JP2019215155A - Refrigeration cycle device - Google Patents

Abstract

To provide a refrigerant cycle device for suppressing the lowering of performance resulting from an increase in pressure loss, and preventing the lowering of the operation performance of the refrigeration cycle device as a whole.SOLUTION: A refrigeration cycle device has a compressor, a condenser, a decompressor, and an evaporator, and has a refrigeration cycle in which a refrigerant circulates. The refrigerant is a mixed refrigerant of R32, R1234yf and R125, and satisfies all of a condition that a ratio X(wt%) of the R32 is 64<X<70, a condition that a ratio X(wt%) of the R1234yf is 23<X<29, a condition that a ratio X(wt%) of the R125 is 4<X<10, and a condition that a total sum of the ratio Xof the R32, the ratio Xof the R1234yf, and the ratio Xof the R125 is 100. A refrigerant saturation temperature difference ΔT(°C) between an inlet and an outlet of the evaporator satisfies a condition of 0.7<ΔT<2.1.SELECTED DRAWING: Figure 1

Description

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

  2. Description of the Related Art Conventionally, as in Patent Document 1, an evaporator, a compressor (corresponding to a compressor of the present invention), a condenser, and a capillary tube (corresponding to a decompression device of the present invention) are arranged such that a refrigerant circulates therethrough. And a refrigeration cycle device for circulating the refrigerant.

  Further, a non-azeotropic mixed refrigerant in which a plurality of types of refrigerants are mixed may be used as the refrigerant used in the refrigeration cycle apparatus. 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. Has a higher temperature than the inlet side of the evaporator. Also, the cooling performance of the evaporator is improved as the temperature of the refrigerant sent to the evaporator is lower, but if the temperature of the refrigerant sent to the evaporator falls below 0 ° C., a problem of frost formation on the evaporator occurs. Therefore, when a non-azeotropic mixed refrigerant is used in the refrigeration cycle apparatus, the temperature of the refrigerant flowing through the evaporator is high, and the cooling performance may be insufficient.

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

JP 2015-34659A

  However, if the pressure loss in the pipe during the evaporation process, that is, the pressure loss of the evaporator, is increased, the pressure of the refrigerant in the evaporator decreases, and the heat exchange performance of the evaporator decreases. For this reason, depending on the pressure loss of the evaporator, the performance decrease due to the increase in the pressure loss is larger than the performance improvement due to the suppression of the temperature rise in the evaporation process, and the entire refrigeration cycle apparatus However, there is a problem that the driving performance of the vehicle decreases. In particular, when using low-temperature equipment with low evaporation temperature, hot water supply / heating operation in cold regions, and refrigerant with low operating pressure, the drop in refrigerant pressure of the evaporator greatly affects the performance of the refrigeration cycle device, This problem appears remarkably.

  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 device that prevents a decrease in the operation performance of the entire refrigeration cycle device.

The refrigeration cycle device of the present invention includes a compressor, a condenser, a decompression device, and an evaporator, and includes a refrigeration cycle in which a refrigerant circulates, and the refrigerant is a mixture of R32, R1234yf, and R125. a refrigerant, and conditions the ratio _X R32 of the R32 (wt%) is _{64 <X} R32 _<70, a condition which is the ratio of _{R1234yf X} R1234yf (wt%) is _{23 <X R1234yf <29,} wherein R125 and conditions the ratio of _X R125 (wt%) is _{4 <X R125} <10, and conditions the sum of the percentage _{X R125} proportion _{X R1234yf} and the R125 in the the ratio _{X R32} of the R32 R1234yf is 100, the a refrigerant satisfying all, the evaporator inlet and outlet of the refrigerant saturation temperature difference [Delta] T _sat (° C.) is 0.7 <Article of [Delta] T _sat <2.1 It is characterized by satisfying.

In the refrigeration cycle of the present invention, since 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.), performance due to an increase in pressure loss is obtained. Can be suppressed, and a decrease in the operation performance of the entire refrigeration cycle apparatus can be prevented.

FIG. 2 is a schematic diagram illustrating a configuration of an entire refrigeration cycle device according to Embodiment 1. FIG. 2 is a schematic diagram illustrating a configuration of an evaporator of the refrigeration cycle device according to Embodiment 1. FIG. 3 is a cross-sectional view of the evaporator of the refrigeration cycle device according to Embodiment 1 taken along the line AA shown in FIG. 2. 9 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 Embodiment 2, and an average flow rate u of refrigerant flowing through the evaporator. 9 is a graph of a temperature difference generated between a cooling target and a refrigerant due to heat transfer performance of a refrigeration cycle device according to Embodiment 2, and an average flow velocity u of the refrigerant flowing through an evaporator. ₉ is a graph showing a refrigerant saturation temperature T _{sat_out at} an outlet of an evaporator 4 of the refrigeration cycle apparatus according to Embodiment 2, and an average flow velocity u of the refrigerant flowing through the evaporator.

Embodiment 1 FIG.
FIG. 1 is a schematic diagram illustrating a configuration of the entire refrigeration cycle apparatus according to Embodiment 1. FIG. 2 is a schematic diagram illustrating a configuration of an evaporator of the refrigeration cycle device according to Embodiment 1. 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 the drawings, the same or corresponding parts are denoted 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. A refrigeration cycle in which the refrigerant circulates inside the compressor 2, the expansion valve 3, the evaporator 4, and the compressor 1 in this order. Of the refrigerant pipes 5, the one connecting the compressor 1 and the condenser 2 is the refrigerant pipe 5a, the one connecting the condenser 2 and the expansion valve 3 is the refrigerant pipe 5b, the expansion valve 3 and the evaporator 4 are What is connected is called refrigerant pipe 5c, and what connects evaporator 4 and compressor 1 is called refrigerant pipe 5d. The refrigerant circulating in the refrigeration cycle apparatus 100 according to Embodiment 1 is not particularly limited, but is determined to be any one refrigerant depending on the use of the refrigeration cycle apparatus 100 and the like.

  The compressor 1 sucks in the refrigerant, compresses it, and discharges it in a high-temperature, high-pressure gas state. The compressor 1 may be configured such that the number of revolutions is controlled by, for example, an inverter circuit, and the discharge amount of the refrigerant can be adjusted by controlling the number of revolutions.

  The refrigerant compressed into the high-pressure and high-pressure gas state by the compressor 1 flows into the condenser 2, and performs heat exchange 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, and brine. In the first embodiment, the heat source of the condenser 2 is outside air, which is outdoor air, and the condenser 2 exchanges heat between the outside air and the refrigerant. Further, in the first embodiment, in order to promote heat exchange in the condenser 2, the condenser blower 6 that blows outside air to the condenser 2 when the refrigerant is circulating in the refrigeration cycle apparatus 100 is provided. It is preferable that the condenser blower 6 is configured to be capable of adjusting the air volume.

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

  The evaporator 4 receives the low-temperature and low-pressure liquid state refrigerant decompressed and expanded by the expansion valve 3 and performs heat exchange between the refrigerant and the object to be cooled. To cool. When the object to be cooled is cooled, the refrigerant evaporates to a high-temperature, low-pressure gas state. In the first embodiment, the object to be cooled is indoor air, 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 for blowing indoor air to the evaporator 4 when the refrigerant is circulating in the refrigeration cycle apparatus 100 is provided. . The evaporator blower 7 may be configured to be capable of adjusting the air volume.

  A specific configuration of the evaporator 4 will be described with reference to FIGS. The evaporator 4 is a plate fin tube heat exchanger including 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 five and the number of the fins is 28. However, the number and the number are only 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 a refrigerant flows, and is made of a metal having high heat conductivity such as aluminum or copper. The plurality of heat transfer tubes 41 are respectively arranged in parallel, and are connected to the refrigerant distributor 43 and the header 44. The flow passage 41a of the heat transfer tube 41 according to the first embodiment has a circular cross section as shown in FIG.

  The fin 42 is a thin metal plate having a high thermal conductivity such as aluminum or copper. The plurality of fins 42 are arranged at predetermined intervals perpendicular to the axial direction of the heat transfer tube 41 (the direction of arrow X in FIG. 2). Further, the heat transfer tubes 41 and the fins 42 are joined by a method such as brazing, so that heat is transmitted from the heat transfer tubes 41 to the fins 42.

  The refrigerant distributor 43 has one inlet and a plurality of outlets, and is a device that distributes the refrigerant flowing from the inlet to the plurality of outlets 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 is provided with a plurality of inflow ports and one outflow port, and is a device that combines the refrigerant flowing in from the inflow ports and flows out from one outflow port. The inlet of the header 44 is connected to the heat transfer tube 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 and 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 flowing in from the inlet of the refrigerant distributor 43 is distributed, and flows from the respective outlets of the refrigerant distributor 43 to the heat transfer tubes 41. The refrigerant flowing into the heat transfer tubes 41 flows along the axial direction of the heat transfer tubes 41 (the direction of the arrow X in FIG. 2). The indoor air to be cooled is blown by the evaporator blower 7 on the surfaces of the heat transfer tubes 41 and the fins 42, 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. To absorb the heat of 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 in the header 44, and flows to the compressor 1 from the outlet of the header 44.

Next, the refrigerant flow path length L (m), the refrigerant flow path diameter d (m), the pipe outer heat transfer area A _o (m ² ), and the pipe inner heat transfer area A of the evaporator 4 used in describing the present invention. Define _i (m ² ). In the description of the first embodiment, the content in parentheses after each symbol indicates the unit of the symbol. As shown in FIG. 2, the coolant flow path length L is changed from a joint 41 b between the heat transfer tube 41 and the fin 42 a located at the most upstream side to a joint 41 c between the heat transfer tube 41 and the fin 42 b located at the most downstream side. Length. The coolant flow path diameter d is the inner diameter of the heat transfer tube 41 as shown in FIG. Abluminal heat transfer area A _o is the sum of a plurality of outer surface area and the surface area of the plurality of fins 42 of the heat transfer tube 41 from the joint 41b to joint 41c. Tube-side heat transfer area A _i is the sum of the inner areas of the plurality of heat transfer tubes 41 from the joint 41b to joint 41c. The refrigerant passage length L, a refrigerant passage diameter d, abluminal heat transfer area A _o and the tube-side heat transfer area A _i are the numerical value determined at the time of constructing a refrigeration cycle apparatus 100, a constant.

  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 operating performance. As a method of reducing the compression power of the compressor 1, there is a reduction in the compression ratio. The compression ratio is reduced by increasing the suction pressure of the compressor 1 or decreasing the discharge pressure. The present invention focuses on increasing the suction pressure of the compressor 1.

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

  As described above, by raising the refrigerant saturation temperature at the outlet of the evaporator 4, the outlet pressure of the evaporator 4 increases, the suction pressure of the compressor 1 increases, the compression ratio of the compressor 1 decreases, and the compressor 1 It can be seen that the compression power of the refrigeration cycle apparatus 100 is finally improved and the operation performance of the refrigeration cycle apparatus 100 is improved. That is, it is understood that the operating 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 determined by subtracting the refrigerant saturation temperature difference ΔT _sat at the outlet from the inlet of the evaporator 4 from the refrigerant saturation temperature T _{sat_in at} the inlet of the evaporator 4 as in Expression 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 by Expression 3 using the refrigerant saturation temperature T _{sat_in} at the inlet of the evaporator 4. Here, 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. The refrigerant saturation temperature T _{sat_in at} the inlet of the evaporator 4 is _calculated based on the following Equation 4. 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, corresponds to the temperature of indoor air. α _o is the heat transfer coefficient (W / m ² K) of the object to be cooled by the evaporator 4, which corresponds to the heat transfer coefficient of indoor air in the first embodiment. α _i is the heat transfer coefficient (W / m ² K) of the refrigerant. Here, the temperature T _{air of the} object to be cooled, the heat exchange amount Q of the evaporator 4, and the heat transfer coefficient α _o of the object to be cooled are values determined when the state of the object to be cooled and the configuration of the evaporator 4 are determined. In the present invention, it is treated as a constant. Since respect to the _{A o} and _{A i} in Equation 3 and Equation 4, respectively abluminal heat transfer area defined by the above _A o ^{(m 2)} and the tube-side heat transfer area _A i ^{(m 2),} where I omit the definition.

  The heat transfer coefficient αi of the refrigerant can be derived based on the following Expression 5. Here, the symbols used in Expression 5 are defined. k is the average thermal conductivity (W / mK) of the refrigerant, and is determined by taking the average value of the thermal conductivity of the refrigerant in the saturated gas state and the thermal conductivity of the refrigerant in the saturated liquid state. 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 determined when the type of the refrigerant is determined, they are treated as constants in the present invention. Note 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 the following Expression 6. Here, the symbols used in Expression 6 are defined. ρ is the average density of the refrigerant (kg / m ³ ), and is determined by taking the 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 rate (m / s) of the refrigerant flowing through the evaporator 4. Hereinafter, when simply referred to as the average flow velocity u, the average flow velocity u (m / s) of the refrigerant flowing through the evaporator 4 is referred to. μ is the average viscosity of the refrigerant (Pa · s), which 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 determined when the type of the refrigerant is determined, they are treated as constants in the present invention. The average flow velocity u is a value that can be easily changed by adjusting the rotation speed of the compressor 1, the opening degree of the expansion valve 3, and the like, and is treated as a variable in the present invention. Note that d in Expression 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, the following Equation 7 is obtained. By substituting Equation 6 into Equation 7, and rearranging, Equation 8 is obtained. Here, all symbols other than the average flow velocity u in Equation 7 are treated as constants, and the first and second terms in Equation 8 are also treated as constants. The third term in Equation 8 is the power of 0.8 of the average flow velocity u. Is inversely proportional to Therefore, it can be seen that the smaller the value of the average flow velocity u is, the larger the value of the third term of Expression 8 is, and the smaller the refrigerant saturation temperature T _{sat_in at} the inlet of the evaporator 4 is.

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 the refrigerant and the pressure loss ΔP (Pa) of the evaporator 4. In particular, 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 greater the pressure loss of the evaporator 4, the larger the difference between the refrigerant saturation temperature ΔT and the inlet and the outlet of the evaporator 4. _sat increases. Although the detailed relation 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, the detailed evaporator is determined once the type of refrigerant is determined. equation 4 of the pressure loss ΔP between the refrigerant saturation temperature difference [Delta] T _sat an inlet and an outlet of the evaporator 4 also uniquely determined, the evaporator refrigerant saturation temperature difference [Delta] T _sat an inlet and an outlet of the evaporator 4 as in equation 9 4 can be expressed as a function T _sat (ΔP) of the pressure loss ΔP. Further, no matter what kind of refrigerant is used, the relation that 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 Expression 10. Here, the symbols used in Expression 10 are defined. Φ is an average correction coefficient of the two-phase flow pressure loss and is a dimensionless number. The average correction coefficient Φ of the two-phase flow pressure loss can be generally derived by the Lockhart-Martinelli calculation method. λ is the coefficient of friction loss of the refrigerant, which is a dimensionless number. Further, the average correction coefficient Φ of the two-phase flow pressure loss is a value determined once the type of the refrigerant to be used is determined, and is therefore treated as a constant. In addition, regarding L, d, ρ, and u in Equation 10, the refrigerant flow path length L (m), the refrigerant flow path diameter d (m), and the average density ρ of the refrigerant (kg / m ³ ) are defined above. And the average flow velocity u (m / s), so the definition here is omitted.

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

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

If the type of the refrigerant is determined as described above, the detailed relation 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 is also uniquely determined. Are all 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 Expression 14 when the type of the 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. There is a relationship that λ increases. From these two relationships, as the value of the average flow velocity u increases, 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.

Then, when Equation 8 is substituted for the refrigerant saturation temperature T _{sat_in} at the inlet of the evaporator 4 in Equation 2, and Equation 14 is substituted for the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4, Equation 15 is obtained. In Equation 15, since all symbols other than the average flow velocity u are treated as constants, the first and second terms of Equation 15 are constants. The third term and the fourth term in Equation 15 each take different values depending on the value of the average flow velocity u, and as shown in Equation 16, the average flow velocity which derives the amount of temperature decrease in the evaporator 4 due to the average flow velocity u. _Equation (17) can be obtained by substituting Equation (16) into Equation (15). From Equation 17, it can be seen that the function T _{sat_u} (u) shown in Equation 16 should be minimized in order to maximize the refrigerant saturation temperature T _{sat_out} at the outlet of the evaporator 4. Further, from Expression 16, it can be seen that the average flow velocity u at which the refrigerant saturation temperature at the outlet of the evaporator 4 becomes the maximum value T _{sat_out_max} is the average flow velocity u at which the function T _{sat_u} (u) becomes the minimum value T _{sat_u_min} . For this reason, in the refrigeration cycle apparatus 100 of the first embodiment, the refrigerant flows through the evaporator 4 at an average flow velocity u such that the function T _{sat —} u (u) becomes 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. The first term of Expression 16 is inversely proportional to the average flow velocity u raised to the 0.8th power, and it is understood that the average flow velocity u increases as the average flow velocity u decreases. Further, as described above, 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 indicates that the average flow velocity u is large. It can be seen that the larger the average flow velocity u, the larger the flow velocity. Therefore, the function T _{sat —} u (u) in Equation 16 is an addition of a term that increases as the average flow velocity u _decreases and a term that increases as the average flow velocity u increases. It can be _seen that there is a value T _{sat —} u — min and that the function T _{sat —} u (u) can be minimized. In addition, when the average flow velocity u is 0, the first term of Expression 16 diverges to infinity. _Therefore , the minimum value T _{sat_u_min} of the function T _{sat —} u (u) exists when the average flow velocity u is larger than 0. You can see that

As described above, in the refrigeration cycle apparatus 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) becomes the minimum value T _{sat —} u — min. The saturation temperature can be maximized, and as a result, the operation performance of the refrigeration cycle device becomes maximum, and a decrease in the operation performance of the entire refrigeration cycle device can be prevented. Furthermore, low-temperature equipment having a low evaporation temperature, hot water supply / heating operation in cold regions, and refrigerant having a low operating pressure are particularly effective because the suction pressure of the compressor has a large effect on the operation performance.

In order to prevent a decrease in the operation performance of the entire refrigeration cycle apparatus 100, the average flow velocity u does not necessarily need to be such that the function T _{sat —} u (u) becomes the minimum value T _{sat —} u — min. The average flow velocity u may have a range to such an extent that the operation performance is not extremely reduced. For example, it is known that, in a generally used refrigerant, when the refrigerant saturation temperature T _{sat_out at} the outlet of the evaporator 4 is reduced by 0.3 ° C. or more, the operating performance is reduced by 1%, which has a considerable 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 Expression 16 is less than T _{sat —} u — _min +0.3 (° C.), the refrigeration cycle apparatus The effect of preventing a decrease in the driving performance of the whole 100 is exhibited.

  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 Expression 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. Incidentally, [rho formula 18, _{A i} is the mean density of the refrigerant as defined above ρ (kg / ^{m 3),} the tube-side heat transfer area of the evaporator for a _A i ^{(m 2),} here I omit the definition.

  From Expression 18, it can be seen that the average flow rate u of the refrigerant flowing through the evaporator 4 can be controlled by controlling the mass flow rate Gr of the refrigerant. That is, means for controlling the average flow velocity u include means for controlling the number of revolutions of the compressor 1, controlling the opening of the expansion valve 3, and the like. 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 reduced by decreasing the rotation speed. In the expansion valve 3, the average flow velocity u can be increased by lowering the opening degree, and the average flow velocity u can be reduced by increasing the opening degree.

  In addition, as a method of obtaining the average flow velocity u of the refrigerant flowing through the evaporator 4, a method of directly measuring the average flow velocity u of the refrigerant flowing through the evaporator 4 using a flow velocity sensor can be used. Furthermore, since the mass flow rate Gr of the refrigerant and the average flow rate u of the refrigerant are proportional to each other according to Expression 18, the mass flow rate Gr flowing through the evaporator 4 measured using the flow meter or the set value of the rotation speed of the compressor is different from the mass flow rate Gr. The average flow velocity u of the refrigerant can be obtained based on two set values, that is, the set value of the opening degree of the expansion valve.

The control device calculates, 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 function T _{sat —} u (u) shown in Expression 16 becomes the minimum value T _{sat —} u — min. By comparison, the average flow velocity u may be controlled so that the measured value approaches 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 determines in advance the rotation speed of the compressor or the opening degree of the evaporator at the start of the operation of the refrigeration cycle device 100. If the temperature is controlled in such a range that 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 the _bisection method, the compressor is operated at the rotation speed of the compressor or the opening degree of the evaporator. Good.

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

Embodiment 2. FIG.
The refrigeration cycle apparatus 100 according to the second embodiment is different from the refrigeration cycle apparatus 100 according to the first embodiment in that the composition of the refrigerant circulating in the refrigeration cycle apparatus 100 is limited. The other configuration of the refrigeration cycle device 100 itself is the same as that of the first embodiment, and thus the description is omitted.

The refrigerant circulating in the refrigeration cycle apparatus 100 according to the second embodiment is a mixed refrigerant of R32, R1234yf, and R125. The ratio XR32 of _R32 is 67 (wt%), and the ratio X _{R1234yf of R1234yf} is 26 (wt%). ), The ratio of _R125 X _R125 is 7 (wt%). By using such a refrigerant, the following effects can be obtained.

  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, since the operating pressure is high, the effect of performance degradation due to pressure loss can be reduced, and the refrigeration capacity can be improved even at a low temperature such as a supermarket showcase. In addition, R1234yf has a global warming potential of 0, and thus can reduce environmental impact. Since the physical properties of R125 are nonflammable, the flammability of R32 and R1234yf can be reduced and safety can be improved. Therefore, the above-mentioned mixed refrigerant has little effect on the global environment and can simultaneously improve safety and performance.

  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. When a non-azeotropic mixed refrigerant is used, the temperature of the outlet side of the evaporator 4 becomes higher than that of the inlet side of the evaporator 4. In particular, since the refrigerant mixture of R32 improves the refrigerating capacity even at low temperatures, the refrigeration cycle apparatus 100 is often used for low-temperature equipment such as a showcase having a saturation temperature below the freezing point. When the evaporator 4 is used in a low temperature device, frost is formed on the evaporator 4. If a refrigerant having a temperature gradient is used in the evaporation process for a device in which frost occurs, cooling of the air is promoted at the inlet side of the evaporator 4 where the temperature of the refrigerant is low, so that the amount of frost is large and the temperature of the refrigerant is high. At the outlet side of the vessel 4, the cooling of the air does not proceed and the amount of frost is reduced, and the frost formation of the evaporator 4 is biased. When the frost is unevenly distributed, the space between the fins 42 on the inlet side of the evaporator 4 is clogged by frost at an early stage, even if the entire amount of frost is the same as compared to the case where the frost is uniform. It occurs and heat exchange becomes impossible.

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. The temperature 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 reduced, a pressure difference is generated between the inlet and the outlet of the evaporator 4. The difference is uniquely determined according to the composition of the non-azeotropic refrigerant mixture. For this reason, the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 is made substantially the same as 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. Thus, the uneven formation of frost on the evaporator 4 can be suppressed.

However, depending on the type of the refrigerant, the difference between the refrigerant saturation temperature at the inlet and the outlet of the evaporator 4 when the temperature of the refrigerant flowing through the evaporator 4 is made uniform by the difference between the refrigerant saturation temperatures ΔT _sat at the inlet and the outlet of the evaporator 4 If the functions are substantially the same, the function T _{sat —} u (u) shown in Expression 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 Expression 16 becomes T _Even if it is less than _{sat_u_min} + 0.3 (° C.), the numerical value such as the refrigerant flow path length L (m) or the refrigerant flow path diameter d (m) of the evaporator 4 becomes an unrealistic value, and as an actual refrigeration cycle apparatus May not hold.

Here, in the refrigeration cycle apparatus 100 of the second embodiment, the difference between the refrigerant saturation temperature ΔT _sat at the inlet and the outlet of the evaporator 4 is determined by the difference between 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 is substantially the same as the refrigerant saturation temperature difference at the outlet will be described. In the following description, the values of the physical properties of the mixed refrigerant and the calculation results of the function T _sat (u) are shown as graphs. These values are represented by software “Refprop” created by NIST (National Institute of Standards and Technology). Ver. The simulation was performed using 9.1 and the mixing rule that is standardly installed in this software.

Next, among the symbols described in Expressions 2 to 17, symbols that depend on the refrigerant will be described with specific numerical values when the above-described mixed refrigerant is used. First, the average density ρ (kg / m ³ ) of the refrigerant is such that 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 1126.8 (kg / m ³ ). ³ ), it is 572.2 (kg / m ³ ) from the average value of both. The average viscosity μ (Pa · s) of the refrigerant is such that 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 ⁻⁴ (Pa · s), which is about 9.2 × 10 ⁻⁵ (Pa · s) from the average value of both. The average thermal conductivity k (W / mK) of the refrigerant is such that 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. (W / mK), which 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 Equations 2 to 17, symbols that depend on the structure of the refrigeration cycle apparatus 100 will be described with specific numerical values. 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 heat transfer area A _o (m ² ) is 29.4 (m ² ), the heat transfer area inside the tube A _i (m ² ) is 4.8 (m ² ), and the heat exchange amount Q (W) is 20,000 (W). Note that these numerical values are realistic values when configuring the refrigeration cycle apparatus 100, and are realized as an actual refrigeration cycle apparatus.

Further, among the symbols described in Equations 2 to 17, symbols that depend on the object to be cooled will be described with specific numerical values. T _air is the temperature T _air (° C.) of the object to be cooled by the evaporator 4 is 0.7 (° C.), and the heat transfer coefficient α _o (W / m ² K) of the object to be cooled is 100 (W / m ² K). ).

Further, in the above-described mixed refrigerant, a difference in refrigerant saturation temperature 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. For this reason, 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., an increase in the temperature during 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 Embodiment 2, 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). As described above, as the value of the average flow velocity u increases, the value of the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 increases. It turns out that they are relationships. Also, 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). For this reason, in the refrigeration cycle apparatus 100 according to 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 uniformed And the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 becomes equal, and the temperature of the refrigerant flowing through the evaporator 4 can be made uniform.

The function T _sat (u) is derived from the relationship between the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 as shown in Expressions 9 to 14 from the pressure loss ΔP of the evaporator 4. The 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. For this reason, the pressure change 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 optimal method is to change the average flow velocity u according to the opening degree of the valve 3 and the number of branches of the refrigerant flow path to change the pressure loss.

FIG. 5 is a graph of a temperature difference between a cooling target and a refrigerant due to heat transfer performance of the refrigeration cycle apparatus according to Embodiment 2, and an average flow velocity u of the refrigerant flowing through the evaporator. From FIG. 5, it can be seen that the smaller the value of the average flow velocity u, the larger the value of the temperature difference generated between the cooling target and the refrigerant due to the heat transfer performance. The reason will be described below. First, the temperature difference generated between the cooling target and the refrigerant due to the heat transfer performance of the refrigeration cycle device is expressed in other words as a difference T _air -T _{sat_in} between the temperature T _{air of the} cooling target and the refrigerant saturation temperature T _{sat_in} at the inlet of the evaporator 4. . Also, by transforming Expression 8, Expression 19 can be derived. _T air _{-T sat_in} from Equation 19 is inversely proportional to the 0.8 power of the average flow velocity u, _T air _{-T sat_in} as the value of the average flow velocity u is small it can be seen that the increase.

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 maximum value T _{sat_out_max} exists at the refrigerant saturation temperature T _{sat_out} at the outlet of the evaporator 4. From the relationship between Expressions 15 and 17, the average flow speed 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} is the average flow speed u at which the function T _{sat_u} (u) becomes the minimum value T _{sat_u_min.} But also.

According to 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 at which 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 through the evaporator 4 is made uniform. And the average flow velocity u at which the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 becomes equal. For this reason, the refrigeration cycle apparatus 100 according to Embodiment 2 makes the temperature of the refrigerant flowing through the evaporator 4 uniform to prevent uneven frost formation, and at the same time, reliably maximizes the operation performance of the entire refrigeration cycle apparatus. Can be.

  As described above, in the case where the refrigerant circulating in the refrigeration cycle apparatus is a mixed refrigerant having a composition of 67% by weight of R32, 26% by weight of R1234yf, and 7% by weight of R125 as in Embodiment 2, an actual refrigeration cycle is used. As long as the device can be realized, the temperature of the refrigerant flowing through the evaporator 4 can be made uniform to suppress frost formation, and at the same time, the operating performance of the entire refrigeration cycle device can be reliably maximized.

In order to prevent a decrease in the operating performance of the entire refrigeration cycle apparatus 100 and to suppress the uneven formation of frost on the evaporator 4, the average flow velocity u must be set such that the refrigerant saturation temperature T _{sat_out at} the outlet of the evaporator 4 is the maximum value. 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 are equal. There is no. For the same reason as in the first embodiment, if the refrigerant flows through the evaporator at an average flow rate u such that the function T _{sat —} u (u) shown in Equation 16 is less than T _{sat —} u — _min +0.3 (° C.), The effect of preventing a decrease in the operation performance of the entire cycle device 100 is exerted, and the uneven formation of frost on the evaporator 4 can also be suppressed. Under the conditions of the second embodiment, the average flow velocity u that function _{T sat_u} (u) is the _{T sat_u_min} +0.3 below (° C.), i.e. the refrigerant saturation temperature _{T Sat_out} the outlet of the evaporator 4 is greater than -10.3 ° C. As shown in FIG. 6, 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). When the average flow velocity u satisfies the condition of 0.85 (m / s) <u <1.6 (m / s), the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 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 that satisfies the condition is at least when the average flow velocity u is 1.6 (m / s) or more. As a result, the difference between 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 through the evaporator 4 is made uniform is reduced. It can be seen that the bias of frost formation on the vessel 4 can be suppressed.

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 _{reaches the} maximum value T ^sat_out_max and the temperature of the refrigerant flowing through the evaporator 4 are made uniform. The average flow velocity u at which 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 become substantially equal, but is not limited to this. . 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 4 becomes equal to the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator. If the indicated function T _{sat —} u (u) is less than T _{sat —} u — _min +0.3 (° C.), the effect of preventing a decrease in the operation performance of the entire refrigeration cycle apparatus is exhibited, and the unevenness of frost formation on the evaporator is also suppressed. it can.

Further, in the second embodiment, a refrigerant mixture of R32, R1234yf, and R125, the ratio XR32 of _R32 is 67 (wt%), the ratio XR1234yf of _R1234yf is 26 (wt%), and the ratio _XR125 of _R125. Is 7 (wt%), but is not limited thereto, and is a non-azeotropic mixed refrigerant, and the inlet and outlet of the evaporator 4 when the temperature of the refrigerant flowing through the evaporator 4 is made uniform. long as the refrigerant saturation temperature difference between the evaporator inlet and the refrigerant saturation temperature difference [Delta] T _sat outlet equals the average function shown in equation 16 in flow velocity _{u T} sat_u (u) becomes less than _{T sat_u_min} +0.3 (℃) For example, an effect of preventing a decrease in the operation performance of the entire refrigeration cycle device is exerted, and an effect of suppressing uneven frost formation of the evaporator is exhibited.

In particular, the refrigerants in which the respective compositions of the mixed refrigerants of the second embodiment differ in a range of less than ± 3 wt%, that is, the condition that the ratio X _R32 (wt%) of _R32 is 64 <X _R32 <70, and the ratio X of R1234yf and conditions _R1234yf (wt%) _{23 <a X R1234yf} <29, and conditions the proportion of R125 _X R125 (wt%) is _{4 <X R125 <10,} the sum of _{X R32} and _{X R1234yf} and _{X R125} 100 The refrigerant satisfying all of the following conditions is a mixed refrigerant in which the ratio R32 of _R32 is 67 (wt%), the ratio X1234yf of _R1234yf is 26 (wt%), and the ratio XR125 of _R125 is 7 (wt%). As compared with the above, the various physical properties, the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 and the function T _sat (u) are not significantly changed. As shown in FIG. 4, the refrigerant saturation temperature difference ΔT _sat between the inlet and the outlet of the evaporator 4 satisfies the condition of 0.7 (° C.) <ΔT _sat <2.1 (° C.). It has the effect of preventing the reduction of the frost and the effect of suppressing the uneven frost formation of the evaporator.

  DESCRIPTION OF SYMBOLS 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 tube, 41a flow path, 41b joint, 41c joint 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, and an evaporator, including a refrigeration cycle in which a refrigerant circulates,
The refrigerant, the R32, and R1234yf, a mixed refrigerant of R125, and conditions the ratio of R32 _X R32 (wt%) is _{64 <X} R32 _<70, the proportion _{X R1234yf} (wt%) of the R1234yf is 23 _<a condition is _{X R1234yf} <29, the proportion _X R125 (wt%) of the R125 is _{4 <and} conditions are _{X R125} <10, the ratio _{X R32} of the R32 and the ratio _{X R1234yf} of the R1234yf of the R125 A condition that the sum of the proportions X _R125 is 100, and
A refrigeration cycle apparatus that satisfies the condition of 0.7 <ΔT _sat <2.1, wherein a difference between the refrigerant saturation temperatures ΔT _sat (° C.) between the inlet and the outlet of the evaporator. In the refrigerant, the ratio X _{R32 of the R32} is 67 (wt%), the ratio XR1234yf of the _R1234yf is 26 (wt%), and the ratio XR125 of the _R125 is 7 (wt%),
2. The refrigeration cycle apparatus according to claim 1, wherein the refrigerant saturation temperature difference ΔT _sat (° C.) between the inlet and the outlet of the evaporator is 1.3 ° C. 3.

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