, since the refrigerating capacity of the mixed refrigerant of R32 is improved even at low temperatures, the refrigeration cycle apparatus 100 is often used in low-temperature equipment whose saturation temperature is below freezing, such as a showcase, and thus the saturation temperature is below freezing. When used in the low temperature equipment, frost is formed on the evaporator 4. When a refrigerant having a temperature gradient in the evaporation process is used for a device in which frost is generated, cooling of the air is promoted on the inlet side of the evaporator 4 having a low refrigerant temperature, the amount of frost is large, and the refrigerant temperature is high. On the outlet side of the evaporator 4, the cooling of air does not proceed and the amount of frost is reduced, so that the frost of the evaporator 4 is uneven. When the frost is unevenly distributed, the space between the fins 42 on the inlet side of the evaporator 4 is clogged with frost early even if the amount of frost is the same as compared with the case where the frost is even. It occurs and it becomes impossible to exchange heat.

The non-azeotropic mixed refrigerant can make the temperature of the refrigerant flowing through the evaporator 4 uniform by continuously lowering the pressure inside the evaporator 4 at a predetermined gradient, and the uneven frost formation in the evaporator 4 can be prevented. This can be suppressed by making the temperature of the refrigerant flowing through the evaporator 4 uniform. When the pressure inside the evaporator 4 is gradually decreased, a pressure difference occurs between the inlet and the outlet of the evaporator 4, so that the refrigerant saturation temperature at the inlet and the outlet of the evaporator 4 when the temperature of the refrigerant flowing to the evaporator 4 is made uniform. The difference is uniquely determined by the composition of the non-azeotropic refrigerant mixture. Therefore, the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 is made substantially equal to the refrigerant saturation temperature difference between the inlet and the outlet of the evaporator 4 when the temperature of the refrigerant flowing through the evaporator 4 is made uniform. As a result, it is possible to suppress uneven frost formation on the evaporator 4.

However, depending on the type of refrigerant, the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 and the refrigerant saturation temperature difference between the inlet and the outlet of the evaporator 4 when the temperature of the refrigerant flowing in the evaporator 4 is made uniform. If substantially the same, the function T _{sat_u} (u) shown in Formula 16 at the average flow velocity (m/s) does not become less than T _{sat_u_min} +0.3 (° C.), or the function T _{sat_u} (u) shown in Formula 16 is T _{Even if} it becomes less than _{sat_u_min} +0.3 (° C.), the numerical values such as the refrigerant flow path length L (m) or the refrigerant flow path diameter d (m) of the evaporator 4 become unrealistic values, and as an actual refrigeration cycle device May not hold.

Here, in the refrigeration cycle apparatus 100 of Embodiment 2, the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 is _equal to the inlet of the evaporator 4 when the temperature of the refrigerant flowing through the evaporator 4 is made uniform. The case where the difference in the saturated temperature of the refrigerant at the outlet is approximately the same will be described. In the following description, a graph of the values of the refrigerant physical properties of the mixed refrigerant and the calculation results of the function T _sat (u) are shown. Of Ver. The simulation was performed using 9.1 and the mixing rule installed as standard in this software.

Next, among the symbols described in Formulas 2 to 17, specific numerical values in the case where the above mixed refrigerant is used will be described for symbols depending on the coolant. First, regarding the average density ρ (kg/m ³ ) of the refrigerant, the density of the refrigerant in the saturated gas state is about 17.6 (kg/m ³ ), and the density of the refrigerant in the saturated liquid state is about 116.8 (kg/m ³ ). ³ ), the average value of both is 572.2 (kg/m ³ ). As for the average viscosity μ (Pa·s) of the refrigerant, the viscosity of the refrigerant in the saturated gas state is about 1.1×10 ⁻⁵ (Pa·s), and the viscosity of the refrigerant in the saturated liquid state is about 1.72×10. ^{Since it is −4} (Pa·s), it is about 9.2×10 ⁻⁵ (Pa·s) from the average value of both. Regarding the average thermal conductivity k (W/mK) of the refrigerant, the thermal conductivity of the refrigerant in the saturated gas state is about 0.011 (W/mK), and the thermal conductivity of the refrigerant in the saturated liquid state is about 0.126. Since it is (W/mK), it is about 0.069 (W/mK) from the average value of both. The Prandtl number Pr is about 1.766. The average correction factor Φ for the two-phase flow pressure loss is about 3.

Next, among the symbols described in Formulas 2 to 17, specific numerical values will be described for the symbols that depend on the structure of the refrigeration cycle apparatus 100. The refrigerant flow path length L (m) of the evaporator 4 is 9 (m), the refrigerant flow path diameter d (m) is 9.52×10 ⁻³ (m), and the tube outer side heat transfer area A _o (m ² ) is 29.4 (m ² ), the tube inner heat transfer area A _i (m ² ) is 4.8 (m ² ), and the heat exchange amount Q(W) is 20000 (W). It should be noted that these numerical values are realistic values in constructing the refrigeration cycle apparatus 100, and are established as an actual refrigeration cycle apparatus.

Further, among the symbols described in Formulas 2 to 17, specific numerical values will be described for the symbols depending on the cooling target. As for T _air , the temperature T _air (° C.) to be cooled by the evaporator 4 is 0.7 (° C.), and the heat transfer coefficient α _o (W/m ² K) to be cooled is 100 (W/m ² K). ).

Further, in the above-mentioned mixed refrigerant, the refrigerant saturation temperature difference between the inlet and the outlet of the evaporator 4 when the temperature of the refrigerant flowing through the evaporator 4 is made uniform is about 1.3°C. Therefore, if the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 is set to about 1.3° C., the temperature rise in the evaporation process is suppressed and the temperature of the refrigerant flowing through the evaporator 4 becomes uniform.

FIG. 4 is a graph of the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator of the refrigeration cycle apparatus according to the second embodiment and the average flow velocity u of the refrigerant flowing through the evaporator. In other words, the graph of FIG. 4 is a graph of the function T _sat (u), and as described above, the value of the refrigerant saturation temperature difference ΔT _sat at the inlet and the outlet of the evaporator 4 increases as the value of the average flow velocity u increases. It turns out to be a relationship. Moreover, from the graph of FIG. 4, the value of the average flow velocity u at which the value of the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 is 1.3° C. is about 1.2 (m/s). Therefore, in the refrigeration cycle apparatus 100 in the second embodiment, if the value of the average flow velocity u is about 1.2 (m/s), the evaporator 4 when the temperature of the refrigerant flowing through the evaporator 4 is made uniform. The refrigerant saturation temperature difference between the inlet and the outlet of the evaporator and the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 become equal, and the temperature of the refrigerant flowing through the evaporator 4 can be made uniform.

Further, the function T _sat (u) is derived from the relationship of the pressure loss ΔP of the evaporator 4 as the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 as shown in Formulas 9 to 14. This pressure loss ΔP is a pressure loss due to friction between the refrigerant and the pipe, and increases continuously from the inlet to the outlet of the evaporator 4. Therefore, the change in the pressure in the evaporator 4 can be continuously reduced with a gradient from the inlet to the outlet of the evaporator 4. Therefore, in order to suppress the uneven frost formation of the evaporator 4, the evaporator 4 is a plate fin tube heat exchanger as shown in FIG. The method of changing the pressure loss by changing the average flow velocity u according to the opening degree of the valve 3 and the number of branches of the refrigerant flow path is optimal.

FIG. 5 is a graph of the temperature difference between the cooling target and the refrigerant due to the heat transfer performance of the refrigeration cycle apparatus according to the second embodiment, and the average flow rate u of the refrigerant flowing through the evaporator. It can be seen from FIG. 5 that the smaller the value of the average flow velocity u, the larger the value of the temperature difference between the cooling target and the refrigerant due to the heat transfer performance. The reason for this will be described below. First, in other words, the temperature difference between the cooling target and the refrigerant due to the heat transfer performance of the refrigeration cycle device is expressed by the difference T _air -T _{sat_in} between the cooling target temperature T _air and the refrigerant saturation temperature T _{sat_in} at the inlet of the evaporator 4. It Further, by modifying the formula 8, the formula 19 can be derived. It can be seen from Equation 19 that T _air −T _{sat_in} is inversely proportional to the _0.8th power of the average flow velocity u, and T _air −T _{sat_in} increases as the value of the average flow velocity u decreases.

FIG. 6 is a graph of the refrigerant saturation temperature T _{sat_out at} the outlet of the evaporator 4 of the refrigeration cycle apparatus according to Embodiment 2 and the average flow rate u of the refrigerant flowing through the evaporator. In other words, the graph of FIG. 6 is a graph of Expression 15, and it can be seen that the refrigerant saturation temperature T _{sat_out} at the outlet of the evaporator 4 has the maximum value T _{sat_out_max} . Further, from the relationship between Formula 15 and Formula 17, the average flow velocity u at which the refrigerant saturation temperature T _{sat_out at} the outlet of the evaporator 4 has the maximum value T _{sat_out_max} is the average flow velocity u at which the function T _{sat_u} (u) has the minimum value T _{sat_u_min.} But also.

From the graph of FIG. 6, the maximum value T _{sat_out_max} of the refrigerant saturation temperature T _{sat_out} at the outlet of the evaporator 4 is about −10.0° C. Further, the average flow velocity u that is the maximum value T _{sat_out_max} of the refrigerant saturation temperature T _{sat_out} at the outlet of the evaporator 4 is about 1.2 (m/s), and when the temperature of the refrigerant flowing to the evaporator 4 is made uniform. The average flow velocity u at which the difference in refrigerant saturation temperature between the inlet and the outlet of the evaporator 4 and the difference in refrigerant saturation temperature ΔT _sat between the inlet and the outlet of the evaporator 4 are equal to each other. Therefore, the refrigeration cycle apparatus 100 according to the second embodiment can make the temperature of the refrigerant flowing through the evaporator 4 uniform to prevent uneven frost formation, and at the same time reliably maximize the operation performance of the entire refrigeration cycle apparatus. You can

As described above, as in the second embodiment, when a mixed refrigerant having a composition of R32 67 wt%, R1234yf 26 wt% and R125 7 wt% is used as the refrigerant circulating in the refrigeration cycle device, the actual refrigeration cycle In the range that can be realized as a device, the temperature of the refrigerant flowing through the evaporator 4 can be made uniform to suppress the unevenness of frost formation, and at the same time, the operation performance of the entire refrigeration cycle device can be reliably maximized.

In order to prevent the deterioration of the operation performance of the entire refrigeration cycle apparatus 100 and suppress the uneven frost formation of the evaporator 4, the average flow velocity u is always _set to the maximum value of the refrigerant saturation temperature T _{sat_out at} the outlet of the evaporator 4. It becomes T _{sat_out_max} , and it is necessary that the refrigerant saturation temperature difference between the inlet and the outlet of the evaporator 4 and the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 when the temperature of the refrigerant flowing through the evaporator 4 is made uniform. There is no. For the same reason as in the first embodiment, if the refrigerant flows through the evaporator at the average flow rate u such that the function T _{sat_u} (u) shown in Formula 16 is less than T _{sat_u_min} +0.3 (° C.), the refrigerant is frozen. The effect of preventing the deterioration of the operation performance of the entire cycle device 100 is exhibited, and the uneven frost formation of the evaporator 4 can be suppressed. Under the conditions of the second embodiment, the average flow velocity u at which the function T _{sat_u} (u) is less than T _{sat_u_min} +0.3 (°C), that is, the refrigerant saturation temperature T _{sat_out at} the outlet of the evaporator 4 is higher than -10.3°C. The average flow velocity u may satisfy the condition that the average flow velocity u is 0.85 (m/s)<u<1.6 (m/s) according to FIG. Further, the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 when the average flow velocity u satisfies the condition of 0.85 (m/s)<u<1.6 (m/s) is shown in FIG. The condition of 0.7 (° C.)<ΔT _sat <2.1 (° C.) is satisfied, and the refrigerant saturation temperature difference ΔT _sat satisfying this condition is at least as compared with the average flow velocity u of 1.6 (m/s) or more. The difference between the refrigerant saturation temperature difference ΔT _sat at the inlet and the outlet of the evaporator 4 and the difference between the refrigerant saturation temperature difference at the inlet and the outlet of the evaporator 4 when the temperature of the refrigerant flowing through the evaporator 4 is made uniform. It can be seen that the uneven frost formation on the container 4 can be suppressed.

Further, in the refrigeration cycle apparatus 100 of the second embodiment, the average flow velocity u at which the refrigerant saturation temperature T _{sat_out at} the outlet of the evaporator 4 becomes the maximum value T _{sat_out_max} and the temperature of the refrigerant flowing through the evaporator 4 are made uniform. The difference in the refrigerant saturation temperature between the inlet and the outlet of the evaporator 4 and the average flow rate u at which the difference in the refrigerant saturation temperature ΔT _sat between the inlet and the outlet of the evaporator 4 are equal to each other, but are not limited to this. .. When the temperature of the refrigerant flowing through the evaporator 4 is made uniform, the equation 16 at the average flow rate u at which the difference between the refrigerant saturation temperature at the inlet and the outlet of the evaporator 4 and the difference between the refrigerant saturation temperature ΔT _{sat at} the inlet and the outlet of the evaporator are equal to If the indicated function T _{sat_u} (u) is less than T _{sat_u_min} +0.3 (°C), the effect of preventing the deterioration of the operating performance of the entire refrigeration cycle device is exerted, and the uneven frosting of the evaporator is suppressed. it can.

Further, in the second embodiment, it is a mixed refrigerant of R32, R1234yf, and R125, and the ratio XR32 of _R32 is 67 (wt%), the ratio X of _R1234yf X _R1234yf is 26 (wt%), and the ratio _XR125 of _R125. However, the refrigerant is a non-azeotropic mixed refrigerant, and the inlet and outlet of the evaporator 4 when the temperature of the refrigerant flowing to the evaporator 4 is made uniform The function T _{sat_u} (u) shown in Formula 16 at the average flow velocity u at which the refrigerant saturation temperature difference ΔT _sat and the refrigerant saturation temperature difference ΔT _{sat at} the inlet and the outlet of the evaporator are less than T _{sat_u_min} +0.3 (° C.) For example, it exerts the effect of preventing the deterioration of the operation performance of the entire refrigeration cycle apparatus, and also exerts the effect of suppressing the uneven frosting of the evaporator.

In particular, the respective compositions different refrigerant in a range of less than ± 3 wt%, i.e. the ratio of R32 _X R32 (wt%) is _{64 <X} R32 _<70 conditions of the mixed refrigerant of the second embodiment, the proportion of R1234yf X _R1234yf (wt%) is 23<X _R1234yf <29, R125 ratio X _R125 (wt%) is 4<X _R125 <10, and the sum of X _R32 , X _R1234yf, and X _R125 is 100. Is a mixed refrigerant in which the ratio X of _R32 X _R32 is 67 (wt %), the ratio X of _R1234yf X _R1234yf is 26 (wt %), and the ratio X of _R125 X _R125 is 7 (wt %). Compared with, the various physical properties, the difference ΔT _sat of the refrigerant saturation temperature between the inlet and the outlet of the evaporator 4 and the function T _sat (u) do not change much, so the difference ΔT _sat of the refrigerant saturation temperature between the inlet and the outlet of the evaporator 4 is As shown in FIG. 4, if the condition of 0.7 (° C.)<ΔT _sat <2.1 (° C.) is satisfied, the effect of preventing the deterioration of the operation performance of the entire refrigeration cycle device is exerted, and the frost formation of the evaporator The effect of suppressing the bias of is also exerted.

1 Compressor, 2 Condenser, 3 Expansion valve, 4 Evaporator, 5 Refrigerant piping, 5a-5d Refrigerant piping, 6 Condenser blower, 7 Evaporator blower, 41 Heat transfer pipe, 41a Flow path, 41b Joint part, 41c Joining Location, 42 fins, 42a fins, 42b fins, 43 refrigerant distributor, 44 header, 100 refrigeration cycle device

Claims (2)

A compressor, a condenser, a decompression device, an evaporator, a refrigeration cycle in which a single refrigerant circulates, and a controller,
A flow rate sensor is attached to the evaporator,
The heat exchange amount of the evaporator is Q(W), the refrigerant flow path diameter of the evaporator is d(m), the heat transfer area inside the tube of the evaporator is Ai(m ² ), and the average thermal conductivity of the refrigerant is k (W/mK), the average density of the refrigerant is ρ (kg/m ³ ), the average viscosity of the refrigerant is μ (Pa·s), the Prandtl number of the refrigerant is Pr, and the refrigerant flowing through the evaporator is An average flow velocity u (m/s), a function of the average flow velocity u defined by the type of the refrigerant and showing a refrigerant saturation temperature difference between the inlet and the outlet of the evaporator is T _sat (u),

The minimum of the function _{T Sat_u} represented by the formula (u) as _{T sat_u_min,}
The control device is
The measured value of the flow velocity sensor,
The rotation speed of the compressor is controlled so that the refrigerant flows through the evaporator so as to approach the average flow velocity u such that the function T _{sat_u} (u) becomes less than T _{sat_u_min} +0.3 (°C). Refrigeration cycle device. The refrigeration cycle according to claim 1, wherein the refrigerant circulates in the refrigeration cycle such that the average flow velocity u of the refrigerant flowing through the evaporator is such that the function T _{sat_u} (u) is T _{sat_u_min} (°C). apparatus.

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