JP5927415B2 - Refrigeration cycle equipment - Google Patents

Description

  The present invention relates to a refrigeration cycle apparatus, and particularly relates to a heat exchanger.

  Conventionally, in this type of refrigeration cycle device, in order to use heat exchangers more efficiently with minimum loss, when used as an evaporator, the refrigerant flow rate was reduced by using multiple passes to reduce pressure loss. Better. However, since it is less necessary to consider the pressure loss when used as a condenser, the heat transfer coefficient of the refrigerant can be increased and the operation can be efficiently performed by reducing the number of passes.

  For example, FIG. 22 shows a conventional refrigeration cycle apparatus described in Patent Document 1, but it is possible to switch the number of passes by combining a plurality of electromagnetic two-way valves or check valves in accordance with switching between an evaporator and a condenser. Yes.

Japanese Patent Laid-Open No. 10-170081

Hiroshi Seshita, Masao Fujii “Compact Heat Exchanger”, Nikkan Kogyo Shimbun, P85

  However, the configuration of the conventional refrigeration cycle apparatus has the following problems.

  The electromagnetic two-way valve arranged for switching the number of passes is expensive, and since a plurality (three) of the electromagnetic valves are arranged, it is difficult to adopt it for a product in consideration of the manufacturing cost. Further, as described in the embodiment, since a part of the heat exchanger cannot be used in the switching method using an inexpensive check valve, the heat exchanger cannot be used to the maximum efficiency. Further, the suction pipe connecting the heat exchanger and the compressor is not particularly described, and the efficiency improvement effect cannot be completely ensured.

  The present invention is directed to a refrigeration cycle apparatus that solves the above-described conventional problems and can switch the number of passes with an inexpensive configuration.

  In order to solve the conventional problems, in the refrigeration cycle apparatus of the present invention, the compressor 1, the four-way valve 2, the outdoor heat exchanger 3, the expansion valve 7, the indoor heat exchanger 12, the suction pipes 8 and 9 Are connected in a ring, and the check valves 20 and 21 are arranged in the outdoor heat exchanger 3, and when used as a condenser for cooling operation, the heat exchangers are connected in series, used for heating operation and as an evaporator In such a case, the heat exchangers are connected in parallel. With such a configuration, the flow rate of the refrigerant increases as the condenser, and the heat transfer coefficient increases. In addition, the pressure loss of the evaporator is reduced and the efficiency is improved. And it becomes the structure which becomes more than a fixed ratio with respect to the maximum internal cross-sectional area or the minimum cross-sectional area at the time of making the internal cross-sectional area of a suction pipe into a multipass as an evaporator.

According to this configuration, the newly added parts are only inexpensive check valves, and in general air conditioners, high-pressure refrigerant (R410A) with a small ozone layer depletion coefficient is used, so that valve pressure loss can be ignored and inexpensive. Thus, an efficient heat exchanger can be configured. In addition, the copper pipe cost, the amount of refrigerant retained, and the pressure loss of the refrigerant can be optimized by setting the suction pipe only to the optimum pipe diameter.

  According to the refrigeration cycle apparatus of the present invention, it is possible to use the heat exchanger with high efficiency only with an inexpensive check valve.

The figure which shows the structure of the heat exchanger in the case of consisting of the three heat exchanger blocks in Embodiment 1 of this invention. The figure which shows the structure of the heat exchanger in the case of consisting of n heat exchanger blocks in Embodiment 1 of this invention. The figure which shows the structure at the time of applying the heat exchanger which consists of three heat exchanger blocks in Embodiment 1 of this invention as a part of outdoor heat exchanger. The figure which shows the structure at the time of applying the heat exchanger which consists of three heat exchanger blocks in Embodiment 1 of this invention to the whole outdoor heat exchanger. The figure which shows the structure at the time of applying the heat exchanger which consists of three heat exchanger blocks in Embodiment 1 of this invention as a part of indoor heat exchanger. The figure which shows the structure at the time of applying the heat exchanger which consists of three heat exchanger blocks in Embodiment 1 of this invention to the whole indoor heat exchanger. Diagram showing the relationship between heat exchanger capacity and pressure loss in evaporators and condensers The figure which shows the optimal composition in a conventional heat exchanger model and a condenser and an evaporator The figure which compared the Example applied to the outdoor heat exchanger in Embodiment 1 of this invention, and a conventional structure. Relationship diagram between heat exchanger capacity and pressure loss when operating capacity changes when applied to outdoor heat exchanger in Embodiment 1 of the present invention Schematic diagram of an example of unifying and unifying the pipe length in the case of consisting of three heat exchanger blocks in Embodiment 2 of the present invention Comparison diagram of pressure loss in outdoor heat exchanger according to Embodiment 1 of the present invention The figure which shows the structure of the heat exchanger in case it consists of two heat exchanger blocks in Embodiment 3 of this invention. The figure which shows the structure of the heat exchanger in case it becomes 3 passes at the time of the evaporator in Embodiment 4 of this invention. FIG. 7 is a relationship diagram between the Reynolds number (Re) at the inlet portion of the heat exchanger and the Nusselt number (Nu) representing the refrigerant side heat exchange efficiency in Embodiment 4 of the present invention. Refrigeration cycle configuration diagram in the case where the evaporator inlet portion of the outdoor heat exchanger in Embodiment 4 of the present invention has three passes Refrigeration cycle configuration diagram in the case where the evaporator inlet portion of the outdoor heat exchanger in the fourth embodiment of the present invention has 5 passes and 7 passes The figure which shows the heating capability at the time of changing the number of passes of the heat exchanger inlet_port | entrance part at the time of the evaporator in Embodiment 4 of this invention with 3-7. The comparison figure which compared the condensation capability at the time of changing the number of piping of 1 pass part in Embodiment 4 of this invention with 2-8, and a refrigerant | coolant pressure loss. The figure which compared arrangement | positioning of the 1-pass part in Embodiment 5 and 6 of this invention The comparison figure which compared the condensation capability at the time of changing arrangement | positioning of the 1-pass part in Embodiment 5 of this invention, and refrigerant | coolant pressure loss Refrigeration cycle block diagram with thick suction pipe from outdoor heat exchanger to compressor accumulator in Embodiment 7 of the present invention Cross-sectional area distribution diagram of refrigerant piping in outdoor heat exchanger-suction piping in Embodiment 7 of the present invention Refrigeration cycle block diagram with thick suction pipe from outdoor heat exchanger to compressor accumulator in Embodiment 7 of the present invention Refrigeration cycle configuration diagram in which the suction pipe is thickened only in the portion from the outdoor heat exchanger to the four-way valve in Embodiment 8 of the present invention. Cross-sectional area distribution diagram of refrigerant pipe in outdoor heat exchanger to suction pipe in embodiment 8 of the present invention Pressure distribution diagram in outdoor heat exchanger to suction pipe in embodiment 8 of the present invention Refrigeration cycle configuration diagram directly connected via an electromagnetic two-way valve from the outdoor heat exchanger outlet to the compressor accumulator in Embodiment 9 of the present invention System diagram in conventional invention

The first invention includes a refrigerant circuit formed by sequentially connecting at least a compressor, a four-way valve, an outdoor heat exchanger, an expansion device, and an indoor heat exchanger, and at least one of the two heat exchangers includes a plurality of refrigerant circuits. In the heat exchanger in which the refrigerant heat exchanger block is arranged in parallel and the refrigerant inlet and outlet are directly connected to each other, the heat exchanger inlet pipe at the time of the condenser is the outermost heat exchanger block and the adjacent one. connected between the second heat exchanger block, a check valve which flows only in the direction of the heat exchanger inlet pipe in the first branch pipe directly connected to the refrigerant inlet mouth of the heat exchanger block of the heat exchanger block disposed between the even-numbered odd and adjacent heat exchanger outlet pipe side, only in a direction opposite to the heat exchanger outlet pipe in a second branch pipe that directly connects the refrigerant exit of the heat exchanger block Even number of heat exchanger block with flowing check valve The heat exchanger outlet pipe at the time of the condenser and the heat exchanger block arranged on the outermost side opposite to the heat exchanger inlet pipe are adjacent to the odd number on the outlet pipe side adjacent to the heat exchanger inlet pipe. By connecting between the heat exchanger blocks, the condenser is configured in series, the evaporator is configured in parallel, and when used as an evaporator, the Reynolds number at the evaporator inlet becomes 3000 or more. Thus, by determining the number of passes, the heat exchanger capacity can be utilized most efficiently.

The second invention includes a refrigerant circuit formed by sequentially connecting at least a compressor, a four-way valve, an outdoor heat exchanger, a throttling device, and an indoor heat exchanger, and at least one of the indoor heat exchanger and the outdoor heat exchanger. , is directly connected across the heat exchanger, or a portion thereof are two only to the heat exchanger block and condensed at the outlet direction a pipe containing a check valve Ru passed through a refrigerant are arranged in parallel and coolant inlet and outlet, respectively In the heat exchanger, the heat exchanger inlet pipe at the time of the condenser is connected between the heat exchanger block arranged farthest from the heat exchanger outlet pipe and the adjacent heat exchanger block, and the heat exchange during a check valve flow only in the direction of the heat exchanger inlet pipe in the first branch pipe directly connected to the refrigerant inlet mouth vessel blocks of the second 1st and adjacent heat exchanger outlet pipe side of the heat exchanger block placed straight refrigerant exit of the heat exchanger block It was the heat exchanger outlet pipe in a second branch pipe disposed between the pipe containing the check valve to the second and the adjacent check valve a heat exchanger block which flows only in the opposite direction, when the condenser When the heat exchanger outlet pipe is connected between the heat exchanger block arranged on the opposite side of the heat exchanger inlet pipe and the pipe including the check valve adjacent thereto , the evaporator is used as an evaporator. By setting the Reynolds number at the inlet to be 3000 or more , the refrigerant flows in parallel through the two heat exchanger blocks at the time of the evaporator, the pressure loss is reduced, and the refrigerant flows in series at the time of the condenser. By increasing the heat transfer rate, the heat exchanger performance can be maximized.

  In the third aspect of the invention, in particular, the check valve of the first or second aspect of the invention is an electromagnetic two-way valve, and an optimum heat exchanger capacity is obtained by selecting valve opening / closing according to the circulation amount, air conditioning conditions, etc. Can do.

  In the fourth aspect of the invention, in particular, in any one of the first to third aspects, the heat exchanger for switching the plurality of heat exchanger blocks in series and in parallel is arranged at the refrigerant inlet portion at the time of the evaporator, so that only the minimum block is provided. However, the maximum efficiency improvement can be achieved by simply applying this system.

  In the fifth aspect of the invention, in particular, in the first to fourth aspects of the invention, the number of pipes of the plurality of heat exchanger blocks is made the same with respect to the refrigerant flow direction, thereby reducing the refrigerant drift and improving the maximum heat exchanger efficiency. Can be planned.

  In the sixth aspect of the invention, in particular, in the first to fifth aspects of the invention, by making the diameter of the check valve arranged on the downstream side of the check valve arranged upstream of the refrigerant flow at the time of the evaporator equal to or greater than, Reduction of refrigerant pressure loss can be achieved.

In the seventh aspect of the invention, in particular, when the pipe of the variable path heat exchanger has a diameter of 7 mm, the number of passes when used as an evaporator is 6 or less, and when the diameter is 6.35 mm, the number of passes when used as an evaporator is When the number of passes is 7 or less and the diameter is 5 mm, the number of passes when used as an evaporator is 12 or less. When the diameter is 7.94 mm or more, the number of passes when used as an evaporator is 4 or less and heat exchange is performed. The most efficient use of vessel capacity.

In the eighth aspect of the invention, in particular, when the variable path heat exchanger is used as a condenser, the number of pipes in one path section is four or six when the pipe has a diameter of 7 mm, and eight condenser outlet pipes are heat exchangers. The heat exchanger capacity can be utilized most efficiently by arranging it on the windward side.

According to a ninth aspect of the invention, the variable path heat exchanger reduces the number of pipes in the refrigerant flow direction of the inner heat exchanger block than either or both of the number of pipes in the refrigerant flow direction of the heat exchanger blocks at both ends. Can improve the shunt characteristics.

The tenth aspect of the invention relates to the suction pipe connecting the outdoor heat exchanger and the four-way valve having a system in which the number of passes is switched according to the cooling operation and the heating operation in the first to ninth inventions. Maximum cross-sectional area of the piping inside the unit) x 1.2> (Cross-sectional area of the suction piping) ≥ (Maximum cross-sectional area of the piping inside the outdoor heat exchanger) x 0.8 By determining the pressure loss in the suction pipe due to the gas refrigerant that has flowed out of the outdoor heat exchanger can be suppressed to an appropriate level, it is possible to achieve compatibility with problems such as an increase in manufacturing cost and an increase in the amount of charged refrigerant.

In an eleventh aspect of the invention, in particular, in the first to ninth aspects of the invention, there is provided a suction pipe for connecting an outdoor heat exchanger having a system in which the number of passes is switched according to cooling operation and heating operation and a compressor suction portion via a four-way valve. Except for the four-way valve and connecting part, (maximum cross-sectional area of piping in outdoor heat exchanger) x 1.0> (cross-sectional area of suction pipe) ≥ (maximum cross-sectional area of piping in outdoor heat exchanger) x 0 By determining the diameter of the suction pipe so as to satisfy the condition of .6, the pressure loss in the suction pipe due to the gas refrigerant flowing out of the outdoor heat exchanger is suppressed to an appropriate level, and the manufacturing cost is increased or filled. It can be compatible with problems such as an increase in the amount of refrigerant.

In a twelfth aspect of the invention, in particular, in the first to ninth aspects of the invention, there is provided a suction pipe for connecting the outdoor heat exchanger having a system in which the number of passes is switched according to the cooling operation and the heating operation and the four-way valve (in the outdoor heat exchanger). The diameter of the suction pipe is determined so as to satisfy the condition of (the maximum cross-sectional area of the pipe) × 1.0> (the cross-sectional area of the suction pipe) ≧ (the minimum cross-sectional area of the pipe in the outdoor heat exchanger) × 1.1. Thus, while suppressing the pressure loss in the suction pipe due to the gas refrigerant flowing out of the outdoor heat exchanger to an appropriate level, it is possible to achieve compatibility with problems such as an increase in manufacturing cost and an increase in the amount of refrigerant charged.

The thirteenth invention is the tenth to twelfth invention, particularly when a φ7 mm pipe and six branches are used as a heat exchanger pipe, and a suction pipe when a quintuple pipe and four branches are used as a suction pipe. As a quadrant, the pressure loss in the suction pipe due to the gas refrigerant flowing out of the outdoor heat exchanger is suppressed to an appropriate level, while at the same time satisfying issues such as an increase in manufacturing costs and an increase in the amount of refrigerant charged. Can do.

The fourteenth invention is the invention according to the tenth to twelfth inventions, in particular when a φ5 mm pipe and 12 branches are used as a heat exchanger pipe, a five-part pipe as a suction pipe, and a quadrant pipe as a suction pipe when eight branches are used. As a result, the pressure loss in the suction pipe due to the gas refrigerant that has flowed out of the outdoor heat exchanger can be suppressed to an appropriate level, and problems such as an increase in manufacturing costs and an increase in the amount of refrigerant charged can be achieved.

A fifteenth aspect of the invention is that in the first to ninth aspects of the invention, there is provided a suction pipe that connects the outdoor heat exchanger having a system in which the number of passes is switched according to the cooling operation and the heating operation and the four-way valve. When the condition of (area) <(maximum cross-sectional area of the pipe in the outdoor heat exchanger) × 0.8 is satisfied, a pipe directly connected from the heat exchanger outlet to the four-way valve or the compressor suction is arranged, and the middle part By having an electromagnetic two-way valve, the pressure circuit can be optimally controlled by opening and closing the bypass circuit as necessary.

A sixteenth aspect of the invention is the first to ninth aspects of the invention, wherein the suction pipe for connecting the outdoor heat exchanger having a system in which the number of passes is switched according to the cooling operation and the heating operation and the compressor suction portion via the four-way valve is provided. Except for the four-way valve and connecting part, when the condition of (cross-sectional area of suction pipe) <(maximum cross-sectional area of pipe in outdoor heat exchanger) x 0.6, the four-way valve from the heat exchanger outlet Alternatively, by arranging a pipe directly connected to the compressor suction part and having an electromagnetic two-way valve in the middle part thereof, the bypass circuit can be opened and closed as necessary, and the pressure loss can be optimally controlled.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings. Note that the present invention is not limited to the embodiments.

(Embodiment 1)
FIG. 1 shows a case where the heat exchanger of the refrigeration cycle apparatus according to the first embodiment of the present invention is composed of three heat exchanger blocks.

  When the heat exchanger is used as an evaporator in FIG. 1, a gas-liquid two-phase low-pressure refrigerant flows from the heat exchanger inlet pipe 5. In this case, since the refrigerant can pass through the check valve 21, the refrigerant can pass in parallel through the three heat exchanger blocks 22, 23, 24 arranged in parallel. The refrigerant that has passed through the stop valve 20 merges and flows out to the outlet pipe.

  Further, since the pressure loss experienced by the refrigerant when passing through the copper pipe constituting the heat exchanger is generally proportional to the square of the refrigerant flow rate, the flow rate can be reduced to one third. In Form 1, the pressure loss can be reduced to 1/9. As the refrigerant flow rate decreases, the heat transfer coefficient on the refrigerant side also decreases. In general, pressure loss has a direct effect on the input of the compressor alone, and thus has a greater influence than a decrease in heat transfer coefficient. Therefore, in this invention, it becomes the structure which can aim at the power consumption reduction of the whole refrigerating-cycle apparatus.

  Next, when the heat exchanger is used as a condenser, the refrigerant flows from the heat exchanger outlet pipe 4 as a high-pressure gas refrigerant or a gas-liquid two-phase refrigerant. In this case, since the refrigerant cannot pass through the check valve 20, the refrigerant first passes through the heat exchanger block 22 to exchange heat. Next, since the refrigerant cannot pass through the other check valve 21, the heat flows through the heat exchanger block 23 in the reverse direction to exchange heat. There is no problem because the heat exchanger has no adverse effects depending on the direction.

  Since the check valve 20 does not flow in the reverse direction because of the pressure, the refrigerant passes through the heat exchanger block 24 and is heat-exchanged. Finally, the check valve 21 does not pass due to the pressure, passes through the heat exchanger inlet pipe 5 and flows to the next refrigeration cycle process. In the case of a condenser, since the refrigerant passes through the heat exchanger blocks arranged in series in this way, the refrigerant flow rate can be increased and the heat transfer rate can be improved. Although the pressure loss of the refrigerant also increases at the same time, the high pressure refrigerant has a small refrigerant density, so the pressure loss is sufficiently small and the influence on the refrigeration cycle apparatus is small.

  By configuring as described above, the heat exchanger efficiency can be increased both as an evaporator and as a condenser. In the first embodiment of the present invention, as can be seen from FIG. 1, there are only two additional parts (a fraction of the price of the electromagnetic two-way valve), and a configuration as proposed previously ( Unlike the electromagnetic two-way valve and four-way valve), not only can the structure be inexpensive, but the problem is small because the check valve has a pipe shape on the main body. There is no need to add electrical parts.

  In terms of reliability, there is no problem because there is a sufficient record of use in such a refrigerant state. Even if the check valve is completely closed or normally opened, the refrigerant passes through the heat exchanger and does not enter an unsafe state.

  At this time, there is no particular limitation on the number of rows, the number, etc. of the internal configuration of the heat exchanger blocks 22, 23, 24, and an optimal configuration can be achieved. For example, the heat exchanger block 22 may be four in one row, and the heat exchanger blocks 23 and 24 may be six in two rows. However, in order to make the best use of the performance of the heat exchanger, it is better to make the flow split even, and as described later, it is better to limit the number of copper tubes.

FIG. 2 expands the case of the three heat exchanger blocks described above, and heat exchanger blocks B ₁ to B ₁ .
The B _n is a diagram illustrating a case of arranging the n pieces parallel. Thus, the number of heat exchanger blocks is not limited to three and can be increased.

In such a case, the refrigerant inlet pipe 4 at the time of the condenser is connected between the heat exchanger block B ₁ disposed on the outermost side and the second heat exchanger block B ₂ adjacent thereto, and the heat in piping directly connected to the refrigerant inlet pipe exchanger blocks the inlet pipe disposed between the even-numbered outlet pipe side adjacent the check valve 20 ₁ which flows only in a direction opposite to the odd-numbered heat exchanger block, the the check valve 21 ₁ flows only in the outlet pipe direction in the piping directly connected to the refrigerant outlet pipe of the heat exchanger block is placed between the odd-numbered outlet pipe side and an adjacent even-numbered heat exchanger block.

In the same manner, the check valves are sequentially arranged, and the heat exchanger outlet pipe at the time of the condenser is arranged at the outermost side opposite to the inlet pipe and the heat exchanger block B _n-1 and the adjacent heat exchange. It becomes the structure connected between unit block _Bn .

  With this arrangement, all the heat exchanger blocks can be arranged in parallel in the evaporator and in series in the condenser. Of course, it is also possible to employ a configuration in which a check valve is not arranged only in a part of the interior and switching in series and parallel cannot be performed.

  Embodiment 1 of the present invention will be described below including actual application examples and effects.

  When the outdoor heat exchanger is operated for heating, it becomes an evaporator, and the pressure loss of the refrigerant causes a large increase in power consumption. In this case, it is a common method to reduce the pressure loss by increasing the number of passes of the heat exchanger and reducing the refrigerant speed in the copper tube, but when such a method is adopted, When switching to cooling operation and using as a condenser, an excessive decrease in the refrigerant flow rate leads to a decrease in heat transfer coefficient, leading to a decrease in heat exchanger efficiency.

  Conventionally, as shown in FIG. 8, a configuration in which a compromise is made between the optimum configuration as a condenser and the optimum configuration as an evaporator has been adopted, but the efficiency of the heat exchanger is not necessarily the maximum. FIG. 7 illustrates this relationship. The optimum number of passes as a condenser having a smaller influence of pressure loss than the optimum number of passes as an evaporator is small, and generally about one third.

FIG. 3 shows a case where the present invention is applied to a part of the outdoor heat exchanger of the refrigeration cycle apparatus.
Since the configuration of the heat exchanger is generally composed of a compromise between the evaporator and the condenser as described above, the configuration of the refrigerant inlet pipe portion at the time of the evaporator becomes a configuration with a small pipe cross-sectional area. Most of the pressure loss is concentrated in this area. The graph in FIG. 12 shows the pressure movement when the refrigerant passes through the heat exchanger. After flowing from the inlet pipe, the pressure suddenly increases in the narrow tube section of the 1-pass part, 2-pass part, and 3-pass part. After that, since the number of passes is increased (six passes), the flow velocity is lowered and the pressure is not lowered so much.

  That is, it can be seen that most of the pressure loss is concentrated in about one third of the vicinity of the heat exchanger inlet. Such concentration of pressure loss is to secure the condensing capacity when it becomes a condenser as already described, and has been inevitable as a heat exchanger. However, in the present invention, as shown in FIG. 3, by applying the variable path system only to the heat exchanger inlet portion at the time of the evaporator, the refrigerant flow velocity around the inlet is reduced as shown by the dashed line in FIG. Therefore (in the case of this figure, the inlet portion has five paths), the pressure loss is greatly reduced.

  By setting it as such a structure, the reduction of the power consumption by reduction of the compressor motive power at the time of heating operation is realizable. In addition, since three heat exchanger blocks are connected in series during the cooling operation, the heat transfer efficiency is improved, the pressure is reduced, and the subcooling (supercooling degree) is increased, so that the efficiency of the heat exchanger is increased = the power consumption is reduced. it can.

  FIG. 9 shows a detailed configuration example of the heat exchanger in this case. Conventionally, an outdoor heat exchanger is generally configured as shown in FIG. During the condensing operation, the refrigerant flows in from the 6-pass leeward side, and the pipe cross-sectional area is decreased in inverse proportion to the increase in the refrigerant density in order of 3 passes, 2 passes, and 1 pass. In the evaporating operation, the pipe cross-sectional area is increased in inverse proportion to the refrigerant density. The vicinity of the evaporator inlet is 100% liquid refrigerant in the condensing operation, but about 80% in the evaporating operation. Only liquid ratio. The low pressure refrigerant has a large pressure loss, and the optimum piping configuration has not been achieved. FIG. 9B shows an embodiment of the present invention, in which only the evaporator inlet section is divided into three blocks, and a circuit and check valve that flows from between the refrigerant inlets to the heat exchanger block (four lines in one row) at the bottom of the heat exchanger. The refrigerant flows into two heat exchanger blocks (each in two rows and eight), and then flows through the heat exchanger block in five passes and then merges through the check valve. Thereafter, the refrigerant branches again and flows out to the upper 6-pass block.

  As a result, when the evaporator is used, the number of passes can be five, the refrigerant flow rate can be reduced to about one third, and the pressure loss can be reduced to about one ninth. At the same time, unlike the conventional example where only two passes can be arranged during the condensation operation, the pressure loss during the evaporation operation does not increase even if four are arranged, so the heat exchanger capacity during the condensation operation is increased. be able to.

  FIG. 10 summarizes the heat exchanger capacity and pressure loss in each operation mode in this case. It can be seen that the pressure loss is reduced in proportion to the circulation rate while securing substantially the same performance as the evaporation capacity as compared with the conventional configuration. Since this pressure loss directly affects the power of the compressor having the largest power consumption in the refrigeration cycle apparatus, the power consumption of the entire apparatus can be greatly reduced. As for the condensing capacity, the pressure loss is slightly increased, but the influence as an absolute value is small, and the condensing capacity is further increased. Therefore, the power of the compressor can be reduced by reducing the refrigerant circulation amount, and the power consumption of the entire refrigeration cycle apparatus can be reduced.

(Embodiment 2)
FIG. 4 shows a case where the present invention is applied to the entire outdoor heat exchanger. Even when the entire heat exchanger is blocked, the same effect can be obtained. 5 and 6 show a case where the present invention is applied to an indoor heat exchanger. In such a case, the effects already described can be obtained.

  Furthermore, an electromagnetic two-way valve can be used instead of the check valve if the problem of manufacturing cost is not taken into consideration. In that case, it is necessary to consider the power consumption of the solenoid valve.

  FIG. 11 is an example showing a specific arrangement configuration of the heat exchanger to be blocked in the second embodiment of the present invention. FIG. 11A shows the same number (four) in order to make the refrigerant flow rate of each block uniform in the case of the embodiment of the present invention. However, in FIG. 11B, the number is changed. In this case, the amount of refrigerant passing through each copper tube is likely to be unbalanced, and it becomes difficult to sufficiently exhibit the heat exchanger performance. The evaporating capacity shown at the bottom of the figure is a numerical value when compared at the rated operation, and a difference of about 18 W occurs.

(Embodiment 3)
FIG. 13 is a diagram showing a configuration of a heat exchanger block according to Embodiment 3 of the present invention. At the time of the evaporator, the refrigerant flows in as a gas-liquid two-phase low-pressure refrigerant from the heat exchanger inlet 5 at the bottom of the figure. Since the refrigerant cannot pass through the check valve 25, it passes through the check valve 21, branches into two heat exchanger blocks 22 and 23, passes in parallel, and exchanges heat. Thereafter, the refrigerant that has passed through the heat exchanger block 22 merges with the refrigerant that has passed through the heat exchanger block 23 and the check valve 20, and flows out through the heat exchanger outlet pipe 4 to the downstream refrigeration cycle. At the time of the condenser, since it flows as a high-pressure gas-liquid two-phase or gas refrigerant from the heat exchanger outlet pipe 4 and cannot pass through the check valve 20, heat is exchanged while passing through the entire heat exchanger block 22, and the check valve 21 Therefore, heat exchange is performed while passing through the heat exchanger block 23. Since the check valve 25 can pass, it flows out to the refrigeration cycle apparatus on the downstream side through the heat exchanger inlet 5.

  Since the refrigerant flows in this way, the pressure loss is reduced by two-pass parallel flow at the time of the evaporator, and the heat transfer coefficient is increased by serial connection at the time of the condenser, so that the efficiency as a heat exchanger can be maximized. It has become. The third embodiment of the present invention has one check valve more than the first embodiment, but the number of passes at the time of the evaporator is reduced, so the performance deteriorates due to a decrease in heat transfer coefficient. This is an effective configuration when the is large.

(Embodiment 4)
FIG. 14 shows a case where the number of passes in the heat exchanger of the refrigeration cycle apparatus according to the second embodiment of the present invention, particularly when the evaporator is three passes. FIG. 16 is a refrigeration cycle diagram showing a case where the heat exchanger of FIG. 14 is applied to an outdoor heat exchanger. Embodiment 4 of the present invention assumes an air conditioner having a cooling standard capacity of 4.0 kW and a heating standard capacity of 5.0 kW using a heat exchanger having a pipe diameter of 7 mm.

  In FIG. 16, the high-temperature and high-pressure gas refrigerant discharged from the compressor 1 (in the present invention, R410A is used as the refrigerant) flows to the indoor heat exchanger 12 through the four-way valve 2 and is condensed. The liquefied refrigerant is decompressed by the expansion valve 7 or capillary mounted in the outdoor unit, and enters the outdoor heat exchanger as a gas-liquid two-phase refrigerant. In the conventional heat exchanger, the inlet portion has a narrow channel cross-sectional area such as one pass or two passes. However, in the present invention, as shown in FIG. 16 or FIG. Thus, the refrigerant flows through three passes.

  This means that the cross-sectional area of the flow path is tripled, the refrigerant flow rate is 1/3, and the refrigerant pressure loss is 1/9 compared to the case of one pass. However, reducing the refrigerant flow rate simultaneously reduces the heat transfer coefficient on the refrigerant side, so it was not clear how far the optimum number of passes was increased. (A) and (b) of FIG. 17 have shown the heat exchanger structure at the time of using 5 passes and 7 passes. FIG. 18 shows a comparison of outdoor heat exchanger capacity (evaporator capacity) when rated heating operation (corresponding to a heating capacity of 5 kW) is performed in these three configurations. As can be seen from FIG. 18, when the number of passes is changed from 3 to 5, the performance is improved, but when the number is 7 passes, the performance is reduced.

  FIG. 18 also shows the refrigerant Reynolds number at the inlet of the evaporator at the same time, but it can be seen that only in the case of 7 passes, it is not a turbulent region but a turbulent transition region. FIG. 15 is a table showing general distinction (relationship with Reynolds number) of laminar flow, turbulent flow, and turbulent transition regions extracted from Non-Patent Document 1. In this case, no clear relational expression has been found for the Nusselt number representing the refrigerant heat transfer coefficient (from Non-Patent Document 1).

  In other words, FIG. 15 and FIG. 18 show that if the number of passes is increased to 7 passes, a turbulent flow transition region is obtained, and there is a possibility that extreme heat exchanger capacity may be lowered. Further, FIG. 15 also shows that when the optimum Reynolds number, that is, the number of passes in the turbulent region, is the lowest Reynolds number in the turbulent region, the pressure loss is low and the decrease in heat transfer coefficient is most suppressed. ing.

To explain this point in more detail, as described above, an increase in the number of passes causes a decrease in the refrigerant flow rate = a decrease in the Reynolds number, but a heat transfer coefficient decreases in proportion to the Reynolds number = the 0.8th power of the refrigerant flow rate. To do. However, since the pressure loss of the refrigerant increases in proportion to the square of the refrigerant flow rate, it is more efficient to reduce the refrigerant flow rate to the maximum. In other words, it is better to reduce the refrigerant flow speed by increasing the number of passes and reduce the pressure loss as much as possible.However, if the transition to the laminar flow region is lost and the heat transfer coefficient decreases rapidly, the most in the turbulent flow region. The most efficient configuration is to determine the number of passes so that the low Reynolds number is 3000.

  Form 4 of the present invention is not limited to the pipe diameter of the heat exchanger and the size of the cooling / heating capacity.

  For example, when the heating capacity (evaporator capacity) is 5.0 kW, the heat exchanger pipe has a diameter of 6.35 mm (2 split pipes), the number of paths when used as an evaporator is 7 paths or less, and the diameter is 5 mm If the number of passes when used as an evaporator is 12 or less and the diameter is 7.94 mm (2.5 split pipes) or more, the number of passes when used as an evaporator is 4 or less and the refrigerant Reynolds described above is used. The number is 3000 or more, and an optimum configuration can be obtained.

(Embodiment 5)
In the configuration in which the conventional variable path system is not installed, the number of pipes in one path portion can be freely selected. However, when the variable path system is installed as in the present invention, the number of pipes in the evaporator is limited to some extent by the length of one path part. For example, assuming that the number of one path is eight and the number of paths at the heat exchange inlet in the evaporator is five, this heat exchanger block requires 8 × 5 = 40 pipes.

  In general, the number of pipes for heat exchange can be charged only about 60 to 70 in two rows of heat exchanges using φ7 mm pipes due to restrictions such as the pitch between pipes, cost, and ventilation resistance. If 40 of these are fixed, it becomes difficult to optimize the heat exchanger as a whole due to the deterioration of the diversion performance and the storage capacity of the piping. In other words, it is difficult to improve the efficiency of the heat exchanger unless the number of pipes in one pass portion is optimized.

  Then, Embodiment 5 of this invention aims at clarifying the optimal structure as a condenser. As an example, a 7 mm outdoor heat exchanger is assumed. When the variable path configuration is adopted at the outlet of the condenser, the configuration is as shown in FIGS. 14 and 16. However, in order to improve the condensing capacity, the refrigerant having a high liquid density at the outlet of the condenser has a pressure loss. It is necessary to increase the heat exchange capacity by decreasing the pipe cross-sectional area and increasing the refrigerant flow rate while not increasing so much.

  The optimal outlet configuration, especially the number of pipes and the location of one pass, is important, but there is a table comparing the heat exchanger capacity when the number of passes is 2, 4, 6, and 8. FIG. The number of passes at the time of the evaporator is all unified at 5 passes. As can be seen from this table, the condensing performance varies depending on the number of pipes in one pass, and the performance is improved relatively for 4, 6, and 8 pipes with the number of pipes being increased compared to the case of two 1 pass. You can see that

  However, even if the number is increased, the capacity does not improve in proportion to the number, but the upper limit is reached around 4 to 6, and instead, the pressure loss rapidly increases. Since the increase in the pressure loss of the refrigerant causes an increase in the power of the compressor 1 and causes a large reduction in the efficiency of the refrigeration cycle apparatus, it can be understood that the number of one-pass number is four or six for the highest condensing efficiency.

  Therefore, while the number of one pass is four, eight condenser outlet pipes including one pass portion are arranged on the windward side with better heat exchanger efficiency, and the anti-AC effect is used. FIG. 20A shows an example in which the pipe is not arranged on the windward side, and FIG. 20B shows a configuration of a heat exchanger in which eight outlet-side pipes including four 1-pass are arranged on the windward side. FIG. 21 shows the difference in condensing capacity between the case where the condenser outlet is disposed on the windward side of the heat exchanger and the case where the condenser outlet is not. 20 and 21, it can be seen that arranging the condenser outlet pipe including four one-pass on the windward side is the configuration in which the heat exchanger efficiency is exhibited to the maximum even as the condenser.

(Embodiment 6)
FIG.20 (b) shows the structure of the heat exchanger of the refrigerating-cycle apparatus of Embodiment 6 of this invention. The example which made the entrance part at the time of the evaporator of a heat exchanger 5 passes is shown.

  First, in the variable path configuration composed of three blocks, the refrigerant passes through the first heat exchanger block 42 having four pipes in two passes, and dissipates heat. Thereafter, the refrigerant enters the second heat exchanger block 43 arranged on the windward side, and the refrigerant passes through two pipes arranged two by two. Usually, the subcool is removed and liquefied in the middle of the pipe. After joining, the third heat exchanger block 46, which is a one-pass part, enters the third heat exchanger block 46, and the subcooling is sufficiently ensured by four pipes whose heat transfer rate is increased by increasing the refrigerant flow rate in the windward and one pass.

  Next, the evaporator will be described. In the evaporator, the refrigerant flows in the opposite direction to that in the condenser, and the refrigerant flowing in from the refrigerant inlet pipe 47 flows into the four pipes of the third heat exchanger block 46 that is a one-pass portion in the condenser. After passing through the check valve 21, the refrigerant flow 2 flowing to the first and second heat exchanger blocks 42 and 43 is divided by the branch pipe 48.

  First, the refrigerant flow is branched by a branch pipe 49 and divided into first and second heat exchanger blocks 42 and 43. The first heat exchanger block 22 evaporates through a total of eight pipes with two 2 passes, merges at the junction 50 without passing through the check valve, and flows to the next outdoor heat exchanger 3. To go. The second heat exchanger block 43 is composed of a total of four pipes with two two paths, where the refrigerant evaporates and merges with the refrigerant flow evaporated in the third heat exchanger block 46. After the merged refrigerant passes through the check valve 20, it merges at the merge section 50, and then flows to the next outdoor heat exchanger 3.

  As can be seen from this, the refrigerant passes through the check valve twice only when passing through the second heat exchanger block 23. If the check valve is a high-pressure refrigerant such as R410A, the pressure loss is at a negligible level, but it is not zero. Therefore, the refrigerant passing through the second heat exchanger block 43 tends to decrease in circulation amount as much as it passes through the check valve. In Embodiment 3 of the present invention, only the second heat exchanger block 43 in which the refrigerant circulation amount is likely to decrease is reduced to two pipes in the refrigerant flow direction as compared with the other heat exchanger blocks. As a result, the path balance is not greatly deteriorated in spite of the multi-pass heat exchanger of 5 passes in all regions from the low frequency during heating operation to the rated frequency to the maximum frequency during heating low temperature.

  Further, from the viewpoint of piping design, as described above, the refrigerant entering from the evaporator inlet is branched into two, but since the current heat exchanger is φ7 mm, the first heat exchanger block 42 which is a one-pass portion is used. This pipe is φ7 mm. On the other hand, the second and third heat exchanger blocks 43 and 46 have a total of four passes and therefore require four times the amount of refrigerant circulation. To obtain this amount of refrigerant circulation, a φ14 mm pipe is required. Is done. However, such a thick pipe may not be accommodated due to the structure of the outdoor unit, resulting in an increase in manufacturing cost and a decrease in performance. Therefore, in the third embodiment of the present invention, the second heat exchanger block 43 reduces the necessary amount of refrigerant circulation by reducing the number of pipes to two, as already described. Even when a φ9.54 mm pipe is used), storage efficiency can be improved while suppressing the occurrence of refrigerant drift and pressure loss.

(Embodiment 7)
FIG. 22 shows a refrigeration cycle configuration diagram of Embodiment 7 of the present invention. In this embodiment, the heat exchanger (φ7.00 mm pipe, outlet 6-path branch) equipped with the variable path system already described is described as an outdoor heat exchanger. However, it is not limited to the illustrated example.

As shown in FIG. 22, in order to reduce the evaporator pressure loss during heating and improve the heat exchanger efficiency, the evaporator inlet portions (22, 23, 24) are multipassed. Since the gas-liquid two-phase refrigerant decompressed and expanded by the expansion valve 7 branches and flows in five paths as shown in the figure, almost no refrigerant pressure loss occurs in this portion. Once merged, it branches again into 6 passes, but no refrigerant pressure loss occurs in this part (heat exchanger 3). However, there are two types of cross-sectional areas in the pipe as seen from the refrigerant side as the inside of the heat exchanger, existing in the minimum pipe cross-sectional area (A) and the latter half (3) that exist in the first five passes (22, 23, 24). There is a maximum cross-sectional area (B). For example each case of using the φ7.00mm tube becomes ^{A = 171mm 2, B = 205mm} 2 as pipe of the outdoor heat exchanger.

However, the refrigerant flowing out of the heat exchanger increases the flow velocity, passes through the suction pipe 8 and returns to the accumulator in the compressor 1 suction portion. At this time, the accumulator and the four-way valve 2 arranged in the middle generally use a quadrant pipe (φ12.70 mm) (in a general-use RAC air conditioner), and the surrounding suction pipe 8 is inevitably also used. A quadrant (φ12.70 mm) is used. Here, when the internal cross-sectional area (C) of the quadrant (φ12.70 mm) is calculated, it is only C = 119 mm ² .

  Compared to the cross-sectional areas A and B inside the heat exchanger, the heat exchanger (A and B) has only about a half, and the suction pipe portion where the volume flow rate of the refrigerant increases and pressure loss tends to increase. Pipes with a larger cross-sectional area should be used, but they are getting thinner. In this case, a large pressure loss of the refrigerant may occur as a disadvantage. In addition, in Embodiments 1 to 6 of the present invention, although the pressure loss inside the heat exchanger is reduced, there is a portion where the resistance in the thin suction pipe is rate limiting and a sufficient effect cannot be obtained. Therefore, by optimizing the relationship between the cross-sectional areas A, B, and C, a configuration that can maximize the heat exchanger efficiency is clarified.

  FIG. 23 compares the pipe cross-sectional area of the outdoor heat exchanger shown in FIG. 22 from the heat exchanger inlet to the compressor. As can be seen from this figure, the dryness of the refrigerant increases compared to the heat exchanger inlet, and the suction pipe 8 from the heat exchanger outlet to the compressor having a high volumetric flow velocity has decreased from the inside of the heat exchanger. ing.

Here, when the pressure loss of the refrigerant is Y and the volume flow rate of the refrigerant is X, the following relationship is satisfied. Y = X ² (Formula 1)
Further, when the cross-sectional area in the tube is Z, the following relationship is satisfied.
X = (1 / Z) (Formula 2)
In other words, the pressure loss of the refrigerant has increased about three times by the suction pipe. However, the suction pipe 8 is in the outdoor unit machine room, and it is necessary to make a comprehensive determination in consideration of limited storage space, an increase in manufacturing costs, an increase in the amount of refrigerant retained in the pipe, and the like. is there.

  Therefore, in the case where the suction pipe 8 is a quadrant (φ12.75 mm), a quadrant (φ15.875 mm), and a 6-segment pipe (φ19.05 mm), the copper pipe mass (g) per 1.6 m, the retained refrigerant amount ( g) The refrigerant pressure loss (MPa) was compared by calculating the heating rated operating condition (Q = 5000 W, refrigerant R410A).

  From this calculation result, it is possible to suppress an increase in the amount of refrigerant retained and the mass of copper while sufficiently reducing the pressure loss if the quintuple pipe is used. Not only the effect does not increase so much at 6 or more pipes, but it cannot be said that the production problems such as bending work as a piping block are greatly effective.

  FIG. 24 shows the result of comparison using a pipe having a cross-sectional area equivalent to a quintuple pipe with the same amount of refrigerant in the actual machine. As can be seen from this figure, the refrigerant pressure loss was reduced by 0.006 MPa. At this time, the effect of reducing the refrigerant pressure loss of the suction pipe 8 appears at the inlet of the heat exchanger due to the rate-determining condition of the refrigeration cycle, the compressor suction pressure is the same, the compressor input hardly increases, and the evaporator temperature When the temperature difference from the air side increased (about 2%), the capacity increased by about 20 W (about 0.4%), and the efficiency of the refrigeration cycle was improved.

  In addition, during the heating intermediate capacity operation (Q = 2500 W), the refrigerant pressure in the compressor suction portion increased by 0.003 MPa, and the efficiency of the refrigeration cycle was improved. Furthermore, under the low-temperature heating conditions (outdoor low-temperature conditions: 2/1 ° C.), the compressor operates at 100 Hz or more, so the efficiency improvement effect is large, the compressor suction pressure rises, and the heating capacity is about 90 W while the compressor input remains the same. Increased (about 1%).

  Summarizing the above results, the cross-sectional area C of the suction pipe 8 is larger than the maximum cross-sectional area (B).

The maximum cross-sectional area in the pipe (B) × 1.2> The cross-sectional area in the pipe C ≧ the maximum cross-sectional area in the pipe (B) × 0.8 was optimal. Also, considering the relationship to the minimum cross-sectional area (A),

Therefore, as an optimum configuration, the cross-sectional area in the pipe C ≧ the cross-sectional area in the minimum pipe (A) × 1.1.

In addition, when a φ5.00 mm tube is used as a heat exchanger, 12 passes are required to suppress the increase in the pressure loss of the heat exchanger to the present level. By using a quintuple pipe as the connecting pipe 8, the maximum cross-sectional area (B) x 1.2> the cross-sectional area C within the pipe C ≥ the maximum cross-sectional area (B) x 0.8
Pipe cross-sectional area C ≧ Minimum pipe cross-sectional area (A) × 1.1
It can be set as the structure which satisfies these. In the case of 8 passes, the quadrant is optimal and satisfies the above relational expression.

(Embodiment 8)
FIG. 25 shows a refrigeration cycle configuration diagram according to Embodiment 8 of the present invention. In the eighth embodiment, unlike the above-described seventh embodiment, only the pipes from the outdoor heat exchanger 3 to the four-way valve 2 in the suction pipe are divided into five, and the four-way pipe from the four-way valve 2 to the compressor 1 is a quarter pipe as before. This is an example. FIG. 26 shows the cross-sectional area in the tube at this time. Since only the five-way pipe from the heat exchanger outlet to the four-way valve is used, the cross-sectional area increases from 119 to 188 mm ² . By changing only this portion, the refrigerant pressure loss reduction effect in the evaporator can be obtained as shown in FIG.
The configuration at this time satisfies the following relational expression.

Maximum pipe cross-sectional area (B) × 1.0> Pipe cross-sectional area C ≧ Maximum pipe cross-sectional area (B) × 0.6, and Pipe cross-sectional area C ≧ Minimum pipe cross-sectional area (A) × 1.1
According to the eighth embodiment, although the effect is small as compared with the seventh embodiment, there is an advantage that the configuration of the piping block is easy and the manufacturing problems are small.

(Embodiment 9)
FIG. 28 shows a refrigeration cycle configuration diagram in Embodiment 9 of the present invention. Unlike Embodiments 7 and 8 described above, Embodiment 9 creates a bypass circuit 10 that bypasses the refrigerant that has flowed out of the outdoor heat exchanger 3 without passing through the suction pipe 8, and only during heating. By opening the electromagnetic two-way valve 9 (blocking during cooling), the effect of substantially increasing the cross-sectional area of the suction pipe is obtained. Specifically, a predetermined effect can be obtained by using a 2.5 branch pipe (φ7.94 mm, cross-sectional area 44.7 mm ² ) as a bypass circuit.

  In FIG. 28, the bypass is directly bypassed from the outlet of the heat exchanger to the compressor. However, although the efficiency improvement effect is reduced, it is possible to connect to the front and rear of the four-way valve due to structural limitations.

  As described above, the refrigeration cycle apparatus according to the present invention can be configured in series in the case of a condenser and in parallel in the case of an evaporator with a simple configuration, and can maximize the performance of the heat exchanger. Therefore, it can be applied to various refrigeration cycle apparatuses including an air conditioner.

DESCRIPTION OF SYMBOLS 1 Compressor 2 Four-way valve 3 Outdoor heat exchanger 4 Heat exchanger exit pipe 5 Heat exchanger inlet pipe 7 Expansion valve 8 Suction piping 9 Electromagnetic two-way valve 10 Bypass circuit 12 Indoor heat exchanger 20 Check valve 21 Reverse Stop valve 22 First heat exchanger block 23 Second heat exchanger block 24 Third heat exchanger block 25 Check valve 42 2 pass part 43 2 pass part 44 1 pass part 46 1 pass part Heat exchanger 47 Refrigerant inlet pipe for evaporator 48 Branch pipe 49 Branch pipe 50 Junction section 101 Pass 1
102 Pass 2
103 On-off valve 1
104 On-off valve 2
105 On-off valve 3

Claims (16)

At least a compressor, a four-way valve, an outdoor heat exchanger, a throttling device, and a refrigerant circuit formed by sequentially connecting an indoor heat exchanger,
At least one of the indoor heat exchanger and the outdoor heat exchanger, in whole or in part the heat exchanger, the odd number of heat exchanger blocks arranged in parallel, the heat exchanger block inlet at the evaporator A first branch pipe directly connected to each refrigerant inlet , and a second branch pipe directly connected to each refrigerant outlet serving as an outlet when the heat exchanger block is an evaporator ,
The heat exchanger includes a heat exchanger inlet pipe that serves as an inlet when the heat exchanger is a condenser, and a heat exchanger outlet pipe that serves as an outlet when the heat exchanger is a condenser.
The heat exchanger inlet pipe is connected to the second branch pipe between a heat exchanger block arranged on the outermost side and a second heat exchanger block adjacent thereto,
The heat exchanger outlet pipe is connected to the first branch pipe between the heat exchanger block arranged on the outermost side opposite to the heat exchanger inlet pipe and the adjacent heat exchanger block. ,
Wherein the second branch pipe, the check valve to flow only in the direction of the heat exchanger inlet pipe is disposed between the even-numbered of the heat exchanger outlet pipe side and an adjacent odd-numbered heat exchanger block ,
In the first branch pipe, a check valve that flows only in the direction opposite to the heat exchanger outlet pipe is between the even number of the heat exchanger block and the odd number of the adjacent heat exchanger outlet pipe. A variable path heat exchanger located in
A refrigeration cycle apparatus , wherein the Reynolds number of the refrigerant at the inlet of the variable path heat exchanger is 3000 or more when the variable path heat exchanger is an evaporator during rated operation . At least a compressor, a four-way valve, an outdoor heat exchanger, a throttling device, and a refrigerant circuit formed by sequentially connecting an indoor heat exchanger,
At least one of the indoor heat exchanger and the outdoor heat exchanger, in whole or in part the heat exchanger, and two heat exchangers blocks arranged in parallel, are arranged in parallel with the heat exchanger block a first branch pipe to which the heat exchanger is directly connected with piping including a check valve Ru passed through a refrigerant only in the exit direction at the time of condensation, the refrigerant inlet the heat exchanger block is an inlet at the evaporator, respectively And a second branch pipe in which the heat exchanger block is directly connected to a refrigerant outlet that is an outlet at the time of the evaporator ,
The heat exchanger inlet pipe is connected to the second branch pipe between a heat exchanger block disposed farthest from the heat exchanger outlet pipe and the adjacent heat exchanger block,
The heat exchanger outlet pipe is connected to the first branch pipe between a heat exchanger block disposed on the opposite side of the heat exchanger inlet pipe and a pipe including the check valve adjacent thereto,
Wherein the second branch pipe, the check valve to flow only in the direction of the heat exchanger inlet pipe is disposed between the second of said heat exchanger outlet pipe side adjacent to the first heat exchanger block ,
Wherein the first branch pipe, said refrigerant outlet pipe check valve to flow only in the opposite direction, is arranged between the pipe comprising the second and the check valve in which the adjacent heat exchanger block Variable path heat exchanger,
A refrigeration cycle apparatus , wherein the Reynolds number of the refrigerant at the inlet of the variable path heat exchanger is 3000 or more when the variable path heat exchanger is an evaporator during rated operation . The refrigeration cycle apparatus according to claim 1 or 2, wherein an electromagnetic two-way valve is arranged instead of the check valve. The refrigeration cycle apparatus according to any one of claims 1 to 3, wherein the variable path heat exchanger has the heat exchanger block arranged at a refrigerant inlet when an evaporator is used. . In the refrigeration cycle apparatus according to any one of the claims 1-4, refrigeration cycle apparatus is characterized in that the same distribution pipe length of the heat exchanger block. The refrigeration cycle apparatus according to claim 1 or 2, wherein the variable path heat exchanger has a check valve having a valve diameter equal to or greater than a check valve disposed upstream of the refrigerant flow during the evaporator. A refrigeration cycle apparatus having a valve disposed downstream. When the pipe of the variable path heat exchanger has a diameter of 7 mm, the number of passes when the heat exchanger is used as an evaporator is 6 or less, and when the diameter is 6.35 mm, the number of passes when used as an evaporator is When the number of passes is 7 or less and the diameter is 5 mm, the number of passes when used as an evaporator is 12 or less, and when the diameter is 7.94 mm or more, the number of passes when used as an evaporator is 4 or less. The refrigeration cycle apparatus according to any one of claims 1 to 6 . When the variable path heat exchanger is used as a condenser, the number of pipes in one pass section is four or six when the pipe has a diameter of 7 mm, and eight pipes at the outlet of the condenser including the one pass section are heated. The refrigeration cycle apparatus according to any one of claims 1 to 7, wherein the refrigeration cycle apparatus is disposed on the windward side of the exchanger. The variable pass heat exchanger, piping in the direction of flow of the refrigerant of the block on either or remote inside by piping the number in the direction of flow of the refrigerant in both heat exchanger block of the heat exchanger block of the outer ends The refrigeration cycle apparatus according to any one of claims 1 to 8, wherein the number is small. The outdoor heat exchanger is the variable path heat exchanger, and the suction pipe connecting the outdoor heat exchanger and the four-way valve is (maximum cross-sectional area of the pipe in the outdoor heat exchanger) × 1.2> (suction (Cross sectional area of piping) ≥ (Maximum cross sectional area of piping in outdoor heat exchanger) x 0.8
The refrigeration cycle apparatus according to any one of claims 1 to 9, wherein the condition is satisfied. The outdoor heat exchanger is the variable pass heat exchanger,吸接piping connecting the suction part of the compressor via the four-way valve and the outdoor heat exchanger, except the four-way valve and connecting portion (Maximum cross-sectional area of piping in outdoor heat exchanger) × 1.0> (cross-sectional area of suction pipe) ≧ (maximum cross-sectional area of piping in outdoor heat exchanger) × 0.6
The refrigeration cycle apparatus according to any one of claims 1 to 9, wherein the condition is satisfied. Wherein an outdoor heat exchanger wherein the variable pass heat exchanger, (maximum cross-sectional area of the outdoor heat exchanger pipe)吸接pipe that connects the four-way valve and the outdoor heat exchanger × 1.0> ( (Cross sectional area of suction pipe) ≧ (Minimum cross sectional area of pipe in outdoor heat exchanger) × 1.1
The refrigeration cycle apparatus according to any one of claims 1 to 9, wherein the condition is satisfied. 5 min tube as the吸接pipe when the pipe of the variable pass heat exchanger is diameter 7m m 及 beauty six passes, in the case of 4-pass characterized by using a 4-minute tube as the吸接pipe The refrigeration cycle apparatus according to any one of claims 10 to 12 . Claims, characterized in that the pipe of the variable pass heat exchanger 5 minutes tube as the吸接pipe when the diameter 5 m m 及 beauty 12 pass, with quarter pipe as said吸接pipe if it is eight-pass Item 13. The refrigeration cycle apparatus according to any one of Items 10 to 12 . Maximum wherein an outdoor heat exchanger wherein the variable pass heat exchanger,吸接piping that connects the four-way valve and the outdoor heat exchanger (cross-sectional area of吸接piping) <(the outdoor heat exchanger in the pipe (Cross sectional area) x 0.8
In the case conditions are satisfied, according to the piping directly connected to the suction portion of the four-way valve or the compressor from the outlet portion of the outdoor heat exchanger is disposed, and having a solenoid two-way valve in the middle part Item 10. The refrigeration cycle apparatus according to any one of Items 1 to 9 . The outdoor heat exchanger is the variable pass heat exchanger,吸接piping connecting the suction part of the compressor via the four-way valve and the outdoor heat exchanger, except the four-way valve and connecting portion (Cross sectional area of suction pipe) <(Maximum cross sectional area of pipe in outdoor heat exchanger) x 0.6
In the case conditions are satisfied, according to the piping directly connected to the suction portion of the four-way valve or the compressor from the outlet portion of the outdoor heat exchanger is disposed, and having a solenoid two-way valve in the middle part Item 10. The refrigeration cycle apparatus according to any one of Items 1 to 9 .