JP2017203620A - Air conditioner - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide an air conditioner including a heat exchanger of high performance.SOLUTION: Heat transfer pipes 26b1, 26b2 extend to one end portion 21A from the other end portion 21B in an intermediate column L2 respectively, and are coupled at one end portion 21A to be heat transfer pipes 26c1, 26c2, and the heat transfer pipes 26c1, 26c2 are constituted to make one reciprocation between one end portion 21A and the other end portion 21B in an uppermost column L1. Heat transfer pipes 26b3, 26b4 extend to one end portion 21A from the other end portion 21B in the intermediate column L2 respectively, and are coupled at one end portion 21A to be heat transfer pipes 26c3, 26c4, and the heat transfer pipes 26c3, 26c4 are constituted to make one reciprocation between one end portion 21A and the other end portion 21B in the uppermost column L1. The heat transfer pipe 26c2 extending to one end portion 21A from the other end portion 21B and the heat transfer pipe 26c4 extending to one end portion 21A from the other end portion 21B are arranged in adjacent to each other.SELECTED DRAWING: Figure 22

Description

  The present invention relates to an air conditioner including a highly efficient heat exchanger.

  In the heat exchanger of an air conditioner, the refrigerant flow rate in the heat transfer tube is optimized to optimize the balance between the pressure loss on the refrigerant side and the heat transfer coefficient, thereby improving the performance of the heat exchanger. That is, in order to exhibit the heat exchanger performance, the design is performed in consideration of the flow path inner diameter of the heat transfer tube and the number of refrigerant flow paths.

  In the heat exchanger in which three rows of heat transfer tubes are arranged, the heat transfer tube diameter D1 on the most windward side is made the smallest and the range of D1 = 3 to 4 mm, and the heat transfer tube diameter D2 in the middle, leeward It has been proposed that the relationship with the side heat transfer relationship D3 is D1 <D2 = D3, 4 mm ≦ D3 ≦ 10 mm, and 0.6 ≦ D1 / D2 <1 (see, for example, Patent Document 1). With this configuration, heat exchange performance is improved while suppressing an increase in pressure loss.

  In addition, the heat transfer tube of the fin connected to the liquid side distributor or the gas side distributor is reciprocated once and branched and connected to two heat transfer tubes of adjacent fins, and one path of the heat transfer tube is constituted of 2 reciprocations. This has also been proposed (see, for example, Patent Document 2). With this configuration, increasing the liquid-side flow rate increases the pressure loss in the heat transfer tube, while improving the surface heat transfer coefficient.

JP 2011-122819 A JP 2010-78287 A

  However, in the configuration described in Patent Document 1, a manufacturing apparatus for each heat transfer tube diameter difference is required, so the number of manufacturing steps for the heat exchanger increases. Furthermore, in the windward row where the small diameter heat transfer tubes are arranged, the heat transfer area inside the heat transfer tubes is reduced, and the performance of the overall heat exchanger is lowered.

  Further, when the heat exchanger disclosed in Patent Document 2 is operated by a condenser, the heat conduction through the fins between the upper and lower adjacent heat transfer tubes is affected by the temperature change in the supercooling region. As a result, heat loss in the supercooling region occurs due to internal heat exchange.

  The present invention has been made in view of the above-described problems, and an object thereof is to provide an air conditioner including a high-performance heat exchanger.

  In order to solve the above problems, an air conditioner according to an aspect of the present invention is an air conditioner including a plurality of heat transfer tubes through which a refrigerant flows and having a heat exchanger that exchanges heat with air. The heat exchanger has one end and the other end, and the plurality of heat transfer tubes have the one end and the other end in a state of being aligned in a direction intersecting with the direction of air flow. The rows of the plurality of heat transfer tubes arranged so as to reciprocate and arranged in the intersecting direction are arranged so as to be arranged in at least two rows along the direction in which the air flows, and the two rows A first row located at the uppermost stream in the flow direction and a second row located next to the first row in the air flow direction, wherein the plurality of heat transfer tubes are mutually connected in the second row. Adjacent first heat transfer tube and second heat transfer tube, the first heat transfer tube and the second heat transfer tube The heat transfer tubes respectively extend the second row from the other end portion to the one end portion, and are combined at the one end portion to become a first connecting tube, and the first connecting tube is connected to the one end at the uppermost row. Each of the plurality of heat transfer tubes further includes a third heat transfer tube and a fourth heat transfer tube that are adjacent to each other in the second row, The third heat transfer tube and the fourth heat transfer tube are disposed adjacent to the first heat transfer tube and the second heat transfer tube, and extend from the other end to the one end, respectively, in the second row. The second coupling pipe is coupled at the one end, and the second coupling pipe is configured to reciprocate between the one end and the other end in the uppermost row, and A portion of one coupling tube extending from the other end to the one end, and the front The extending portion to the one end from the other end of the second coupling tube, are arranged next to each other.

  In order to solve the above problems, an air conditioner according to an aspect of the present invention is an air conditioner including a plurality of heat transfer tubes through which a refrigerant flows and having a heat exchanger that exchanges heat with air. The heat exchanger has one end and the other end, and the plurality of heat transfer tubes have the one end and the other end in a state of being aligned in a direction intersecting with the direction of air flow. The rows of the plurality of heat transfer tubes arranged so as to reciprocate and arranged in the intersecting direction are arranged so as to be arranged in at least two rows along the direction in which the air flows, and the two rows A first row located at the uppermost stream in the flow direction and a second row located next to the first row in the air flow direction, wherein the plurality of heat transfer tubes are mutually connected in the second row. Adjacent first heat transfer tube and second heat transfer tube, the first heat transfer tube and the second heat transfer tube Each of the heat transfer tubes extends from the other end portion to the one end portion of the second row, and is combined at the one end portion to form a single heat transfer tube, and the one heat transfer tube is the first row, It is comprised so that it may reciprocate between one end part and the said other end part, and the said refrigerant | coolant is a refrigerant | coolant which contains 70 weight% or more of R32 or R32.

ADVANTAGE OF THE INVENTION According to this invention, an air conditioner provided with a high performance heat exchanger can be provided.

1 shows a refrigeration cycle of an air conditioner according to the present invention. It is the figure which showed the refrigerating cycle at the time of heating operation when R410A was used as a refrigerant | coolant and when R32 was used as a refrigerant | coolant on the Mollier diagram. It is a figure which shows the influence with respect to the pressure loss of a heat exchanger tube of the amount of refrigerant circulation. It is a figure which shows the influence with respect to the surface heat transfer rate of a heat exchanger tube of the amount of refrigerant circulation. It is a cross-sectional view of a ceiling-embedded indoor unit. It is a longitudinal cross-sectional view of a ceiling-embedded indoor unit. It is a figure which shows the structure of the heat exchanger tube and fin of an indoor heat exchanger. It is a longitudinal cross-sectional view of an indoor heat exchanger. It is sectional drawing along the IX-IX line of FIG. It is a figure which shows the structure of the heat exchanger tube and fin of the conventional indoor heat exchanger. It is a figure which shows the relationship between the subcooling degree of an indoor heat exchanger at the time of heating operation, and COP. It is a figure which shows the influence with respect to COP of the supercooling degree at the time of heating operation in the air conditioner which uses R32 for a refrigerant | coolant. It is a figure which shows the influence with respect to COP of the supercooling degree at the time of heating operation in the air conditioner which uses R410A for a refrigerant | coolant. It is a figure which shows the influence with respect to COP of the refrigerant | coolant circulation amount at the time of air_conditionaing | cooling operation in the air conditioner which uses R32 for a refrigerant | coolant. It is a figure which shows the influence with respect to COP of the refrigerant | coolant circulation amount at the time of air_conditionaing | cooling operation in the air conditioner which uses R410A for a refrigerant | coolant. It is a figure which shows the relationship between the mass flux at the time of evaporation, a pipe | tube heat transfer rate, and a pressure loss. It is a figure which shows the relationship between the mass flux at the time of condensation, the heat transfer coefficient in a pipe | tube, and a pressure loss. It is explanatory drawing of the performance influence of the air conditioner by the heat exchanger tube outer diameter. It is explanatory drawing of the performance influence of the air conditioner by the up-and-down pitch of the heat exchanger tube of a heat exchanger. It is explanatory drawing of the performance influence of the air conditioner by the horizontal pitch of the heat exchanger tube of a heat exchanger. It is explanatory drawing of the performance influence of the air conditioner by the fin board thickness t and the fin pitch Pf of a heat exchanger. It is a figure which shows the modification of the row | line | column structure of the heat exchanger tube of an indoor heat exchanger. It is an external view which shows a trifurcated vent. It is a figure which shows the other example of a change of the row | line | column structure of the heat exchanger tube of an indoor heat exchanger. It is a figure which shows the case where the row | line | column structure of the heat exchanger tube of an indoor heat exchanger is two rows. It is a figure which shows the case where the row | line | column structure of the heat exchanger tube of an indoor heat exchanger is 4 rows.

  Hereinafter, an air conditioner according to an embodiment of the present invention will be described with reference to the drawings. FIG. 1 shows a refrigeration cycle of an air conditioner 1 according to an embodiment of the present invention.

  The air conditioner 1 includes an outdoor unit 10 and an indoor unit 20. The outdoor unit 10 and the indoor unit 20 are connected by a gas connection pipe 2 and a liquid connection pipe 3. In the present embodiment, the outdoor unit 10 and the indoor unit 20 are connected on a one-to-one basis. However, a plurality of outdoor units may be connected to a single indoor unit, or a single outdoor unit. Alternatively, a plurality of indoor units may be connected.

  The outdoor unit 10 includes a compressor 11, a four-way valve 12, an outdoor heat exchanger 13, an outdoor fan 14, an outdoor expansion valve 15, and an accumulator 16. The outdoor heat exchanger 13 is provided with an outdoor gas side refrigerant distributor 17 and an outdoor liquid side refrigerant distributor 18.

  The compressor 11 compresses the refrigerant and discharges it to the piping. By switching the four-way valve 12, the flow of the refrigerant changes, and the cooling operation and the heating operation are switched. The outdoor heat exchanger 13 exchanges heat between the refrigerant and the outside air. The outdoor fan 14 supplies outside air to the outdoor heat exchanger 13. The outdoor expansion valve 15 depressurizes the refrigerant to a low temperature. The accumulator 16 is provided to store the liquid return at the time of transition, and adjusts the refrigerant to an appropriate dryness.

  The indoor unit 20 includes an indoor heat exchanger 21, an indoor fan 22, and an indoor expansion valve 23. The indoor heat exchanger 21 exchanges heat between the refrigerant and the inside air. The indoor fan 22 supplies outside air to the outdoor heat exchanger 21. The indoor expansion valve 23 can change the flow rate of the refrigerant flowing through the indoor heat exchanger 21 by changing the throttle amount. The indoor heat exchanger 21 is provided with an indoor gas side refrigerant distributor 24 and an indoor liquid side refrigerant distributor 25.

  In the air conditioner 1 according to the present embodiment, R32 is used alone (100% by weight) or R32 is used as a refrigerant that is enclosed in a refrigeration cycle and serves to carry heat energy during cooling operation and heating operation. Mixed refrigerant containing more than wt% is used.

  Next, the operation of the refrigeration cycle of the air conditioner 1 will be described.

  First, the cooling operation in the air conditioner 1 will be described. In the cooling operation, the four-way valve 12 communicates the discharge side of the compressor 11 and the outdoor heat exchanger 13 and communicates the suction side of the compressor 11 and the gas connection pipe 2 as indicated by a solid line.

  The high-temperature and high-pressure gas refrigerant discharged from the compressor 11 passes through the four-way valve 12 and flows into the outdoor heat exchanger 13. The high-temperature gas refrigerant that has entered the outdoor heat exchanger 13 is condensed by exchanging heat with outdoor air supplied by the outdoor fan 14 and becomes liquid refrigerant. The liquid refrigerant passes through the outdoor expansion valve 15 and the liquid connection pipe 3 and flows into the indoor unit 20. The liquid refrigerant flowing into the indoor unit 20 is decompressed by the indoor expansion valve 22 and becomes a low-temperature and low-pressure gas-liquid mixed refrigerant. This low-temperature and low-pressure refrigerant flows into the indoor heat exchanger 21, exchanges heat with the indoor air supplied by the indoor fan 22, evaporates, and becomes a gas refrigerant. At this time, the indoor air is cooled by the latent heat of vaporization of the refrigerant, and the cool air is sent into the room. Thereafter, the gas refrigerant is returned to the outdoor unit 10 through the gas connection pipe 2.

  The gas refrigerant that has returned to the outdoor unit 10 passes through the four-way valve 12 and the accumulator 16, is sucked into the compressor 11, and is compressed again by the compressor 11, thereby forming a series of refrigeration cycles.

  Next, the heating operation in the air conditioner 1 will be described. In the heating operation, the switching valve 12 causes the discharge side of the compressor 11 and the gas connection pipe 2 to communicate with each other and the suction side of the compressor 11 and the outdoor heat exchanger 13 communicate with each other, as indicated by a dotted line.

  The high-temperature and high-pressure gas refrigerant discharged from the compressor 11 is sent to the gas connection pipe 2 through the four-way valve 12 and flows into the indoor heat exchanger 21 of the indoor unit 20. The high-temperature and high-pressure gas refrigerant that has flowed into the indoor heat exchanger 21 is condensed by exchanging heat with the indoor air supplied by the indoor fan 23 and becomes high-pressure liquid refrigerant. At this time, the room air is heated by the refrigerant, and the warm air is sent into the room. Thereafter, the liquefied refrigerant passes through the indoor expansion valve 22 and the liquid connection pipe 3 and is returned to the indoor unit 10.

  The liquid refrigerant returned to the outdoor unit 10 is decompressed by the outdoor expansion valve 15 and becomes a low-temperature and low-pressure gas-liquid mixed refrigerant. The decompressed refrigerant flows into the outdoor heat exchanger 13, exchanges heat with the outside air supplied by the outdoor fan 14, is evaporated, and becomes a low-pressure gas refrigerant. The gas refrigerant flowing out of the outdoor heat exchanger 13 passes through the accumulator 16, is sucked into the compressor 11, and is compressed again by the compressor 11, thereby forming a series of refrigeration cycles.

  Next, the characteristic of R32 used with the air conditioner 1 of this embodiment is demonstrated. Specifically, the difference when these refrigerants are used will be described due to the difference in refrigerant physical properties between R32 and R410A. FIG. 2 is a diagram illustrating a refrigeration cycle during heating operation on the Mollier diagram when R410A (broken line) is used as a refrigerant and when R32 (solid line) is used as a refrigerant. In addition, R410A is a refrigerant | coolant used conventionally, and is a refrigerant | coolant with a high GWP (global warming potential) compared with R32.

  Since R32 has a characteristic that the latent heat of vaporization is larger than that of R410A, the specific enthalpy difference in the evaporator or condenser indicated by Δhe_R32 and Δhc_R32 of R32 is larger than Δhe_R410A and Δhc_R410A of R410A. Therefore, R32 can reduce the amount of refrigerant circulation necessary for generating the same capacity than R410A.

  Here, Δhe indicates the specific enthalpy difference in the evaporator, Δhc indicates the specific enthalpy difference in the condenser, and the suffixes _R410A and _R32 indicate the states in the refrigerants R410A and R32, respectively.

  When R32 is used as the refrigerant, the amount of refrigerant circulation can be reduced, so the pressure loss when the refrigerant passes through the flow path of the heat exchangers 13 and 21 is reduced, and the differential pressure between the high pressure and the low pressure is reduced. . Therefore, the required compression power in the compressor 11 can be reduced, and the COP (Coefficient of Performance: coefficient of performance) improvement effect of the air conditioner 1 can be obtained. On the other hand, a decrease in the refrigerant flow rate in the heat transfer tubes of the heat exchangers 13 and 21 may cause a decrease in the surface heat transfer coefficient on the refrigerant side, resulting in a decrease in efficiency of the heat exchangers 13 and 21.

  FIG. 3 is a diagram showing the influence of the refrigerant circulation amount on the pressure loss of the heat transfer tube, and FIG. 4 is a diagram showing the influence of the refrigerant circulation amount on the surface heat transfer coefficient of the heat transfer tube.

  As shown in FIGS. 3 and 4, when R32 is used in a condenser rather than an evaporator, the pressure loss is relatively small and the surface heat transfer coefficient is small. For this reason, in the air conditioner 1 that switches between cooling and heating, the refrigerant circulation amount per one flow path (one heat transfer tube 26 (FIG. 7)) of the heat exchangers 13 and 21 is set to both the cooling and heating. It is necessary to set the flow rate with a good balance.

  In order to adjust the refrigerant circulation amount per flow path of the heat exchangers 13 and 21, for example, at the refrigerant inlet of the indoor heat exchanger 21, the indoor gas side refrigerant distributor 24 and the indoor liquid side refrigerant distributor 25 (FIG. 7). Is used. The distributors 24 and 25 are branched into a plurality of flow paths (a plurality of heat transfer tubes 26) so that the refrigerant flows through the indoor heat exchanger 21.

  Next, the configuration of the ceiling-embedded four-way blowing indoor unit 20 in the present embodiment will be described in detail. FIG. 5 shows a transverse section of the indoor unit 20 of the air conditioner 1, and FIG. 6 shows a longitudinal section of the indoor unit 20.

  As shown in FIGS. 5 and 6, the indoor heat exchanger 21 and the indoor fan 22 are accommodated in the housing 28 of the indoor unit 20, and the indoor heat exchanger 21 is disposed so as to surround the indoor fan 22. . As described above, the indoor unit 20 in the present embodiment is a ceiling-embedded four-way blowing indoor unit.

  As shown in FIG. 5, the indoor heat exchanger 21 has a shape (substantially B-shaped) that surrounds and surrounds the indoor fan 22 and has one end 21A and the other end 21B. Therefore, since the length of the indoor heat exchanger 21 is increased, when the indoor heat exchanger 21 is divided into a plurality of flow paths, branching and assembly are possible only at both ends of the indoor heat exchanger 21. There are many restrictions on road division. An indoor gas side refrigerant distributor 24 and an indoor liquid side refrigerant distributor 25 are connected to one end 21 </ b> A of the indoor heat exchanger 21.

  Also, as shown in FIG. 6, the air introduced from the room by the indoor fan 22 is configured to exchange heat in the indoor heat exchanger 21 and to be sent into the room through the outlet.

  FIG. 7 shows the configuration of the heat transfer tubes 26 and the fins 27 of the indoor heat exchanger 21 in the present embodiment. The arrows in FIG. 7 indicate the flow of the refrigerant flowing through the heat transfer tube 26 during the heating operation. As shown in FIG. 7, the plurality of heat transfer tubes 26 are inserted through a plurality of metal plate-like fins 27. The plurality of heat transfer tubes 26 have a row configuration including three rows along the airflow direction F, in which the plurality of heat transfer tubes 26 are formed side by side in a direction intersecting the airflow direction F of indoor air by the indoor fan 22.

  By configuring in three rows, when acting as a condenser, when the refrigerant passage is configured in a direction opposite to the air flow, the temperature difference from the intake air can be kept relatively uniform. It is possible to divide the fins of the heat exchanger every time the refrigerant temperature level differs from the supercooling zone, saturation zone, and superheat zone in the 1st row, 2nd row, 3rd row with respect to the air flow, It is an advantage. Furthermore, it is advantageous in terms of ventilation resistance and mounting space.

  The row configuration includes an uppermost row (first row) L1 located at the uppermost stream in the airflow direction F, a lowermost row (third row) L3 located at the most downstream side in the airflow direction F, and the uppermost row L1 and the lowermost row L3. And an intermediate row (second row) L2 located between the two. Here, the heat transfer tube constituting the lowermost row L3 is referred to as a heat transfer tube 26a, the heat transfer tube constituting the intermediate row L2 is referred to as a heat transfer tube 26b, and the heat transfer tube constituting the uppermost row L1 is referred to as a heat transfer tube 26c. In addition, in each row | line | column L1-L3, the heat exchanger tube 26 is located in a line in the up-down direction.

  The heat transfer tube 26 c constituting the uppermost row L 1 is connected to the indoor liquid side refrigerant distributor 25, and the heat transfer tube 26 a constituting the lowermost row L 3 is connected to the indoor gas side refrigerant distributor 24. The heat transfer tubes 26a in the lowermost row L3 extend from the one end 21A of the indoor heat exchanger 21 to the other end 21B, make a U-turn at the other end 21B, and end up at one end of the indoor heat exchanger 21 in the intermediate row L2. Return to part 21A. Then, at one end portion 21A of the indoor heat exchanger 21, two adjacent heat transfer tubes 26b in the intermediate row L2 are joined together, and the heat transfer tube 26c that is joined to become one end portion in the uppermost row L1. The heat transfer tube 26c extending so as to make one reciprocation between 21A and the other end 21B and returning to the one end 21A is connected to the indoor liquid side refrigerant distributor 25.

  In other words, the heat transfer tubes 26 (first heat transfer tubes) extend from one end 21A of the indoor heat exchanger 21 to the other end 21B in the lowermost row (third row) L3, and in the intermediate row (second row) L2. The other end portion 21B of the indoor heat exchanger 21 extends from the other end portion 21B to the one end portion 21A and is coupled to another heat transfer tube 26 (second heat transfer tube) adjacent to each other at the upper and lower ends in the one end portion 21A. The heat pipe 26 is configured to reciprocate once between the one end 21A and the other end 21B of the indoor heat exchanger 21 in the uppermost row (first row) L1. Further, the three-way bend 28 that connects the two heat transfer tubes 26b in the intermediate row L2 and the heat transfer tubes 26c in the uppermost row L1 is connected to the heat transfer tube 26c substantially in the middle in the vertical direction of the two heat transfer tubes 26b. Shape. That is, when viewed from the airflow direction F, the heat transfer tube 26c connected to the trifurcated vent 28 is located between the two heat transfer tubes 26b.

  Since the heat transfer tube 26 of the indoor heat exchanger 21 is configured as described above, when functioning as a condenser during heating operation, the refrigerant that is R32 is the indoor gas side refrigerant distribution as shown by the arrows in FIG. The refrigerant 24 flows into the plurality of heat transfer tubes 26 and merges through the lowermost row L3 and the intermediate row L2, and the merged refrigerant is reciprocated once in the uppermost row L1 and discharged to the indoor liquid side refrigerant distributor 25.

  FIG. 8 shows a longitudinal sectional view of the indoor heat exchanger 21. As shown in FIG. 8, the diameter D of the heat transfer tube 26 is 4 ≦ D ≦ 6 mm, and the vertical pitch Pt (distance between the centers of the heat transfer tubes 26) of the adjacent heat transfer tubes 26 in the upper and lower directions is 11 ≦ Pt ≦ The horizontal pitch PL of the heat transfer tubes 26 (the distance between the straight lines passing through the centers of the heat transfer tubes 26 constituting each row) is 7 ≦ PL ≦ 11 mm.

  9 is a cross-sectional view taken along line IX-IX in FIG. As shown in FIG. 8, the fin 27 is provided with slits 27A and 27B. The plate thickness t [mm] of the fins 27 and the pitch Pf [mm] of the adjacent fins 27 are configured to satisfy 0.06 ≦ t / Pf ≦ 0.12. Further, the slit cut and raised widths Hs1 and Hs2 [mm] take into consideration heat transfer performance and ventilation resistance, for example, 1.2 ≦ Hs1 / Hs2 ≦ 1. 6 is configured.

  As described above, the heat transfer tubes 26 extend from the one end 21A of the indoor heat exchanger 21 to the other end 21B in the lowermost row L3, and from the other end 21B of the indoor heat exchanger 21 to one end 21A in the intermediate row L2. One end 21A of the indoor heat exchanger 21 and the other end 21B of the indoor heat exchanger 21 are connected to the other heat transfer tubes 26 that are adjacent to each other vertically at the one end 21A. Are configured to reciprocate once.

  Therefore, by combining the refrigerants flowing through the two heat transfer tubes 26 to flow through the single heat transfer tube 26, the flow rate of the refrigerant can be increased, and the surface heat transfer coefficient can be increased.

  Moreover, in this Embodiment, since R32 is used for a refrigerant | coolant, the refrigerant | coolant circulation amount to be used can be decreased. Therefore, even if the refrigerant flow rates are merged as described above, the pressure loss can be suppressed because the refrigerant flow rate is relatively small.

  On the other hand, in the configuration of the conventional heat exchanger 121 shown in FIG. 10, the heat transfer pipe 126 connected to the indoor gas side refrigerant distributor 24 reciprocates the rows L1 to L3 in total 1.5 times to distribute the indoor liquid side refrigerant. It is configured to be connected to the device 25. In this case, when the heat exchanger 121 is used as a condenser, the number of refrigerant channels flowing out from the indoor gas side refrigerant distributor 24 and the number of refrigerant channels flowing into the indoor liquid side refrigerant distributor 25 are the same. It becomes.

  Therefore, in order to reduce the number of refrigerant flow paths, it is necessary to reduce the number of heat transfer tubes 126 of the heat exchanger 121. If the number of heat transfer tubes 126 is decreased, the heat transfer area in the tube is reduced. This does not lead to an improvement in the performance of the heat exchanger 121.

  Further, as the condensation process proceeds, the density of the refrigerant increases and the refrigerant flow rate in the heat transfer pipe 126 decreases as the density advances from the lowermost row L3 to the intermediate row L2 and the uppermost row L1. Since the surface heat transfer coefficient deteriorates, the efficiency of the heat exchanger 121 cannot be maximized.

  Next, in the air conditioner 1 using R32, the relationship between the degree of supercooling of the indoor heat exchanger 21 that functions as a condenser during heating operation and COP will be described with reference to FIG. As a comparison of R32, the relationship between the degree of subcooling of the indoor heat exchanger 21 and COP when R410A is used as the refrigerant of the air conditioner 1 is also shown. In any case where R410A and R32 are used, it can be seen that there is a peak where the COP is maximum with respect to the degree of supercooling. When the degree of subcooling is smaller in R32 than in C410 peak R1 in R410A, COP shows peak P2.

  This is related to the fact that the specific enthalpy difference is larger in R32 as shown in the refrigeration cycle on the Mollier diagram of FIG.

  The contribution to the capacity of the condenser outlet supercooling is the increase in the specific enthalpy difference indicated by Δhsc_R410A and Δhsc_R32 in FIG. Since R32 has a large specific enthalpy difference Δhc_R32 in the original condenser, the capacity increase rate due to the subcooling amount Δhsc_R32 tends to be smaller than that of R410A.

  Moreover, since it is necessary to increase the compression power by increasing the condensing pressure with respect to the increase in capacity due to the increase in the degree of supercooling, there is a point where the COP reduction exceeds. Thus, R32 has the largest COP during heating in that the degree of supercooling is smaller.

  This can exhibit a special effect by using R32 in the configuration of the indoor heat exchanger 21 of the present embodiment shown in FIG. That is, by reducing the degree of supercooling at the outlet of the condenser, the temperature difference between the adjacent heat transfer tubes 26 in the uppermost row L1 through which the liquid refrigerant flows in the indoor heat exchanger 21 can be reduced. That is, heat loss between adjacent heat transfer tubes 26 can be suppressed, the surface heat transfer rate can be improved, and the performance of the indoor heat exchanger 21 can be improved.

  Moreover, as shown in FIG. 11, a larger COP can be obtained when R32 is used than when R410A is used.

  FIG. 12 and FIG. 13 are the results of verifying the above effects. FIG. 12 shows the effect of the degree of supercooling on the COP during heating operation in an air conditioner using R32 as the refrigerant and R410A as the refrigerant in FIG. Is shown. C1 and C3 in FIGS. 12 and 13 indicate the influence of the degree of supercooling on the COP when R32 and R410A are used in the air conditioner 1 including the indoor heat exchanger 21 of the present embodiment illustrated in FIG. C2 and C4 show the influence of the degree of supercooling on the COP when R32 and R410A are used in the air conditioner including the indoor heat exchanger 121 shown in FIG.

  As shown in FIG. 12, the COP of C1 is increased because of the above effect. On the other hand, when R410A is used as a refrigerant in the air conditioner 1 of the present embodiment as shown in FIG. 13, the performance (COP) is lowered as shown in C3.

  14 and 15 show the influence of the refrigerant circulation amount on the COP during the cooling operation in the air conditioner using R32 or R410A as the refrigerant. C5 and C7 in FIGS. 14 and 15 show the influence of the refrigerant circulation amount on the COP when R32 and R410A are used in the air conditioner 1 including the indoor heat exchanger 21 of the present embodiment shown in FIG. C6 and C8 show the influence of the refrigerant circulation amount on the COP when R32 and R410A are used in the air conditioner including the indoor heat exchanger 121 shown in FIG.

  Since there is no heat loss effect in the supercooling region during cooling operation, the effect of the refrigerant flow rate is dominant. Therefore, it can be seen that due to the difference in physical properties between R410A and R32, C5 and C7 using R32 and R410A in the air conditioner 1 including the indoor heat exchanger 21 of the present embodiment have a higher COP especially in the cooling intermediate capacity region. .

  In order to explain this in detail, FIG. 16 shows the relationship between the mass flux during evaporation, the heat transfer coefficient in the tube, and the pressure loss. The mass flux, the heat transfer coefficient in the pipe, and the pressure loss are shown as average values over the entire length.

  FIG. 16 shows the operation state at the intermediate cooling capacity, and shows the heat transfer coefficient and pressure loss in the pipe due to the mass flux at the time of evaporation in comparison with R32 and R410A. Specifically, in both R32 and R410A, the arrangement of the heat transfer tubes 126 in the conventional heat exchanger 121 shown in FIG. 10 (hereinafter referred to as a conventional arrangement) and the heat exchange of the present embodiment shown in FIG. The operation state in the arrangement | sequence (henceforth this application arrangement | sequence) of the heat exchanger tube 26 in the apparatus 21 is each shown by the point.

  When the arrangement is changed from the conventional arrangement to the present arrangement in R410A, the increase rate of the heat loss is small although the increase in the pressure loss is large. Even in this case, the rate of increase in pressure loss is small, and the rate of increase in heat transfer coefficient is large. Therefore, it can be said that it is more effective for improving the performance of R32 during cooling.

  Note that FIG. 17 shows the heat transfer coefficient and pressure loss in the tube due to the mass flux at the time of condensation in comparison with R32 and R410A. Although the absolute value of the influence due to the change in mass flux is different even during condensation, it is the same as during evaporation, that is, using the present arrangement for R32 is more effective for improving the performance during heating.

  Further, as described above, since the outer diameter D of the heat transfer tube 26 is 4 ≦ D ≦ 6 mm, as shown in FIG. 18, the increase in the ventilation resistance is suppressed and the heat transfer tube pitch (Pt, PL) is reduced. Therefore, the efficiency of the air conditioner 1 (APF: Annual Performance Factor) can be improved. That is, the decrease from the peak of APF can be suppressed within 3%.

  Moreover, the vertical pitch Pt of the heat transfer tubes 26 adjacent in the vertical direction is 11 ≦ Pt ≦ 17 mm. If it is this range, as shown in FIG. 19, the efficiency of the air conditioner 1 can be improved, reducing the heat loss influence of the upper and lower heat exchanger tubes due to the heat conduction of the fins.

  That is, the smaller the vertical pitch Pt, the greater the loss due to heat conduction of the fins. FIG. 19 shows the effect of the vertical pitch on the APF. When the vertical pitch is 11 mm or less, the effect of heat conduction through the fins increases, so the APF decreases. Conversely, when the vertical pitch is 17 mm or more, the heat transfer tube The reduction in the number of 26 mounted leads to a reduction in the heat transfer area in the tube and the fin efficiency, and a decrease in APF. Therefore, it is desirable to set 11 mm ≦ Pt ≦ 17 mm as the range of the upper and lower pitches Pt that can ensure the rate of decrease within 3% from the peak of APF.

  Further, since the horizontal pitch PL of the heat transfer tubes 26 is 7 ≦ PL ≦ 11 mm, the efficiency of the air conditioner 1 can be improved by optimizing the balance between the heat transfer area and the ventilation resistance as shown in FIG. it can. That is, the decrease from the peak of APF can be suppressed within 3%.

  Further, since the relationship between the plate thickness t [mm] and the fin pitch Pf [mm] of the fin 27 is 0.06 ≦ t / Pf ≦ 0.12, the heat loss in the supercooling region is reduced as shown in FIG. While obtaining the effect, the APF of the air conditioner 1 can be increased. That is, as the plate thickness of the fin 27 is thicker and the number of fins is larger, the effect of heat loss on the adjacent heat transfer tube 26 due to the heat conduction effect through the fin 27 is more likely to occur. Loss impact is mitigated. In consideration of this influence, if t / Pf is small when the fin pitch Pf is constant, the performance is reduced due to a decrease in fin efficiency, and if t / Pf is large, the influence of heat loss becomes large. Therefore, it is desirable to set 0.06 ≦ t / Pf ≦ 0.12 as a range in which the APF of the air conditioner 1 has a performance within 3% from the peak.

  In addition, since the fins 27 are provided with the slits 27A and 27B, the surface heat transfer rate is increased and the fin efficiency is relatively low. Therefore, the influence of heat conduction on the adjacent heat transfer tubes 26 can be suppressed.

  In addition, this invention is not limited to the Example mentioned above. A person skilled in the art can make various additions and changes within the scope of the present invention.

  For example, the effect of the path of the heat transfer pipe 26 of the indoor heat exchanger 21 is particularly large in the ceiling-embedded indoor unit 20 because of the large influence of the supercooling area in heating and the degree of freedom of arrangement of the heat transfer pipes 26. large. That is, in the ceiling-mounted indoor unit, as shown in FIGS. 5 and 6, the indoor heat exchanger 21 is arranged around the blower (indoor fan 22) so as to surround substantially one round, and the indoor heat exchanger 21. There are restrictions on the depth and height. Therefore, the improvement of the performance of the indoor heat exchanger 21 due to the high density arrangement of the heat transfer tubes 26 is effective, and in addition to the refrigerant passage of the present embodiment that can reduce the mounting space of the refrigerant distributors 24 and 25, heat transfer. By setting the tube diameter, the vertical pitch, and the horizontal pitch within the above ranges, the high-performance air conditioner 1 that makes the most of the characteristics of R32 can be realized.

  However, the effect can be exhibited even when used in another indoor form or the outdoor unit 10, and the form is not limited. Therefore, the configuration of the path of the heat transfer tube 26 may be used for other indoor configurations or the outdoor heat exchanger 13 of the outdoor unit 10.

  Further, although the slits 27A and 27B are provided in the fin 27, a louver may be provided. Further, in the above-described embodiment, the case where R32 is used alone as the refrigerant has been described, but the same effect can be obtained even if a mixed refrigerant containing 70% by weight or more of R32 is used.

  Further, the row configuration of the heat transfer tubes of the indoor heat exchanger may be the row configuration of the heat transfer tubes 26 shown in FIG. That is, as shown in FIG. 22, the two heat transfer tubes 26b1 and 26b2 in the intermediate row L2 may be connected to the heat transfer tube 26c1 in the uppermost row L1 located above the heat transfer tube 26b1. . The two heat transfer tubes 26b3 and 26b4 adjacent to the two heat transfer tubes 26b1 and 26b2 and the heat transfer tube 26c3 in the uppermost row L1 are connected in the same manner as in the above embodiment. Here, as shown in FIG. 23, the trifurcated vent 128 that connects the two heat transfer tubes 26b1, 26b2 and the heat transfer tube 26c1 has a position where it is connected to the heat transfer tube 26c1 in the uppermost row L1, in the middle row L2. It is comprised so that it may become above the position connected with the two heat exchanger tubes 26b. In addition, the trifurcated vent 128 is configured such that the refrigerant collides at a branching portion at the time of cooling operation and branches, so that the gas-liquid two-phase flow is substantially equally distributed.

  The heat transfer tubes (first coupling tubes) 26c1 and 26c2 in which the two heat transfer tubes 26b1 and 26b2 are coupled to each other, the heat transfer tube 26c1 extends from one end 21A (FIG. 5) to the other end 21B (FIG. 5). A heat transfer tube 26c2 is disposed below the heat transfer tube 26c1 so as to extend from the other end 21B to the one end 21A. In addition, the heat transfer tubes (second connection tubes) 26c3 and 26c4 in which the two heat transfer tubes 26b3 and 26b4 are connected to each other. 26c4 is arranged to extend from the other end 21B to the one end 21A. Therefore, the heat transfer tube 26b2 and the heat transfer tube 26b4 extending from the other end portion 21B to the one end portion 21A are arranged adjacent to each other.

  Therefore, in the row configuration of the heat transfer tubes 26 shown in FIG. 22, the heat transfer tube 26b2 and the heat transfer tube 26b4 extending from the other end portion 21B to the one end portion 21A are disposed adjacent to each other. Since heat is continuous up and down, heat loss hardly occurs at temperatures close to each other. Thereby, there exists an effect which further reduces a heat loss by half, and APF of the air conditioner 1 can be raised further.

  The row configuration of the heat transfer tubes of the indoor heat exchanger may be the row configuration of the heat transfer tubes 26 shown in FIG. As shown in FIG. 24, in heat transfer tubes 26c5 and 26c6 in which a plurality of sets of two heat transfer tubes 26b in the intermediate row L2 are coupled, the heat transfer extends from one end 21A (FIG. 5) to the other end 21B (FIG. 5). The heat tubes 26c5 are collected and arranged on the upper side, and the heat transfer tubes 26c6 extending from the other end 21B to the one end 21A are collected and arranged on the lower side. In other words, the heat transfer tubes 26c5 extending from the one end 21A to the other end 21B are arranged adjacent to each other, and the heat transfer tubes 26c6 extending from the other end 21B to the one end 21A are arranged adjacent to each other.

  According to this configuration, the heat loss between the adjacent heat transfer tubes 26 in the vertical direction in the supercooling region when acting as a condenser is further reduced with respect to the row configuration of the heat transfer tubes 26 shown in FIG. The indoor heat exchanger 21 with higher efficiency can be provided, and the APF of the air conditioner 1 can be increased.

  In the above embodiment, the case where the row structure of the heat transfer tubes of the indoor heat exchanger is constituted by three rows has been described. However, as shown in FIG. 25, as shown in FIG. ) Even in a two-row configuration consisting of only the heat transfer tubes 26b and 26c of L1 and the intermediate row (second row) L2, the effect of heat loss in the supercooling zone when acting as a condenser, and the flow velocity on the liquid side The effect of this embodiment of improving the heat transfer coefficient due to the increase can be exhibited. In other words, the lowermost row L3 may be eliminated, and only the uppermost row L1 and the intermediate row L2 may be arranged. In this case, the indoor gas side refrigerant distributor 24 is provided on the other end 21 </ b> A of the indoor heat exchanger 21. In addition, the air-conditioning apparatus having a relatively small capacity in two rows can optimize the balance between performance and cost.

  Furthermore, as shown in FIG. 26, the row configuration of the heat transfer tubes of the indoor heat exchanger may be constituted by four rows. That is, the additional row L4 may be provided on the downstream side in the airflow direction F from the lowermost row L3. The heat transfer tubes 26d constituting the additional row L4 are each connected to the indoor liquid side refrigerant distributor 25, and extend through the additional row L4 from the other end portion 21B of the indoor heat exchanger 21 to the one end portion 21A. Are connected to the heat transfer tubes 26a constituting the lowermost row L3. Even with such a configuration, the effect of the present embodiment can be exhibited, that is, the effect of reducing the heat loss in the supercooling zone when acting as a condenser and the improvement of the heat transfer coefficient by increasing the flow velocity on the liquid side. In addition, in the structure with four or more heat transfer tubes 26, the heat transfer area can be increased, so that further performance improvement can be realized.

1: air conditioner, 20: indoor unit, 21: indoor heat exchanger, 21A: one end, 21B: other end, 26: heat transfer tube, 27: fin, 27A, 27B: slit, L1: top row, L2 : Intermediate row, L3: Bottom row

Claims (4)

An air conditioner having a plurality of heat transfer tubes through which a refrigerant flows and having a heat exchanger for exchanging heat with air,
The heat exchanger has one end and the other end,
The plurality of heat transfer tubes are arranged so as to reciprocate between the one end and the other end in a state in which the plurality of heat transfer tubes are arranged in a direction intersecting with a direction in which air flows, and the plurality of the heat transfer tubes arranged in the intersecting direction. The rows of heat transfer tubes are configured to be arranged in at least two rows along the direction in which the air flows,
The two rows have a first row located in the uppermost stream in the direction in which the air flows, and a second row located next to the first row in the direction in which the air flows,
The plurality of heat transfer tubes include a first heat transfer tube and a second heat transfer tube which are adjacent to each other in the second row, and the first heat transfer tube and the second heat transfer tube are respectively in the second row. Extending from the other end to the one end, and joined at one end to form a single heat transfer tube, the one heat transfer tube being connected between the one end and the other end in the first row. It is configured to make one round trip,
The refrigerant is R32, or a refrigerant containing R32 in an amount of 70% by weight or more,
The heat exchanger includes a plurality of fins attached around the plurality of heat transfer tubes, and the fins are provided with slits or louvers,
The plurality of fins are plate-like, and 0.06 ≦ t / pf ≦ 0.12 where the plate thickness of the fins is t [mm] and the interval between adjacent fins is pf [mm]. Air conditioner. Has a ceiling-embedded cassette type indoor unit,
The air conditioner according to claim 1, wherein the heat exchanger is used in the indoor unit.   The air conditioner according to claim 1 or 2, wherein a distance between centers of the heat transfer tubes adjacent in the intersecting direction in each row is 11 mm or more and 17 mm or less.   The air conditioner according to any one of claims 1 to 3, wherein a distance between straight lines passing through the centers of the heat transfer tubes constituting each row is 7 mm or more and 11 mm or less.

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