WO2017149642A1 - Dispositif à cycle frigorifique - Google Patents

Dispositif à cycle frigorifique Download PDF

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Publication number
WO2017149642A1
WO2017149642A1 PCT/JP2016/056199 JP2016056199W WO2017149642A1 WO 2017149642 A1 WO2017149642 A1 WO 2017149642A1 JP 2016056199 W JP2016056199 W JP 2016056199W WO 2017149642 A1 WO2017149642 A1 WO 2017149642A1
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Prior art keywords
refrigerant
evaporator
sat
refrigeration cycle
outlet
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PCT/JP2016/056199
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English (en)
Japanese (ja)
Inventor
智隆 石川
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三菱電機株式会社
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Application filed by 三菱電機株式会社 filed Critical 三菱電機株式会社
Priority to CN201680082443.7A priority Critical patent/CN108700340B/zh
Priority to PCT/JP2016/056199 priority patent/WO2017149642A1/fr
Priority to JP2018502900A priority patent/JP6715918B2/ja
Publication of WO2017149642A1 publication Critical patent/WO2017149642A1/fr

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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F25—REFRIGERATION OR COOLING; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS; MANUFACTURE OR STORAGE OF ICE; LIQUEFACTION SOLIDIFICATION OF GASES
    • F25B—REFRIGERATION MACHINES, PLANTS OR SYSTEMS; COMBINED HEATING AND REFRIGERATION SYSTEMS; HEAT PUMP SYSTEMS
    • F25B1/00—Compression machines, plants or systems with non-reversible cycle

Definitions

  • the present invention relates to a refrigeration cycle apparatus used for applications such as air conditioning and hot water supply.
  • Patent Document 1 an evaporator, a compressor (corresponding to the compressor of the present invention), a condenser, and a capillary tube (corresponding to the decompression device of the present invention) through which the refrigerant circulates.
  • a compressor corresponding to the compressor of the present invention
  • a condenser corresponding to the condenser
  • a capillary tube corresponding to the decompression device of the present invention
  • a non-azeotropic refrigerant mixture in which a plurality of types of refrigerants are mixed may be used as a refrigerant used in the refrigeration cycle apparatus.
  • the temperature of the non-azeotropic refrigerant mixture changes due to a phase change under the same pressure, and the temperature on the downstream side becomes higher than that on the upstream side in the evaporation process. The temperature becomes higher than the inlet side of the evaporator.
  • the cooling performance of an evaporator improves, so that the refrigerant
  • Patent Document 1 the pressure loss in the pipeline in the evaporation process is increased, and the pressure in the pipeline in the evaporation process is gradually reduced so that the temperature of the refrigerant in the evaporator becomes constant.
  • An refrigeration cycle apparatus is described. In such a refrigeration cycle apparatus, an increase in temperature during the evaporation process is suppressed, and both improvement in cooling performance and prevention of frost formation can be achieved.
  • the present invention has been made in view of the above-described problems, and an object of the present invention is to provide a refrigeration cycle apparatus that prevents a decrease in operating performance of the entire refrigeration cycle apparatus.
  • the refrigeration cycle apparatus 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.
  • the amount of heat exchange of the evaporator is Q (W) and evaporation
  • the refrigerant flow path diameter of the evaporator is d (m)
  • the heat transfer area inside the pipe 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 ⁇ (kg / kg).
  • the average viscosity of the refrigerant is ⁇ (Pa ⁇ s)
  • the Prandtl number of the refrigerant is Pr
  • the average flow velocity of the refrigerant flowing through the evaporator is u (m / s)
  • the refrigerant flowing through the evaporator is determined by the type of refrigerant.
  • T sat (u) is a function that derives the refrigerant saturation temperature difference between the inlet and outlet of the evaporator from the average flow velocity u of Assuming that the minimum value of the function T sat_u (u) represented by the equation is T sat_u_min , the refrigerant flows through the evaporator at an average flow rate u at which the function T sat_u (u) is less than T sat_u_min +0.3 (° C.). It is characterized by.
  • FIG. 3 is a schematic diagram illustrating a configuration of an evaporator of the refrigeration cycle apparatus 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. It is a graph of the refrigerant
  • DELTA coolant saturation temperature difference
  • FIG. FIG. 1 is a schematic diagram illustrating a configuration of the entire refrigeration cycle apparatus according to Embodiment 1.
  • FIG. FIG. 2 is a schematic diagram illustrating the configuration of the evaporator of the refrigeration cycle apparatus according to Embodiment 1.
  • 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.
  • symbol is attached
  • the refrigeration cycle apparatus 100 according to Embodiment 1 includes a compressor 1, a condenser 2, an expansion valve 3, and an evaporator 4 connected by a refrigerant pipe 5.
  • a refrigeration cycle in which the refrigerant circulates in the order of the unit 2, the expansion valve 3, the evaporator 4, and the compressor 1 is configured.
  • the refrigerant pipe 5 a is connected to the compressor 1 and the condenser 2
  • the refrigerant pipe 5 b is connected to the condenser 2 and the expansion valve 3, and the expansion valve 3 and the evaporator 4 are connected.
  • What is connected is referred to as a refrigerant pipe 5c
  • what is connected to the evaporator 4 and the compressor 1 is referred to as a refrigerant pipe 5d.
  • coolant which circulates through the refrigerating cycle apparatus 100 which concerns on Embodiment 1 is although it does not specifically limit, According to the use etc. of the refrigerating cycle apparatus 100, it determines as arbitrary one refrigerant
  • Compressor 1 sucks in refrigerant, compresses it, and discharges it in a high-temperature and high-pressure gas state.
  • the compressor 1 may be configured such that the rotation speed is controlled by, for example, an inverter circuit and the discharge amount of the refrigerant can be adjusted by controlling the rotation speed.
  • the condenser 2 flows into the high-temperature and high-pressure gas state that is compressed by the compressor 1 and performs heat exchange between the refrigerant and the heat source to cool the refrigerant to a low-temperature and high-pressure liquid state.
  • the heat source include air, water, brine, and the like.
  • the heat source of the condenser 2 is outside air that is outdoor air, and the condenser 2 performs heat exchange between the outside air and the refrigerant.
  • the condenser blower 6 that blows outside air to the condenser 2 when the refrigerant circulates in the refrigeration cycle apparatus 100 is provided.
  • the condenser blower 6 may be configured with an air flow rate adjustable.
  • the expansion valve 3 flows in the 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 rate control means such as an electronic expansion valve and a temperature-sensitive expansion valve, a capillary tube (capillary tube), and the like.
  • the expansion valve 3 corresponds to the decompression device of the present invention.
  • a low-temperature and low-pressure liquid refrigerant decompressed and expanded by the expansion valve 3 flows in, performs heat exchange between the refrigerant and the object to be cooled, and absorbs the heat of the object to be cooled by the refrigerant. Cool down. When the object to be cooled is cooled, the refrigerant evaporates into a high temperature and low pressure gas state.
  • the air to be cooled is indoor air, and the evaporator 4 performs heat exchange between the indoor air and the refrigerant.
  • an evaporator blower 7 that blows indoor air to the evaporator 4 when the refrigerant circulates in the refrigeration cycle apparatus 100 is provided.
  • the evaporator blower 7 may be configured with an air volume adjustable.
  • the evaporator 4 is a plate fin tube heat exchanger that includes a plurality of heat transfer tubes 41, a plurality of fins 42, a refrigerant distributor 43, and a header 44.
  • the number of heat transfer tubes 41 is five and the number of fins 42 is 28.
  • these numbers and numbers are merely examples, and the present invention is not limited to these numbers and numbers.
  • the heat transfer tube 41 is a tube constituting a flow path through which a refrigerant flows, and a metal having high thermal conductivity such as aluminum or copper is used.
  • 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 Embodiment 1 has a circular cross section as shown in FIG.
  • the fin 42 is a thin metal plate having high thermal conductivity such as aluminum or copper.
  • the plurality of fins 42 are arranged perpendicularly to the axial direction of the heat transfer tube 41 (the direction of the arrow X in FIG. 2) with a predetermined interval.
  • the heat transfer tubes 41 and the fins 42 are joined by a method such as brazing, so that heat is transferred from the heat transfer tubes 41 to the fins 42.
  • the refrigerant distributor 43 includes a single inflow port and a plurality of outflow ports, and is a device that distributes the refrigerant flowing in from the inflow port to the plurality of outflow ports and flows it out.
  • the inlet of the refrigerant distributor 43 is connected to the expansion valve 3 via the refrigerant pipe 5 c, 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 inlets and one outlet, and is a device that causes the refrigerant flowing in from the inlet to join and flow out from the one outlet.
  • 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.
  • the refrigerant flow in the evaporator 4 will be described.
  • the low-temperature and low-pressure liquid refrigerant decompressed and expanded by the expansion valve 3 flows 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 that has flowed into the heat transfer tube 41 flows along the axial direction of the heat transfer tube 41 (the arrow X direction in FIG. 2).
  • the surfaces of the heat transfer tubes 41 and the fins 42 are supplied with indoor air to be cooled 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. To absorb the heat of indoor air.
  • the refrigerant that has exchanged heat with indoor air in the heat transfer pipe 41 flows in from the inlet of the header 44, merges in the header 44, and flows from the outlet of the header 44 to the compressor 1.
  • the refrigerant flow path length L (m), the refrigerant flow path diameter d (m), the tube outer heat transfer area A o (m 2 ), and the tube inner heat transfer area A of the evaporator 4 used when explaining the present invention Define i (m 2 ).
  • the content in parentheses after each symbol indicates the unit of the symbol.
  • the refrigerant flow path length L is changed from a joint 41 b between the heat transfer tube 41 and the fin 42 a located on the most upstream side to a joint 41 c between the heat transfer tube 41 and the fin 42 b located on the most downstream side. Is the length.
  • the refrigerant 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.
  • the operation performance of the refrigeration cycle apparatus 100 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 operation performance.
  • the present invention focuses on increasing the suction pressure of the compressor 1.
  • Raising the suction pressure of the compressor 1 is, in other words, raising the outlet pressure of the evaporator 4.
  • the refrigerant pressure can be indicated by the refrigerant saturation temperature, and in the following description, the refrigerant pressure is indicated by the refrigerant saturation temperature. If a refrigerant whose temperature changes due to a phase change under the same pressure, such as a non-azeotropic 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. The temperature difference is the refrigerant saturation temperature difference in the pressure difference at the same enthalpy.
  • the refrigerant saturation temperature T sat_out at the outlet of the evaporator 4 is obtained by subtracting the refrigerant saturation temperature difference ⁇ T 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 Equation 2. .
  • the refrigerant saturation temperature T sat_in at the inlet of the evaporator 4 will be described.
  • the heat exchange amount Q (W) of the evaporator 4 can be expressed as Equation 3 using the refrigerant saturation temperature T sat_in at the inlet of the evaporator 4.
  • Equation 3 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 based on Equation 4 below. Can be derived.
  • symbols used in Equation 3 and Equation 4 are defined.
  • T air is the temperature (° C.) to be cooled by the evaporator 4, and corresponds to the temperature of indoor air in the first embodiment.
  • ⁇ o is a heat transfer coefficient (W / m 2 K) to be cooled by the evaporator 4, and corresponds to the heat transfer coefficient of indoor air in the first embodiment.
  • ⁇ i is the heat transfer coefficient (W / m 2 K) of the refrigerant.
  • the temperature T air to be cooled, the heat exchange amount Q of the evaporator 4 and the heat transfer coefficient ⁇ o of the cooling object are values determined if the state of the cooling object 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 The definition of is omitted.
  • Equation 5 The heat transfer coefficient ⁇ i of the refrigerant can be derived based on Equation 5 below.
  • 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 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.
  • Equation 5 is the refrigerant flow path diameter d (m) defined above, the definition here is omitted.
  • the Reynolds number Re can be derived based on Equation 6 below.
  • ⁇ 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 an 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.
  • ⁇ 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 ⁇ of the refrigerant and the average viscosity ⁇ of the refrigerant are values determined when the type of the refrigerant is determined, they are treated as constants in the present invention. In addition, the average flow velocity u is a value that can be easily changed by adjusting the number of rotations of the compressor 1 or the opening of the expansion valve 3, and is therefore treated as a variable in the present invention. Since d in Equation 6 is the refrigerant flow path diameter d (m) defined above, the definition here is omitted.
  • the refrigerant saturation temperature difference ⁇ T sat between the inlet and outlet of the evaporator 4 is determined by the type of refrigerant and the pressure loss ⁇ P (Pa) of the evaporator 4.
  • the refrigerant saturation temperature difference ⁇ T between the inlet and outlet of the evaporator 4 increases as the pressure loss of the evaporator 4 increases. The sat increases.
  • the detailed relation between the pressure loss ⁇ P of the evaporator 4 and the refrigerant saturation temperature difference ⁇ T sat between the inlet and outlet of the evaporator 4 varies depending on the type of refrigerant, but if the type of refrigerant is determined, the detailed evaporator 4 and the refrigerant saturation temperature difference ⁇ T sat between the inlet and outlet of the evaporator 4 are also uniquely determined, and the refrigerant saturation temperature difference ⁇ T sat between the inlet and outlet of the evaporator 4 is expressed by the following equation (9). 4 as a function T sat ( ⁇ P) of the pressure loss ⁇ P. Further, no matter what refrigerant is used, the relationship in which the refrigerant saturation temperature difference ⁇ T sat between the inlet and outlet of the evaporator 4 increases as the pressure loss ⁇ P of the evaporator 4 increases.
  • the pressure loss ⁇ P of the evaporator 4 can be derived based on Equation 10 below.
  • ⁇ is an average correction factor for two-phase flow pressure loss and is a dimensionless number.
  • the average correction coefficient ⁇ of the two-phase flow pressure loss can be generally derived by the Lockhart-Martinelli calculation method.
  • ⁇ is a friction loss coefficient of the refrigerant and is a dimensionless number.
  • the average correction coefficient ⁇ of the two-phase flow pressure loss is a value that is determined when the type of refrigerant to be used is determined, and thus is treated as a constant.
  • Equation 10 the refrigerant flow path length L (m) defined above, the refrigerant flow path diameter d (m), and the average density ⁇ (kg / m 3 ) of the refrigerant. Since this is the average flow velocity u (m / s), the definition here is omitted.
  • the friction loss coefficient ⁇ of the refrigerant can be derived based on the following formula 11.
  • Re in Expression 11 is the Reynolds number
  • Expression 12 is obtained by substituting Expression 6 into the Reynolds number Re in Expression 11.
  • Equation 12 Substituting Equation 12 into the friction loss coefficient ⁇ of the refrigerant in Equation 10, the following Equation 13 is obtained.
  • symbols other than the average flow velocity u in Equation 13 are all treated as constants.
  • the friction loss coefficient ⁇ of the refrigerant is proportional to the average flow velocity u to the 1.75th power. It can be seen that the friction loss coefficient ⁇ increases.
  • the refrigerant saturation temperature difference ⁇ T sat between the inlet and outlet of the evaporator 4 can be expressed as a function T sat (u) of the average flow velocity u as shown in Equation 14. it can.
  • Equation 13 there is a relationship in which the refrigerant saturation temperature difference ⁇ T sat between the inlet and outlet of the evaporator 4 increases as the pressure loss of the evaporator 4 increases, and the friction loss coefficient of the refrigerant increases as the average flow velocity u increases as shown in Equation 13.
  • ⁇ has a relationship of increasing. From these two relationships, the value of the refrigerant saturation temperature difference ⁇ T sat between the inlet and 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.
  • Expression 15 is obtained by substituting Expression 8 into the refrigerant saturation temperature T sat_in at the inlet of the evaporator 4 and Expression 14 into the refrigerant saturation temperature difference ⁇ T sat between the inlet and outlet of the evaporator 4 in Expression 2.
  • the first and second terms of Expression 15 are constants.
  • the third term and the fourth term of Equation 15 are different depending on the value of the average flow velocity u, and as shown in Equation 16, the average flow velocity for deriving the amount of temperature decrease in the evaporator 4 caused by the average flow velocity u.
  • Equation 17 It can be expressed as a function T sat — u (u) of u, and when Expression 16 is substituted into Expression 15, Expression 17 is obtained. From Equation 17, it can be seen that in order to maximize the refrigerant saturation temperature T sat — out at the outlet of the evaporator 4, the function T sat — u (u) shown in Equation 16 should be minimized. 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 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 . Therefore, in the refrigeration cycle device 100 of the first embodiment, the refrigerant, the average flow velocity u as a function T sat_u (u) is a minimum value T Sat_u_min, flowing through the evaporator 4.
  • the function T sat — u (u) shown in Expression 16 can be minimized.
  • the first term of Equation 16 is inversely proportional to the average flow velocity u to the 0.8th power, and increases as the average flow velocity u decreases.
  • the second term of Expression 16 is that, as described above, the value of the function T sat (u) indicating the relationship between the refrigerant saturation temperature difference ⁇ T sat at the inlet and the outlet of the evaporator 4 and the average flow velocity u is increased by the average flow velocity u. It can be seen that the larger the average flow velocity u is, the larger it is.
  • the function T sat — u (u) in Expression 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. Therefore , the function T sat — u (u) is the minimum. It can be seen that the value T sat_u_min exists and the function T sat_u (u) can be minimized. Further, when the average flow velocity u is 0, the first term of Formula 16 diverges infinitely, so 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. I understand that
  • 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.
  • the saturation temperature can be maximized.
  • the operation performance of the refrigeration cycle apparatus is maximized, and a decrease in the operation performance of the entire refrigeration cycle apparatus can be prevented.
  • low temperature equipment having a low evaporation temperature, hot water supply / heating operation in a cold region, a refrigerant having a low operating pressure, and the like are particularly effective because the influence of the suction pressure of the compressor on the operating performance is great.
  • the average flow velocity u In order to prevent a decrease in the operating performance of the entire refrigeration cycle apparatus 100, it is not always necessary for the average flow velocity u to have the function T sat_u (u) be the minimum value T sat_u_min .
  • a range may be given to the average flow velocity u to such an extent that the operating performance does not extremely decrease. For example, in a commonly used refrigerant, it is known that when the refrigerant saturation temperature T sat_out at the outlet of the evaporator 4 is lowered by 0.3 ° C. or more, the operating performance is reduced by 1%, which is not negligible.
  • 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 expressed by the following mathematical formula 18.
  • 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 rotational speed of the compressor 1 and the opening degree of the expansion valve 3.
  • ⁇ and Ai in Equation 18 are defined here because the average density ⁇ (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.
  • the mass flow rate Gr of the refrigerant may be controlled in order to control the average flow velocity u of the refrigerant flowing through the evaporator 4. That is, means for controlling the average flow velocity u include means for controlling the rotational speed 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 degree, and the average flow velocity u can be decreased by increasing the opening degree.
  • 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 proportional to each other from Equation 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 refrigerant flow velocity u can be acquired based on two set values, that is, the set value of the opening of the expansion valve.
  • the control device sets the measured value of the flow rate sensor attached to the heat transfer tube 41 of the evaporator 4 and the set value of the average flow rate u at which the function T sat_u (u) expressed in advance in Equation 16 becomes the minimum value T sat_u_min.
  • the average flow velocity u may be controlled so that the measured value approaches the set value.
  • the control device predetermines the rotation speed of the compressor or the opening of the evaporator at the start of the operation of the refrigeration cycle apparatus 100.
  • the structure of the refrigeration cycle apparatus 100 of the first embodiment is an example, and the structure of the refrigeration cycle apparatus of the present invention is not limited to the structure of the first embodiment.
  • the cross section of the flow path of the heat transfer tube 41 of Embodiment 1 is circular, the cross section of the heat transfer tube is not limited to this, and may be elliptical, flat, or rectangular.
  • the equivalent diameter may be used as the refrigerant flow path diameter d (m) when the cross section of the flow path of the heat transfer tube is elliptical, flat or rectangular.
  • FIG. The refrigeration cycle apparatus 100 according to Embodiment 2 is different from the refrigeration cycle apparatus 100 according to Embodiment 1 in that the composition of the refrigerant circulating in the refrigeration cycle apparatus 100 is limited.
  • description is omitted since the structure of the other refrigerating-cycle apparatus 100 itself is the same as that of Embodiment 1, description is omitted.
  • the refrigerant circulating through the refrigeration cycle apparatus 100 according to the second embodiment is a mixed refrigerant of R32, R1234yf, and R125.
  • the R32 ratio X R32 is 67 (wt%), and the R1234yf ratio X R1234yf is 26 (wt%). ),
  • the ratio X R125 of R125 is 7 (wt%).
  • 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.
  • the physical properties of R32 since the operating pressure is high, it is possible to reduce the performance degradation effect due to pressure loss, and the refrigeration capacity can be improved even at low temperatures such as in supermarket showcases.
  • a physical property of R1234yf since a global warming potential is 0, an environmental influence can be reduced. Since the physical properties of R125 are nonflammable, the combustibility of R32 and R1234yf can be reduced and safety can be increased. Therefore, the above-described mixed refrigerant has little influence on the global environment and can simultaneously improve safety and performance.
  • 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 the upstream side in the evaporation process, and the refrigerant of the refrigeration cycle apparatus 100 If a non-azeotropic refrigerant mixture is used, the temperature on the outlet side of the evaporator 4 becomes higher than that on the inlet side of the evaporator 4. In particular, the mixed refrigerant of R32 improves the refrigerating capacity even when it is used at a low temperature.
  • the refrigeration cycle apparatus 100 is often used for a low-temperature apparatus such as a showcase where the saturation temperature is lower than freezing.
  • the saturation temperature is lower than freezing.
  • frosting occurs in the evaporator 4.
  • a refrigerant having a temperature gradient in the evaporation process is used in a device in which frost formation occurs, on the inlet side of the evaporator 4 where the refrigerant temperature is low, cooling of the air is promoted so that the amount of frost formation is large and the refrigerant temperature is high.
  • the cooling of the air does not proceed and the amount of frost formation decreases, and the frost formation of the evaporator 4 is biased.
  • frost formation is uneven, the space between the fins 42 on the inlet side of the evaporator 4 is quickly clogged by frost even if the overall frost formation amount is the same as compared with the case where the frost formation is uniform. It occurs and falls into heat exchange.
  • the non-azeotropic refrigerant mixture can make the temperature of the refrigerant flowing through the evaporator 4 uniform by continuously lowering the pressure inside the evaporator 4 with a certain predetermined gradient. It can suppress by making the temperature of the refrigerant
  • a pressure difference is generated 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 through the evaporator 4 is made uniform.
  • the difference is uniquely determined according to the composition of the non-azeotropic refrigerant mixture.
  • the refrigerant saturation temperature difference ⁇ T sat between the inlet and outlet of the evaporator 4 is made substantially the same as the refrigerant saturation temperature difference between the inlet and outlet of the evaporator 4 when the temperature of the refrigerant flowing through the evaporator 4 is made uniform. Thereby, the bias of the frost formation of the evaporator 4 can be suppressed.
  • the function T sat_u (u) shown in Expression 16 at the average flow velocity (m / s) is not less than T sat_u_min +0.3 (° C.), or the function T sat_u (u) shown in Expression 16 is T Even if it becomes less than sat_u_min + 0.3 (° C.), 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 as an actual refrigeration cycle apparatus. May not hold.
  • the refrigerant saturation temperature difference ⁇ T sat between the inlet and 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 difference is substantially the same as the refrigerant saturation temperature difference at the outlet.
  • the refrigerant physical property values of the mixed refrigerant and the graph of the calculation result of the function T sat (u) are shown. These values are the 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 standard in this software.
  • the average density ⁇ (kg / m 3 ) of the refrigerant is approximately 17.6 (kg / m 3 ) in the saturated gas state and approximately 1126.8 (kg / m) in the saturated liquid state. 3 ), it is 572.2 (kg / m 3 ) from the average value of both.
  • the average viscosity ⁇ (Pa ⁇ s) of the refrigerant is about 1.1 ⁇ 10 ⁇ 5 (Pa ⁇ s) in the saturated gas state, and about 1.72 ⁇ 10 6 in the saturated liquid state.
  • the average thermal conductivity k (W / mK) of the refrigerant is about 0.011 (W / mK) for the refrigerant in the saturated gas state, and about 0.126 for the refrigerant in the saturated liquid state. 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 ⁇ for the two-phase flow pressure loss is about 3.
  • 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)
  • the outside heat transfer area A o (m 2 ) is 29.4 (m 2 )
  • the pipe inner heat transfer area A i (m 2 ) is 4.8 (m 2 )
  • the heat exchange amount Q (W) is 20000 (W).
  • T air is 0.7 (° C.) for the temperature T air (° C.) of the cooling target cooled by the evaporator 4, and 100 (W / m 2 K) for the heat transfer coefficient ⁇ o (W / m 2 K) of the cooling target. ).
  • the above-mentioned mixed refrigerant has a refrigerant saturation temperature difference of about 1.3 ° C. between the inlet and outlet of the evaporator 4 when the temperature of the refrigerant flowing in the evaporator 4 is made uniform.
  • the refrigerant saturation temperature difference ⁇ T sat between the inlet and outlet of the evaporator 4 is 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 ⁇ T sat between the inlet and 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.
  • the graph of FIG. 4 is a graph of the function T sat (u).
  • T sat the function of the refrigerant saturation temperature difference ⁇ T sat between the inlet and outlet of the evaporator 4 increases as the average flow velocity u increases. It turns out that it is a relationship. Further, from the graph of FIG.
  • the value of the average flow velocity u at which the refrigerant saturation temperature difference ⁇ T sat between the inlet and outlet of the evaporator 4 becomes 1.3 ° C. is about 1.2 (m / s).
  • the average flow velocity u is about 1.2 (m / s)
  • the refrigerant saturation temperature difference between the inlet and the outlet of the evaporator 4 and the refrigerant saturation temperature difference ⁇ T 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.
  • the function T sat (u) is derived from the relationship between the pressure loss ⁇ P of the evaporator 4 as shown in Expressions 9 to 14 and the refrigerant saturation temperature difference ⁇ T sat between the inlet and outlet of the evaporator 4.
  • This pressure loss ⁇ P is a pressure loss due to friction between the refrigerant and the pipe, and continuously increases from the inlet to the outlet of the evaporator 4. For this reason, the change of the pressure in the evaporator 4 can be continuously lowered 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. 3 as in the first embodiment. A 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 generated between the cooling target and the refrigerant due to the heat transfer performance of the refrigeration cycle apparatus according to Embodiment 2, and the average flow velocity 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 generated between the cooling target and the refrigerant due to the heat transfer performance. The reason for this will be described below.
  • the temperature difference between the cooling target and the refrigerant due to the heat transfer performance of the refrigeration cycle apparatus is expressed in other words 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.
  • Formula 19 can be derived by transforming Formula 8. It can be seen from Equation 19 that T air ⁇ T sat —in is inversely proportional to the average flow velocity u to the 0.8th power, and that 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 velocity u of the refrigerant flowing through the evaporator.
  • 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 in the refrigerant saturation temperature T sat_out at the outlet of the evaporator 4.
  • 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 is the average flow velocity u at which the function T sat_u (u) becomes the minimum value T sat_u_min. But there 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 ⁇ 10.0 ° C.
  • 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
  • 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 operating performance of the entire refrigeration cycle apparatus. Can do.
  • the actual refrigeration cycle The temperature of the refrigerant flowing through the evaporator 4 can be made uniform within a range that can be established as an apparatus to suppress uneven frost formation, and at the same time, the operation performance of the entire refrigeration cycle apparatus can be reliably maximized.
  • 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.
  • T sat_out_max and the refrigerant saturation temperature difference between the inlet and outlet of the evaporator 4 and the refrigerant saturation temperature difference ⁇ T sat between the inlet and outlet of the evaporator 4 when the temperature of the refrigerant flowing through the evaporator 4 is uniform need to be equal.
  • 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.)
  • T sat_u_min +0.3 (° C.) the average flow rate 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 greater 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).
  • the refrigerant saturation temperature difference ⁇ T sat between the inlet and 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.) ⁇ T sat ⁇ 2.1 (° C.) is satisfied, and the refrigerant saturation temperature difference ⁇ T sat satisfying the condition satisfies at least the average flow velocity u of 1.6 (m / s) or more.
  • the difference between the refrigerant saturation temperature difference ⁇ T sat between the inlet and outlet of the evaporator 4 and the refrigerant saturation temperature difference between the inlet and 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 uneven frosting of the vessel 4 can be suppressed.
  • 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 between the refrigerant saturation temperature difference between the inlet and the outlet of the evaporator 4 and the average flow velocity u at which the refrigerant saturation temperature difference ⁇ T sat between the inlet and the outlet of the evaporator 4 are substantially equal.
  • the present invention is not limited to this. .
  • Equation 16 in the average flow velocity u at which the refrigerant saturation temperature difference between the inlet and outlet of the evaporator 4 and the refrigerant saturation temperature difference ⁇ T sat between the inlet and outlet of the evaporator when the temperature of the refrigerant flowing through the evaporator 4 is made uniform are equal. If the function T sat — u (u) shown is less than T sat — u — min +0.3 (° C.), the effect of preventing a decrease in the operating performance of the entire refrigeration cycle apparatus will be exhibited, and the uneven frosting of the evaporator will also be suppressed it can.
  • the present invention is not limited to this, and it is a non-azeotropic refrigerant, and the inlet and outlet of the evaporator 4 when the temperature of the refrigerant flowing in the evaporator 4 is made uniform.
  • T sat_u (u) shown in Equation 16 at an average flow velocity u at which the refrigerant saturation temperature difference between the refrigerant and the refrigerant saturation temperature difference ⁇ T sat at the inlet and outlet of the evaporator is equal is less than T sat_u_min +0.3 (° C.).
  • the mixed refrigerants of the second embodiment are different in composition in the 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
  • the refrigerant satisfying all the above conditions is a mixed refrigerant in which the ratio X R32 of R32 is 67 (wt%), the ratio X R1234yf of R1234yf is 26 (wt%), and the ratio XR125 of R125 is 7 (wt%).
  • the refrigerant saturation temperature difference ⁇ T sat and the function T sat (u) at the inlet and outlet of the evaporator 4 are not significantly changed. If the refrigerant saturation temperature difference ⁇ T sat between the inlet and outlet of the evaporator 4 satisfies the condition of 0.7 (° C.) ⁇ T sat ⁇ 2.1 (° C.) from FIG. 4, the operation performance of the entire refrigeration cycle apparatus The effect which prevents the fall of this is exhibited, and the effect which also suppresses the bias

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  • Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • Mechanical Engineering (AREA)
  • Thermal Sciences (AREA)
  • General Engineering & Computer Science (AREA)
  • Other Air-Conditioning Systems (AREA)
  • Devices That Are Associated With Refrigeration Equipment (AREA)
  • Air Conditioning Control Device (AREA)

Abstract

L'invention concerne un dispositif à cycle frigorifique permettant de supprimer toute diminution de performance due à une augmentation de perte de pression, et d'empêcher toute diminution de la performance d'entraînement du dispositif à cycle frigorifique dans son ensemble. Le dispositif à cycle frigorifique 100 selon la présente invention comprend un cycle frigorifique qui comprend un évaporateur 4 et à travers lequel circule un fluide frigorigène, le fluide frigorigène traversant l'évaporateur 4 à une vitesse d'écoulement moyenne u telle qu'une fonction Tsat_u (u) est inférieure à Tsat_u_min + 0,3 (°C) : où Q (W) est la quantité de chaleur échangée par l'évaporateur 4; d (m) est le diamètre du passage d'écoulement du fluide frigorigène de l'évaporateur 4; Ai (m2) est la surface de chauffage à l'intérieur du tuyau de l'évaporateur 4; k (W/mK) est la conductivité thermique moyenne du fluide frigorigène; ρ (kg/m3) est la densité moyenne du fluide frigorigène; μ (Pa⋅s) est la viscosité moyenne du fluide frigorigène; Pr est le nombre de Prandtle du fluide frigorigène; u (m/s) est la vitesse d'écoulement moyenne du fluide frigorigène traversant l'évaporateur 4; Tsat (u) est une fonction déterminée par le type de fluide frigorigène et qui est utilisée pour dériver une différence de température de saturation du fluide frigorigène entre une entrée et une sortie de l'évaporateur 4 à partir de la vitesse d'écoulement moyenne u du fluide frigorigène traversant l'évaporateur; et Tsat_u_min est la valeur minimale de la fonction Tsat_u (u) exprimée par l'expression (1).
PCT/JP2016/056199 2016-03-01 2016-03-01 Dispositif à cycle frigorifique WO2017149642A1 (fr)

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PCT/JP2016/056199 WO2017149642A1 (fr) 2016-03-01 2016-03-01 Dispositif à cycle frigorifique
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WO2019021364A1 (fr) * 2017-07-25 2019-01-31 三菱電機株式会社 Dispositif frigorifique et procédé de fonctionnement de dispositif frigorifique
WO2022004895A1 (fr) * 2020-07-03 2022-01-06 ダイキン工業株式会社 Utilisation en tant que fluide frigorigène dans un compresseur, compresseur et appareil à cycle de réfrigération

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JPH08145490A (ja) * 1994-11-17 1996-06-07 Matsushita Refrig Co Ltd ヒートポンプエアコン用熱交換器
JPH09105560A (ja) * 1995-08-04 1997-04-22 Mitsubishi Electric Corp 冷凍装置
JPH09133433A (ja) * 1995-11-07 1997-05-20 Sanyo Electric Co Ltd 熱交換器
JP2002323272A (ja) * 2001-04-24 2002-11-08 Denso Corp 蒸発器
WO2009154149A1 (fr) * 2008-06-16 2009-12-23 三菱電機株式会社 Mélange non azéotropique et dispositif à cycle de réfrigération
JP2011122819A (ja) * 2009-11-04 2011-06-23 Daikin Industries Ltd 熱交換器及びそれを備えた室内機
JP2012237543A (ja) * 2011-04-25 2012-12-06 Panasonic Corp 冷凍サイクル装置

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Publication number Priority date Publication date Assignee Title
JPH06272998A (ja) * 1993-03-18 1994-09-27 Toshiba Corp 冷凍装置
JPH08145490A (ja) * 1994-11-17 1996-06-07 Matsushita Refrig Co Ltd ヒートポンプエアコン用熱交換器
JPH09105560A (ja) * 1995-08-04 1997-04-22 Mitsubishi Electric Corp 冷凍装置
JPH09133433A (ja) * 1995-11-07 1997-05-20 Sanyo Electric Co Ltd 熱交換器
JP2002323272A (ja) * 2001-04-24 2002-11-08 Denso Corp 蒸発器
WO2009154149A1 (fr) * 2008-06-16 2009-12-23 三菱電機株式会社 Mélange non azéotropique et dispositif à cycle de réfrigération
JP2011122819A (ja) * 2009-11-04 2011-06-23 Daikin Industries Ltd 熱交換器及びそれを備えた室内機
JP2012237543A (ja) * 2011-04-25 2012-12-06 Panasonic Corp 冷凍サイクル装置

Cited By (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
WO2019021364A1 (fr) * 2017-07-25 2019-01-31 三菱電機株式会社 Dispositif frigorifique et procédé de fonctionnement de dispositif frigorifique
JPWO2019021364A1 (ja) * 2017-07-25 2020-02-27 三菱電機株式会社 冷凍装置及び冷凍装置の運転方法
WO2022004895A1 (fr) * 2020-07-03 2022-01-06 ダイキン工業株式会社 Utilisation en tant que fluide frigorigène dans un compresseur, compresseur et appareil à cycle de réfrigération
JP2022013930A (ja) * 2020-07-03 2022-01-18 ダイキン工業株式会社 圧縮機における冷媒としての使用、圧縮機、および、冷凍サイクル装置

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JP6715918B2 (ja) 2020-07-01

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