JP2013134024A - Refrigeration cycle device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a refrigeration cycle device, in which performance of a heat exchanger functioning as an evaporator can be improved.SOLUTION: The refrigeration cycle device includes a main circuit having the heat exchanger 14 which functions at least as the evaporator. The heat exchanger 14 includes a refrigerant path 4 extending from a first opening 4a on an expansion means side to a second opening 4b on a compressor side. A branch path 3 branched from the main circuit is connected midway of the refrigerant path 4. The refrigerant path 4 includes a first path 41 forming the first opening 4a and a plurality of second paths 42 divided from the first path 41, and a merging path 44 to which a downstream end of the branch path 3 is connected. An inner diameter D and a length L of a pipe configuring the first path 41 are set to satisfy two inequations on the liquid phase side and a gas phase side.

Description

  The present invention relates to a refrigeration cycle apparatus including a heat exchanger that functions as an evaporator for evaporating a refrigerant.

  Conventionally, in an air conditioner, a heat exchanger that functions as an evaporator and a condenser has been used. For example, Patent Document 1 discloses a heat exchanger 100 that functions as a condenser as shown in FIG.

  The heat exchanger 100 includes a refrigerant path 120 having an inlet 121 and an outlet 122, and fins 110 that are attached in large numbers orthogonal to the refrigerant path 120 and function as an enlarged heat transfer surface. The vicinity of the outlet 122 in the refrigerant path 120 is constituted by a relatively thin pipe (thin tube part), and the other part is constituted by a relatively thick pipe (standard part).

  A high-temperature and high-pressure gas refrigerant discharged from a compressor (not shown) flows into the refrigerant path 120 from the inlet 121, exchanges heat with the wind indicated by the arrow W supplied to the heat exchanger 100, and gradually radiates the refrigerant path 120. It is condensed by flowing through the standard part. The high-temperature and high-pressure liquid refrigerant flows through a thin tube portion attached to a portion of the refrigerant path 120 on the windward side of the fin 110. Thereby, the liquid refrigerant is further cooled. The liquid refrigerant that has flowed out of the outlet 122 of the refrigerant path 120 becomes a low-temperature and low-pressure gas refrigerant through a heat exchanger that functions as a decompressor and an evaporator (not shown), and the low-temperature and low-pressure gas refrigerant is sucked into the compressor again.

  As shown in FIG. 6, if the vicinity of the outlet 121 of the refrigerant path 120 is configured by a thin tube, the heat transfer coefficient between the liquid refrigerant and the air is improved, the degree of cooling of the liquid refrigerant is increased, and the performance of the air conditioner Will improve. In addition, it is possible to improve the problem of an increase in the amount of refrigerant encapsulated, which has been caused by increasing the cooling degree of the liquid refrigerant, and to obtain an air conditioner with high performance and high reliability. Moreover, since the heat transfer coefficient of the heat exchange part on the liquid refrigerant side is high, the size of the entire heat exchanger can be maintained at the same level as the conventional one.

Japanese Examined Patent Publication No. 62-13574

  However, the heat exchanger 100 shown in FIG. 6 is not assumed to be used as an evaporator. Rather, when the heat exchanger 100 is used as an evaporator, the performance of the heat exchanger 100 is degraded by the narrow tube portion located on the refrigerant inflow side.

  An object of this invention is to provide the refrigerating-cycle apparatus which can improve the performance of the heat exchanger which functions as an evaporator in view of such a situation.

In order to solve the above-described problems, a refrigeration cycle apparatus of the present invention includes a heat exchanger that functions as at least an evaporator having a refrigerant path extending from a first opening on the expansion means side to a second opening on the compressor side. A circuit, and a branch path branched from the main circuit on the upstream side or downstream side of the expansion means when the heat exchanger functions as an evaporator, and connected in the middle of the refrigerant path, the refrigerant path Includes a first path forming the first opening, a plurality of second paths split from the first path, and a merging path to which downstream ends of the branch paths are connected, and the refrigerant from the first opening. The mass flow rate of the refrigerant flowing into the path is G [kg / (m ² · s)], the length of the first path is L [m], the inner diameter of the pipe constituting the first path is D [m], The density of the liquid phase of the refrigerant flowing into the refrigerant path from the first opening Each fine viscosity ρ _l [kg / m ^3] and μ _l [μPa · s], the first opening, respectively the density and viscosity of the gas phase of the refrigerant flowing into the refrigerant passage from ρ _g [kg / m ^3] and When μ _g [μPa · s] is set, 1.60 × 10 ³ ≦ ((μ _l / (G · D)) ^0.25 · (2 · G ² / (D · ρ _l ))) · L ≦ 4. 64 × 10 ⁴ and 1.55 × 10 ⁴ ≦ ((μ _g / (G · D)) ^0.25 · (2 · G ² / (D · ρ _g ))) · L ≦ 4.17 × 10 ⁵ Is true.

  According to said structure, since a part of refrigerant | coolant which circulates through a main circuit is injected into the middle of a heat exchanger through a branch path, the flow volume of the refrigerant | coolant which flows in into a refrigerant path from 1st opening reduces. In addition, by appropriately setting the inner diameter D and the length L of the pipe constituting the first path, the performance of the heat exchanger functioning as an evaporator is improved as compared with the case where no branch path is provided. be able to.

The block diagram of the refrigerating-cycle apparatus which concerns on one Embodiment of this invention. Configuration of indoor heat exchanger when viewed from the side Expression expressing pressure loss when R410A in gas-liquid two-phase state flows in the pipe The figure which shows the relationship between the refrigerant | coolant dryness in a refrigerant | coolant path | route, and a refrigerant | coolant state The graph which shows the evaporator performance when changing the diameter D of the piping which comprises the 1st path | route, and the length L of the 1st path | route. A perspective view of a conventional heat exchanger

  Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings. However, the present invention is not limited to the following embodiments.

  FIG. 1 shows a refrigeration cycle apparatus 1 according to an embodiment of the present invention. The refrigeration cycle apparatus 1 is an air conditioner. However, the refrigeration cycle apparatus of the present invention is not necessarily an air conditioner, and may be a refrigerator, a heater, a hot water heater, or the like.

  Specifically, the refrigeration cycle apparatus 1 includes a main circuit 2 that circulates refrigerant, and a branch path 3 that is parallel to a part of the main circuit 2. As the refrigerant, chlorofluorocarbon (CFC), hydrochlorofluorocarbon (HCFC), or hydrofluorocarbon (HFC) can be used. Among these, R410A is suitable for the refrigeration cycle apparatus 1. However, it is also possible to use refrigerants such as R407C, R134a, R32, isobutane, propane and carbon dioxide.

  The main circuit 2 includes a compressor 11, a four-way valve 15, an indoor heat exchanger 14, an expansion valve 13 and an outdoor heat exchanger 12. The four-way valve 15 is connected to the suction port and the discharge port of the compressor 11 by flow paths 21 and 26, and is connected to the outdoor heat exchanger 12 and the indoor heat exchanger 14 by flow paths 22 and 25. The outdoor heat exchanger 12 and the indoor heat exchanger 14 are connected to the expansion valve 13 by flow paths 23 and 24.

  The four-way valve 15 switches the refrigerant flow direction between the heating operation and the cooling operation. During the heating operation, the refrigerant discharged from the compressor 11 is guided to the indoor heat exchanger 14, and during the cooling operation, the refrigerant discharged from the compressor 11 is guided to the outdoor heat exchanger 12. That is, during the heating operation, the indoor heat exchanger 14 functions as a condenser, and the outdoor heat exchanger 12 functions as an evaporator. On the other hand, during the cooling operation, the outdoor heat exchanger 12 functions as a condenser, and the indoor heat exchanger 14 functions as an evaporator. In the present embodiment, the indoor heat exchanger 14 is configured such that the refrigerant is injected through the branch path 3.

  The expansion valve 13 is an example of the expansion means of the present invention. As the expansion means of the present invention, in addition to the expansion valve 13, an expander that recovers power from the expanding refrigerant can be employed.

  As shown in FIG. 2, the indoor heat exchanger 14 has a refrigerant path 4 extending from the first opening 4a on the expansion means 13 side to the second opening 4b on the compressor 11 side. That is, during the cooling operation, the first opening 4a serves as the refrigerant inlet, and the second opening 4b serves as the refrigerant outlet.

  The branch path 3 branches from a flow path 24 between the expansion valve 13 and the indoor heat exchanger 14 and is connected to the refrigerant path 4 in the middle. The branch path 3 is provided with a check valve 31 so that the refrigerant flows through the branch path 3 only during the cooling operation. That is, a part of the refrigerant circulating in the main circuit 2 is injected into the middle of the indoor heat exchanger 14 through the branch path 3 during the cooling operation, and the refrigerant is prevented from flowing out from the middle of the indoor heat exchanger 14 during the heating operation. .

  If a flow rate control valve is installed in the branch path 3, the injection flow rate (refrigerant flow rate through the branch path 3) can be optimally controlled under all conditions. On the other hand, considering cost, it is preferable not to install a flow control valve and to configure the branch path 3 with a capillary tube or the like. However, when a capillary tube is used, it cannot be freely controlled. Therefore, a capillary tube having specifications optimized in advance under optimum control conditions is installed, and the injection flow rate is likely to occur under other conditions.

  Next, the configuration of the indoor heat exchanger 14 will be described in detail with reference to FIG. In the following, for convenience of explanation, the first opening 4a side is referred to as the upstream side and the second opening 4b side is referred to as the downstream side based on the refrigerant flow direction during the cooling operation.

  The refrigerant path 4 includes a first path 41 that forms the first opening 4a, a first path group 4A, an intermediate path 44, and a second path group 4B. The first path group 4A and the second path group 4B are connected by an intermediate path 44. The intermediate path 44 is provided with a dehumidifying valve 6 that depressurizes the refrigerant. The downstream end of the branch path 3 is connected to the intermediate path 44 on the upstream side of the dehumidifying valve 6. That is, the intermediate path 44 corresponds to the merge path of the present invention.

  The first path group 4A includes a plurality of (two in the illustrated example) second paths 42 split from the first path 41, and a plurality of (four in the illustrated example) third paths further split from each second path 42. 43. The third path 43 converges at the upstream end of the intermediate path 44. Instead of omitting the third path 43, the number of the second paths 42 may be increased, and the second path 42 may converge at the upstream end of the intermediate path 44.

  The second path group 4B includes a plurality of (three in the illustrated example) end paths 45 split from the intermediate path 44, and the end paths 45 converge to the second opening 4b.

  The first path 41, the second path 42, the third path 43, and the end path 45 are mainly composed of a plurality of hairpin tubes 51, 52, 53, and 55 that pass through fans arranged at predetermined intervals (not shown). Is done. In each path, the ends of the hairpin tubes are connected by a bend tube. All the hairpin tubes 51, 52, 53, 55 are preferably grooved tubes in which a plurality of spiral grooves are formed on the inner peripheral surface.

  The indoor heat exchanger 14 has the sirocco fan 7, and the hairpin tube 53 constituting the third path 43 and the hairpin tube 55 constituting the terminal path 45 are arranged so as to surround the sirocco fan 7. The hairpin tube 52 constituting the second path 42 is arranged on the windward side of the hairpin tube 53 constituting the third path 43 in the wind direction generated by the sirocco fan 7, and the hairpin tube constituting the first path 41. 51 is arranged further on the windward side of the hairpin tube 52 constituting the second path 42.

  The pipe constituting the first path 41 may be the same thickness as the pipe constituting the second path 42 or may be thicker than that. However, it is preferable that the pipe constituting the first path 41 is thinner than the pipe constituting the second path 42. Moreover, it is preferable that the piping which comprises the 1st path | route 41 is thinner than the piping which comprises the flow path 24 connected to the 1st opening 4a.

  Next, a mechanism for improving the evaporator performance by the injection through the branch path 3 will be described.

  FIG. 3 shows the heat transfer mode of forced convection boiling in the tube. The heat transfer mode of forced convection boiling can be divided into three regions by dryness X as shown in this figure. The first region is a low dryness region where the proportion of subcool and gas phase is small, and heat transfer is mainly controlled by nucleate boiling on the heat transfer surface. The second region has a thin liquid film on the wall surface, and is a high dryness region that is an annular flow in which steam flows through the center of the tube. When the heating heat flux is not high, the liquid film that flows along the heat transfer surface This is the region where convective heat transfer and evaporation from the liquid film surface dominate. The third region is a spray flow region after dry-out where the wall surface is dry, and the gas phase contains a large amount of droplets. The third spray flow region has a very poor heat transfer coefficient as shown in this figure. Therefore, in practice, the first and second regions are used.

  From the relationship between the refrigerant dryness and the heat transfer coefficient, a part of the refrigerant supplied to the indoor heat exchanger 14 is injected into the middle of the indoor heat exchanger 14 so that the branch path 3 in the refrigerant path 4 becomes the refrigerant path 4. The flow rate of the main-stream refrigerant flowing upstream from the merge position M connected to is reduced, and the state becomes easier to evaporate. As a result, the dryness of the mainstream refrigerant at the merge position M increases and the refrigerant heat transfer coefficient increases. However, since the heat transfer coefficient rapidly decreases after dryout, the refrigerant heat transfer coefficient decreases if the injection flow rate is increased too much. Therefore, it is important to control to an optimal injection flow rate. Moreover, since the pressure loss of the mainstream refrigerant | coolant in the merge position M can also be reduced by injecting, the temperature difference between air and a refrigerant | coolant can be enlarged and evaporation performance can be improved. In order to realize this, it is important to appropriately set the inner diameter D and the length L of the pipes constituting the first path 41. When the pipe is a grooved pipe, the inner diameter D of the pipe is 4 times the cross-sectional area of the inner space of the grooved pipe, and the wetting edge length (the wall in contact with the fluid when the pipe is viewed in cross section) The length divided by the so-called hydraulic diameter (also called the equivalent diameter).

  Furthermore, when the piping which comprises the 1st path | route 41 is thinner than the piping which comprises the 2nd path | route 42, both condensation performance and evaporation performance can be improved. In the heat exchanger 100 functioning as the condenser shown in FIG. 6 disclosed in Patent Document 1, the heat transfer coefficient between the liquid refrigerant and the air can be increased by configuring the vicinity of the outlet 122 of the refrigerant path 120 with a thin tube. This improves the cooling degree of the liquid refrigerant and improves the performance of the air conditioner. However, when the heat exchanger 100 functions as an evaporator, the pressure loss of the refrigerant increases due to the thin tubes located on the refrigerant inflow side, and the refrigerant evaporation temperature decreases. Therefore, there was a problem that the compressor power was increased and the performance of the entire refrigeration cycle apparatus was deteriorated. However, even when a thin tube is used at the evaporator inlet, the refrigerant pressure loss at the evaporator inlet, which has been a conventional problem, can be reduced by combining the branch path 3, and the heat transfer coefficient on the refrigerant side is improved. Can also be expected.

  FIG. 5 shows a result of a trial calculation of the evaporator performance when the diameter D of the pipe constituting the first path 41 and the length L of the first path 41 are changed. The reason for focusing only on the evaporator performance is as follows.

  When a thin tube is used at the outlet of the condenser, the outlet of the condenser is a liquid refrigerant single phase, so that the refrigerant pressure is reduced, but the refrigerant temperature is hardly changed. Therefore, there is almost no decrease in condensation performance due to an increase in refrigerant pressure loss, and the condensation performance improves as the diameter of the pipe is reduced. Therefore, it is not necessary to pay attention to the condensation performance. On the other hand, if the diameter of the evaporator inlet pipe is reduced, the refrigerant pressure loss in the evaporator increases and the evaporation performance decreases. The purpose of the present invention is to improve the performance more than before by injection into the evaporator, and this is the reason why attention is paid only to the evaporator performance.

The calculation conditions are: dry bulb temperature of air: 27 [° C.], wet bulb temperature: [19 ° C.], speed of wind supplied to the indoor heat exchanger 14: [1.5 m / s], refrigerant: R410A, overall Refrigerant circulation rate: 84 [kg / h]. The temperature of the refrigerant in the first opening 4a is 17.5 [° C.] (cooling rated condition), the dryness X is 0.12, and the physical property value of the liquid phase and the physical property of the gas phase in the first opening 4a at that time The values were calculated using general-purpose software Refprop7 (NIST). As a result, the liquid phase density ρ _l = 1063 [kg / m ³ ], the liquid phase viscosity μ _l = 123 [μPa · s], the gas phase density ρ _g = 64 [kg / m ³ ], The viscosity was μ _g = 14 [μPa · s].

  In the calculation, three models 1 to 3 having outer diameters D ′ of the pipes constituting the first path 41 being different from 7 mm, 5 mm, and 4 mm were created. In Models 1 to 3, the optimum refrigerant path 4 configuration and injection flow rate were set according to the outer diameter D '. In Models 1 to 3, the number of hairpin tubes in the first path 41 was 4, 5, and 6, respectively.

Specifically, the wall thickness t [m], the inner diameter D [m], and the length L [m] of the pipe constituting the first path 41, and the mass flow rate G of the refrigerant flowing into the refrigerant path 4 from the first opening 4a. [Kg / (m ² · s)] was as shown in Table 1.

  As shown in FIG. 5, if the inner diameter D and the length L of the pipe constituting the first path 41 are optimally designed, the evaporator performance is the highest. When the inner diameter D is decreased and the length L is increased from the optimum state, the evaporator performance gradually decreases, and when the situation A is reached, it becomes the same level as the current product in which the branch path 3 is not provided. On the other hand, when the inner diameter D is increased and the length L is decreased from the optimum state, the evaporator performance gradually decreases, and when the situation B is reached, it becomes the same level as the current product in which the branch path 3 is not provided.

In situation A, F (l) = 1.60 × 10 ³ and F (g) = 1.55 × 10 ⁴ . In situation B, F (l) = 4.64 × 10 ⁴ and F (g) = 4.17 × 10 ⁵ . F (l) and F (g) are defined by the following equations.

F (l) = ((μ _l / (G · D)) ^0.25 · (2 · G ² / (D · ρ _l ))) · L
F (g) = ((μ g / (G · D)) 0.25 · (2 · G 2 / (D · ρ g))) · L

  The above equation is obtained by extracting the terms of the equation representing the pressure loss when R410A in the gas-liquid two-phase state shown in FIG. 4 flows in the pipe, and the physics related to the refrigerant pressure loss in the liquid phase and the gas phase, respectively. Amount. The formula shown in FIG. 4 is cited from the following paper.

Title: Carbon Dioxide and R410a Flow Boiling Heat Transfer,
Pressure Drop, and Flow Pattern in Horizontal Tubes
at Low Temperatures
Author: CY Park and PS Hrnjak
Open Date: January 2007
Affiliation: Air Conditioning and Refrigeration Center
University of Illinois
Department of Mechanical Science & Engineering

That is, the inner diameter D and the length L are 1.60 × 10 ³ ≦ F (l) ≦ 4.64 × 10 ⁴ and 1.55 × 10 ⁴ ≦ F (g) ≦ 4.17 × 10 ^5. If it is designed to hold, the evaporator performance can be improved as compared with the prior art. Thereby, for example, even when a thin tube is used at the outlet of the condenser, the performance of the evaporator can be improved.

  For example, when CFC, HCFC, or HFC is used as the refrigerant, the inner diameter D and the length L can be selected so that 738 ≦ L / D ≦ 2075 is satisfied. In addition, the lower limit value and the upper limit value of the above formula were calculated using L and D of model 1 and L and D of model 3, respectively. Further, the outer diameter D ′ of the pipes constituting the first path 41 is, for example, not less than 4 mm and not more than 7 mm.

When the refrigerant is changed from R410A to R134a, the physical properties at the first opening 4a are as follows: liquid phase density ρ _l = 1510 [kg / m ³ ], liquid phase viscosity μ _l = 197 [μPa · s], gas phase Density ρ _g = 32 [kg / m ³ ], and the viscosity of the gas phase μ _g = 12 [μPa · s]. When the refrigerant is changed from R410A to R32, the physical property values at the first opening 4a are as follows: liquid phase density ρ _l = 964 [kg / m ³ ], liquid phase viscosity μ _l = 115 [μPa · s], gas phase Density ρ _g = 46 [kg / m ³ ] and the viscosity of the gas phase μ _g = 13 [μPa · s]. The mass flow rates G in Models 1 to 3 when using R134a and R32 are as shown in Table 2.

  In the above embodiment, the target heat exchanger for improving the evaporator performance is an indoor heat exchanger of an air conditioner that functions as a condenser or functions as an evaporator depending on operating conditions. The heat exchanger of the invention is not limited to this. For example, the heat exchanger of the present invention may be an eco-cute (heat pump type water heater) heat exchanger that functions only as an evaporator. Even in this case, the effect of improving the evaporator performance as compared with the prior art can be obtained as in the above embodiment.

<Modification>
The merging path of the present invention does not necessarily need to be the intermediate path 44. For example, the downstream end of the branch path 3 may be branched and connected to the third path 43, and the third path 43 may serve as the merge path of the present invention.

  The branch path 3 may branch from the flow path 23 between the expansion valve 13 and the outdoor heat exchanger 12. In this case, the liquid-phase refrigerant before being decompressed by the expansion valve 13 is injected into the indoor evaporator 14. Even in this case, the refrigerant dryness can be increased on the upstream side of the merging position M due to the decrease in the flow rate of the refrigerant flowing into the refrigerant path 4 from the first opening 4a.

DESCRIPTION OF SYMBOLS 1 Refrigeration cycle apparatus 11 Compressor 12 Outdoor heat exchanger 13 Expansion valve (expansion means)
14 indoor heat exchanger 2 main circuit 3 branch path 4 refrigerant path 4A first path group 4B second path group 41 first path 42 second path 43 third path 44 intermediate path (merging path)
45 End route 6 Dehumidification valve

Claims (6)

A main circuit including at least a heat exchanger functioning as an evaporator, having a refrigerant path extending from the first opening on the expansion means side to the second opening on the compressor side;
A branch path branched from the main circuit on the upstream side or downstream side of the expansion means when the heat exchanger functions as an evaporator, and connected to the middle of the refrigerant path,
The refrigerant path includes a first path that forms the first opening, a plurality of second paths split from the first path, and a merging path to which the downstream ends of the branch paths are connected,
The mass flow rate of the refrigerant flowing into the refrigerant path from the first opening is G [kg / (m ² · s)], the length of the first path is L [m], and the pipes constituting the first path The inner diameter is D [m], the density and viscosity of the liquid phase of the refrigerant flowing into the refrigerant path from the first opening are ρ _l [kg / m ³ ] and μ _l [μPa · s], respectively, and the refrigerant from the first opening When the density and viscosity of the gas phase of the refrigerant flowing into the path are ρ _g [kg / m ³ ] and μ _g [μPa · s], respectively,
1.60 × 10 ³ ≦ ((μ _l / (G · D)) ^0.25 · (2 · G ² / (D · ρ _l ))) · L ≦ 4.64 × 10 ⁴ and 1.55 × _{^{10 4 ≦ ((μ g /}} (G · D)) 0.25 · (2 · G 2 / (D · ρ g))) · L ≦ 4.17 × 10 5 holds, the refrigeration cycle according to claim 1 apparatus. The refrigerant is chlorofluorocarbon, hydrochlorofluorocarbon or hydrofluorocarbon,
The refrigeration cycle apparatus according to claim 1, wherein 738 ≦ L / D ≦ 2075 is established.   The refrigeration cycle apparatus according to claim 2, wherein an outer diameter of the pipe constituting the first path is 4 mm or more and 7 mm or less.   The refrigeration cycle apparatus according to claim 2 or 3, wherein the refrigerant is R410A.   The refrigeration according to any one of claims 1 to 4, wherein the heat exchanger is an indoor heat exchanger of an air conditioner that operates in an environment of a dry bulb temperature: 27 ° C and a wet bulb temperature: 19 ° C. Cycle equipment. The refrigerant path includes a first path group including the plurality of second paths, and a second path group including a plurality of terminal paths that converge to the outlet,
The merging path is an intermediate path provided with a dehumidification valve that connects the first path group and the second path group.
The refrigeration cycle apparatus according to claim 5, wherein a downstream end of the branch path is connected to the intermediate path on the upstream side of the dehumidification valve when the heat exchanger functions as an evaporator.