EP3115730B1

EP3115730B1 - Refrigeration cycle device

Info

Publication number: EP3115730B1
Application number: EP14884296.6A
Authority: EP
Inventors: Takashi Okazaki; Yuki UGAJIN; Akira Ishibashi; Shinya Higashiiue; Daisuke Ito; Takumi NISHIYAMA; Shigeyoshi MATSUI
Original assignee: Mitsubishi Electric Corp
Current assignee: Mitsubishi Electric Corp
Priority date: 2014-03-07
Filing date: 2014-03-07
Publication date: 2020-05-27
Anticipated expiration: 2034-03-07
Also published as: JP6141514B2; EP3115730A4; EP3115730A1; WO2015132968A1; JPWO2015132968A1

Description

Technical Field

The present invention relates to a refrigeration cycle apparatus.

Background Art

In Patent Literature 1, there is disclosed a cross fin-and-tube heat exchanger including heat transfer tubes with inner grooves, through which an R32-based refrigerant is caused to pass. In Patent Literature 1, in consideration of brazeability or heat transfer performance, there are set numerical ranges of an outer diameter of the heat transfer tube with inner grooves, the sectional area of each inner groove, a depth of the inner groove, a bottom thickness in a grooved portion, a lead angle of the inner groove with respect to a tube axis, an apex angle of an inner fin, the number of the inner grooves, and other factors.

Citation List

Patent Literature

Patent Literature 1: Japanese Patent No. 4761719
JP2012-167912-A discloses an air conditioner.
JP2010-019489-A discloses a heat transfer pipe with inner helical groove for an evaporator.
JP10-267578-A discloses a heating tube, and heat exchanger using the same.

Summary of Invention

Technical Problem

In recent years, as refrigerant for a refrigeration cycle apparatus, in air-conditioning use, consideration has been given to use of high-pressure refrigerant having a relatively low critical point for the purpose of improving the performance in a case of using HFO-1234yf being a refrigerant having a relatively low pressure or other purposes. Therefore, the numerical ranges in Patent Literature 1 on the assumption of use of the R32-based refrigerant are not always appropriate in order to enhance the heat transfer performance of the heat exchanger in a case of using the high-pressure refrigerant having a low critical point. Thus, there is a problem in that the energy efficiency of the refrigeration cycle apparatus may not be able to be enhanced.
The present invention has been made to solve the above-mentioned problem, and has an object to provide a refrigeration cycle apparatus, which is capable of enhancing the energy efficiency in a case of using high-pressure refrigerant having a relatively low critical point.

Solution to Problem

The present invention provides a refrigeration cycle apparatus according to claim 1. The characterizing feature of the invention is that both the apex angles of the inner fins of the load-side heat exchanger and the heat source side heat exchanger have to satisfy a given relationship and at least the lead angles of the inner grooves or the heights of the inner fins as well.

Advantageous Effects of Invention

According to one embodiment of the present invention, in the case of using the high-pressure refrigerant having a relatively low critical point, the heat transfer performance of the heat exchanger can be enhanced, thereby being capable of enhancing the energy efficiency of the refrigeration cycle apparatus.

Brief Description of Drawings

[Fig. 1] Fig. 1 is a schematic view for illustrating an overall configuration of a refrigeration cycle apparatus according to Embodiment 1 of the present invention.
[Fig. 2] Fig. 2 is a sectional view for illustrating a partial sectional configuration of a heat transfer tube 22 to be used in a heat exchanger of the refrigeration cycle apparatus according to Embodiment 1 of the present invention.
[Figs. 3] Figs. 3 are views for illustrating a configuration of the heat transfer tube 22 to be used in the heat exchanger of the refrigeration cycle apparatus according to Embodiment 1 of the present invention.
[Fig. 4] Fig. 4 is a sectional view for illustrating a state of an inner surface of the heat transfer tube 22 of the heat exchanger serving as a condenser in the refrigeration cycle apparatus according to Embodiment 1 of the present invention.
[Fig. 5] Fig. 5 is a graph for showing a relationship between a ratio P/H of a pitch P to a height H in an inner fin 24 and an evaporation heat transfer coefficient in a case of using refrigerant having a relatively low critical point in the refrigeration cycle apparatus according to Embodiment 1 of the present invention.
[Fig. 6] Fig. 6 is a sectional view for illustrating a partial sectional configuration of a heat transfer tube 22 to be used in a heat exchanger of a refrigeration cycle apparatus according to Embodiment 2 of the present invention.
[Fig. 7] Fig. 7 is a perspective view for illustrating a schematic configuration of a heat exchanger of a refrigeration cycle apparatus according to Embodiment 3 of the present invention.

Description of Embodiments

Embodiment

1

A refrigeration cycle apparatus according to Embodiment 1 of the present invention is described. The refrigeration cycle apparatus of this embodiment is used, for example, in an air-conditioning apparatus. Fig. 1 is a schematic view for illustrating an overall configuration of the refrigeration cycle apparatus of this embodiment. In Fig. 1, a flow direction of refrigerant during a cooling operation is indicated by the solid-line arrows, a flow direction of the refrigerant during a heating operation is indicated by the broken-line arrows, and a flow direction of air is indicated by the thick outlined arrows. Note that, in the figures including Fig. 1 referred to below, the dimensional relationship between components, the shape of each of the components, or other factors may be different from that of actual components.
As illustrated in Fig. 1, the refrigeration cycle apparatus of this embodiment includes a refrigerant circuit 10 through which refrigerant is circulated. The refrigerant circuit 10 includes a compressor 11, a four-way valve 12, a heat source-side heat exchanger 13, an expansion valve 14 (example of an expansion device), and a load-side heat exchanger 15. The compressor 11, the four-way valve 12, the heat source-side heat exchanger 13, the expansion valve 14, and the load-side heat exchanger 15 are connected to one another via refrigerant pipes.
Passages in the four-way valve 12 are connected as indicated by the solid lines during the cooling operation, and are connected as indicated by the broken lines during the heating operation. With this, during the cooling operation, the refrigerant is allowed to flow the compressor 11, the heat source-side heat exchanger 13, the expansion valve 14, and the load-side heat exchanger 15 in the stated order. On the other hand, during the heating operation, the refrigerant is allowed to flow the compressor 11, the load-side heat exchanger 15, the expansion valve 14, and the heat source-side heat exchanger 13 in the stated order. The heat source-side heat exchanger 13 serves as a condenser (radiator) during the cooling operation, and serves as an evaporator during the heating operation. The load-side heat exchanger 15 serves as the evaporator during the cooling operation, and serves as the condenser (radiator) during the heating operation.
During the cooling operation, refrigerant discharged from the compressor 11 passes through the four-way valve 12 to flow into the heat source-side heat exchanger 13. The refrigerant flowing into the heat source-side heat exchanger 13 rejects heat to outdoor air sent by an outdoor fan 16 to be condensed and liquefied, and is allowed to flow out of the heat source-side heat exchanger 13. The refrigerant flowing out of the heat source-side heat exchanger 13 is decompressed in the expansion valve 14, and is allowed to flow into the load-side heat exchanger 15. The refrigerant flowing into the load-side heat exchanger 15 removes heat from indoor air sent by an indoor fan 17 to be evaporated, and is allowed to flow out of the load-side heat exchanger 15. The refrigerant flowing out of the load-side heat exchanger 15 passes through the four-way valve 12 again to be sucked into the compressor 11.
During the heating operation, the refrigerant discharged from the compressor 11 passes through the four-way valve 12 to flow into the load-side heat exchanger 15. The refrigerant flowing into the load-side heat exchanger 15 rejects heat to indoor air sent by the indoor fan 17 to be condensed and liquefied, and is allowed to flow out of the load-side heat exchanger 15. The refrigerant flowing out of the load-side heat exchanger 15 is decompressed in the expansion valve 14, and is allowed to flow into the heat source-side heat exchanger 13. The refrigerant flowing into the heat source-side heat exchanger 13 removes heat from outdoor air sent by the outdoor fan 16 to be evaporated, and is allowed to flow out of the heat source-side heat exchanger 13. The refrigerant flowing out of the heat source-side heat exchanger 13 passes through the four-way valve 12 again to be sucked into the compressor 11.
As the refrigerant, there is used a HFO-based refrigerant having low GWP and a low critical point (for example, a critical point of less than 70 degrees C) or a mixed refrigerant thereof. In the case of the mixed refrigerant, for example, R32 or HFO-1234yf may be used as a refrigerant to be mixed.
The heat source-side heat exchanger 13 and the load-side heat exchanger 15 are each, for example, a cross fin heat exchanger as illustrated in Fig. 7 described later. The cross fin heat exchanger includes a plurality of heat transfer fins 21 laminated to one another, and a plurality of heat transfer tubes 22 being arranged in parallel to one another and passing through each heat transfer fin 21. As the heat transfer tubes 22, there are used tubes with inner grooves, in which inner grooves for promoting heat transfer and inner fins each being formed between the inner grooves are formed at an inner surface (inner peripheral surface) thereof. The heat transfer tubes 22 of this embodiment are manufactured through drawing, rolling, or other methods. The inner grooves extend obliquely with respect to a direction of a tube axis of each of the heat transfer tubes 22, and are each formed into, for example, a helical shape. For example, a plurality of inner grooves are formed in each of the heat transfer tubes 22. The heat transfer tubes 22 each have, for example, a circular tube outer shape.
Fig. 2 is a sectional view for illustrating a partial sectional configuration of the heat transfer tube 22. As illustrated in Fig. 2, an inner fin 24 formed between adjacent inner grooves 23 has, for example, a triangular cross-sectional shape having a height larger than a width. In Fig. 2, H represents a height of the inner fin 24, θ represents an apex angle of the inner fin 24, and P represents a pitch of the inner fins 24 (for example, a pitch of distal ends (apexes) of the inner fins 24). As the height H, the apex angle θ, and the pitch P, for example, averages of respective values measured at a plurality of points in the heat transfer tube 22 may be used. A distance between a bottom surface of the inner groove 23 and an outer peripheral surface of the heat transfer tube 22 indicates a bottom thickness of the heat transfer tube 22. The height H of the inner fin 24 is, for example, from 0.1 mm to 0.5 mm, the apex angle θ of the inner fin 24 is, for example, from 5 degrees to 50 degrees, and the pitch P of the inner fins 24 is, for example, from 0.1 mm to 0.5 mm.
Figs. 3 are views for illustrating a configuration of the heat transfer tube 22. Fig. 3(a) is a developed view for illustrating the heat transfer tube 22 illustrated in Fig. 3(b) cut along the broken line part parallel to the tube axis. In Fig. 3(a) and Fig. 3(b), the thick outlined arrows indicate the flow direction of the refrigerant. As illustrated in Fig. 3(a), the inner grooves 23 extend in one oblique direction with respect to the direction of the tube axis indicated by the alternate long and short dash line. That is, the inner grooves 23 are each formed into a helical shape in the inner surface of the heat transfer tube 22. α represents an angle (lead angle of each of the inner grooves 23) formed by an extending direction of the inner grooves 23 each having a helical shape and the direction of the tube axis. The lead angle α of each of the inner grooves 23 is, for example, from 10 degrees to 50 degrees. Note that, in the heat transfer tube 22 of this embodiment, all the inner grooves 23 extend in the one direction in the developed view, but the inner grooves 23 may be each formed into a V-like shape, a W-like shape, or other shapes in the developed view.
In a case of using refrigerant having a relatively low critical point (for example, refrigerant having a critical point of less than 70 degrees C), in the heat exchanger serving as the condenser, the operation is performed in a region in which the refrigerant is near the critical point, and hence a ratio of a gas single-phase region or a liquid single-phase region is increased. With this, the heat transfer performance of the condenser is degraded. Therefore, in the case of using the refrigerant having a relatively low critical point, in particular, the performance of the heat exchanger serving as the condenser needs to be enhanced.
Fig. 4 is a sectional view for illustrating a state of the inner surface of the heat transfer tube 22 of the heat exchanger serving as the condenser. In Fig. 4, the adjacent two inner fins 24 and the inner groove 23 formed therebetween are illustrated. As illustrated in Fig. 4, when the heat exchanger serves as the condenser, liquid condensed at a distal end of each of the inner fins 24 of the heat transfer tube 22 is allowed to flow downward along a side surface (inclined surface) of each of the inner fins 24 to be accumulated in the bottom portion of the inner groove 23 so that the liquid forms a liquid film 25. Therefore, at least the apex angle θ and the height H of the inner fin 24 serve as parameters determining a phenomenon.
As the apex angle θ of the inner fin 24 is reduced, the degree of the inclination of the side surface is increased, and hence the dischargeability of the liquid condensed at the distal end of each of the inner fins 24 is enhanced, thereby enhancing the performance. Therefore, it is preferred that the apex angle θ be reduced in order to enhance the performance. However, when the apex angle θ is excessively small, the inner groove 23 is collapsed when expanding the tube in assembly of the heat exchanger. As a result, the close contact between the heat transfer fins 21 and the heat transfer tubes 22 may be degraded. Therefore, there is an appropriate apex angle θ.
Further, as the height H of the inner fin 24 is increased, the distal end of each inner fin 24 can be protruded toward a vapor core region, thereby being capable of avoiding such a situation that the distal end of each inner fin 24 is covered with the condensed liquid. Therefore, it is preferred that the height H be increased in order to enhance the performance. However, when the height H is excessively increased, the sectional area of the passage in the heat transfer tube 22 is reduced. As a result, a pressure loss is increased so that the performance of the heat exchanger may be degraded.
Further, as the lead angle α of the inner groove 23 is increased, stirring of the refrigerant passing through the heat transfer tube 22 is promoted, thereby enhancing the performance. Therefore, it is preferred that the lead angle α be increased in order to enhance the performance. However, when the lead angle is excessively large, the pressure loss is increased so that the performance of the heat exchanger may be degraded.
As described above, in order to enhance the performance of the heat exchanger, basically, it is preferred that the apex angle θ of the inner fin 24 be reduced, that the height H of the inner fin 24 be increased, and that the lead angle α of the inner groove 23 be increased. However, when those parameters are set placing priority on the performance of the heat exchanger, the manufacturing cost of the heat exchanger may be increased. For example, when the lead angle α is set larger, the drawing speed when drawing the heat transfer tube 22 is decreased. As a result, the yields of the heat transfer tubes 22 and the heat exchangers per unit time are reduced, thereby increasing the manufacturing cost of the heat exchanger. Therefore, it is preferred that the lead angle α of the inner groove 23 be reduced in order to increase the drawing speed for the heat transfer tube 22 and reduce the manufacturing cost of the heat exchanger. Further, even when the apex angle θ is set smaller or the height H is set larger, the drawing speed is decreased similarly, and hence the yields of the heat transfer tubes 22 and the heat exchangers per unit time are reduced, thereby increasing the manufacturing cost of the heat exchanger. Therefore, it is preferred that the apex angle θ of the inner fin 24 be increased, and that the height H of the inner fin 24 be reduced in order to increase the drawing speed for the heat transfer tube 22 and reduce the manufacturing cost of the heat exchanger.
Meanwhile, the required specifications vary between the load-side heat exchanger 15 and the heat source-side heat exchanger 13. In a case of a room air-conditioner (RAC) or a package air-conditioner (PAC), the contribution to an annual performance factor (APF) is more significant in the load-side heat exchanger 15 (heat exchanger serving as the condenser during the heating operation) than in the heat source-side heat exchanger 13. Therefore, in this embodiment, in the load-side heat exchanger 15 contributing significantly to the APF, the respective parameters (the apex angle θ, the height H, and the lead angle α) of the heat transfer tube 22 are set placing higher priority on the enhancement of the performance. In the heat source-side heat exchanger 13 contributing less significantly to the APF, the respective parameters of the heat transfer tube 22 are set placing higher priority on the reduction of the manufacturing cost. With this, the energy efficiency of the refrigeration cycle apparatus (for example, the APF) can be enhanced, and the manufacturing cost of the refrigeration cycle apparatus can be reduced, thereby being capable of enhancing the cost effectiveness of the refrigeration cycle apparatus.
As a specific example, α1 and α2 are set so as to satisfy the following relationship (first relationship): $α 1 > α 2$
where α1 represents the lead angle of the inner groove 23 of the load-side heat exchanger 15, and α2 represents the lead angle of the inner groove 23 of the heat source-side heat exchanger 13. With this, the performance of the load-side heat exchanger 15 can be enhanced, and the manufacturing cost of the heat source-side heat exchanger 13 can be reduced. Therefore, the energy efficiency can be enhanced, and the manufacturing cost can be reduced. In consideration of actual manufacture management, the mass production management including variations may be difficult unless α1 and α2 have a difference of about 10%, and hence it is preferred that α1 and α2 satisfy the following relationship: $(α 1 - α 2) / α 1 > 0.10$
As another example, H1 and H2 are set so as to satisfy the following relationship (second relationship): $H 1 > H 2$
where H1 represents the height of the inner fin 24 of the load-side heat exchanger 15, and H2 represents the height of the inner fin 24 of the heat source-side heat exchanger 13. With this, the performance of the load-side heat exchanger 15 can be enhanced, and the manufacturing cost of the heat source-side heat exchanger 13 can be reduced. Therefore, the energy efficiency can be enhanced, and the manufacturing cost can be reduced. In consideration of actual manufacture management, the mass production management including variations may be difficult unless H1 and H2 have a difference of about 10%, and hence it is preferred that H1 and H2 satisfy the following relationship: $(H 1 - H 2) / H 1 > 0.10$
As yet another example, θ1 and θ2 are set so as to satisfy the following relationship (third relationship): $θ 1 < θ 2$
where θ1 represents the apex angle of the inner fin 24 of the load-side heat exchanger 15, and θ2 represents the apex angle of the inner fin 24 of the heat source-side heat exchanger 13. With this, the performance of the load-side heat exchanger 15 can be enhanced, and the manufacturing cost of the heat source-side heat exchanger 13 can be reduced. Therefore, the energy efficiency can be enhanced, and the manufacturing cost can be reduced. In consideration of actual manufacture management, the mass production management including variations may be difficult unless θ1 and θ2 have a difference of about 20%, and hence it is preferred that θ1 and θ2 satisfy the following relationship: $(θ 1 - θ 2) / θ 1 < - 0.20$
This embodiment is constructed so as to satisfy at least two (more preferably, all the three) relationships of the above-mentioned three expressions (1), (2), and (3). Alternately, this embodiment is constructed so as to satisfy at least two (more preferably, all the three) relationships of the above-mentioned three expressions (1-2), (2-2), and (3-2). With this, the performance of the load-side heat exchanger 15 contributing significantly to the energy efficiency can be enhanced, thereby being capable of enhancing the energy efficiency of the refrigeration cycle apparatus. Further, the manufacturing cost of the heat source-side heat exchanger 13 can be reduced, thereby being capable of reducing the manufacturing cost of the refrigeration cycle apparatus.
Next, a case where the heat exchanger serves as the evaporator is described. In a case of using refrigerant having a critical temperature lower than that of a general HFC-based refrigerant, the surface tension is likely to be lowered as compared to the general HFC-based refrigerant. In a case of using the general HFC-based refrigerant, a meniscus is formed on a liquid film inside the inner groove 23 so that a thin liquid film is formed along a side surface of each of the inner fins 24, thereby promoting evaporation of the liquid refrigerant. On the other hand, in the case of using the refrigerant having a low critical point, a meniscus is less likely to be formed on the liquid film inside the inner groove 23. Thus, when the pitch P is set smaller than in the case of using the general HFC-based refrigerant, the wetted area of the liquid film is increased, thereby enhancing the performance. However, when the pitch P is set excessively small, the thin liquid film is not formed at the time of evaporation, and the distal end of each of the inner fins 24 with high heat transfer performance is not exposed from the liquid film also at the time of condensation, thereby degrading the performance of the heat exchanger. Therefore, the pitch P also has a lower limit value.
In this embodiment, the size of the pitch P is evaluated with a ratio P/H of the pitch P to the height H in the inner fin 24. Fig. 5 is a graph for showing a relationship between the ratio P/H of the pitch P to the height H in the inner fin 24 and an evaporation heat transfer coefficient in the case of using the refrigerant having a relatively low critical point. The horizontal axis in the graph denotes the ratio P/H, and the vertical axis in the graph denotes the evaporation heat transfer coefficient. As shown in Fig. 5, in a range where a relationship of 0.5<P/H<3.5 is satisfied, a higher evaporation heat transfer coefficient is obtained than in the other ranges. Therefore, the ratio P/H in the inner fin 24 is set so as to satisfy the relationship of 0.5<P/H<3.5, thereby obtaining a heat transfer tube 22 suitable for the refrigerant having a relatively low critical point. With this, in the case of using the refrigerant having a relatively low critical point, the heat transfer performance of the heat exchanger can be enhanced, thereby being capable of enhancing the energy efficiency of the refrigeration cycle apparatus.

Embodiment 2

A refrigeration cycle apparatus according to Embodiment 2 of the present invention is described. Fig. 6 is a sectional view for illustrating a partial sectional configuration of the heat transfer tube 22 to be used in the heat exchanger (at least one of the heat source-side heat exchanger 13 and the load-side heat exchanger 15) of the refrigeration cycle apparatus according to this embodiment. Note that, components having the same functions and effects as those of Embodiment 1 are denoted by the same reference symbols, and description thereof is omitted herein.
As illustrated in Fig. 6, at least one of the side surfaces (in this embodiment, both the side surfaces) of the inner fin 24 of the heat transfer tube 22 is changed in inclination angle in at least a part of the inner fin 24 in a height direction thereof so as to be convexed outward. In this embodiment, an angle β formed on the inner fin 24 side by a surface 26a forming a root 24a of the inner fin 24 and a surface 26b forming a distal end 24b of the inner fin 24 is less than 180 degrees. Further, in this embodiment, the surface 26a and the surface 26b are inclined in directions opposite to each other with reference to a radial direction of the heat transfer tube 22. That is, the surface 26a at the root 24a is inclined so as to face an outer side in the radial direction of the heat transfer tube 22, and the surface 26b at the distal end 24b is inclined so as to face an inner side in the radial direction of the heat transfer tube 22. Further, the inner fin 24 of this embodiment has a constriction at the root 24a, and has a thick portion having a width larger than that of the root 24a in at least a part of the inner fin 24 in the height direction.
The inner fin 24 is formed as described above, and hence a sludge receiving space 27 having a relatively large width is formed at the bottom portion of the inner groove 23. The HFO-based refrigerant liable to generate sludge generally has low stability, and hence generates sludge by reacting with air mixed in the refrigerant circuit or constituent substances in refrigerating machine oil. In this embodiment, when sludge is generated in the refrigeration cycle apparatus, the generated sludge can be received in the sludge receiving space 27 formed at the bottom portion of the inner groove 23, thereby being capable of preventing deposition of the sludge at the distal end 24b of each of the inner fins 24. Therefore, according to this embodiment, even when the refrigeration cycle apparatus (for example, an air-conditioning apparatus) using the HFO-based refrigerant liable to generate sludge is used for a long period of time, high heat transfer performance can be always maintained in the heat exchanger, thereby being capable of enhancing the energy efficiency of the refrigeration cycle apparatus.

Embodiment 3

A refrigeration cycle apparatus according to Embodiment 3 of the present invention is described. Fig. 7 is a perspective view for illustrating a schematic configuration of the heat exchanger (at least one of the heat source-side heat exchanger 13 and the load-side heat exchanger 15) of the refrigeration cycle apparatus of this embodiment. In Fig. 7, a flow direction of refrigerant at the time when the heat exchanger serves as the condenser is indicated by the solid-line arrows, and a flow direction of air is indicated by the thick outlined arrows.
As illustrated in Fig. 7, the heat exchanger according to this embodiment includes the plurality of heat transfer fins 21 laminated to one another, and the plurality of heat transfer tubes 22 (tubes with inner grooves) being arranged in parallel to one another and passing through each heat transfer fin 21. In Fig. 7, there is exemplified twelve heat transfer tubes 22 arrayed in a single row in a plane intersecting with the flow direction of the air. The twelve heat transfer tubes 22 may be hereinafter referred to as heat transfer tubes 22a, 22b,... , 22l in the order from the top. An end of the heat transfer tube 22a on the far side of Fig. 7, which is the first heat transfer tube from the top, and an end of the heat transfer tube 22b on the far side of Fig. 7, which is the second heat transfer tube from the top and is located immediately below the heat transfer tube 22a, are connected to each other by a U-shaped tube (not shown). Similarly, ends of the heat transfer tubes 22c, 22e, 22g, 22i, and 22k on the far side of Fig. 7, which are the odd-numbered heat transfer tubes from the top, and ends of the heat transfer tubes 22d, 22f, 22h, 22j, and 22l on the far side of Fig. 7, which are the even-numbered heat transfer tubes from the top and are located immediately below the heat transfer tubes 22c, 22e, 22g, 22i, and 22k, are connected to each other by U-shaped tubes, respectively. Note that, two heat transfer tubes adjacent vertically to each other may be obtained by bending a single heat transfer tube 22 into a hair-pin shape.
Now, passages in the heat exchanger are described on the assumption of the flow direction of the refrigerant at the time when the heat exchanger serves as the condenser. A bifurcation portion 32 is connected to an inlet portion 31 being an inlet of the refrigerant. With this, a passage of the refrigerant flowing into the inlet portion 31 branches into two passages. Bifurcation portions 33 and 34 are connected to the two passages branching off at the bifurcation portion 32, respectively. A total of four passages branching off at the bifurcation portions 33 and 34, respectively, are connected to the ends of the heat transfer tubes 22a, 22c, 22e, and 22g of the heat exchanger on the near side of Fig. 7. That is, in the flow of the refrigerant at the time when the heat exchanger serves as the condenser, the number of the passages on the inlet side in the heat exchanger (number of the passages connected to the inlet portion 31) is four.
The passage extending through the heat transfer tube 22a is turned back at the end thereof on the far side to pass through the heat transfer tube 22b located below the heat transfer tube 22a, and is returned to an end of the heat transfer tube 22b on the near side. Similarly, the three passages extending through the heat transfer tubes 22c, 22e, and 22g are turned back at the ends thereof on the far side to pass through the heat transfer tubes 22d, 22f, and 22h located below the heat transfer tubes 22c, 22e, and 22g, and are returned to ends of the heat transfer tubes 22d, 22f, and 22h on the near side.
A bifurcation portion 35 is connected to the end of the heat transfer tube 22b on the near side and the end of the heat transfer tube 22d on the near side. With this, the two passages extending through the heat transfer tubes 22b and 22d join a single passage. The single passage joined at the bifurcation portion 35 is connected to the end of the heat transfer tube 22k on the near side. A bifurcation portion 36 is connected to the end of the heat transfer tube 22f on the near side and the end of the heat transfer tube 22h on the near side. With this, the two passages extending through the heat transfer tubes 22f and 22h join a single passage. The single passage joined at the bifurcation portion 36 is connected to the end of the heat transfer tube 22i on the near side.
The passage extending through the heat transfer tube 22i is turned back at the end thereof on the far side to pass through the heat transfer tube 22j located below the heat transfer tube 22i, and is returned to an end of the heat transfer tube 22j on the near side. Similarly, the passage extending through the heat transfer tube 22k is turned back at the end thereof on the far side to pass through the heat transfer tube 221 located below the heat transfer tube 22k, and is returned to an end of the heat transfer tube 221 on the near side. A bifurcation portion 37 is connected to the end of the heat transfer tube 22j on the near side and the end of the heat transfer tube 221 on the near side. With this, the two passages extending through the heat transfer tubes 22j and 221 join an outlet portion 38 being an outlet of the refrigerant. That is, in the flow of the refrigerant at the time when the heat exchanger serves as the condenser, the number of the passages on the outlet side in the heat exchanger (number of the passages connected to the outlet portion 38) is two. As described above, in the heat exchanger of this embodiment, in the flow of the refrigerant at the time when the heat exchanger serves as the condenser, the number of the passages is reduced on the way, and the number of the passages on the outlet side is 1/2 or less of (in this embodiment, 1/2 of) the number of the passages on the inlet side.
In a condenser of a refrigeration cycle apparatus using high-pressure refrigerant having a low critical point, the ratio of the liquid single-phase region is increased, thereby generally degrading the performance. However, in the heat exchanger of this embodiment, in the flow of the refrigerant at the time when the heat exchanger serves as the condenser, the number of the passages on the outlet side is reduced to be 1/2 or less of the number of the passages on the inlet side. Thus, the flow rate in the latter half part of the passage in the condenser, that is, the flow rate in the liquid single-phase region is increased, thereby being capable of promoting the heat transfer. Therefore, according to this embodiment, the heat transfer performance of the heat exchanger can be enhanced, thereby being capable of enhancing the energy efficiency of the refrigeration cycle apparatus. In other words, according to this embodiment, the performance degradation unique to the refrigeration cycle apparatus using the high-pressure refrigerant having a low critical temperature can be suppressed.

Other Embodiments

The present invention is not limited to the above-mentioned embodiments, and various modifications may be made thereto.
For example, in the above-mentioned embodiments, the cross fin heat exchanger is given as an example, but the present invention is applicable to other heat exchangers.
Further, each of the above-mentioned embodiments or modifications may be carried out in combination.

Reference Signs List

10 refrigerant circuit 11 compressor 12 four-way valve 13 heat source-side heat exchanger 14 expansion valve 15 load-side heat exchanger 16 outdoor fan 17 indoor fan 21 heat transfer fin 22, 22a-22l heat transfer tube 23 inner groove 24 inner fin 24a root 24b distal end 25 liquid film 26a, 26b surface 27 sludge receiving space 31 inlet portion 32, 33, 34, 35, 36, 37 bifurcation portion 38 outlet portion

Claims

A refrigeration cycle apparatus, comprising a refrigerant circuit (10) configured to circulate refrigerant,
the refrigerant circuit (10) comprising a compressor (11), a load-side heat exchanger (15), an expansion device (14), and a heat source-side heat exchanger (13),
as the refrigerant, the refrigeration cycle apparatus employing an HFO-based refrigerant having a critical point of less than 70 degrees C or a mixed refrigerant including the HFO-based refrigerant having the critical point of less than 70 degrees C in the refrigerant circuit,
the load-side heat exchanger (15) and the heat source-side heat exchanger (13) each comprising tubes (22) with inner grooves, each of the tubes (22) having the inner grooves (23) extending obliquely with respect to a direction of an axis of each of the tubes (22), and inner fins (24) each being formed between the inner grooves (23),
wherein a first relationship of θ1 <θ2 is satisfied
and wherein at least one of a second relationship, and a third relationship described below is also satisfied:
the second relationship of α1>α2; and

the third relationship of H1 >H2;

where α1 and α2 represent a lead angle of each of the inner grooves of the load-side heat exchanger (15) and a lead angle of each of the inner grooves of the heat source-side heat exchanger (13), respectively, H1 and H2 represent a height of each of the inner fins of the load-side heat exchanger (15) and a height of each of the inner fins of the heat source-side heat exchanger (13), respectively, and θ1 and θ2 represent an apex angle of each of the inner fins of the load-side heat exchanger (15) and an apex angle of each of the inner fins of the heat source-side heat exchanger (13), respectively.
The refrigeration cycle apparatus of claim 1,
wherein the first relationship is (θ1-θ2)/θ1 <-0.20
wherein the second relationship is (α1-α2)/α1>0.10,
wherein the third relationship is (H1-H2)/H1 >0.10.
The refrigeration cycle apparatus of claim 1 or 2,
wherein, in at least one of the load-side heat exchanger (15) and the heat source-side heat exchanger (13), a relationship of 0.5<P/H<3.5 is satisfied, where P represents a pitch between the inner fins, and H represents a height of each of the inner fins.
The refrigeration cycle apparatus of any one of claims 1 to 3,
wherein in at least one of the load-side heat exchanger (15) and the heat source-side heat exchanger (13), an angle (β) formed on an inner fin side by a surface forming a root (24a) and a surface forming a distal end (24b) of at least one of side surfaces of each of the inner fins is less than 180 degrees.
The refrigeration cycle apparatus of claim 4, wherein each of the inner fins has a thick portion having a width larger than a width of the root in at least a part of each of the inner fins in a height direction of each of the inner fins.
The refrigeration cycle apparatus of any one of claims 1 to 5, wherein in at least one of the load-side heat exchanger (15) and the heat source-side heat exchanger (13), a number of passages on an outlet side is 1/2 or less of a number of passages on an inlet side in a flow of the refrigerant when the at least one of the load-side heat exchanger (15) and the heat source-side heat exchanger (13) serves as a condenser.